2020 ASHRAE Handbook: HVAC Systems and Equipment, SI Edition 9781947192539

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English Pages 984 Year 2020

Table of contents :
2020 ASHRAE Handbook: HVAC Systems and Equipment
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SI Table of Contents
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ASHRAE TECHNICAL COMMITTEES, TASK GROUPS, AND TECHNICAL RESOURCE GROUPS
ASHRAE Research: Improving the Quality of Life
Preface
CHAPTERS
--- CHAPTER 01: HVAC SYSTEM ANALYSIS AND SELECTION ---
1. Selecting a System
Additional Goals
Equipment and System Constraints
Constructability Constraints
Narrowing the Choices
2. HVAC Systems and Equipment
Centralized System Characteristics
Air Distribution Systems
3. Space Requirements
Equipment Rooms
Fan Rooms
Horizontal Distribution
Vertical Shafts
Rooftop Equipment
Equipment Access
4. Air Distribution
Air Terminal Units
8. Maintenance Management
9. Building System Commissioning
References
Bibliography
Tables
Table 1 Sample HVAC System Analysis and Selection Matrix (0 to 10 Score)
Figures
Fig. 1 Process Flow Diagram
--- CHAPTER 02: DECENTRALIZED COOLING AND HEATING ---
1. System Characteristics
Advantages
Disadvantages
2. Design Considerations
Air-Side Economizer
Advantages
Disadvantages
Water-Side Economizer
Advantages
Disadvantages
3. Window-Mounted and Through-the- Wall Room HVAC Units
Advantages
Disadvantages
Design Considerations
4. Water-Source Heat Pump Systems
Advantages
Disadvantages
Design Considerations
5. Multiple-Unit Systems
Advantages
Disadvantages
Design Considerations
6. Residential and Light Commercial Split Systems
Advantages
Disadvantages
Design Considerations
7. Commercial Self-Contained (Floor- by-Floor) Systems
Advantages
Disadvantages
Design Considerations
8. Commercial Outdoor Packaged Systems
Advantages
Disadvantages
Design Considerations
9. Single-Zone VAV Systems
Advantages
Disadvantages
Design Considerations
10. Automatic Controls and Building Management Systems
11. Maintenance Management
12. Building System Commissioning
Bibliography
Figures
Fig. 1 Multiple-Unit Systems: Single-Zone Unitary HVAC Equipment for Interior and Packaged Terminal Air Conditioners (PTACs) for Perimeter
Fig. 2 Vertical Self-Contained Unit
Fig. 3 Dedicated Outdoor-Air-Conditioning Unit
Fig. 4 Commercial Self-Contained Unit with Discharge Plenum
--- CHAPTER 03: CENTRAL COOLING AND HEATING PLANTS ---
1. System Characteristics
Advantages
Disadvantages
2. Design Considerations
Cooling and Heating Loads
Security
System Flow Design
Energy Recovery and Thermal Storage
3. Equipment
Primary Refrigeration Equipment
Ancillary Refrigeration Equipment
Primary Heating Equipment
Ancillary Heating Equipment
4. Distribution Systems
5. Sound, Vibration, Seismic, and Wind Considerations
Sound and Vibration
Seismic and Wind Issues
6. Space Considerations
Location of Central Plant and Equipment
Central Plant Security
7. Automatic Controls and Building Management Systems
Instrumentation
8. Maintenance Management Systems
9. Building System Commissioning
10. System Replacements and Expansions
References
Bibliography
Figures
Fig. 1 Primary Variable-Flow System
Fig. 2 Primary (Limited) Variable-Flow System Using Distribution Pressure Control
Fig. 3 Primary/Secondary Pumping Chilled-Water System
Fig. 4 Primary/Secondary Pumping Hot-Water System
--- CHAPTER 04: AIR HANDLING AND DISTRIBUTION ---
Advantages of All-Air Systems
Disadvantages of All-Air Systems
Heating and Cooling Calculations
Zoning
Space Heating
Air Temperature Versus Air Quantity
Space Pressure
Other Considerations
First, Operating, and Maintenance Costs
Energy in Air Handling
1. AIR-HANDLING UNITS
1.1 Air-Handling Unit Psychrometric Processes
Cooling
Heating
Humidification
Dehumidification
Air Mixing or Blending
1.2 Air-Handling Unit Components
Return Air Fan
Relief Air Fan
Automatic Dampers
Relief Openings
Return Air Dampers
Outdoor Air Intakes
Economizers
Mixing Plenums
Static Air Mixers
Filter Section
Preheat Coil
Cooling Coil
Reheat Coil
Humidifiers
Dehumidifiers
Energy Recovery Devices
Sound Control Devices
Supply Air Fan
Miscellaneous Components
1.3 Air Distribution
Ductwork Design
Primary Equipment
Air-Handling Equipment
Central Mechanical Equipment Rooms (MERs)
Decentralized MERs
Fans
2. AIR-HANDLING SYSTEMS
2.1 Single-Duct Systems
Constant Volume
Variable Air Volume (VAV)
2.2 Dual-Duct Systems
Constant Volume
Variable Air Volume
2.3 Multizone Systems
2.4 Special Systems
Primary/Secondary
Dedicated Outdoor Air
Underfloor Air Distribution
Wetted Duct/Supersaturated
Compressed-Air and Water Spray
Low-Temperature
Smoke Control
2.5 Air Terminal Units
Constant-Volume Reheat
Variable Air Volume
Terminal Humidifiers
Terminal Filters
2.6 Air Distribution System Controls
2.7 Automatic Controls and Building Management Systems
2.8 Maintenance Management System
2.9 Building System Commissioning
References
Bibliography
Figures
Fig. 1 Typical Air-Handling Unit Configurations
Fig. 2 Direct-Expansion or Chilled-Water Cooling and Dehumidification
Fig. 3 Direct Spray of Water in Airstream Cooling
Fig. 4 Supersaturated Evaporative Cooling
Fig. 5 Steam, Hot-Water, and Electric Heating, and Direct and Indirect Gas- and Oil-Fired Heat Exchangers
Fig. 6 Direct Spray Humidification
Fig. 7 Steam Injection Humidification
Fig. 8 Chemical Dehumidification
Fig. 9 Constant-Volume System with Reheat
Fig. 10 Variable-Air-Volume System with Reheat and Induction and Fan-Powered Devices
Fig. 11 Single-Fan, Dual-Duct System
Fig. 12 Dual-Fan, Dual-Duct System
Fig. 13 Multizone System
Fig. 14 Primary/Secondary System
Fig. 15 Underfloor Air Distribution
Fig. 16 Supersaturated/Wetted Coil
--- CHAPTER 05: IN-ROOM TERMINAL SYSTEMS ---
1. System Characteristics
Advantages
Disadvantages
Heating and Cooling Calculations
Space Heating
Central (Primary-Air) Ventilation Systems
Central Plant Sizing
Building Pressurization
First, Operating, and Maintenance Costs
Energy
Life-Cycle Costs
2. System Components and Configurations
Components
Configurations
3. Secondary-Water Distribution
4. Piping Arrangements
Four-Pipe Distribution
Two-Pipe Distribution
Three-Pipe Distribution
Condenser Water Systems with Heat Pump Terminal Units
5. Fan-Coil Unit and Unit Ventilator Systems
Types and Location
Ventilation Air Requirements
Selection
Wiring
Condensate
Capacity Control
Maintenance
6. Variable-Refrigerant-Flow (VRF) Units
7. Chilled-Beam Systems
Types and Location
Ventilation Air Requirements
Selection
Wiring
Condensate
Capacity Control
Maintenance
Other Concerns
8. Radiant-Panel Heating Systems
Types and Location
Ventilation Air Requirements
Selection
Wiring
Capacity Control
Maintenance
9. Radiant-Floor Heating Systems
Types and Location
Ventilation Air Requirements
Selection
Wiring
Capacity Control
Maintenance
10. Induction Unit Systems
11. Supplemental Heating Units
12. Primary-Air Systems
13. Performance Under Varying Load
14. Changeover Temperature
15. Two-Pipe Systems with Central Ventilation
Critical Design Elements
Changeover Temperature Considerations
Nonchangeover Design
Zoning
Room Control
Evaluation
Electric Heat for Two-Pipe Systems
16. Four-Pipe Systems
Zoning
Room Control
Evaluation
17. Automatic Controls and Building Management Systems
18. Maintenance Management Systems and Building System Commissioning
References
Bibliography
Figures
Fig. 1 Typical Fan-Coil Unit
Fig. 2 Passive and Active Chilled-Beam Operation
Fig. 3 Primary-Air System
Fig. 4 Solar Radiation Variations with Seasons
Fig. 5 Capacity Ranges of In-Room Terminal Operating on Two-Pipe System
Fig. 6 Primary-Air Temperature Versus Outdoor Air Temperature
Fig. 7 Psychrometric Chart, Two-Pipe System, Off-Season Cooling
Fig. 8 Typical Changeover System Temperature Variation
Fig. 9 Typical Nonchangeover System Variations
Fig. 10 Fan-Coil Unit Control
--- CHAPTER 06: RADIANT HEATING AND COOLING ---
1. PRINCIPLES OF RADIANT SYSTEMS
1.1 Heat Transfer
Heat Transfer by Thermal Radiation
Heat Transfer by Natural Convection
Combined Heat Flux (Thermal Radiation and Natural Convection)
1.2 Factors Affecting Heat Transfer
Panel Thermal Resistance
Effect of Floor Coverings
Panel Heat Losses or Gains
Panel Performance
1.3 Panel Design
Special Cases
Examples
2. General Design Considerations
2.1 Hybrid Systems
3. RADIANT HEATING AND COOLING SYSTEMS
3.1 Hydronic Ceiling Panels
3.2 Embedded Systems with Tubing in Ceilings, Walls, or Floors
Hydronic Wall Panels
Hydronic Floor Panels
3.3 Electrically Heated Radiant Systems
Electric Ceiling Panels
Electric Wall Heating
Electric Floor Heating
4. DESIGN PROCEDURE
Sensible Cooling
Sensible Heating
Other Steps Common for Sensible Heating and Cooling
4.1 Controls
Sensible Cooling Controls
Heating Slab Controls
References
Bibliography
Tables
Table 1 Thermal Resistance of Ceiling Panels
Table 2 Thermal Conductivity of Typical Tube Material
Table 3 Thermal Resistance of Floor Coverings
Table 4 Characteristics of Typical Electric Panels
Figures
Fig. 1 Radiation Heat Flux at Heated Ceiling, Floor, or Wall Panel Surfaces
Fig. 2 Heat Removed by Radiation at Cooled Ceiling or Wall Panel Surface
Fig. 3 Natural-Convection Heat Transfer at Floor, Ceiling, and Wall Panel Surfaces
Fig. 4 Empirical Data for Heat Removal by Ceiling Cooling Panels from Natural Convection
Fig. 5 Relation of Inside Surface Temperature to Overall Heat Transfer Coefficient
Fig. 6 Inside Surface Temperature Correction for Exposed Wall at Dry-Bulb Air Temperatures Other Than 21°C
Fig. 7 Cooled Ceiling Panel Performance in Uniform Environment with No Infiltration and No Internal Heat Sources
Fig. 8 Downward and Edgewise Heat Loss Coefficient for Concrete Floor Slabs on Grade
Fig. 9 Design Graph for Sensible Heating and Cooling with Floor and Ceiling Panels
Fig. 10 Design Graph for Heating with Aluminum Ceiling and Wall Panels
Fig. 11 Typical Residential Hybrid HVAC System
Fig. 12 Metal Ceiling Panels Attached to Pipe Laterals
Fig. 13 Metal Ceiling Panels Bonded to Copper Tubing
Fig. 14 Extruded Aluminum Panels with Integral Copper Tube
Fig. 15 Permitted Design Ceiling Surface Temperatures at Various Ceiling Heights
Fig. 16 Coils in Structural Concrete Slab
Fig. 17 Coils in Plaster Above Lath
Fig. 18 Coils in Plaster Below Lath
Fig. 19 Coils in Floor Slab on Grade
Fig. 20 Embedded Tube in Thin Slab
Fig. 21 Tube in Subfloor
Fig. 22 Tube Under Subfloor
Fig. 23 Electric Heating Panels
Fig. 24 Electric Heating for Wet Plaster Ceiling
Fig. 25 Electric Heating Cable in Concrete Slab
Fig. 26 Primary/Secondary Water Distribution System with Mixing Control
Fig. 27 Split Panel Piping Arrangement for Two-Pipe and Four-Pipe Systems
--- CHAPTER 07: COMBINED HEAT AND POWER SYSTEMS ---
1. Terminology
2. CHP System Concepts
2.1 Custom-Engineered Systems
2.2 Packaged and Modular Systems
2.3 Load Profiling and Prime Mover Selection
2.4 Peak Load Shaving
2.5 Continuous-Duty Standby
2.6 Power Plant Incremental Heat Rate
3. Performance Parameters
3.1 Heating Value
3.2 CHP Electric Effectiveness
Power and Heating Systems
3.3 Fuel Energy Savings
4. Fuel-to-Power Components
4.1 Reciprocating Engines
Types
Performance Characteristics
Fuels and Fuel Systems
Combustion Air
Lubricating Systems
Starting Systems
Cooling Systems
Exhaust Systems
Emissions
Instruments and Controls
Noise and Vibration
Installation Ventilation Requirements
Operation and Maintenance
4.2 Combustion Turbines
Types
Advantages
Disadvantages
Gas Turbine Cycle
Components
4.3 Performance Characteristics
Fuels and Fuel Systems
Combustion Air
Lubricating Systems
Starting Systems
Exhaust Systems
Emissions
Instruments and Controls
Noise and Vibration
Operation and Maintenance
4.4 Fuel Cells
Types
5. Thermal-to-Power Components
5.1 Steam Turbines
Types
Performance Characteristics
Fuel Systems
Lubricating Oil Systems
Power Systems
Exhaust Systems
Instruments and Controls
Operation and Maintenance
5.2 Organic Rankine Cycles
5.3 Expansion Engines/Turbines
5.4 Stirling Engines
Types
Performance Characteristics
Fuel Systems
Power Systems
Exhaust Systems
Coolant Systems
Operation and Maintenance
6. Thermal-to-Thermal Components
6.1 Thermal Output Characteristics
Reciprocating Engines
Combustion Turbines
6.2 Heat Recovery
Reciprocating Engines
Combustion Turbines
Steam Turbines
6.3 Thermally Activated Technologies
Heat-Activated Chillers
Desiccant Dehumidification
Hot Water and Steam Heat Recovery
Thermal Energy Storage Technologies
7. Electrical Generators and Components
7.1 Generators
8. System Design
8.1 CHP Electricity-Generating Systems
Thermal Loads
Prime Mover Selection
Air Systems
Hydronic Systems
Service Water Heating
District Heating and Cooling
Utility Interfacing
Power Quality
Output Energy Streams
8.2 CHP Shaft-Driven HVAC and Refrigeration Systems
Engine-Driven Systems
Combustion-Turbine-Driven Systems
Steam-Turbine-Driven Systems
9. Codes and Installation
9.1 General Installation Parameters
9.2 Utility Interconnection
9.3 Air Permits
9.4 Building, Zoning, and Fire Codes
Zoning
Building Code/Structural Design
Mechanical/Plumbing Code
Fire Code
Electrical Connection
10. Economic Evaluation
CHP Application Assessment
Types and Scope of CHP Studies
CHP System Modeling Techniques
CHP Feasibility Study for New Facilities
Tools and Software for Feasibility Study
10.1 Load Profiles and Load Duration Curves
Load Duration Curve Analysis
Two-Dimensional Load Duration Curve
Analysis by Simulations
References
Bibliography
Tables
Table 1 Applications and Markets for DG/CHP Systems
Table 2 Values of  for Conventional Thermal Generation Technologies
Table 3 Summary of Results from Examples 1 to 5
Table 4 Summary of Results Assuming 33% Efficient Combustion Turbine
Table 5 Typical  Values
Table 6 Summary of Fuel Energy Savings for 25% Power Generator in Examples 1 to 5
Table 7 Summary of Fuel Energy Savings for 33% Power Generator in Examples 1 to 5
Table 8 Reciprocating Engine Types by Speed (Available MW Ratings)
Table 9 Line Regulator Pressures
Table 10 Ventilation Air for Engine Equipment Rooms
Table 11 Exhaust Pipe Diameter
Table 12 Recommended Engine Maintenance
Table 13 Overview of Fuel Cell Characteristics
Table 14 Theoretical Steam Rates For Turbines at Common Conditions, kgkWh
Table 15 NEMA Classification of Speed Governors
Table 16 Temperatures Normally Required for Various Heating Applications
Table 17 Full-Load Exhaust Mass Flows andTemperatures for Various Engines
Table 18 Generator Control Functions
Table 19 Coefficient of Performance (COP) of Engine-Driven Chillers
Table 20 Typical Efficiency of Engine-Driven Refrigeration Equipment (Ammonia Screw Compressor)
Figures
Fig. 1 CHP Cycles
Fig. 2 Dual-Service Applications
Fig. 3 Conventional Boiler for Example 1
Fig. 4 Power-Only Generator for Example 1
Fig. 5 Performance Parameters for Combined Systemfor Example 2
Fig. 6 CHP Power and Heating Energy Boundary Diagram for Example 2
Fig. 7 Performance Parameters for Example 3
Fig. 8 CHP Power and Direct Heating Energy Boundary Diagram for Example 3
Fig. 9 Performance Parameters for Example 4
Fig. 10 CHP Power and HRSG Heating Without Duct Burner Energy Boundary Diagram for Example 4
Fig. 11 Cofiring Performance Parameters for Example 5
Fig. 12 CHP Power and HRSG Heating with Duct Burner Energy Boundary Diagram for Example 5
Fig. 13 Electric Effectiveness E Versus Overall Efficiency O
Fig. 14 Efficiency (HHV) of Spark Ignition Engines
Fig. 15 Heat Rate (HHV) of Spark Ignition Engines
Fig. 16 Thermal-to-Electric Ratio of Spark Ignition Engines (Jacket and Exhaust Energy)
Fig. 17 Part-Load Heat Rate (HHV) of 1430, 425, and 85 kW Gas Engines
Fig. 18 Part-Load Thermal-to-Electric Ratio of 1430, 425, and 85 kW Gas Engines
Fig. 19 Typical Reciprocating Engine Exhaust Noise Curves
Fig. 20 Typical Attenuation Curves for Engine Silencers
Fig. 21 Temperature-Entropy Diagram for Brayton Cycle
Fig. 22 Simple-Cycle Single-Shaft Turbine
Fig. 23 Simple-Cycle Dual-Shaft Turbines
Fig. 24 Turbine Engine Performance Characteristics
Fig. 25 Gas Turbine Refrigeration System Using Exhaust Heat
Fig. 26 CHP System Boundary
Fig. 27 PAFC
Fig. 28 SOFC
Fig. 29 MCFC
Fig. 30 PEMFC
Fig. 31 AFC
Fig. 32 Basic Types of Axial Flow Turbines
Fig. 33 Isentropic Versus Actual Turbine Process
Fig. 34 Efficiency of Typical Multistage Turbines
Fig. 35 Effect of Inlet Pressure and Superheat on Condensing Turbine
Fig. 36 Effect of Exhaust Pressure on Noncondensing Turbine
Fig. 37 Single-Stage Noncondensing Turbine Efficiency
Fig. 38 Effect of Extraction Rate on Condensing Turbine
Fig. 39 Oil Relay Governor
Fig. 40 Part-Load Turbine Performance Showing Effect of Auxiliary Valves
Fig. 41 Multivalve Oil Relay Governor
Fig. 42 Cutaway Core of a Kinematic Stirling Engine
Fig. 43 Cutaway Core of a Free-Piston Stirling Engine
Fig. 44 Heat Balance for Naturally Aspirated Engine
Fig. 45 Heat Balance for Turbocharged Engine
Fig. 46 Hot-Water Heat Recovery
Fig. 47 Hot-Water Engine Cooling with Steam Heat Recovery (Forced Recirculation)
Fig. 48 Engine Cooling with Gravity Circulation and Steam Heat Recovery
Fig. 49 Lubricant and Aftercooler System
Fig. 50 Exhaust Heat Recovery with Steam Separator
Fig. 51 Effect of Soot on Energy Recovery from Flue Gas Recovery Unit on Diesel Engine
Fig. 52 Automatic Boiler System with Overriding Exhaust Temperature Control
Fig. 53 Combined Exhaust and Jacket Water Heat Recovery System
Fig. 54 Effect of Lowering Exhaust Temperature below 150°C
Fig. 55 Back-Pressure Turbine
Fig. 56 Integration of Back-Pressure Turbine with Facility
Fig. 57 Condensing Automatic Extraction Turbine
Fig. 58 Automatic Extraction Turbine CHP System
Fig. 59 Performance Map of Automatic Extraction Turbine
Fig. 60 Hybrid Heat Recovery Absorption Chiller-Heater
Fig. 61 Typical Generator Efficiency
Fig. 62 Typical Heat Recovery Cycle for Gas Turbine
Fig. 63 Performance Curve for Typical 350 kW, Gas-Engine-Driven, Reciprocating Chiller
Fig. 64 Typical Gas Turbine Refrigeration Cycle
Fig. 65 Condensing Turbine-Driven Centrifugal Compressor
Fig. 66 Combination Centrifugal-Absorption System
Fig. 67 Hypothetical Steam Load Profile
Fig. 68 Load Duration Curve
Fig. 69 Load Duration Curve with Multiple Generators
Fig. 70 Hypothetical Peaking Generator
Fig. 71 Example of Two-Dimensional Load Duration Curve
--- CHAPTER 08: COMBUSTION TURBINE INLET COOLING ---
1. Advantages
Economic Benefits
Environmental Benefits
2. Disadvantages
3. Definition and Theory
4. System Types
Evaporative Systems
Chiller Systems
LNG Vaporization Systems
Hybrid Systems
5. Calculation of Power Capacity Enhancement and Economics
References
Bibliography
Tables
Table 1 Typical Combined-Cycle, Simple-Cycle, and Steam Turbine Systems
Figures
Fig. 1 Effect of Ambient Temperature on CT Output
Fig. 2 Effect of Ambient Temperature on CT Heat Rate
Fig. 3 Effects of Ambient Temperature on Thermal Energy, Mass Flow Rate and Temperature of CT Exhaust Gases
Fig. 4 Typical Hourly Power Demand Profile
Fig. 5 Example of Daily System Load and Electric Energy Pricing Profiles
Fig. 6 Schematic Flow Diagram of Typical Combustion
Turbine System
Fig. 7 Psychrometric Chart Showing Direct and Indirect Inlet Air Cooling Processes
Fig. 8 Typical Wetted Media
Fig. 9 Water Fog Created for Fogging System
Fig. 10 Cooling Coil Used for Indirect Cooling Chiller Systems
Fig. 11 Examples of Effect of CTIC Technology and Ambient Condition on Capacity Enhancement Potential
Fig. 12 Example of Effects of CTIC Technology and Ambient Condition on Unitized Capital Cost of
Capacity Enhancement
--- CHAPTER 09: APPLIED HEAT PUMP AND HEAT RECOVERY SYSTEMS ---
1. TERMINOLOGY
2. APPLIED HEAT PUMP SYSTEMS
2.1 Heat Pump Cycles
2.2 Heat Sources and Sinks
Air
Water
Ground
Solar Energy
2.3 Types of Heat Pumps
2.4 Heat Pump Components
Compressors
Heat Transfer Components
Refrigeration Components
Controls
Supplemental Heating
2.5 Industrial Process Heat Pumps
Closed-Cycle Systems
Open-Cycle and Semi-Open-Cycle Heat Pump Systems
Heat Recovery Design Principles
3. APPLIED HEAT RECOVERY SYSTEMS
3.1 Waste Heat Recovery
General Considerations
Applications of Waste Heat Recovery
Alternative Heat Sources
Locating the Heat Recovery Heat Pump
Specific Considerations of Condenser-Side Recovery
Specific Considerations of Evaporator-Side Recovery
Special Considerations of Double-Bundle Heat Recovery
Selecting a Compressor Type
Pumping Considerations
HRHP Selection
Example
3.2 Water-Loop Heat Pump Systems
Description
Design Considerations
Controls
Advantages of a WLHP System
Limitations of a WLHP System
3.3 Balanced Heat Recovery Systems
Definition
Heat Redistribution
Heat Balance Concept
Heat Balance Studies
General Applications
Multiple Buildings
3.4 Heat Pumps in District Heating and Cooling Systems
References
Bibliography
Tables
Table 1 Heat Pump Sources and Sinks
Figures
Fig. 1 Closed Vapor Compression Cycle
Fig. 2 Mechanical Vapor Recompression Cycle with
Heat Exchanger
Fig. 3 Open Vapor Recompression Cycle
Fig. 4 Heat-Driven Rankine Cycle
Fig. 5 Heat Pump Types
Fig. 6 Comparison of Parallel and Staged Operation
for Air-Source Heat Pumps
Fig. 7 Suction Line Separator for Protection Against Liquid Floodback
Fig. 8 Liquid Subcooling Coil in Ventilation Air Supply
to Increase Heating Effect and Heating COP
Fig. 9 Typical Increase in Heating Capacity Resulting from
Using Liquid Subcooling Coil
Fig. 10 Dehumidification Heat Pump
Fig. 11 Air-to-Water Heat Pump
Fig. 12 Cooling Tower Heat Recovery Heat Pump
Fig. 13 Effluent Heat Recovery Heat Pump
Fig. 14 Refrigeration Heat Recovery Heat Pump
Fig. 15 Condensing Ammonia Heat Pump
Fig. 16 Closed-Cycle Vapor Compression System
Fig. 17 Recompression of Boiler-Generated Process Steam
Fig. 18 Single-Effect Heat Pump Evaporator
Fig. 19 Multiple-Effect Heat Pump Evaporator
Fig. 20 Distillation Heat Pump System
Fig. 21 Heat Recovery Heat Pump System in a Rendering Plant
Fig. 22 Semi-Open-Cycle Heat Pump in a Textile Plant
Fig. 23 Possible Heat Recovery Heat Pump Locations
Fig. 24 Primary/Secondary, Equal Loading
Fig. 25 Variable Primary Flow Example
Fig. 26 HRHP Application Flowchart
Fig. 27 Heat Balance Chart
Fig. 28 Selection of Simultaneous
Heating and Cooling HRHP
Fig. 29 Operating Areas for Simultaneous HRHP
Fig. 30 Heat Recovery System Using Water-to-Air Heat
Pumps in a Closed Loop
Fig. 31 Closed-Loop Heat Pump System with Thermal
Storage and Optional Solar-Assist Collectors
Fig. 32 Secondary Heat Recovery from WLHP System
Fig. 33 Cooling Tower with Heat Exchanger
Fig. 34 Major Load Components
Fig. 35 Composite Plot of Loads in Figure 34
Fig. 36 Non-Heat-Recovery System
--- CHAPTER 10: SMALL FORCED-AIR HEATING AND COOLING SYSTEMS ---
1. Components
Heating and Cooling Units
Ducts
Accessory Equipment
Controls
2. Common System Problems
3. System Design
Estimating Heating and Cooling Loads
Locating Outlets, Returns, Ducts, and Equipment
Selecting Heating and Cooling Equipment
Determining Airflow Requirements
Finalize Duct Design and Size
Selecting Supply and Return Grilles and Registers
4. Detailed Duct Design
Detailing the Duct Configuration
Detailing the Distribution Design
Duct Design Recommendations
Zone Control for Small Systems
Duct Sizing for Zone Damper Systems
Box Plenum Systems Using Flexible Duct
Embedded Loop Ducts
5. Small Commercial Systems
Air Distribution in Small Commercial Buildings
Controlling Airflow in New Buildings
6. Testing for Duct Efficiency
Data Inputs
Data Output
Standards
References
Bibliography
Tables
Table 1 General Characteristics of Supply Outlets
Table 2 Recommended Division of Duct Pressure Loss
Figures
Fig. 1 Heating and Cooling Components
Fig. 2 Sample Floor Plans for Locating Ductwork in Second Floor of (A) Two-Story House and (B) Townhouse
Fig. 3 Sample Floor Plans for One-Story House with (A) Dropped Ceilings, (B) Ducts in Conditioned Spaces, and(C) Right-Sized Air Distribution in Conditioned Spaces
Fig. 4 (A) Ducts in Unconditioned Spaces and (B) Standard Air Distribution System in Unconditioned Spaces
Fig. 5 Entrance Fittings to Eliminate Unstable Airflow in Box Plenum
Fig. 6 Dimensions for Efficient Box Plenum
--- CHAPTER 11: STEAM SYSTEMS ---
1. Advantages
2. Fundamentals
3. Effects of Water , Air , and Gases
4. Heat Transfer
5. Basic Steam System Design
6. Steam Source
Boilers
Heat Recovery and Waste Heat Boilers
Heat Exchangers
7. Boiler Connections
Supply Piping
Return Piping
8. Design Steam Pressure
9. Piping
Supply Piping Design Considerations
Terminal Equipment Piping Design Considerations
Return Piping Design Considerations
10. Condensate Removal from Temperature-Regulated Equipment
11. Steam Traps
Thermostatic Traps
Mechanical Traps
Kinetic Traps
12. Pressure-Reducing Valves
Installation
Valve Size Selection
13. Terminal Equipment
Selection
Natural Convection Units
Forced-Convection Units
14. Convection Steam Heating
One-Pipe Steam Heating Systems
Two-Pipe Steam Heating Systems
15. Steam Distribution
16. Temperature Control
17. Heat Recovery
Flash Steam
Direct Heat Recovery
18. Combined Steam and Water Systems
19. Commissioning
References
Bibliography
Tables
Table 1 Properties of Saturated Steam
Table 2 Pressure Differential Temperature Control
Figures
Fig. 1 Exhaust Heat Boiler
Fig. 2 Typical Boiler Connections
Fig. 3 Boiler with Gravity Return
Fig. 4 Method of Dripping Steam Mains
Fig. 5 Trap Discharging to Overhead Return
Fig. 6 Trapping Strainers
Fig. 7 Trapping Multiple Coils
Fig. 8 Recommended Steam Trap Piping
Fig. 9 Trapping Temperature-Regulated Coils
Fig. 10 Steam Traps
Fig. 11 Pressure-Reducing Valve Connections: Single-Stage
Fig. 12 Two-Stage Pressure-Reducing Valve
Fig. 13 One-Third/Two-Thirds Pressure-Reducing Valve Station
Fig. 14 One-Pipe System
Fig. 15 Two-Pipe System
Fig. 16 Inlet Orifice
Fig. 17 Orifice Capacities for Different Pressure Differentials
Fig. 18 Flash Steam
Fig. 19 Vertical Flash Tank
Fig. 20 Flash Tank Diameters
--- CHAPTER 12: DISTRICT HEATING AND COOLING ---
Applicability
Components
Environmental Benefits
1. SYSTEM MASTER PLANNING
1.1 Economic Considerations
Consumer Economics
Producer Economics
District Energy Economic Comparison
2. CENTRAL PLANT
2.1 Heating and Cooling Production
Heating Medium
Steam and Hot Water Generation
Chilled-Water Generation
Thermal Storage
Auxiliaries
2.2 Chilled-Water Distribution Design Considerations
Constant Flow
Variable Flow
Chilled-Water System Design Guidelines
3. DISTRIBUTION SYSTEM
3.1 Hydraulic Considerations
Objectives of Hydraulic Design
Water Hammer
Pressure Losses
Pipe Sizing
Network Calculations
Condensate Drainage and Return in Steam Systems
3.2 Thermal Considerations
Thermal Design Conditions
Thermal Properties of Pipe Insulation and Soil
3.3 Methods of Heat Transfer Analysis
Calculation of Undisturbed Soil Temperatures
Convective Heat Transfer at Ground Surface
Uninsulated Buried Pipe
Insulated Buried Pipe
Buried Pipe in Conduit with Air Space
Buried Pipe with Composite Insulation
Two Pipes Buried in Common Conduit with Air Space
Two Buried Pipes or Conduits
Pipes in Buried Trenches or Tunnels
Pipes in Shallow Trenches
Buried Pipes with Other Geometries
Pipes in Air
Economical Thickness for Pipe Insulation
3.4 Expansion Provisions
Pipe Supports, Guides, and Anchors
3.5 Distribution System Construction
Tables
Table 1 Summary of Economic Analysis Factors
Table 2 Annual Utility Consumption Summary for Central Plant Alternatives
Table 3 Estimates of Annual Maintenance and Utility Costsand Commercial Tower Costs
Table 4 LLC Example of District Cooling Evaluation
Table 5 Chiller Technology
Table 6 Summary Table of Chiller Characteristics
Table 7 Comparison of Commonly Used Insulations in Underground Piping Systems
Table 8 Effect of Moisture on Underground Piping System Insulations
Table 9 Soil Thermal Conductivities
Table 10 Effect of Interior Insulation Thickness on Exterior Insulation Temperature (Example 6)
Table 11 Effects of Soil Thermal Conductivity and Burial Depth on Exterior Insulation Temperature (Example 6)
Table 12 Relative Merits of Piping Materials Commonly Used for District Cooling Distribution Systems
Table 13 Relative Merits of Direct and Indirect Consumer Interconnection
Table 14 Conversion Suitability of Heating System by Type
Table 15 Measuring Points and Derivative Parametersfor Remote Monitoring and Control of Indirect
Table 16 Flowmeter Characteristics
Figures
Fig. 1 Major Components of District Heating System
Fig. 2 Master Planning Pyramid
Fig. 3 Breakdown of Life-Cycle Cost by Components for Example 1
Fig. 4 Layout for Hot-Water/Chilled-Water Plant
Fig. 5 Constant-Flow Primary Distribution with Secondary Pumping
Fig. 6 Variable-Flow Primary/Secondary Systems
Fig. 7 Uninsulated Buried Pipe
Fig. 8 Insulated Buried Pipe with Air Space
Fig. 9 Two Pipes Buried in Common Conduit with Air Space
Fig. 10 Two Buried Pipes or Conduits
Fig. 11 Pipes in Buried Trenches or Tunnels
Fig. 12 Slide and Guide Detail
Fig. 13 Relative Costs for Piping Alone, Uninsulated
Fig. 14 Walk-Through Tunnel
Fig. 15 Concrete Surface Trench
Fig. 16 Deep-Bury Small Tunnel (Boxed Conduit)
Fig. 17 Poured Insulation System
Fig. 18 Field-Installed, Direct-Buried Rigid Closed-CellInsulation System
Fig. 19 Conduit System Components
Fig. 20 Corrosion Rate in Aggressive Environment Similar to Mild Steel Casings in Soil
Fig. 21 Conduit System with Annular Air Space and Single Carrier Pipe
Fig. 22 Conduit System with Two Carrier Pipes and Annular Air Space
Fig. 23 Conduit System with Single Carrier Pipe and No Air Space (WSL)
Fig. 24 Conduit Casing Temperature Versus Soil Thermal Conductivity
Fig. 25 Direct Connection of Building System to District Chilled Water with Building Pumps
Fig. 26 Direct Connection of Building System to District Chilled Water Without Building Pumps
Fig. 27 Direct Connection of Building System to District Hot Water
Fig. 28 Indirect Connection of Building System to District Chilled Water
Fig. 29 Basic Cascading Indirect Heating-System Schematic
Fig. 30 District/Building Interconnection with Heat Recovery Steam System
Fig. 31 District/Building Interconnection with Heat Exchange Steam System
Fig. 32 Plate Heat-Exchanger Performance with Constant Flow on Customer Side and Customer-Side Supply Temperature of 5.6°C
Fig. 33 Plate Heat-Exchanger Performance with Variable Flow on Customer Side and Customer-Side Supply Temperature of 5.6°C
--- CHAPTER 13: HYDRONIC HEATING AND COOLING ---
Principles
1. TEMPERATURE CLASSIFICATIONS
2. CLOSED WATER SYSTEMS
2.1 Method of Design
2.2 Thermal Components
Loads
Load Devices
Source
Expansion Chamber
2.3 Hydraulic Components
Pump or Pumping System
Variable-Speed Pumping Application
Pump Connection
Distribution System
Expansion Chamber
2.4 Piping Circuits
2.5 Capacity Control of Load System
Sizing Control Valves
Alternatives to Control Valves
2.6 Low-Temperature Heating Systems
Nonresidential Heating Systems
2.7 Chilled-Water Systems
2.8 Dual-Temperature Systems
Two-Pipe Systems
Two-Pipe Dual-Temperature Chilled-Water Systems
Four-Pipe Common Load Systems
Four-Pipe Independent Load Systems
2.9 Other Design Considerations
Makeup and Fill Water Systems
Safety Relief Valves
Air Elimination
Drain and Shutoff
Balance Fittings
Pitch
Strainers
Thermometers
Flexible Connectors and Pipe Expansion Compensation
Gage Cocks
Insulation
Condensate Drains
Common Pipe
2.10 Other Design Procedures
Preliminary Equipment Layout
Final Pipe Sizing and Pressure Drop Determination
Freeze Prevention
2.11 Antifreeze Solutions
Effect on Heat Transfer and Flow
Effect on Heat Source or Chiller
Effect on Terminal Units
Effect on Pump Performance
Effect on Piping Pressure Loss
Installation and Maintenance
References
Bibliography
Tables
Table 1 Chilled-Water Coil Performance
Figures
Fig. 1 Fundamental Components of Hydronic System
Fig. 2 Henry’s Constant Versus Temperature for Air and Water
Fig. 3 Solubility Versus Temperature and Pressure for Air/Water Solutions
Fig. 4 Example of Manufacturer’s Published Pump Curve
Fig. 5 Pump Curve and System Curve
Fig. 6 Shift of System Curve Caused by Circuit Unbalance
Fig. 7 General Pump Operating Condition Effects
Fig. 8 Operating Conditions for Parallel-Pump Installation
Fig. 9 Operating Conditions for Series-Pump Installation
Fig. 10 Compound Pumping (Primary/Secondary Pumping)
Fig. 11 Example of Variable-Speed Pump System Schematic
Fig. 12 Example of Variable-Speed Pump and System Curves
Fig. 13 System Curve with System Static Pressure(Control Area)
Fig. 14 Typical System Curves for Closed System
Fig. 15 Tank Pressure Related to System Pressure
Fig. 16 Effect of Expansion Tank Location with Respect to Pump Pressure
Fig. 17 Flow Diagram of Simple Series Circuit
Fig. 18 Series Loop System
Fig. 19 One-Pipe Diverting Tee System
Fig. 20 Series Circuit with Load Pumps
Fig. 21 Direct- and Reverse-Return Two-Pipe Systems
Fig. 22 Load Control Valves
Fig. 23 System Flow with Two-Way and Three-Way Valves
Fig. 24 Chilled-Water Coil Heat Transfer Characteristic
Fig. 25 Equal-Percentage Valve Characteristic with Authority
Fig. 26 Control Valve and Coil Response, Inherent and 50% Authority
Fig. 27 Control Valve and Coil Response, 33% Authority
Fig. 28 Coil Valve and Coil Response, 10% Authority
Fig. 29 Load Pumps with Valve Control
Fig. 30 Schematic of Variable-Speed Pump Coil Control
Fig. 31 Example of Series-Connected Loading
Fig. 32 Heat Emission Versus Flow Characteristic of Typical Hot-Water Heating Coil
Fig. 33 Generic Chilled-Water Coil Heat Transfer Characteristic
Fig. 34 Recommendations for Coil Flow Tolerance to Maintain 97% Design Heat Transfer
Fig. 35 Constant-Flow Chilled-Water System
Fig. 36 Variable-Flow Chilled-Water System
Fig. 37 Simplified Diagram of Two-Pipe System
Fig. 38 Dual-Temperature Chilled-Water Systems Using Sources with Different Design Chilled-Water Temperatures
Fig. 39 Dual-Temperature Chilled-Water System Using Three Chillers
Fig. 40 Example Four-Pipe Common Load System
Fig. 41 Four-Pipe Independent Load System
Fig. 42 Typical Makeup Water and Expansion Tank Piping Configuration for Plain Steel Expansion Tank
Fig. 43 Pressure Increase Resulting from Thermal Expansion as Function of Temperature Increase
Fig. 44 Example of Effect of Aqueous Ethylene Glycol Solutions on Heat Exchanger Output
Fig. 45 Effect of Viscosity on Pump Characteristics
Fig. 46 Pressure Drop Correction for Glycol Solutions
--- CHAPTER 14: CONDENSER WATER SYSTEMS ---
1. Once-Through City Water Systems
2. Open Cooling Tower Systems
Air and Vapor Precautions
Pump Selection and Pressure Calculations
Water Treatment
Freeze Protection and Winter Operation
3. Low-Temperature (Water Economizer) Systems
4. Closed-Circuit Evaporative Coolers
5. Other Sources of Water
6. Overpressure Caused by Thermal Fluid Expansion
Bibliography
Figures
Fig. 1 Condenser Connections for Once-Through City Water System
Fig. 2 Cooling Tower Piping System
Fig. 3 Schematic Piping Layout Showing Static and Suction Pressure
Fig. 4 Cooling Tower Piping to Avoid Freeze-Up
Fig. 5 Closed-Circuit Evaporative Cooler System
--- CHAPTER 15: MEDIUM- AND HIGH-TEMPERATURE WATER HEATING ---
1. System Characteristics
3. Design Considerations
Direct-Contact Heaters (Cascades)
System Circulating Pumps
4. Distribution Piping Design
5. Heat Exchangers
6. Air-Heating Coils
7. Space-Heating Equipment
8. Instrumentation and Controls
9. Water Treatment
10. Heat Storage
11. Safety Considerations
References
Bibliography
Tables
Table 1 Properties of Water, 100 to 210°C
Figures
Fig. 1 Relation of Saturation Pressure and Enthalpy to Water Temperature
Fig. 2 Elements of High-Temperature Water System
Fig. 3 Density and Specific Heat of Water
Fig. 4 Inert-Gas Pressurization for Primary-Pumped System
Fig. 5 Inert-Gas Pressurization for Primary/Secondary-Pumped System
Fig. 6 Inert-Gas Pressurization Using Variable Gas Quantity with Gas Recovery
Fig. 7 Cascade HTW System
Fig. 8 Cascade HTW System Combined with Boiler Feedwater Preheating
Fig. 9 Control Diagram for HTW Generator
Fig. 10 Heat Exchanger Connections
--- CHAPTER 16: INFRARED RADIANT HEATING ---
1. Energy Conservation
2. Infrared Energy Sources
Gas Infrared
Electric Infrared
Oil Infrared
3. System Efficiency
4. Reflectors
5. Controls
6. Precautions
7. Maintenance
8. Design Considerations for Beam Radiant Heaters
References
Bibliography
Tables
Table 1 Characteristics of Typical Gas-Fired Infrared Heaters
Table 2 Characteristics of Four Electric Infrared Elements
Figures
Fig. 1 Types of Gas-Fired Infrared Heaters
Fig. 2 Common Electric Infrared Heaters
Fig. 3 Relative Absorptance and Reflectance of Skin andTypical Clothing Surfaces
Fig. 4 Projected Area Factor for Seated Persons, Nude and Clothed
Fig. 5 Projected Area Factor for Standing Persons, Nude and Clothed
Fig. 6 Radiant Heat Flux Distribution Curve of Typical Narrow-Beam High-Intensity Electric Infrared Heaters
Fig. 7 Radiant Heat Flux Distribution Curve of Typical Broad-Beam High-Intensity Electric Infrared Heaters
Fig. 8 Radiant Heat Flux Distribution Curve of Typical Narrow-Beam High-Intensity Atmospheric Gas-Fired Infrared Heaters
Fig. 9 Radiant Heat Flux Distribution Curve of Typical Broad-Beam High-Intensity Atmospheric Gas-Fired Infrared Heaters
Fig. 10 Calculation of Total ERF from Three Gas-Fired Heaters on Worker Standing at Positions A Through E
--- CHAPTER 17: ULTRAVIOLET LAMP SYSTEMS ---
1. Terminology
2. UVGI Fundamentals
Microbial Dose Response
Susceptibility of Microorganisms to UV Energy
3. Lamps and power supplies
Types of UV-C Lamps
UV-C Lamp Drivers or Ballasts
Germicidal Lamp Cooling and Heating Effects
UV-C Lamp Aging
UV-C Lamp Irradiance
UV-C Photodegradation of Materials
4. Maintenance
Lamp Replacement
Lamp Disposal
Visual Inspection
5. Safety
Hazards of Ultraviolet Radiation to Humans
Sources of UV Exposure
Exposure Limits
UV Radiation Measurements for Upper Air Applications
Safety Design Guidance
Personnel Safety Training
Lamp Breakage
6. Unit Conversions
References
Bibliography
Tables
Table 1 Representative Members of Organism Groups
Table 2 Summary of UV-C Resistance Ratings of Tested Materials (High UV Dose Results)
Table 3 Permissible Exposure Times for Given Effective Irradiance Levels of UV-C Energy at 253.7 nm
Figures
Fig. 1 Relative UV-C Germicidal Efficiency
Fig. 2 General Ranking of Susceptibility to UV-C Inactivation of Microorganisms by Group
Fig. 3 Typical Hot-Cathode UVGI Lamp
Fig. 4 Example of Lamp Output as Function of Cold-Spot Temperature
Fig. 5 Windchill Effect on Typical Low-Pressure UV-C Lamp Efficiency in Moving Airstream at 21°C
Fig. 6 Diagram of Irradiance Calculation
--- CHAPTER 18: VARIABLE REFRIGERANT FLOW ---
System Types
VRF Applications
Zoned Comfort
Indoor Air Quality
Annual Operating Efficiency Characteristics
Local and Remote Monitoring
Life-Cycle Cost Comparison
1. Standards
2. Equipment
Air-Source Outdoor and Water-Source Units
Indoor Unit Types
System Controls
System Expansion or Reconfiguration
3. VRF System Operation
Load Management
Cooling Operation
Heating Operation
Saturation Temperature Reset
Heat Recovery Operation
Defrost Operation
Oil Recovery Management
Humidity Control
High-Heating-Performance Air-Source VRF Units
4. Modeling Considerations
5. Design Considerations
Water-Source VRF Systems
Air-Source VRF Systems
Low External Ambient Heating-Dominant Applications
Integration with Supplemental Heating Sources
Outdoor Air Economizer
Generating Radiant Heating/Cooling and Domestic Hot Water
6. VRF System Design Example
Performing a Load-Profile Analysis
System Type Selection, Zoning, and Potential for Heat Recovery
Accurately Sizing Air-Source Outdoor and Indoor Units
Selecting Indoor Units
Ventilation Air Strategy
Refrigerant Piping
Refrigerant Piping Guidelines
Controls
Safety Considerations for Refrigerants
Fault Tree Analysis
Optimizing VRF Systems to Minimize Environmental Impact
7. Commissioning
References
Bibliography
Tables
Table 1 VRF Multisplit System Classifications
Table 2 Examples of Typical Refrigerant Values
Table 3 Minimum Fan-Cooling Unit Size Requiring Economizer for Comfort Cooling
Table 4 Peak Load Profile
Figures
Fig. 1 Compressor Frequency
Fig. 2 Cooling-Only and Heat Pump VRF System
Fig. 3 Two-Pipe Heat Recovery VRF System
Fig. 4 Three-Pipe Heat Recovery VRF System Examples
Fig. 5 Common VRF Indoor Unit Types
Fig. 6 Controls System Example
Fig. 7 Typical VRF Enthalpy-Pressure Chart
Fig. 8 Two-Pipe VRF Heat Recovery System Operation
Fig. 9 Three-Pipe Heat Recovery System Operation
Fig. 10 Multilayer Heat Recovery Operation
Fig. 11 Flash Injection Schematic
Fig. 12 Staged Compression Cycle Schematic
Fig. 13 Example System Zoning
Fig. 14 Heating Output Derating Chart
Fig. 15 Maximum Refrigerant Piping Length Capacity Correction Factor
Fig. 16 Cooling Output Derating Chart
Fig. 17 Indoor Unit Layout
--- CHAPTER 19: DUCT CONSTRUCTION ---
1. Building Code Requirements
2. Pressure Classifications
3. Duct Cleaning
4. HVAC System Leakage
System Sealing
Sealants
Leakage Testing
Responsibilities
5. Air-Handling Unit Leakage
6. Residential and Commercial Duct Construction
Terminology
Buildings and Spaces
Round, Flat Oval, and Rectangular Ducts
Fibrous Glass Ducts
Phenolic Ducts
Flexible Ducts
Hangers and Supports
Installation
Plenums and Apparatus Casings
Acoustical Treatment
7. Industrial Duct Construction
Materials
Round Ducts
Rectangular Ducts
Construction Details
Hangers
8. Antimicrobial-Treated Ducts
9. Duct Construction for Grease- and Moisture-Laden Vapors
Factory-Built Grease Duct Systems
Site-Built Grease Duct Systems
Duct Systems for Moisture-Laden Air
10. Rigid Plastic Ducts
11. Air Dispersion Systems
Dispersion Types
12. Underground Ducts
13. Ducts Outside Buildings
14. Seismic Qualification
15. Sheet Metal Welding
16. Thermal Insulation
17. Specifications
References
Bibliography
Tables
Table 1 Pressure Classification for Ductworka
Table 2 Recommended Maximum System Leakage Percentages
Table 3 Leakage as Percentage of Flow
Table 4 Example 1 System Parameters
Table 5 Solution for Example 1
Table 6A Galvanized Sheet Thickness
Table 6B Uncoated Steel Sheet Thickness
Table 6C Stainless Steel Sheet Thickness
Figures
Fig. 1 Hierarchy of Building Codes and Standards
Fig. 2 Fabric Duct with Linear Vent Outlet
Fig. 3 Fabric Duct with Nozzle Outlets
Fig. 4 Fabric Duct with Porous Material as Air Outlet
--- CHAPTER 20: ROOM AIR DISTRIBUTION EQUIPMENT ---
1. Systems Overview
All-Air Systems
Decoupled Cooling Systems
Sensible-Only Decoupled Cooling Systems
2. System Classifications
2.1 Fully Mixed Systems
Factors That Influence Selection
Outlet Selection Procedure
2.2 Fully Stratified Systems
Factors that Influence Selection
Outlet Selection Procedure
2.3 Partially Mixed Systems
Factors That Influence Selection
Outlet Selection Procedures
3. EQUIPMENT
3.1 Supply air outlets
3.2 Return and Exhaust Air Inlets
3.3 Grilles
Types
Application-Specific Grilles
3.4 Nozzles and Drum Louvers
3.5 Diffusers
Types
Accessories
3.6 Terminal Units
Single-Duct Terminal Units
Dual-Duct Terminal Units
Air-to-Air Induction Terminal Units
Fan-Powered Terminal Units
3.7 Fan-Coil Units
3.8 Chilled Beams
Beam Types and Configurations
3.9 Air Curtain Units
References
Bibliography
Tables
Table 1 Typical Applications for Supply Air Outlets
Figures
Fig. 1 Designations for Inlet and Outlet
Fig. 2 Classification of Air Distribution Strategies
Fig. 3 Accessories for Supply Air Grilles
Fig. 4 Accessory Controls for Ceiling Diffusers
Fig. 5 Basic Fan-Coil Unit
Fig. 6 Typical Vertical Fan-Coil Unit
Fig. 7 Typical Horizontal Fan-Coil Unit
Fig. 8 (A) Active and (B) Passive Beams
Fig. 9 Non-Recirculating, Horizontal-Mount High-Velocity Air Curtain Unit
Fig. 10 Recirculating, Horizontal-Mount Air Curtain Unit
--- CHAPTER 21: FANS ---
1. Types of Fans
2. Principles of Operation
3. Testing and Rating
4. Field Testing of Fans for Air Performance
5. Fan Laws
6. Fan and System Pressure Relationships
7. AIR Temperature Rise Across Fans
8. Duct System Characteristics
9. System Effects
10. Selection
11. Parallel Fan Operation
12. Series Fan Operation
13. Noise
14. Vibration
Vibration Isolation
15. Arrangement and Installation
16. Fan Control
17. Fan Inlet Cone Instrumented for Airflow Measurement
18. Symbols
References
Bibliography
Tables
Table 1 Types of Fans
Table 2 Fan Laws
Table 3 Fan Application Categories for Balance and Vibration
Table 4 BV Categories and Balance Quality Grades
Table 5 Seismic Vibration Velocity Limits for In Situ Operation
Table 6 Summary of Control Strategies
Figures
Fig. 1 Centrifugal Fan Components
Fig. 2 Axial Fan Components
Fig. 3 Method of Obtaining Fan Performance Curves
Fig. 4 Example Application of Fan Laws
Fig. 5 Pressure Changes for Fan with Equal-Sized Ducted Inlet and Outlet Systems
Fig. 6 Pressure Changes for Fan with Outlet System Only
Fig. 7 Pressure Changes for Fan with Ducted Inlet System Only
Fig. 8 Pressure Changes for Fan with Diverging Cone at Outlet
Fig. 9 Simple Duct System with Resistance to Flow Represented by Three 90° Elbows
Fig. 10 Example System Total Pressure Loss P Curves
Fig. 11 Resistance Added to Duct System of Figure 9
Fig. 12 Resistance Removed from Duct System of Figure 9
Fig. 13 Fan Performance Curve
Fig. 14 Desirable Combination of Ptf andP Curves
Fig. 15 Two (A) Stable and (B) Unstable Fans in Parallel Operation
Fig. 16 Theoretical Characteristic Curve of Two Fans Operating in Series
Fig. 17 Effect of Inlet Vane Control on Backward-Curved Centrifugal Fan Performance
Fig. 18 Effect of Controlled Blade Pitch on Vaneaxial Fan Performance
Fig. 19 Illustration of Instrumented Fan Inlet Cone
--- CHAPTER 22: HUMIDIFIERS ---
1. Environmental Conditions
Health and Comfort
Prevention and Treatment of Disease
Electronic Equipment
Process Control and Materials Storage
Static Electricity
Sound Wave Transmission
Miscellaneous
2. Enclosure Characteristics
Vapor Retarders
Visible Condensation
Concealed Condensation
3. Energy and water Considerations
Load Calculations
Design Conditions
Ventilation Rate
Additional Moisture Losses
Internal Moisture Gains
Supply Water for Humidifiers
Scaling
Potential Bacterial Growth
4. Equipment
Residential Humidifiers for Central Air Systems
Residential Humidifiers for Nonducted Applications
Industrial and Commercial Humidifiers for Central Air Systems
Selecting Humidifiers
5. Controls
Mechanical Controls
Electronic Controls
Control Location
Management Systems
6. Application Considerations
Humidity Control with Direct Space Humidification
Humidity Control with Duct-Mounted Humidification
Humidity Control in Variable-Air-Volume Systems
Commissioning Systems
References
Bibliography
Tables
Table 1 Maximum Relative Humidity in a Space for
No Condensation on Windows
Table 2 Types of Humidifiers
Table 3 Humidifier Advantages and Limitations
Figures
Fig. 1 Optimum Humidity Range for Human Comfort and Health
Fig. 2 Patient Infections at Indoor Relative Humidities
Fig. 3 Mice Survival Rates at 20 and 50% rh
Fig. 4 Mortality of Pneumococcus Bacterium
Fig. 5 Mortality in Mice Exposed to Aerosolized Influenza
Fig. 6 Effect of Relative Humidity on Static Electricityfrom Carpets
Fig. 7 Limiting Relative Humidity for No Window
Condensation
Fig. 8 Effect of Relative Humidity on Static Electricity
from Carpets
Fig. 9 Residential Humidifiers
Fig. 10 Industrial Isothermal (Steam) Humidifiers
Fig. 11 Room Fan Distributor
Fig. 12 Industrial Adiabatic (Atomizing and Evaporative) Humidifiers
Fig. 13 Recommended Humidity Controller Location
--- CHAPTER 23: AIR-COOLING AND DEHUMIDIFYING COILS ---
1. Uses for Coils
2. Coil Construction and Arrangement
Water and Aqueous Glycol Coils
Direct-Expansion Coils
Control of Coils
Flow Arrangement
Applications
3. Coil Selection
Performance and Ratings
4. Airflow Resistance
5. Heat Transfer
6. Performance of Sensible Cooling Coils
7. Performance of Dehumidifying Coils
8. Determining Refrigeration Load
9. Maintenance
10. Symbols
References
Bibliography
Figures
Fig. 1 Typical Water Circuit Arrangements
Fig. 2 Arrangements for Coils with Multiple Thermostatic Expansion Valves
Fig. 3 Typical Coil Hand Designation
Fig. 4 Typical Arrangement of Cooling Coil Assembly in Built-Up or Packaged Central Station Air Handler
Fig. 5 Coil Bank Arrangement with Intermediate Condensate Pan
Fig. 6 Sprayed-Coil System with Air Bypass
Fig. 7 Two-Component Driving Force Between Dehumidifying Air and Coolant
Fig. 8 Surface Temperature Chart
Fig. 9 Thermal Diagram for General Case When Coil Surface Operates Partially Dry
Fig. 10 Leaving-Air Dry-Bulb Temperature Determi
Fig. 11 Typical Total Metal Thermal Resistance of Fin and Tube Assembly
Fig. 12 Typical Air-Side Application Rating Data Determined Experimentally for Cooling and Dehumidifying Water Coils
Fig. 13 Psychrometric Performance of Cooling and Dehumidifying Coil
--- CHAPTER 24: DESICCANT DEHUMIDIFICATION ANDPRESSURE-DRYING EQUIPMENT ---
1. Methods of Dehumidification
2. Desiccant Dehumidification
2.1 Liquid Desiccant Equipment
2.2 Solid-Sorption Equipment
2.3 Rotary Solid-Desiccant Dehumidifiers
2.4 Equipment Ratings
2.5 Equipment Operating Recommendations
2.6 Applications for Atmospheric- Pressure Dehumidification
3. Desiccant Drying at Elevated Pressure
3.1 Equipment Types
3.2 Applications
References
Bibliography
Additional Information
Figures
Fig. 1 Methods of Dehumidification
Fig. 2 Flow Diagram for Liquid-Absorbent Dehumidifier (A) Without and (B) With Extended Surface Contact Medium
Fig. 3 Flow Diagram for Liquid-Absorbent Unit with Onboard Refrigeration System
Fig. 4 Lithium Chloride Equilibrium
Fig. 5 Liquid Desiccant System with Multiple Conditioners
Fig. 6 Liquid Desiccant Regenerator Capacity
Fig. 7 Typical Rotary Dehumidification Wheel
Fig. 8 Effect of Changes in Process Air Velocity on Dehumidifier Outlet Moisture
Fig. 9 Effect of Changes in Process Air Inlet Moisture on Dehumidifier Outlet Moisture
Fig. 10 Effect of Changes in Reactivation Air Inlet Temperature on Dehumidifier Outlet Moisture
Fig. 11 Effect of Changes in Process Air Inlet Moisture on Dehumidifier Outlet Temperature
Fig. 12 Effect of Changes in Reactivation Air Inlet Temperature on Dehumidifier Outlet Temperature
Fig. 13 Typical Performance Data for Rotary Solid Desiccant Dehumidifier
Fig. 14 Typical Peak Moisture Loads for Medium-Sized Retail Store in Atlanta, Georgia
Fig. 15 Predrying Ventilation Air to Dehumidify a Commercial Building
Fig. 16 Typical Rooftop Arrangement for Drying Ventilation Air Centrally, Removing Moisture Load from Cooling Units
Fig. 17 Typical Performance Data for Solid Desiccant Dryersat Elevated Pressures
Fig. 18 Typical Adsorption Dryer for ElevatedPressures
--- CHAPTER 25: MECHANICAL DEHUMIDIFIERS AND RELATED COMPONENTS ---
1. Mechanical Dehumidifiers
Psychrometrics of Dehumidification
Residential Dehumidifiers
General-Purpose Dehumidifiers
DX Dedicated Outdoor Air System (DOAS) Units
Indoor Swimming Pool Dehumidifiers
Ice Rink Dehumidifiers
Industrial Dehumidifiers
Dehumidifiers for Controlled Environment Agriculture
Tunnel Dryer Dehumidifier
2. Controls and Sensors
3. Installation and Service Considerations
4. Wraparound Heat Exchangers
References
Bibliography
Figures
Fig. 1 Dehumidification Process Points
Fig. 2 Psychrometric Diagram of Typical Dehumidification Process
Fig. 3 Psychrometric Diagram of Typical Dehumidification Process with Bypass Air
Fig. 4 Dehumidification Process Points with Bypass Air
Fig. 5 Typical Portable Dehumidifier
Fig. 6 Typical Whole-House Dehumidifier Installation
Fig. 7 Typical General-Purpose Dehumidifier
Fig. 8 DX-DOAS Unit
Fig. 9 DX-DOAS Unit with Exhaust Air Heat/Energy Recovery
Fig. 10 Typical Single-Blower Pool Dehumidifier
Fig. 11 Typical Double-Blower Pool Dehumidifier with DX Coil in Supply Air Section
Fig. 12 Typical Double-Blower Pool Dehumidifier with DX Coil in Return Air Section
Fig. 13 Supply Blower and Double Exhaust Blower Pool Dehumidifier
Fig. 14 Typical Installation of Ice Rink Dehumidifiers
Fig. 15 Typical Tunnel Dryer Dehumidifier
Fig. 16 Schematic of Dehumidification Enhancement with Wraparound Heat Pipe
Fig. 17 Enhanced Dehumidification Processwith Wraparound Heat Pipe
Fig. 18 Slide-in Heat Pipe for Rooftop Air Conditioner Refit
--- CHAPTER 26: AIR-TO-AIR ENERGY RECOVERY EQUIPMENT ---
1. Applications
2. Basic heat or heat and water vapor transfer relations
Effectiveness
Rate of Energy Transfer
Fan Power
3. Types of Air-to-Air Heat Exchangers
Ideal Air-to-Air Energy Exchange
Fixed-Plate Heat Exchangers
Rotary Air-to-Air Energy Exchangers
Coil Energy Recovery (Runaround) Loops
Heat Pipe Heat Exchangers
Thermosiphon Heat Exchangers
Liquid-Desiccant Cooling Systems
Twin-Tower Enthalpy Recovery Loops
Fixed-Bed Regenerators
4. Performance Ratings
Performance Ratings for Air-to-Air Heat or Heat and Mass Exchangers
Performance Ratings for Residential Ventilators with Air-to-Air Heat or Heat and Mass Exchangers
5. Additional technical considerations
Air Leakage
Air Capacity of Ventilator Fans
Pressure Drop
Maintenance
Filtration
Controls
Fouling
Corrosion
Condensation and Freeze-Up
Frost Control Strategies for Air-to-Air Energy Recovery Systems
Indirect Evaporative Air Cooling
Use of Economizer
6. Comparison of Air-to-Air Heat or Heat and Mass exchanger characteristics
7. Use of Air-to-Air Heat or Heat and Mass Exchangers in Systems
Characterizing System Efficiency of Heat or Energy Recovery Ventilators
Selection of Heat or Energy Recovery Ventilators
Systems with Multiple Energy Recovery Exchangers
Using Air-to-Air Heat Exchangers to Modify the Latent Capacity Ratio of Cooling Coils
Dessicant and Heat Wheel Systems
8. Economic Considerations
9. Energy and/or Mass Recovery Calculation Procedure
10. Symbols
References
Bibliography
Tables
Table 1 Typical Applications for Air-to-Air Energy Recovery
Table 2 Types of Liquid-Desiccant Solutions
Table 3 Comparison of Air-to-Air Energy Recovery Devices
Figures
Fig. 1 Airstream Numbering Convention
Fig. 2 Fixed-Plate Cross-Flow Heat Exchanger
Fig. 3 Variation of Pressure Drop and Effectiveness with Airflow Rates for Membrane Plate Exchanger
Fig. 4 Typical Temperature Stratification at Outlets of Cross-Flow Heat Exchanger
Fig. 5 Plate Heat or Heat and Mass Exchanger Airflow Configurations
Fig. 6 Rotary Air-to-Air Energy Exchanger
Fig. 7 Latent and Sensible Effectiveness Versus (A) Wheel Speed and (B) Bypassed Air
Fig. 8 Effectiveness of Counterflow Regenerator as a Function of NTU
Fig. 9 Coil Energy Recovery Loop
Fig. 10 Heat Pipe Assembly
Fig. 11 Heat Pipe Operation
Fig. 12 Heat Pipe Exchanger Effectiveness
Fig. 13 Heat Pipe Heat Exchanger with Tilt Control
Fig. 14 Sealed-Tube Thermosiphons
Fig. 15 Coil-Type Thermosiphon Loops
Fig. 16 Heat Recovered Across Evaporator and Condenser Coils for (A) Level Thermosiphon and (B) Thermosiphon with Elevated Condenser
Fig. 17 Eight-Row, Unidirectional System with 2 mm Fin Spacing; Static Charge of 80%
Fig. 18 Typical Performance of Two-Phase Thermosiphon Loop as Function of Percent Static Charge: Eight Rows, Unidirectional System with 2 mm Fin Spacing; Overall Temperature Difference of 35 K
Fig. 19 Typical Dehumidification of Outdoor Air by Liquid-Desiccant System
Fig. 20 Twin-Tower System Performing Indirect Evaporative Air Cooling: (A) Schematic and (B) Installation in Building
Fig. 21 Double-Core Fixed-Bed Regenerators
Fig. 22 Typical Temperature Profile of Fixed-Bed Regenerator Operating at 60 s Recovery Periods
Fig. 23 Single-Core Fixed-Bed Regenerator
Fig. 24 Air Leakage in Energy Recovery Units
Fig. 25 Common Frost Control Strategies for Fixed-Plate Exchangers
Fig. 26 Examples of Frost Threshold Temperatures for Energy and Heat Recovery Exchangers
Fig. 27 Indirect Evaporative Cooling Recovery (Example 4)
Fig. 28 Economizer with Dessicant Wheel
Fig. 29 Supply Conditions for a Given Load qt and Given SHR
Fig. 30 Multiple Energy Recovery Exchangers in Parallel Mode
Fig. 31 Psychrometric Processes of Exchangers in Parallel Mode
Fig. 32 Multiple Energy Recovery Exchangers in Parallel Mode
Fig. 33 Psychrometric Processes of Exchangers in Series Mode
Fig. 34 Psychrometric Chart: In-Series Energy Recovery
Fig. 35 Wraparound (In-Series) Heat Pipe
Fig. 36 Flat Heat Recovery Device in Wraparound (In-Series) Configuration
Fig. 37 Flat Plate Heat Exchanger in Wraparound (In-Series) Configuration
Fig. 38 Wraparound (In-Series) Thermosiphon with a Valvefor Modulation
Fig. 39 Precooling Air Reheater Dehumidifier (Example 5)
Fig. 40 Heat Pump Augmented by Heat and Dessicant Wheels
Fig. 41 Heat Pump Augmented by Evaporative Cooler and Heat and Dessicant Wheels
Fig. 42 System for Surgery Room
Fig. 43 Dessicant System of Evaporative Coolers and Heatand Dessicant Wheels
Fig. 44 Dessicant and Heat Wheels with Indirect Evaporative Cooler (M-Cyle)
Fig. 45 Dessicant and Heat Wheels with Humidifier for Winter Heating
Fig. 46 Schematic of Liquid Dessicant System and Evaporative Cooler
Fig. 47 Maximum Sensible and Latent Heat from Process A-B
Fig. 48 Sensible Heat Recovery in Winter (Example 6)
Fig. 49 Sensible Heat Recovery in Winter with Condensate (Example 6)
Fig. 50 Total Heat Recovery in Summer (Example 7)
Fig. 51 Total Energy Recovery with EATR ≠ 0 and OACF ≠
1 (Example 8)
Fig. 52 Actual Airflow Rates at Various State Points (Example 8)
--- CHAPTER 27: AIR-HEATING COILS ---
1. Coil Construction and Design
Steam Coils
Water/Aqueous Glycol Heating Coils
Volatile Refrigerant Heat Reclaim Coils
Electric Heating Coils
2. Coil Selection
Coil Ratings
Overall Requirements
3. Installation Guidelines
4. Coil Maintenance
References
Tables
Table 1 Preferred Operating Limits for Continuous-Duty Steam Coil Materials in Commercial andInstitutional Applications
Table 2 Typical Maximum Condensate Loads
--- CHAPTER 28: UNIT VENTILATORS, UNIT HEATERS, AND MAKEUP AIR UNITS ---
1. Unit Ventilators
Application
Selection
Control
2. Unit Heaters
Application
Selection
Control
Piping Connections
Maintenance
3. Makeup Air Units
Description and Applications
Selection
Control
Applicable Codes and Standards
Commissioning
Maintenance
References
Bibliography
Tables
Table 1 Typical Unit Ventilator Capacities
Figures
Fig. 1 Typical Unit Ventilators
Fig. 2 Methods of Preventing Downdraft along Windows
Fig. 3 Typical Unit Heaters
Fig. 4 Hot Water and Steam Connections for Unit Heaters
--- CHAPTER 29: AIR CLEANERS FOR PARTICULATE CONTAMINANTS ---
1. terminology
Definitions
Acronyms
2. Atmospheric Aerosols
3. Aerosol Characteristics
4. Air-Cleaning Applications
5. Mechanisms of Particle Collection
6. Evaluating Air Cleaners
7. Air Cleaner Test Methods
Arrestance Test
Dust-Holding Capacity (DHC) Test
Particle Size Removal Efficiency (PSE) Test
DOP Penetration Test
Leakage (Scan) Tests
ISO Standard 29462
Other Performance Tests
Environmental Tests
AHRI Standards
8. Types of Air Cleaners
9. Filter Types and Performance
Panel Filters
Electronic Air Cleaners
10. Selection and Maintenance
Residential Air Cleaners
VAV Systems
Antimicrobial Treatment of Filter Media
11. Air Cleaner Installation
12. Safety Considerations
References
Bibliography
Tables
Table 1 U.S. EPA Standards for Particulate Matter in Outdoor Air
Table 2 Performance of Renewable Media Filters (Steady-State Values)
Table 3 Cross-Reference and Application Guidelines
Figures
Fig. 1 Collection Mechanisms for Filter Fiber
Fig. 2 Cross Section of Plate-Type Precipitator Air Cleaner
Fig. 3 Electrically Enhanced Air Cleaner
Fig. 4 Typical Filter Locations for HVAC System
--- CHAPTER 30: INDUSTRIAL GAS CLEANING AND AIR
POLLUTION CONTROL ---
Equipment Selection
1. Regulations and Monitoring
Gas-Cleaning Regulations
Measuring Gas Streams and Contaminants
Gas Flow Distribution
Monitors and Controls
2. Particulate Contaminant Control
2.1 Mechanical Collectors
Settling Chambers
Inertial Collectors
2.2 Electrostatic Precipitators
Single-Stage Designs
Two-Stage Designs
2.3 Fabric Filters
Principle of Operation
Pressure-Volume Relationships
Electrostatic Augmentation
Fabrics
Types of Self-Cleaning Mechanisms for Fabric Dust Collectors
2.4 Granular-Bed Filters
Principle of Operation
2.5 Particulate Scrubbers (Wet Collectors)
Principle of Operation
Spray Towers and Impingement Scrubbers
Centrifugal-Type Collectors
Orifice-Type Collectors
Venturi Scrubber
Electrostatically Augmented Scrubbers
Collector Performance
3. Gaseous Contaminant Control
3.1 Spray Dry Scrubbing
Principle of Operation
Equipment
3.2 Wet-Packed Scrubbers
Scrubber Packings
Arrangements of Packed Scrubbers
Pressure Drop
Absorption Efficiency
General Efficiency Comparisons
Liquid Effects
3.3 Adsorption of Gaseous Contaminants
Equipment for Adsorption
Solvent Recovery
Odor Control
Applications of Fluidized Bed Adsorbers
3.4 Incineration of Gases and Vapors
Thermal Oxidizers
Catalytic Oxidizers
Applications of Oxidizers
Adsorption and Oxidation
4. Auxiliary Equipment
4.1 Ducts
Temperature Controls
Fans
4.2 Dust- and Slurry-Handling Equipment
Hoppers
Dust Conveyors
Dust Disposal
Slurry Treatment
5. Operation and Maintenance
Corrosion
Fires and Explosions
References
Bibliography
Tables
Table 1 Intended Duty of Gas-Cleaning Equipment
Table 2 Principal Types of Particulate Control Equipment
Table 3 Measures of Performance for Gas-Cleaning Equipment
Table 4 Collectors Used in Industry
Table 5 Terminal Settling Velocities of Particles, m/s
Table 6 Temperature Limits and Characteristics of Fabric Filter Media
Table 7 Packing Factor F for Various Scrubber Packing Materials
Table 8 Mass Transfer Coefficients (KG a) for Scrubber Packing Materials
Table 9 Relative KGa for Various Contaminants in Liquid-Film-Controlled Scrubbers
Table 10 Relative KGa for Various Contaminants in Gas-Film-Controlled Scrubbers
Figures
Fig. 1 Typical Louver and Baffle Collectors
Fig. 2 Typical Cyclone Collectors
Fig. 3 Cyclone Efficiency
Fig. 4 Typical Single-Stage Electrostatic Precipitator
Fig. 5 Typical Two-Stage Electrostatic Precipitators
Fig. 6 Typical Single-Stage Precipitators
Fig. 7 Condensing Precipitator Systems for Control of Hot Organic Smokes
Fig. 8 Time Dependence of Pressure Drop Across Fabric Filter
Fig. 9 Bag-Type Shaker Collector
Fig. 10 Envelope-Type Shaker Collector
Fig. 11 Pressure Drop Across Shaker Collector Versus Time
Fig. 12 Draw-Through Reverse-Flow Cleaning of Fabric Filter
Fig. 13 Typical Pulse-Jet Fabric Filter
Fig. 14 Pulse Jet Cartridge Filters (Upflow Design with Vertical Filters)
Fig. 15 Typical Granular-Bed Filter
Fig. 16 Fractional Efficiency of Several Wet Collectors
Fig. 17 Efficiency of Venturi Scrubber
Fig. 18 Typical Spray Tower
Fig. 19 Typical Impingement Scrubber
Fig. 20 Typical Orifice-Type Wet Collector
Fig. 21 Typical High-Energy Venturi Scrubber
Fig. 22 Typical Electrostatically Augmented Scrubber
Fig. 23 Typical Packings for Scrubbers
Fig. 24 Flow Arrangements Through Packed Beds
Fig. 25 Typical Countercurrent Packed Scrubber
Fig. 26 Horizontal Flow Scrubber with Extended Surface
Fig. 27 Vertical Flow Scrubber with Extended Surface
Fig. 28 Pressure Drop Versus Gas Rate for Typical Packing
Fig. 29 Generalized Pressure Drop Curves for Packed Beds
Fig. 30 Contaminant Control at Superficial Velocity = 0.6 m/s (Liquid-Film-Controlled)
Fig. 31 Contaminant Control at Superficial
Velocity = 1.2 m/s (Liquid-Film-Controlled)
Fig. 32 Contaminant Control at Superficial Velocity = 1.8 m/s (Liquid-Film-Controlled)
Fig. 33 Contaminant Control at Superficial Velocity = 0.6 m/s (Gas-Film-Controlled)
Fig. 34 Contaminant Control at Superficial Velocity = 1.2 m/s (Gas-Film-Controlled)
Fig. 35 Contaminant Control at Superficial Velocity = 1.8 m/s (Gas-Film-Controlled)
Fig. 36 Adsorption Isotherms on Activated Carbon
Fig. 37 Fluidized-Bed Adsorption Equipment
Fig. 38 Schematic of Two-Unit Fixed Bed Adsorber
Fig. 39 Moving-Bed Adsorber
Fig. 40 Typical Odor Adsorber
--- CHAPTER 31: AUTOMATIC FUEL-BURNING SYSTEMS ---
1. GENERAL CONSIDERATIONS
1.1 Terminology
1.2 System Application
1.3 Safety
1.4 Efficiency and Emission Ratings
Steady-State and Cyclic Efficiency
Emissions
2. GAS-BURNING APPLIANCES
2.1 Gas-Fired Combustion Systems
Burners
Combustion System Flow
Ignition
Input Rate Control
2.2 Residential Appliances
Boilers
Forced-Air Furnaces
Water Heaters
Combination Space- and Water-Heating Appliances
Pool Heaters
Conversion Burners
2.3 Commercial-Industrial Appliances
Boilers
Space Heaters
Water Heaters
Pool Heaters
2.4 Applications
Location
Gas Supply and Piping
Air for Combustion and Ventilation
Draft Control
Venting
Building Depressurization
Gas Input Rate
Effect of Gas Temperature and Barometric Pressure Changes on Gas Input Rate
Fuel Gas Interchangeability
Altitude
3. OIL-BURNING APPLIANCES
3.1 Residential Oil Burners
3.2 Commercial/Industrial Oil Burners
Pressure-Atomizing Oil Burners
Return-Flow Pressure-Atomizing Oil Burners
Air-Atomizing Oil Burners
Horizontal Rotary Cup Oil Burners
Steam-Atomizing Oil Burners (Register Type)
Mechanical Atomizing Oil Burners (Register Type)
Return-Flow Mechanical Atomizing Oil Burners
3.3 Dual-Fuel Gas/Oil Burners
3.4 Equipment Selection
Fuel Oil Storage Systems
Fuel-Handling Systems
Fuel Oil Preparation System
4. SOLID-FUEL-BURNING APPLIANCES
4.1 Capacity Classification of Stokers
4.2 Stoker Types by Fuel-Feed Methods
Chain and Traveling Grate Stokers
Vibrating Grate Stokers
5. CONTROLS
5.1 Safety Controls and Interlocks
Ignition and Flame Monitoring
Draft Proving
Limit Controls
Other Safety Controls
Prescriptive Requirements for Safety Controls
Reliability of Safety Controls
5.2 Operating Controls
Integrated and Programmed Controls
References
Bibliography
Tables
Table 1 Classification of Atomizing Oil Burners
Table 2 Guide for Fuel Oil Grades Versus Firing Rate
Table 3 Characteristics of Various Types of Stokers (Class 5)
Figures
Fig. 1 Partially Aerated (Bunsen) Burner
Fig. 2 Premix Burner
Fig. 3 Forced-Draft Combustion System
Fig. 4 Induced-Draft Combustion System
Fig. 5 Packaged Power Burner
Fig. 6 Combustion System with Linked Air and Gas Flow
Fig. 7 Pneumatically Linked Gas/Air Ratio Combustion System
Fig. 8 Typical Single-Port Upshot Gas Conversion Burner
Fig. 9 Altitude Effects on Gas Combustion Appliances
Fig. 10 High-Pressure Atomizing Gun Oil Burner
Fig. 11 Details of High-Pressure Atomizing Oil Burner
Fig. 12 Typical Oil Storage Tank (No. 6 Oil)
Fig. 13 Industrial Burner Auxiliary Equipment
Fig. 14 Horizontal Underfeed Stoker with Single Retort
Fig. 16 Chain Grate Stoker
Fig. 17 Vibrating Grate Stoker
Fig. 18 Basic Control Circuit for Fuel-Burning Appliance
Fig. 19 Control Characteristics of Three-Stage System
Fig. 20 Integrated Control System for Gas-Fired Appliance
--- CHAPTER 32: BOILERS ---
1. Classifications
Working Pressure and Temperature
Fuel Used
Construction Materials
Type of Draft
Condensing or Noncondensing
Wall-Hung Boilers
Integrated (Combination) Boilers
Electric Boilers
2. Selection Parameters
3. Efficiency: Input and Output Ratings
4. Performance Codes and Standards
5. Sizing
6. Burner Types
7. Boiler Controls
Operating Controls
Water Level Controls
8. Flame Safeguard Controls
References
Bibliography
Figures
Fig. 1 Residential Boilers
Fig. 2 Cast-Iron Commercial Boilers
Fig. 3 Scotch Marine Commercial Boilers
Fig. 4 Commercial Fire-Tube and Water-Tube Boilers
Fig. 5 Commercial Condensing Boilers
Fig. 6 Effect of Inlet Water Temperature on
Efficiency of Condensing Boilers
Fig. 7 Relationship of Dew Point, Carbon Dioxide, and Combustion Efficiency for Natural Gas
Fig. 8 Boiler Efficiency as Function of Fuel and Air Input
--- CHAPTER 33: FURNACES ---
1. Components
Casing or Cabinet
Heat Exchangers
Heat Sources
Combustion Venting Components
Circulating Blowers and Motors
Filters and Other Accessories
Airflow Variations
Combustion System Variations
Indoor/Outdoor Furnace Variations
2. Heat Source Types
Natural Gas and Propane Furnaces
Oil Furnaces
Electric Furnaces
3. Commercial Equipment
Ducted Equipment
Unducted Heaters
4. Controls and Operating Characteristics
External to Furnace
Internal to Furnace
5. Equipment Selection
Distribution System
Equipment Location
Forced-Air System Primary Use
Fuel Selection
Combustion Air and Venting
Equipment Sizing
Types of Furnaces
Consumer Considerations
Selecting Furnaces for Commercial Buildings
6. Calculations
7. Technical Data
Natural Gas Furnaces
Propane Furnaces
Oil Furnaces
Electric Furnaces
Commercial Furnaces
8. Installation
9. Agency Listings
References
Bibliography
Tables
Table 1 Historical and Typical Values of Efficiency
Figures
Fig. 1 Induced-Draft Gas Furnace
Fig. 2 Upflow Category I Furnace with Induced-Draft Blower
Fig. 3 Downflow (Counterflow) Category I Furnacewith Induced-Draft Blower
Fig. 4 Horizontal Category I Furnace withInduced-Draft Blower
Fig. 5 Basement (Lowboy) Category I Furnace withInduced-Draft Blower
Fig. 6 Terminology Used to Describe Fan-Assisted Combustion
Fig. 7 Electric Forced-Air Furnace
Fig. 8 Standing Floor Furnace
--- CHAPTER 34: RESIDENTIAL IN-SPACE HEATING EQUIPMENT ---
1. GAS IN-SPACE HEATERS
1.1 Controls
Valves
Thermostats
1.2 Vent Connectors
1.3 Sizing Units
Room Heaters
Wall Furnaces
Floor Furnaces
U.S. Minimum Efficiency Requirements
2. OIL AND KEROSENE IN-SPACE HEATERS
Vaporizing Oil Pot Heaters
Powered Atomizing Heaters
Portable Kerosene Heaters
3. ELECTRIC IN-SPACE HEATERS
3.1 Radiant Heating Systems
Heating Panels and Heating Panel Sets
Embedded Cable and Storage Heating Systems
Cord-Connected Portable Heaters
Controls
Wall, Floor, Toe Space, and Ceiling Heaters
Baseboard Heaters
4. SOLID-FUEL IN-SPACE HEATERS
4.1 Fireplaces
Simple Fireplaces
Factory-Built Fireplaces
Freestanding Fireplaces
4.2 Stoves
Conventional Wood Stoves
Advanced-Design Wood Stoves
Fireplace Inserts
Pellet-Burning Stoves
5. GENERAL INSTALLATION PRACTICES
Safety with Solid Fuels
Utility-Furnished Energy
Products of Combustion
Agency Testing
References
Bibliography
Tables
Table 1 Efficiency Requirements in the United States for Gas-Fired Direct Heating Equipment
Table 2 Gas Input Required for In-Space Supplemental Heaters
Table 3 Solid-Fuel In-Space Heaters
Table 4 Chimney Connector Wall Thickness
Figures
Fig. 1 Room Heater
Fig. 2 Wall Furnace
Fig. 3 Floor Furnace
Fig. 4 Oil-Fueled Heater with Vaporizing Pot-Type Burner
--- CHAPTER 35: CHIMNEY, VENT, AND FIREPLACE SYSTEMS ---
1. Terminology
2. Draft Operating Principles
3. Chimney Functions
Start-Up
Air Intakes
Vent Size
Draft Control
Pollution Control
Equipment Location
Wind Effects
Safety Factors
4. Steady-State Chimney Design Equations
Mass Flow of Combustion Products in Chimneys and Vents
Mean Chimney Gas Temperature and Density
Theoretical Draft
System Pressure Loss Caused by Flow
Available Draft
Chimney Gas Velocity
System Resistance Coefficient
Configuration and Manifolding Effects
Input, Diameter, and Temperature Relationships
Volumetric Flow in Chimney or System
5. Steady-State Chimney Design Graphical Solutions
6. Vent and Chimney Capacity Calculation Examples
7. Gas Appliance Venting
Vent Connectors
Masonry Chimneys for Gas Appliances
Type B and Type L Factory-Built Venting Systems
Gas Appliances Without Draft Hoods
Conversion to Gas
8. Oil-Fired Appliance Venting
Condensation and Corrosion
Connector and Chimney Corrosion
Vent Connectors
Masonry Chimneys for Oil-Fired Appliances
Replacement of Appliances
9. Fireplace Chimneys
10. Air Supply to Fuel-Burning Appliances
11. Vent and Chimney Materials
12. Vent and Chimney Accessories
Draft Hoods
Draft Regulators
Vent Dampers
Heat Exchangers or Flue Gas Heat Extractors
13. Draft Fans
14. Terminations: Caps and Wind Effects
15. Codes and Standards
16. Symbols
References
Bibliography
Tables
Table 1 Mass Flow Equations for Common Fuels
Table 2 Typical Chimney and Vent Design Conditionsa
Table 3 Mass Flow for Incinerator Chimneys
Table 4 Mean Chimney Gas Temperature for Various Appliances
Table 5 Overall Heat Transfer Coefficients of Various Chimneys and Vents
Table 6 Approximate Theoretical Draft of Chimneys
Table 7 Altitude Correction
Table 8 Pressure Equations for Dp
Table 9 Resistance Loss Coefficients
Table 10 Underwriters Laboratories Test Standards
Table 11 List of U.S. National Standards Relating to Installationa
Figures
Fig. 1 Graphical Evaluation of Rate of Vent Gas Flow from Percent CO2 and Fuel Rate
Fig. 2 Flue Gas Mass and Volumetric Flow
Fig. 3 Temperature Multiplier Cu for Compensation of Heat Losses in Connector
Fig. 4 Theoretical Draft Nomograph
Fig. 5 Friction Factor for Commercial Iron and Steel Pipe
Fig. 6 Typical Connector Design
Fig. 7 Gas Vent with Lateral
Fig. 8 Draft-Regulated Appliance with 25 Pa Available Draft Required
Fig. 9 Forced-Draft Appliance with Neutral (Zero) Draft (Negative Pressure Lateral)
Fig. 10 Forced-Draft Appliance with Positive Outlet Pressure (Negative Draft)
Fig. 11 Illustration for Example 2
Fig. 12 Illustration for Example 3
Fig. 13 Illustration for Example 4
Fig. 14 Illustration for Example 6
Fig. 15 Typical Fan Operating Data and System Curves
Fig. 16 Eddy Formation
Fig. 17 Effect of Chimney Gas (Combustion Products) Temperature on Fireplace Frontal Opening Velocity
Fig. 18 Permissible Fireplace Frontal Opening Area for Design Conditions (0.24 m/s mean frontal velocity with 0.3 m inside diameter round flue)
Fig. 19 Effect of Area Ratio on Frontal Velocity (for chimney height of 4.6 m with 0.3 m inside diameter round flue)
Fig. 20 Variation of Chimney Flue Gas Temperature with Heat Input Rate of Combustion Products
Fig. 21 Chimney Sizing Chart for Fireplaces
Fig. 22 Estimation of Fireplace Frontal Opening Area
Fig. 23 Building Heating Appliance, Medium-Heat Chimney
Fig. 24 Use of Barometric Draft Regulators
Fig. 25 Draft Inducers
Fig. 26 Wind Eddy and Wake Zones for One- or Two-Story Buildings and Their Effect on Chimney Gas Discharge
Fig. 27 Height of Eddy Currents Around Single High-Rise Buildings
Fig. 28 Eddy and Wake Zones for Low, Wide Buildings
Fig. 29 Vent and Chimney Rain Protection
--- CHAPTER 36: HYDRONIC HEAT-DISTRIBUTING UNITS AND RADIATORS ---
1. Description
Radiators
Pipe Coils
Convectors
Baseboard Units
Finned-Tube Units
Heat Emission
2. Ratings of Heat-Distributing Units
Radiators
Convectors
Baseboard Units
Finned-Tube Units
Other Heat-Distributing Units
Corrections for Nonstandard Conditions
3. Design
Effect of Water Velocity
Effect of Altitude
Effect of Mass
Performance at Low Water Temperatures
Effect of Enclosure and Paint
4. Applications
Radiators
Convectors
Baseboard Radiation
Finned-Tube Radiation
Radiant Panels
References
Bibliography
Tables
Table 1 Small-Tube Cast-Iron Radiators
Table 2 Correction Factors c for Various Types of Heating Units
Figures
Fig. 1 Terminal Units
Fig. 2 Typical Radiators
Fig. 3 Water Velocity Correction Factor for Baseboard and Finned-Tube Radiators
Fig. 4 Effect of Air Density on Radiator Output
--- CHAPTER 37: SOLAR ENERGY EQUIPMENT AND SYSTEMS ---
1. SOLAR HEATING SYSTEMS
1.1 Air-Heating Systems
1.2 Liquid-Heating Systems
Direct and Indirect Systems
Freeze Protection
1.3 Solar Thermal Energy Collectors
Collector Types
Collector Construction
1.4 Row Design
Piping Configuration
Velocity Limitations
Thermal Expansion
1.5 Array Design
Piping Configuration
Shading
1.6 Thermal Collector Performance
Testing Methods
Collector Test Results and Initial Screening Methods
Generic Test Results
1.7 Thermal Energy Storage
Air System Thermal Storage
Liquid System Thermal Storage
Storage Tank Construction
Storage Tank Insulation
Stratification and Short Circuiting
Storage Sizing
1.8 Heat Exchangers
Requirements
Internal Heat Exchanger
External Heat Exchanger
Heat Exchanger Performance
1.9 Controls
Differential Temperature Controllers
Photovoltaically Powered Pumps
Overtemperature Protection
Hot-Water Dump
Heat Exchanger Freeze Protection
2. PHOTOVOLTAIC SYSTEMS
Fundamentals of Photovoltaics
Related Equipment
References
Bibliography
Tables
Table 1 Worldwide Solar Capacity in Operation by Type
Table 2 Average Performance* Parameters for Generic Types of Liquid Flat-Plate Collectors
Table 3 Thermal Performance Ratings* for Generic Types of Liquid Flat-Plate Collectors
Table 4 Insulation Factor fQ/A for Cylindrical Water Tanks
Figures
Fig. 1 Air-Heating Space and Domestic Water Heater System
Fig. 2 Simplified Schematic of Indirect Nonfreezing System
Fig. 3 Simplified Schematic of Indirect Drainback Freeze Protection System
Fig. 4 Solar Flat-Plate Collectors
Fig. 5 Evacuated-Tube Collector
Fig. 6 Plan View of Liquid Collector Absorber Plates
Fig. 7 Cross Sections of Various Solar Air and Water Heater
Fig. 8 Cross Section of Suggested Insulation to Reduce Heat Loss from Back Surface of Absorber
Fig. 9 Collector Manifolding Arrangements for Parallel-Flow Row
Fig. 10 Pressure Drop and Thermal Performance of Collectors with Internal Manifolds Numbers
Fig. 11 Flow Pattern in Long Collector Row with Restrictions
Fig. 12 Reverse-Return Array Piping
Fig. 13 In-Series Array Piping
Fig. 14 Mounting for Drainback Collector Modules
Fig. 15 Direct-Return Array Piping
Fig. 16 Solar Collector Type Efficiencies
Fig. 17 Pressurized Storage with Internal Heat Exchanger
Fig. 18 Multiple Storage Tank Arrangement with Internal Heat Exchangers
Fig. 19 Pressurized Storage System with External Heat Exchanger
Fig. 20 Unpressurized Storage System with External Heat Exchanger
Fig. 21 Typical Tank Support Detail
Fig. 22 Tank Plumbing Arrangements to Minimize Short Circuiting and Mixing
Fig. 23 Schematic of CSHPSS
Fig. 24 Cross Section of Wraparound Shell Heat Exchangers
Fig. 25 Double-Wall Tubing
Fig. 26 Tube Bundle Heat Exchanger with Intermediate Loop
Fig. 27 Double-Wall Protection Using Two Heat Exchangers in Series
Fig. 28 Basic Nonfreezing Collector Loop for Building Service Hot-Water Heating (Nonglycol Heat Transfer Fluid)
Fig. 29 Heat Rejection from Nonfreezing System Using Liquid-to-Air Heat Exchanger
Fig. 30 Nonfreezing System with Heat Exchanger Bypass
Fig. 31 Representative Current-Voltage and Power-Voltage Curves for Typical Photovoltaic Module
Fig. 32 Module IV Curve at STC
Fig. 33 Effect of Insolation on Module Performance
Fig. 34 Effect of Cell Temperature on Module Performance
Fig. 35 PV System Configurations Using (A) String Inverter, (B) Microinverters, and (C) String Inverter with Power Optimizers Source: Natural Resources Canada 2019
--- CHAPTER 38: COMPRESSORS ---
1. POSITIVE-DISPLACEMENT COMPRESSORS
1.1 Performance
Ideal Compressor
Actual Compressor
Compressor Efficiency, Subcooling, and Superheating
1.2 Abnormal Operating Conditions, Hazards, and Protective Devices
Liquid Hazard
Suction and Discharge Pulsations
Noise
Vibration
Shock
Testing and Operating Requirements
1.3 Motors
2. RECIPROCATING COMPRESSORS
Performance Data
Motor Performance
Features
Special Devices
Application
3. ROTARY COMPRESSORS
3.1 Rolling-Piston Compressors
Performance
Features
3.2 Rotary-Vane Compressors
3.3 Screw Compressors
Single-Screw Compressors
Twin-Screw Compressors
3.4 Scroll Compressors
Mechanical Features
Capacity Control
Energy Efficiency
Noise and Vibration
Operation and Maintenance
3.5 Trochoidal Compressors
Description and Performance
4. CENTRIFUGAL COMPRESSORS
Refrigeration Cycle
Angular Momentum
Nondimensional Coefficients
Mach Number
Performance
Surging
System Balance and Capacity Control
4.1 Application
Vibration
Noise
Drivers
Paralleling
Other Specialized Applications
4.2 Mechanical Design
Impellers
Casings
Rotor Dynamics
Lubrication
Bearings
Oil-Free Centrifugal Compressors
Accessories and Controls
4.3 Isentropic Analysis
4.4 Polytropic Analysis
Testing
4.5 Operation and Maintenance
4.6 Symbols
References
Bibliography
Tables
Table 1 Typical Design Features of Reciprocating Compressors
Table 2 Motor-Starting Torques
Table 3 Typical Rolling-Piston Compressor Performance
Figures
Fig. 1 Comparison of Single-Stage Centrifugal,
Reciprocating, and Screw Compressor Performance
Fig. 2 Types of Positive-Displacement Compressors (Classified by Compression Mechanism Design)
Fig. 3 Ideal Compressor Cycle
Fig. 4 Pressure-Enthalpy Diagram for
Ideal Refrigeration Cycle
Fig. 5 Example of Compressor Operating Envelope
Fig. 6 Basic Reciprocating Piston with Reed Valves
Fig. 7 Pumping Cycle of Reciprocating Compressor
Fig. 8 Six-Cylinder Semihermetic Compressor
Fig. 9 Hermetic Reciprocating Compressor
Fig. 10 Capacity and Power-Input Curves for Typical
Semihermetic Reciprocating Compressor
Fig. 11 Modified Oil-Equalizing System
Fig. 12 Tandem Compressors
Fig. 13 Fixed-Vane, Rolling-Piston Rotary Compressor
Fig. 14 Performance Curves for Typical Rolling-Piston Compressor
Fig. 15 Sound Level of Combination Refrigerator-Freezer with Typical Rotary Compressor
Fig. 16 Rotary-Vane Compressor
Fig. 17 Section of Single-Screw Refrigeration Compressor
Fig. 18 Sequence of Compression Process in Single-Screw Compressor
Fig. 19 Radial and Axially Balanced Main Rotor
Fig. 20 Oil and Refrigerant Schematic of Oil Injection System
Fig. 21 Schematic of Oil-Injection-Free Circuit
Fig. 22 Theoretical Economizer Cycle
Fig. 23 Capacity-Control Slide Valve Operation
Fig. 24 Refrigeration Compressor Equipped with Variable-Capacity Slide Valve and Variable-Volume-Ratio Slide Valve
Fig. 25 Capacity Slide in Full-Load Position and Volume Ratio Slide in Intermediate Position
Fig. 26 Capacity Slide in Part-Load Position and Volume Ratio Slide Positioned to Maintain System Volume Ratio
Fig. 27 Part-Load Effect of Symmetrical and Asymmetrical Capacity Control
Fig. 28 Typical Open-Compressor Performance on R-22
Fig. 29 Typical Compressor Performance on R-717 (Ammonia)
Fig. 30 Typical Semihermetic Single-Screw Compressor
Fig. 31 Single-Gate-Rotor Semihermetic Single-Screw Compressor
Fig. 32 Twin-Screw Compressor
Fig. 33 Twin-Screw Compression Process
Fig. 34 Slide Valve Unloading Mechanism
Fig. 35 Lift Valve Unloading Mechanism
Fig. 36 View of Fixed- and Variable-Volume-Ratio (Vi) Slide Valves from Above
Fig. 37 Twin-Screw Compressor Efficiency
Fig. 38 Semihermetic Twin-Screw Compressor with Suction-
Gas-Cooled Motor and Slide Valve Unloading Mechanism
Fig. 39 Semihermetic Twin-Screw Compressor with Motor Housing Used as Economizer; Built-In Oil Separator, and Slide Valve Unloading Mechanism
Fig. 40 Vertical, Discharge-Cooled, Hermetic Twin-Screw Compressor
Fig. 41 Semihermetic Twin-Screw Compressor with Suction-Gas-Cooled Motor, Slide Valve Unloading Mechanism, Economizer Port in Slide Valve, and Built-In Oil Separator
Fig. 42 Semihermetic Twin-Screw Compressor with Suction-Gas-Cooled Motor, Refrigerant-Cooled Frequency Inverter, Means for Vi Control, and Built-In Oil Separator
Fig. 43 Interfitted Scroll Members
Fig. 44 Components of Scroll Compressor
Fig. 45 Scroll Compression Process
Fig. 46 Two-Step Modulation
Fig. 47 Multistep Modulation
Fig. 48 Scroll with Vapor Injection
Fig. 49 Volumetric and Isentropic Efficiency for Scroll Compressors
Fig. 50 Variable Volume Ratio (VVR)
Fig. 51 Scroll Capacity Versus Residence Demand
Fig. 52 Typical Residential Scroll Sound Spectrum
Fig. 53 Possible Versions of Epitrochoidal and
Hypotrochoidal Machines
Fig. 54 Wankel Sealing System for Trochoidal Compressors
Fig. 55 Centrifugal Refrigeration Unit Cross Section
Fig. 56 Simple Vapor Compression Cycle
Fig. 57 Compression Cycle with Flash Cooling
Fig. 58 Compression Cycle with Brazed-Plate Economizer
Fig. 59 Compression Cycle with Power Recovery Expander
Fig. 60 Impeller Exit Velocity Diagram
Fig. 61 Impeller Inlet Velocity Diagram
Fig. 62 Optimal Specific Speed
Fig. 63 Typical Constant-Speed Centrifugal Compressor Performance
Fig. 64 Typical Variable-Speed Centrifugal Compressor Performance
Fig. 65 Typical Compressor Performance Curves
Fig. 66 Typical Compressor Performance with Various Prerotation Vane Settings
Fig. 67 Typical Part-Load Gas Compression Power Input for Speed and Vane Capacity Controls
Fig. 68 Magnetic Bearing Schematic on Oil-Free Centrifugal Compressor
Fig. 69 Ratio of Polytropic to Adiabatic Work
--- CHAPTER 39: CONDENSERS ---
1. WATER-COOLED CONDENSERS
1.1 Heat Removal
1.2 Heat Transfer
Overall Heat Transfer Coefficient
Water-Side Film Coefficient
Refrigerant-Side Film Coefficient
Tube-Wall Resistance
Surface Efficiency
Fouling Factor
1.3 Water Pressure Drop
1.4 Liquid Subcooling
1.5 Water Circuiting
1.6 Types
Shell-and-Tube Condensers
Shell-and-Coil Condensers
Tube-in-Tube Condensers
Brazed-Plate and Plate-and-Frame Condensers
1.7 Noncondensable Gases
1.8 Testing and Rating
Design Pressure
1.9 Operation and Maintenance
2. AIR-COOLED CONDENSERS
2.1 Types
2.2 Fans and Air Requirements
2.3 Heat Transfer and Pressure Drop
2.4 Condensers Remote from Compressor
2.5 Condensers as Part of Condensing Unit
2.6 Water-Cooled Versus Air-Cooled Condensing
2.7 Testing and Rating
2.8 Control
2.9 Installation and Maintenance
Plate-and-Fin
Integral-Fin
Microchannel
3. EVAPORATIVE CONDENSERS
3.1 Heat Transfer
3.2 Condenser Configuration
Coils
Method of Coil Wetting
Airflow
3.3 Condenser Location
3.4 Multiple-Condenser Installations
3.5 Ratings
3.6 Desuperheating Coils
3.7 Refrigerant Liquid Subcoolers
3.8 Multicircuit Condensers and Coolers
3.9 Water Treatment
3.10 Water Consumption
3.11 Capacity Modulation
3.12 Purging
3.13 Maintenance
3.14 Testing and Rating
References
Bibliography
Tables
Table 1 Net Refrigeration Effect Factors for Reciprocating Compressors Used with Air-Cooled and Evaporative Condensers
Table 2 Typical Maintenance Checklist
Figures
Fig. 1 Heat Removed in Condenser
Fig. 2 Effect of Fouling on Condenser
Fig. 3 Effect of Condenser Water Circuiting
Fig. 4 Loss of Refrigerant During Purging at Various Gas Temperatures and Pressures
Fig. 5 Effect of Fouling on Chiller Performance
Fig. 6 Temperature and Enthalpy Changes in Air-Cooled Condenser with R-134a
7 Equal-Sized Condenser Sections Connected in Parallel and for Half-Condenser Operation During Winter
Fig. 8 Unit Condensers Installed in Parallel with Combined Fan Cycling and Damper Control
Fig. 9 Air-Cooled Unit Condenser for Winter Heating and Summer Ventilation
Fig. 10 Functional Views of Evaporative Condenser
Fig. 11 Heat Transfer Diagram for Evaporative Condenser
Fig. 12 Combined Coil/Fill Evaporative Condenser
Fig. 13 Evaporative Condenser Arranged for
Year-Round Operation
Fig. 14 Parallel Operation of Evaporative and Shell-and-Tube Condenser
Fig. 15 Parallel Operation of Two Evaporative Condensers
Fig. 16 Evaporative Condenser with Desuperheater Coil
Fig. 17 Evaporative Condenser with Liquid Subcooling Coil
--- CHAPTER 40: COOLING TOWERS ---
1. Principle of Operation
2. Design Conditions
3. Types of Cooling Towers
Direct-Contact Cooling Towers
Indirect-Contact Cooling Towers
Hybrid Closed-Circuit Cooling Towers
Modular Fluid Coolers with Mixed Operational Mode
Adiabatic Fluid Coolers
4. Materials of Construction
5. Selection Considerations
6. Application
Siting
Piping
Capacity Control
Water-Side Economizer (Free Cooling)
Winter Operation
Sound
Drift
Fogging (Cooling Tower Plume)
Maintenance
Inspections
Water Treatment
White Rust
7. Performance Curves
8. Cooling Tower Thermal Performance
9. Cooling Tower Theory
Counterflow Integration
Cross-Flow Integration
10. Tower Coefficients
Available Coefficients
Establishing Tower Characteristics
11. Additional Information
References
Bibliography
Tables
Table 1 Counterflow Integration Calculations for Example 1
Figures
Fig. 1 Temperature Relationship Between Water and Air in Counterflow Cooling Tower
Fig. 2 Psychrometric Analysis of Air Passing Through Cooling Tower
Fig. 3 Direct-Contact or Open Evaporative Cooling Tower
Fig. 4 Indirect-Contact or Closed-Circuit Evaporative Cooling Towers
Fig. 5 Types of (A) Fill and (B) Coils
Fig. 6 Combined Flow Coil/Fill Evaporative Cooling Tower
Fig. 7 Coil Shed Cooling Tower
Fig. 8 Vertical Spray Cooling Tower
Fig. 9 Horizontal Spray Cooling Tower
Fig. 10 Hyperbolic Tower
Fig. 11 Conventional Mechanical-Draft Cooling Towers
Fig. 12 Factory-Assembled Counterflow Forced-Draft Cooling Tower
Fig. 13 Field-Erected Cross-Flow Mechanical-Draft Cooling Tower
Fig. 14 Combination Wet/Dry Cooling Tower
Fig. 15 Adiabatically Saturated Air-Cooled Heat Exchanger
Fig. 16 Hybrid Cooling Towers in Wet Operational Mode
Fig. 17 Hybrid Cooling Towers in Dry/Wet Operational Mode
Fig. 18 Hybrid Cooling Tower in Adiabatic Operational Mode
Fig. 19 Hybrid Cooling Towers in Dry Operational Mode
Fig. 20 Modular Fluid Coolers Capable of (A) Mixed Operational Mode, (B) Wet Mode, and (C) Dry Mode
Fig. 21 Adiabatic Fluid Cooler in Adiabatic Operational Mode
Fig. 22 Example of Adiabatic Pad Precooling System
Fig. 23 Example of Corrugated Adiabatic Pads
Fig. 24 Example of Adiabatic Spray Precooling System
Fig. 25 Psychrometric Chart Showing Depressed Dry Bulb: Entering Air at 35°C db, 25.6°C wb, and 75% Pad Saturation Efficiency
Fig. 26 Discharge Air Reentering Cooling Tower
Fig. 27 Cooling Tower Fan Power Versus Speed
(White 1994)
Fig. 28 Free Cooling by Auxiliary Heat Exchanger
Fig. 29 Free Cooling by Refrigerant Vapor Migration
Fig. 30 Free Cooling by Interconnection of Water Circuits
Fig. 31 Fog Prediction Using Psychrometric Chart
Fig. 32 Cooling Tower Performance: 100% Design Flow
Fig. 33 Cooling Tower Performance: 67% Design Flow
Fig. 34 Cooling Tower Performance: 133% Design Flow
Fig. 35 Cooling Tower Performance: 167% Design Flow
Fig. 36 Heat and Mass Transfer Relationships Between Water, Interfacial Film, and Air
(Baker and Shryock 1961)
Fig. 37 Counterflow Cooling Diagram
Fig. 38 Water Temperature and Air Enthalpy Variation Through Cross-Flow Cooling Tower (Baker and Shryock 1961)
Fig. 39 Cross-Flow Calculations (Baker and Shryock 1961)
Fig. 40 Cross-Flow Cooling Diagram
Fig. 41 Counterflow Cooling Diagram for Constant Conditions, Variable L/G Ratios
Fig. 42 Tower Characteristic, KaV/L Versus L/G (Baker and Shryock 1961)
Fig. 43 True Versus Apparent Potential Difference (Baker and Shryock 1961)
--- CHAPTER 41: EVAPORATIVE AIR-COOLING EQUIPMENT ---
Tables
Table 1 Heat Recovery Economizer Effectiveness at 70% and 21.1°C Building Return Air Condition
Figures
Fig. 1 Typical Random-Media Evaporative Cooler
Fig. 2 Typical Rigid-Media Air Cooler
Fig. 3 Indirect Evaporative Cooling (IEC) Heat Exchanger
Fig. 4 Indirect Evaporative Cooler Used as Precooler
Fig. 5 Heat Pipe Indirect Evaporative Cooling (IEC) Heat Exchanger Packaged with DX System
Fig. 6 Combination Indirect/Direct Evaporative Cooling Process
Fig. 7 Indirect/Direct Evaporative Cooler with Heat Exchanger (Rotary Heat Wheel or Heat Pipe)
Fig. 8 Three-Stage Indirect/Direct Evaporative Cooler
Fig. 9 Interaction of Air and Water in Air Washer Heat Exchanger
Fig. 10 Schematic Showing Airflow Through VAV Air-Handling Unit with Heat Recovery Economizer (HRE) and Adiabatic Direct Evaporative Cooler/Humidifier (DEC/H) for Winter Hydration of Dry Outdoor Air
Fig. 11 Performance of Heat Recovery Economizer in Cold Climate
--- CHAPTER 42: LIQUID COOLERS ---
Tables
Table 1 Types of Coolers
Figures
Fig. 1 Direct-Expansion Shell-and-Tube Cooler
Fig. 2 Flooded Shell-and-Tube Cooler
Fig. 3 Flooded Plate Cooler
Fig. 4 Baudelot Cooler
Fig. 5 Shell-and-Coil Cooler
Fig. 6 Nucleate Boiling Contribution to Total Refrigerant Heat Transfer
--- CHAPTER 43: LIQUID-CHILLING SYSTEMS ---
Tables
Table 1 Properties of Various Refrigerants
Figures
Fig. 1 Equipment Diagram for Basic Liquid Chiller
Fig. 2 Parallel-Operation High Design Water Leaving Coolers (Approximately 7°C and Above)
Fig. 3 Parallel-Operation Low Design Water Leaving Coolers (Below Approximately 7°C)
Fig. 4 Series Operation
Fig. 5 Approximate Liquid Chiller Availability Range by Compressor Type
Fig. 6 Comparison of Single-Stage Centrifugal, Reciprocating, and Screw Compressor Performance
Fig. 7 Reciprocating Liquid Chiller Performance with Three Equal Steps of Unloading
Fig. 8 Reciprocating Liquid Chiller Control System
Fig. 9 Typical Centrifugal Compressor Performance at Constant Speed
Fig. 10 Typical Variable-Speed Centrifugal Compressor Performance
Fig. 11 Temperature Relations in a Typical Centrifugal Liquid Chiller
Fig. 12 Refrigeration System Schematic
Fig. 13 Typical Screw Compressor Chiller Part-Load Power Consumption
Fig. 14 Typical External Connections for Screw Compressor Chiller
--- CHAPTER 44: CENTRIFUGAL PUMPS ---
1. Centrifugal Pumping
2. Construction Features
3. Pump Types
Circulator Pump
Close-Coupled, Single-Stage, End-Suction Pump
Frame-Mounted, End-Suction Pump on Base Plate
Base-Mounted, Horizontal (Axial) or Vertical, Split-Case, Single-Stage, Double-Suction Pump
Base-Mounted, Horizontal, Split-Case, Multistage Pump
Vertical In-Line Pump
Vertical In-Line Split-Coupled Pump
Vertical Turbine, Single- or Multistage, Sump-Mounted Pump
4. Pump Performance Curves
5. Hydronic System Curves
6. Pump and Hydronic System Curves
7. Pump Power
8. Pump Efficiency
9. Affinity Laws
10. Radial Thrust
11. Net Positive Suction Characteristics
12. Selection of Pumps
13. Arrangement of Pumps
Duty Standby
Parallel Pumping
Series Pumping
Standby Pump
Primary-Secondary Pumping
Variable-Speed Central Pumping
Variable-Speed Distributed Pumping
Differential Pressure Control with Predefined Control Curves
14. Motive Power
15. Energy Conservation in Pumping
16. Installation, Operation, and Commissioning
Commissioning Base-Mounted Centrifugal Pumps
17. Troubleshooting
References
Bibliography
Tables
Table 1 Pump Affinity Laws
Table 2 Flow Redundancy
Table 3 Pumping System Noise Analysis Guide
Table 4 Pumping System Inadequate/No-Flow Analysis Guide
Figures
Fig. 1 Centrifugal Pump
Fig. 2 Cross Section of Typical Overhung-Impeller End-Suction Pump
Fig. 3 Wet-Rotor Circulator Pump
Fig. 4 Circulator Driven by Close-Coupled Motor
Fig. 5 Close-Coupled Single-Stage End-Suction Pump
Fig. 6 Frame-Mounted End-Suction Pump on Base Plate
Fig. 7 Base-Mounted, Horizontal (Axial), Split-Case, Single-Stage, Double-Suction Pump
Fig. 8 Base-Mounted, Vertical, Split-Case, Single-Stage, Double-Suction Pump
Fig. 9 Base-Mounted, Horizontal, Split-Case, Multistage Pump
Fig. 10 Vertical In-Line Split-Coupled Pump (Double Suction Impeller)
Fig. 11 Vertical In-Line Pump (Single Suction Impeller)
Fig. 12 Vertical Turbine Pump
Fig. 13 Typical Pump Performance Curve
Fig. 14 Typical Pump Curve
Fig. 15 Flat Versus Steep Performance Curves
Fig. 16 Typical Pump Manufacturer’s Performance Curve Series
Fig. 17 Typical System Curve
Fig. 18 Typical System Curve with Independent Pressure
Fig. 19 System and Pump Curves
Fig. 20 System Curve Change due to Part-Load Flow
Fig. 21 Pump Operating Points
Fig. 22 System Curve, Constant and Variable Pressure Loss
Fig. 23 Typical Pump Water Power Increase with Flow
Fig. 24 Pump Efficiency Versus Flow
Fig. 25 Pump Efficiency Curves
Fig. 26 Pump Best Efficiency Curves
Fig. 27 Pumping Power, Pressure, and Flow Versus Pump Speed
Fig. 28 Example Application of Affinity Law
Fig. 29 Variable-Speed Pump Operating Points
Fig. 30 Radial Thrust
Fig. 31 Radial Thrust Versus Pumping Rate
Fig. 32 Cavitation in Impeller
Fig. 33 Net Positive Suction Pressure Available
Fig. 34 Pump Performance and NPSR Curves
Fig. 35 Pump Selection Regions
Fig. 36 Pump Curve Construction for Parallel Operation
Fig. 37 Operating Conditions for Parallel Operation
Fig. 38 Construction of Curve for Dissimilar Parallel Pumps
Fig. 39 Typical Piping for Parallel Pumps
Fig. 40 Pump Curve Construction for Series Operation
Fig. 41 Operating Conditions for Series Operation
Fig. 42 Typical Piping for Series Pumps
Fig. 43 Primary-Secondary Pumping
Fig. 44 Variable-Speed Central Pumping
Fig. 45 Variable-Speed Distributed Pumping
Fig. 46 Constant-Pressure Control
Fig. 47 Variable- (Proportional-) Pressure Control
Fig. 48 Variable- (Quadratic-) Pressure Control
Fig. 49 Efficiency Comparison of Four-Pole Motors
Fig. 50 Typical Efficiency Range of Variable-Speed Drives
Fig. 51 Base-Plate-Mounted Centrifugal Pump Installation
Fig. 52 In-Line Pump Installation
--- CHAPTER 45: MOTORS, MOTOR CONTROLS, AND VARIABLE-FREQUENCY DRIVES ---
1. MOTORS
1.1 Alternating-Current Power Supply
1.2 Codes and Standards
1.3 Motor Efficiency
1.4 General-Purpose Motors
Application
1.5 Permanent-Magnet AC Motors
1.6 Hermetic Motors
Application
1.7 Integral Thermal Protection
1.8 Motor Protection and Control
Separate Motor Protection
Protection of Control Apparatus and Branch Circuit Conductors
Three-Phase Motor Starting
Direct-Current Motor Starting
Single-Phase Motor Starting
Operating AC Induction Motors above Nameplate Speed Using Variable-Frequency Drives
VFD-Induced Bearing Currents
Detecting Bearing Currents
Strategies for Mitigating Bearing Currents
2. AIR VOLUME CONTROL
2.1 Variable-Frequency Drives
Power Transistor Characteristics
Motor and Conductor Impedance
Motor Ratings and NEMA Standards
Motor Noise and Drive Carrier Frequencies
Carrier Frequencies and Drive Ratings
2.2 Power Distribution System Effects
VFDs and Harmonics
2.3 performance testing and rating standards
Calculating VFD and Motor Efficiency
VFD-Generated Harmonics
Motor Insulation Stress
References
Bibliography
Tables
Table 1 Motor and Motor Control Equipment Voltages (Alternating Current)
Table 2 Effect of Voltage and Frequency Variation on Induction Motor Characteristics*
Table 3 Motor Types
Table 4 Characteristics of AC Motors (Nonhermetic)
Table 5 Comparison of VAV Energy Consumption with Various Volume Control Techniques
Table 6 Speed/Torque Points for Drive System Efficiency and Motor Stress Insulation Tests*
Figures
Fig. 1 Typical Performance Characteristics of Capacitor-Start/Induction-Run Two-Pole General-Purpose Motor, 0.75 kW
Fig. 2 Typical Performance Characteristics of Resistance-Start Split-Phase Two-Pole Hermetic Motor, 0.2 kW
Fig. 3 Typical Performance Characteristics of Permanent Split-Capacitor Two-Pole Motor, 0.75 kW
Fig. 4 Typical Performance Characteristics of Three-Phase Two-Pole Motor, 4 kW
Fig. 5 Motor Voltage Versus Frequency
Fig. 6 Motor Torque Limits Versus Frequency
Fig. 7 Motor Power Capacity at Maximum Extended Frequency
Fig. 8 Fluted Bearing from VFD-Induced Electrical Bearing Current Discharges
Fig. 9 2.5 V Peak Shaft Voltage, No Discharge
Fig. 10 14.6 V Peak Shaft Voltage, No Discharge
Fig. 11 11.4 V Peak Shaft Voltage, Discharge Detected
Fig. 12 32.2 V Peak Shaft Voltage, Discharge Detected
Fig. 13 Common Mode Toroid Built into Output of VFD
Fig. 14 Typical Common-Mode Toroids for Retrofitting into Output of VFD
Fig. 15 Shaft Grounding Ring Installed on Motor
Fig. 16 Fan Duty Cycle for a VAV System
Fig. 17 Outlet Damper Control
Fig. 18 Variable Inlet Vane Control
Fig. 19 Eddy Current Coupling Control
Fig. 20 AC Drive Control
Fig. 21 Bipolar Versus IGBT PWM Switching
Fig. 22 Motor and Drive Relative Impedance
Fig. 23 Typical Switching Times, Cable Distance, and Pulse Peak Voltage
Fig. 24 Typical Reflected Wave Voltage Levels at Drive and Motor Insulation
Fig. 25 Motor Voltage Peak and Limits
Fig. 26 Damaging Reflected Waves above Motor CIV Levels
Fig. 27 Motor Audible Noise
Fig. 28 Voltage Waveform Distortion by Pulse-Width-Modulated VSD
Fig. 29 Basic Elements of Solid-State Drive
Fig. 30 Remote Complete Drive Module (CDM) Test Setup
Fig. 31 Attached Complete Drive Module (CDM)
--- CHAPTER 46: VALVES ---
1. Fundamentals
Body Ratings
Materials
Flow Factor and Pressure Drop
Cavitation
Water Hammer
Noise
Body Styles
2. Manual Valves
Selection
Globe Valves
Gate Valves
Plug Valves
Ball Valves
Butterfly Valves
3. Automatic Valves
Actuators
Pneumatic Actuators
Electric Actuators
Electronic Hydraulic Actuators
Solenoids
Thermostatic Radiator Valves
Control of Automatic Valves
Two-Way Valves
Three-Way Valves
Special-Purpose Valves
Ball Valves
Butterfly Valves
Pressure-Independent Control Valves
Flow-Limiting Valves
Control Valve Flow Characteristics
Control Valve Sizing
4. Balancing Valves
Manual Balancing Valves
Automatic Flow-Limiting Valves
Balancing Valve Selection
5. Multiple-Purpose Valves
6. Safety Devices
7. Self-Contained Temperature Control Valves
8. Pressure-Reducing Valves
Makeup Water Valves
9. Check Valves
10. Stop-Check Valves
11. Backflow Prevention Devices
Selection
Installation
12. Steam Traps
References
Bibliography
Figures
Fig. 1 Flow Coefficient Test Arrangement
Fig. 2 Valve Cavitation at Sharp Curves
Fig. 3 Globe Valve
Fig. 4 Two Variations of Gate Valve
Fig. 5 Plug Valve
Fig. 6 Ball Valve
Fig. 7 Butterfly Valve
Fig. 8 Two-Way, Direct-Acting Control Valve with Pneumatic Actuator and Positioner
Fig. 9 Two-Way Control Valve with Electric Actuator
Fig. 10 Electric Solenoid Valve
Fig. 11 Thermostatic Valves
Fig. 12 Typical Three-Way Control Applications
Fig. 14 Butterfly Valves, Diverting Tee Application
Fig. 15 Pressure-Independent Control Valve
Fig. 16 Control Valve Flow Characteristics
Fig. 17 Heat Output, Flow, and Stem Travel Characteristics of Equal-Percentage Valve
Fig. 18 Authority Distortion of Linear Flow Characteristics
Fig. 19 Authority Distortion of Equal-Percentage Flow Characteristic
Fig. 20 Manual Balancing Valve
Fig. 21 Automatic Flow-Limiting Valve
Fig. 22 Automatic Flow-Limiting Valve Curve
Fig. 23 Typical Multiple-Purpose Valve (Straight Pattern) on Discharge of Pump
Fig. 24 Typical Multiple-Purpose Valve (Angle Pattern) on Discharge of Pump
Fig. 25 Safety/Relief Valve with Drip-Pan Elbow
Fig. 26 Self-Operated Temperature Control Valve
Fig. 27 Pilot-Operated Steam Valve
Fig. 28 Swing Check Valves
Fig. 29 Backflow Prevention Device
--- CHAPTER 47: HEAT EXCHANGERS ---
1. Fundamentals
2. Types of Heat Exchangers
Shell-and-Tube Heat Exchangers
Tube-in-Tube Heat Exchanger
Plate Heat Exchangers
Double-Wall Heat Exchangers
3. Components
Shell-and-Tube Components
Plate Components
4. Application
5. Selection Criteria
Thermal/Mechanical Design
Cost
Maintenance
Space Requirements
Steam
Water Quality
6. Installation
Additional Resources
Figures
Fig. 1 Temperature Distribution in Counterflow Heat Exchanger
Fig. 2 Counterflow Path in Shell-and-Tube Heat Exchanger
Fig. 3 U-Tube Shell-and-Tube Heat Exchanger with Removable Bundle Assembly and Cast K-Pattern Flanged Head
Fig. 4 U-Tube Tank Heater with Removable Bundle Assembly and Cast Bonnet Head
Fig. 5 U-Tube Tank Suction Heater with Removable Bundle Assembly and Cast Flanged Head
Fig. 6 Straight-Tube Fixed Tubesheet Shell-and-Tube Heat Exchanger with Fabricated Bonnet Heads and Split-Shell Flow Design
Fig. 7 Straight-Tube Floating Tubesheet Shell-and-Tube Heat Exchanger with Removable Bundle Assembly and Fabricated Channel Heads
Fig. 8 Flow Path of Gasketed-Plate Heat Exchanger
Fig. 9 Flow Path of Welded-Plate Heat Exchanger
Fig. 10 Brazed-Plate Heat Exchanger
Fig. 11 Double-Wall Tube
Fig. 12 Double-Wall Plate Cassette for Gasketed Plate and Frame Exchanger
Fig. 13 Double-Wall Plate Assembly with Brazed-Plate Exchanger
Fig. 14 Exploded View of Straight-Tube Heat Exchanger
Fig. 15 Components of a Gasketed Plate Heat Exchanger
--- CHAPTER 48: UNITARY AIR CONDITIONERS AND HEAT PUMPS ---
1. General Design Considerations
User Requirements
Application Requirements
Installation
Service
Sustainability
2. Types of Unitary Equipment
Single-Package Equipment: Types and Installations
Combined Space-Conditioning/Water-Heating Systems
Engine-Driven Heat Pumps and Air Conditioners
3. Equipment and System Standards
Energy Conservation and Efficiency
AHRI Certification Programs
Safety Standards and Installation Codes
4. Air Conditioners
Refrigerant Circuit Design
Air-Handling Systems
Electrical Design
Mechanical Design
Accessories
Heating
5. Air-Source Heat Pumps
Add-On Heat Pumps
Selection
Refrigerant Circuit and Components
System Control and Installation
6. Water-Source Heat Pumps
Systems
Performance Certification Programs
Equipment Design
7. Variable-Refrigerant-Flow Heat Pumps
Application
Categories
Refrigerant Circuit and Components
Heating and Defrost Operation
References
Bibliography
Tables
Table 1 AHRI Standard 210/240 Classification of Unitary Air Conditioners
Table 2 AHRI Standard 210/240 Classification of Air-Source Unitary Heat Pumps
Table 3 Amended Energy Conservation Standards for Residential Central Air Conditioners and Heat Pumps, Effective January 1, 2023
Table 4 Space Requirements for Typical Packaged Water-Source Heat Pumps
Figures
Fig. 1 Typical Rooftop Air-Cooled Single-Package Air
Conditioner
Fig. 2 Single-Package Air Equipment with Variable Air Volume
Fig. 3 Example of Two-Zone Ductless Multisplit System in Typical Residential Installation
Fig. 4 Water-Cooled Single-Package Air Conditioner
Fig. 5 Rooftop Installation of Air-Cooled Single-Package Unit
Fig. 6 Multistory Rooftop Installation of Single-Package Unit
Fig. 7 Through-the-Wall Installation of Air-Cooled Single-Package Unit
Fig. 8 Residential Installation of Split-System Air-Cooled Condensing Unit with Coil and Upflow Furnace
Fig. 9 Outdoor Installations of Split-System Air-Cooled Condensing Units with Coil and Upflow Furnace or with Indoor Blower-Coils
Fig. 10 Outdoor Installation of Split-System Air-Cooled Condensing Unit with Indoor Coil and Downflow Furnace
Fig. 11 Regional Appliance Efficiency Standards Map, Effective January 1, 2015
Fig. 12 Schematic Typical of Air-to-Air Heat Pump System
Fig. 13 Operating Characteristics of Single-Stage Unmodulated Heat Pump
Fig. 14 Schematic of Typical Water-Source Heat Pump System
Fig. 15 Typical Horizontal Water-Source Heat Pump
Fig. 16 Typical Vertical Water-Source Heat Pump
Fig. 17 Water-Source Heat Pump Systems
--- CHAPTER 49: ROOM AIR CONDITIONERS AND PACKAGED TERMINAL AIR CONDITIONERS ---
Tables
Table 1 Minimum Efficiency Standards for Room Air Conditioners
Figures
Fig. 1 Schematic View of Typical Room Air Conditioner
Fig. 2 Sectional Packaged Terminal Air Conditioner
Fig. 3 Integrated Packaged Terminal Air Conditioner
--- CHAPTER 50: THERMAL STORAGE ---
Terminology
Classification of Systems
Storage Media
Basic Thermal Storage Concepts
Benefits of Thermal Storage
Design Considerations
1. Sensible Thermal Storage Technology
Sensible Energy Storage
Temperature Range and Storage Size
Techniques for Thermal Separation in Sensible Storage Devices
Performance of Chilled-Water Storage Systems
Design of Stratification Diffusers
Storage Tank Insulation
Other Factors
Chilled-Water Storage Tanks
Low-Temperature Fluid Sensible Energy Storage
Storage in Aquifers
Chilled-Water Thermal Storage Sizing Examples
2. Latent Cool Storage Technology
Water as Phase-Change Thermal Storage Medium
Internal Melt Ice-On-Coil
3. Chiller and Ice Storage Selection
Operation With Disabled Chiller
Selecting Storage Equipment
External-Melt Ice-On-Coil
Encapsulated Ice
Ice Harvesters
Ice Slurry Systems
Unitary Thermal Storage Systems
Other Phase-Change Materials
4. Heat Storage Technology
Sizing Heat Storage Systems
Service Water Heating
Brick Storage (ETS) Heaters
Pressurized Water Storage Heaters
Underfloor Heat Storage
Building Mass Thermal Storage
Factors Favoring Thermal Storage
Comparative Value of TEC versus Other Energy Storage Technologies
Factors Discouraging Thermal Storage
Typical Applications
5. Sizing Cool Storage Systems
Sizing Strategies
Calculating Load Profiles
Sizing Equipment
6. Application of Thermal Storage Systems
Chilled-Water Storage Systems
Ice (and PCM) Storage Systems
Unitary Thermal Storage Systems (UTSSs)
7. Operation and Control
Operating Modes
Control Strategies
Operating Strategies
Utility Demand Control
Instrumentation Requirements
8. Other Design Considerations
Hydronic System Design for Open Systems
Cold-Air Distribution
Storage of Heat in Cool Storage Units
System Interface
Insulation
9. Cost Considerations
10. Maintenance Considerations
Water Treatment
11. Commissioning
Statement of Design Intent
Commissioning Specification
Required Information
Performance Verification
Sample Commissioning Plan Outline for Chilled-Water Plants with Thermal Storage Systems
12. Good Practices
References
Bibliography
Tables
Table 1 Chilled-Water Density Table
Table 2 Peak Day Full-Storage TES Sizing Calculations
Table 3 Peak-Day Partial-Storage TES Sizing Calculations
Table 4 Design Day Chiller and Storage Load Contributions
and Leaving Coolant Temperatures (LCT)
Table 5 Common Thermal Storage Operating Modes
Table 6 Recommended Accuracies of Instrumentation for Measurement of Cool Storage Capacity
Figures
Fig. 1 Typical Two-Ring Octagonal Slotted Pipe Diffuser
Fig. 2 Typical Temperature Stratification Profile in Storage Tank
Fig. 3 Typical Chilled-Water Storage Profiles
Fig. 4 Radial Disk Diffuser
Fig. 5 Full-Storage TES Tank Peak-Day Operation: Facility Cooling Load Versus Chiller Output
Fig. 6 TES Tank Full-Storage Capacity (kWh) and Chiller Operation Output (kW)
Fig. 7 Cooling Load (kW) and Chiller Load (kW) with Partial-Storage TES Tank
Fig. 8 TES Tank Partial-Storage Capacity (kWh) and Chiller Operation Output (kW)
Fig. 9 Charge and Discharge of Internal-Melt Ice Storage
Fig. 10 Design-Day Cooling Load
Fig. 11 Minimum-Sized Chilled and Storage Contribution to Cooling Load
Fig. 12 Redundancy with One of Two Chillers Disabled
Fig. 13 Charge and Discharge of External-Melt Ice Storage
Fig. 14 Encapsulated Ice: Spherical Container
Fig. 15 Ice-Harvesting Schematic
Fig. 16 Representative Sizing Factor Selection Graph for Residential Storage Heaters
Fig. 17 Typical Storage Heater Performance Characteristics
Fig. 18 Room Storage Heater
Fig. 19 Room Storage Heater Dynamic Discharge and Charge Curves
Fig. 20 Static Discharge from Room Storage Heater
Fig. 21 Pressurized Water Heater
Fig. 22 Underfloor Heat Storage
Fig. 23 Annual Energy Cost Savings from Precooling, Relative to Conventional Controls, as Function of Re
Fig. 24 Annual Energy Cost Savings from Precooling, Relative to Conventional Controls, as Function of Rd
Fig. 25 Typical Sensible Storage Connection Scheme
Fig. 26 Direct Transfer Pumping Interface
Fig. 27 Charge Mode Status of Direct Transfer Pumping Interface
Fig. 28 Indirect Transfer Pumping Interface
Fig. 29 Charge Mode Status of Indirect Transfer Pumping Interface
Fig. 30 Primary/Secondary Chilled-Water Plant with Stratified Storage Tank as Decoupler
Fig. 31 Series Flow, Chiller Upstream
Fig. 32 Series Flow, Chiller Downstream
Fig. 33 Parallel Flow for Chiller and Storage
Fig. 34 Typical UTSS with Packaged DX Equipment
--- CHAPTER 51: DEDICATED OUTDOOR AIR SYSTEMS ---
1. Applications
1.1 Humidity Control
1.2 Energy Impact
1.3 Systems without ventilation capabilities
1.4 First-Cost Reduction
2. Air Distribution
2.1 Direct supply to Each Zone
2.2 Supply to Intake of Local Units
2.3 Delivery to Supply Side of Local Units
2.4 Supply to Plenum Near Local Units
3. Equipment configurations
Tables
Table 1 Example of Dedicated Outdoor Air Unit Control Modes
Figures
Fig. 1 Typical DOAS Configuration for Large Retail Store
Fig. 2 Annual Cumulative Latent (Dehumidification) and Sensible Cooling Loads from Outdoor Air
Fig. 3 Design Extremes
Fig. 4 Typical DOAS Configuration Diagram during Humid-Season Operation
Fig. 5 Demand-Controlled Ventilation Benefits in DOAS
Fig. 6 Schematic of DOAS Supplying Conditioned Outdoor Air Directly to Each Zone
Fig. 7 Common DOAS Equipment Configuration with Total-Energy Wheel
Fig. 8 Common DOAS Equipment Configuration with Heat Pipe
Fig. 9 Example of Dedicated Outdoor Air Unit Control Modes
--- CHAPTER 52: CODES AND STANDARDS ---
--- Additions and Corrections ---
2017 Fundamentals
2019 HVAC Applications
--- COMPOSITE INDEX: ASHRAE HANDBOOK SERIES ---
Abbreviations, F38
Absorbents
Absorption
Acoustics. See Sound
Activated alumina, S24.1, 4, 12
Activated carbon adsorption, A47.9
Adaptation, environmental, F9.17
ADPI. See Air diffusion performance index (ADPI)
Adsorbents
Adsorption
Aeration, of farm crops, A26
Aerosols, S29.1
AFDD. See Automated fault detection and diagnostics (AFDD)
Affinity laws for centrifugal pumps, S44.8
AFUE. See Annual fuel utilization efficiency (AFUE)
AHU. See Air handlers
Air
Air barriers, F25.9; F26.5
Airborne infectious diseases, F10.7
Air cleaners. (See also Filters, air; Industrial exhaust gas cleaning)
Air conditioners. (See also Central air conditioning)
Air conditioning. (See also Central air conditioning)
Air contaminants, F11. (See also Contaminants)
Aircraft, A13
Air curtains
Air diffusers, S20
Air diffusion, F20
Air diffusion performance index (ADPI), A58.6
Air dispersion systems, fabric, S19.11
Air distribution, A58; F20; S4; S20
Air exchange rate
Air filters. See Filters, air
Airflow
Airflow retarders, F25.9
Air flux, F25.2. (See also Airflow)
Air handlers
Air inlets
Air intakes
Air jets. See Air diffusion
Air leakage. (See also Infiltration)
Air mixers, S4.8
Air outlets
Airports, air conditioning, A3.6
Air quality. (See also Indoor air quality [IAQ])
Air terminal units (ATUs)
Airtightness, F37.24
Air-to-air energy recovery, S26
Air-to-transmission ratio, S5.13
Air transport, R27
Air washers
Algae, control, A50.12
All-air systems
Altitude, effects of
Ammonia
Anchor bolts, seismic restraint, A56.7
Anemometers
Animal environments
Annual fuel utilization efficiency (AFUE), S34.2
Antifreeze
Antisweat heaters (ASH), R15.5
Apartment buildings
Aquifers, thermal storage, S50.7
Archimedes number, F20.6
Archives. See Museums, galleries, archives, and libraries
Arenas
Argon, recovery, R47.17
Asbestos, F10.5
ASH. See Antisweat heaters (ASH)
Atriums
Attics, unconditioned, F27.2
Auditoriums, A5.3
Automated fault detection and diagnostics (AFDD), A40.4; A63.1
Automobiles
Autopsy rooms, A9.12; A10.6, 7
Avogadro’s law, and fuel combustion, F28.11
Backflow-prevention devices, S46.14
BACnet®; F7.18
Bacteria
Bakery products, R41
Balance point, heat pumps, S48.9
Balancing. (See also Testing, adjusting, and balancing)
BAS. See Building automation systems (BAS)
Baseboard units
Basements
Bayesian analysis, F19.37
Beer’s law, F4.16
Behavior
BEMP. See Building energy modeling professional (BEMP)
Bernoulli equation, F21.1
Best efficiency point (BEP), S44.8
Beverages, R39
BIM. See Building information modeling (BIM)
Bioaerosols
Biocides, control, A50.13
Biodiesel, F28.8
Biological safety cabinets, A17.5
Biomanufacturing cleanrooms, A19.11
Bioterrorism. See Chemical, biological, radio- logical, and explosive (CBRE) incidents
Boilers, F19.21; S32
Boiling
Brayton cycle
Bread, R41
Breweries
Brines. See Coolants, secondary
Building automation systems (BAS), A41.8; A63.1; F7.14
Building energy modeling professional (BEMP), F19.5
Building energy monitoring, A42. (See also Energy, monitoring)
Building envelopes
Building information modeling (BIM), A41.8; A60.18
Building materials, properties, F26
Building performance simulation (BPS), A65.8
Buildings
Building thermal mass
Burners
Buses
Bus terminals
Butane, commercial, F28.5
CAD. See Computer-aided design (CAD)
Cafeterias, service water heating, A51.12, 21
Calcium chloride brines, F31.1
Candy
Capillary action, and moisture flow, F25.10
Capillary tubes
Carbon dioxide
Carbon emissions, F34.7
Carbon monoxide
Cargo containers, R25
Carnot refrigeration cycle, F2.6
Cattle, beef and dairy, A25.7. (See also Animal environments)
CAV. See Constant air volume (CAV)
Cavitation, F3.13
CBRE. See Chemical, biological, radiological, and explosive (CBRE) incidents
CEER. See Combined energy efficiency ratio (CEER)
Ceiling effect. See Coanda effect
Ceilings
Central air conditioning, A43. (See also Air conditioning)
Central plant optimization, A8.13
Central plants
Central systems
Cetane number, engine fuels, F28.9
CFD. See Computational fluid dynamics (CFD)
Change-point regression models, F19.28
Charge minimization, R1.36
Charging, refrigeration systems, R8.4
Chemical, biological, radiological, and explosive (CBRE) incidents, A61
Chemical plants
Chemisorption, A47.10
Chilled beams, S20.10
Chilled water (CW)
Chillers
Chilton-Colburn j-factor analogy, F6.7
Chimneys, S35
Chlorinated polyvinyl chloride (CPVC), A35.44
Chocolate, R42.1. (See also Candy)
Choking, F3.13
CHP systems. See Combined heat and power (CHP)
Cinemas, A5.3
CKV. See Commercial kitchen ventilation (CVK)
Claude cycle, R47.8
Cleanrooms. See Clean spaces
Clean spaces, A19
Clear-sky solar radiation, calculation, F14.8
Climate change
Climatic design information, F14
Clinics, A9.17
Clothing
CLTD/CLF. See Cooling load temperature differential method with solar cooling load factors (CLTD/CLF)
CMMS. See Computerized maintenance management system (CMSS)
Coal
Coanda effect, A34.23; F20.2, 7; S20.2
Codes, A66. (See also Standards)
Coefficient of performance (COP)
Coefficient of variance of the root mean square error (CV[RMSE]), F19.33
Cogeneration. See Combined heat and power (CHP)
Coils
Colburn’s analogy, F4.17
Colebrook equation
Collaborative design, A60
Collectors, solar, A36.6, 11, 24, 25; S37.3
Colleges and universities, A8.11
Combined energy efficiency ratio (CEER), S49.3
Combined heat and power (CHP), S7
Combustion, F28
Combustion air systems
Combustion turbine inlet cooling (CTIC), S7.21; S8.1
Comfort. (See also Physiological principles, humans)
Commercial and public buildings, A3
Commercial kitchen ventilation (CKV), A34
Commissioning, A44
Comprehensive room transfer function method (CRTF), F19.11
Compressors, S38
Computational fluid dynamics (CFD), F13.1, F19.25
Computer-aided design (CAD), A19.6
Computerized maintenance management system (CMMS), A60.17
Computers, A41
Concert halls, A5.4
Concrete
Condensate
Condensation
Condensers, S39
Conductance, thermal, F4.3; F25.1
Conduction
Conductivity, thermal, F25.1; F26.1
Constant air volume (CAV)
Construction. (See also Building envelopes)
Containers. (See also Cargo containers)
Contaminants
Continuity, fluid dynamics, F3.2
Control. (See also Controls, automatic; Supervisory control)
Controlled-atmosphere (CA) storage
Controlled-environment rooms (CERs), and plant growth, A25.16
Controls, automatic, F7. (See also Control)
Convection
Convectors
Convention centers, A5.5
Conversion factors, F39
Cooking appliances
Coolants, secondary
Coolers. (See also Refrigerators)
Cooling. (See also Air conditioning)
Cooling load
Cooling load temperature differential method with solar cooling load factors (CLTD/CLF), F18.57
Cooling towers, S40
Cool storage, S50.1
COP. See Coefficient of performance (COP)
Corn, drying, A26.1
Correctional facilities. See Justice facilities
Corrosion
Costs. (See also Economics)
Cotton, drying, A26.8
Courthouses, A10.5
Courtrooms, A10.5
CPVC. See Chlorinated polyvinyl chloride (CPVC)
Crawlspaces
Critical spaces
Crops. See Farm crops
Cruise terminals, A3.6
Cryogenics, R47
Curtain walls, F15.6
Dairy products, R33
Dampers
Dampness problems in buildings, A64.1
Dams, concrete cooling, R45.1
Darcy equation, F21.6
Darcy-Weisbach equation
Data centers, A20
Data-driven modeling
Daylighting, F19.26
DDC. See Direct digital control (DDC)
Dedicated outdoor air system (DOAS), F36.12; S4.14; S18.2, 8; S25.4; S51
Definitions, of refrigeration terms, R50
Defrosting
Degree-days, F14.12
Dehumidification, A48.15; S24
Dehumidifiers
Dehydration
Demand control kitchen ventilation (DCKV), A34.20
Density
Dental facilities, A9.17
Desiccants, F32.1; S24.1
Design-day climatic data, F14.12
Desorption isotherm, F26.20
Desuperheaters
Detection
Dew point, A64.8
Diamagnetism, and superconductivity, R47.5
Diesel fuel, F28.9
Diffusers, air, sound control, A49.12
Diffusion
Diffusivity
Dilution
Dining halls, in justice facilities, A10.4
DIR. See Dispersive infrared (DIR)
Direct digital control (DDC), F7.4, 11
Direct numerical simulation (DNS), turbulence modeling, F13.4; F24.13
Dirty bombs. See Chemical, biological, radio- logical, and explosive (CBRE) incidents
Disabilities, A8.23
Discharge coefficients, in fluid flow, F3.9
Dispersive infrared (DIR), F7.10
Display cases
Display cases, R15.2, 5
District energy (DE). See District heating and cooling (DHC)
District heating and cooling (DHC), S12
d-limonene, F31.12
DNS. See Direct numerical simulation (DNS)
DOAS. See Dedicated outdoor air system (DOAS)
Doors
Dormitories
Draft
Drag, in fluid flow, F3.5
Driers, R7.6. (See also Dryers)
Drip station, steam systems, S12.14
Dryers. (See also Driers)
Drying
DTW. See Dual-temperature water (DTW) system
Dual-duct systems
Dual-temperature water (DTW) system, S13.1
DuBois equation, F9.3
Duct connections, A64.10
Duct design
Ducts
Dust mites, F25.16
Dusts, S29.1
Dynamometers, A18.1
Earth, stabilization, R45.3, 4
Earthquakes, seismic-resistant design, A56.1
Economic analysis, A38
Economic coefficient of performance (ECOP), S7.2
Economic performance degradation index (EPDI), A63.5
Economics. (See also Costs)
Economizers
ECOP. See Economic coefficient of performance (ECOP)
ECS. See Environmental control system (ECS)
Eddy diffusivity, F6.7
Educational facilities, A8
EER. See Energy efficiency ratio (EER)
Effectiveness, heat transfer, F4.22
Effectiveness-NTU heat exchanger model, F19.19
Effective radiant flux (ERF), A54.2
Efficiency
Eggs, R34
Electricity
Electric thermal storage (ETS), S50.17
Electronic smoking devices (“e-cigarettes”), F11.19
Electrostatic precipitators, S29.7; S30.7
Elevators
Emissions, pollution, F28.8
Emissivity, F4.2
Emittance, thermal, F25.2
Enclosed vehicular facilities, A16
Energy
Energy and water use and management, A37
Energy efficiency ratio (EER)
Energy savings performance contracting (ESPC), A38.8
Energy transfer station, S12.37
Engines, S7
Engine test facilities, A18
Enhanced tubes. See Finned-tube heat transfer coils
Enthalpy
Entropy, F2.1
Environmental control
Environmental control system (ECS), A13
Environmental health, F10
Environmental tobacco smoke (ETS)
EPDI. See Economic performance degradation index (EPDI)
Equipment vibration, A49.44; F8.17
ERF. See Effective radiant flux (ERF)
ESPC. See Energy savings performance contracting (ESPC)
Ethylene glycol, in hydronic systems, S13.23
ETS. See Environmental tobacco smoke (ETS); Electric thermal storage (ETS)
Evaluation. See Testing
Evaporation, in tubes
Evaporative coolers. (See also Refrigerators)
Evaporative cooling, A53
Evaporators. (See also Coolers, liquid)
Exfiltration, F16.2
Exhaust
Exhibit buildings, temporary, A5.6
Exhibit cases
Exhibition centers, A5.5
Expansion joints and devices
Expansion tanks, S12.10
Explosions. See Chemical, biological, radio- logical, and explosive (CBRE) incidents
Fairs, A5.6
Family courts, A10.4. (See also Juvenile detention facilities)
Fan-coil units, S5.6
Fans, F19.18; S21
Farm crops, drying and storing, A26
Faults, system, reasons for detecting, A40.4
f-Chart method, sizing heating and cooling systems, A36.20
Fenestration. (See also Windows)
Fick’s law, F6.1
Filters, air, S29. (See also Air cleaners)
Finned-tube heat-distributing units, S36.2, 5
Finned-tube heat transfer coils, F4.25
Fins, F4.6
Fire/smoke control. See Smoke control
Firearm laboratories, A10.7
Fire management, A54.2
Fireplaces, S34.5
Fire safety
Fish, R19; R32
Fitness facilities. (See also Gymnasiums)
Fittings
Fixed-guideway vehicles, A12.7. (See also Mass-transit systems)
Fixture units, A51.1, 28
Flammability limits, gaseous fuels, F28.1
Flash tank, steam systems, S11.14
Floors
Flowers, cut
Flowmeters, A39.27; F37.18
Fluid dynamics computations, F13.1
Fluid flow, F3
Food. (See also specific foods)
Food service
Forced-air systems, residential, A1.1
Forensic labs, A10.6
Fouling factor
Foundations
Fountains, Legionella pneumophila control, A50.15
Fourier’s law, and heat transfer, F25.5
Four-pipe systems, S5.5
Framing, for fenestration
Freeze drying, A31.6
Freeze prevention. (See also Freeze protection systems)
Freeze protection systems, A52.19, 20
Freezers
Freezing
Friction, in fluid flow
Fruit juice, R38
Fruits
Fuel cells, combined heat and power (CHP), S7.22
Fuels, F28
Fume hoods, laboratory exhaust, A17.3
Fungi
Furnaces, S33
Galleries. See Museums, galleries, archives, and libraries
Garages
Gases
Gas-fired equipment, S34. (See also Natural gas)
Gas vents, S35.1
Gaussian process (GP) models, F19.30
GCHP. See Ground-coupled heat pumps (GCHP)
Generators
Geothermal energy, A35
Geothermal heat pumps (GHP), A35.1
Glaser method, F25.15
Glazing
Global climate change, and refrigerants, F29.1
Global warming potential (GWP), F29.5
Glossary, of refrigeration terms, R50
Glycols, desiccant solution, S24.2
Graphical symbols, F38
Green design, and sustainability, F35.1
Greenhouses. (See also Plant environments)
Grids, for computational fluid dynamics, F13.4
Ground-coupled heat pumps (GCHP)
Ground-coupled systems, F19.23
Ground-source heat pumps (GSHP), A35.1
Groundwater heat pumps (GWHP), A35.30
GSHP. See Ground-source heat pumps (GSHP)
Guard stations, in justice facilities, A10.5
GWHP. See Groundwater heat pumps (GWHP)
GWP. See Global warming potential (GWP)
Gymnasiums, A5.5; A8.3
HACCP. See Hazard analysis critical control point (HACCP)
Halocarbon
Hartford loop, S11.3
Hay, drying, A26.7
Hazard analysis and control, F10.4
Hazard analysis critical control point (HACCP), R22.4
Hazen-Williams equation, F22.6
HB. See Heat balance (HB)
Health
Health care facilities, A9. (See also specific types)
Health effects, mold, A64.1
Heat
Heat and moisture control, F27.1
Heat balance (HB), S9.23
Heat balance method, F19.3
Heat capacity, F25.1
Heat control, F27
Heaters, S34
Heat exchangers, S47
Heat flow, F25. (See also Heat transfer)
Heat flux, F25.1
Heat gain. (See also Load calculations)
Heating
Heating load
Heating seasonal performance factor (HSPF), S48.6
Heating values of fuels, F28.3, 9, 11
Heat loss. (See also Load calculations)
Heat pipes, air-to-air energy recovery, S26.14
Heat pumps
Heat recovery. (See also Energy, recovery)
Heat storage. See Thermal storage
Heat stress
Heat transfer, F4; F25; F26; F27. (See also Heat flow)
Heat transmission
Heat traps, A51.1
Helium
High-efficiency particulate air (HEPA) filters, A29.3; S29.6; S30.3
High-rise buildings. See Tall buildings
High-temperature short-time (HTST) pasteurization, R33.2
High-temperature water (HTW) system, S13.1
Homeland security. See Chemical, biological, radiological, and explosive (CBRE) incidents
Hoods
Hospitals, A9.3
Hot-box method, of thermal modeling, F25.8
Hotels and motels, A7
Hot-gas bypass, R1.35
Houses of worship, A5.3
HSI. See Heat stress, index (HSI)
HSPF. See Heating seasonal performance factor (HSPF)
HTST. See High-temperature short-time (HTST) pasteurization
Humidification, S22
Humidifiers, S22
Humidity (See also Moisture)
HVAC security, A61
Hybrid inverse change point model, F19.31
Hybrid ventilation, F19.26
Hydrofluorocarbons (HFCs), R1.1
Hydrofluoroolefins (HFOs), R1.1
Hydrogen, liquid, R47.3
Hydronic systems, S35. (See also Water systems)
Hygrometers, F7.9; F37.10, 11
Hygrothermal loads, F25.2
Hygrothermal modeling, F25.15; F27.10
IAQ. See Indoor air quality (IAQ)
IBD. See Integrated building design (IBD)
Ice
Ice makers
Ice rinks, A5.5; R44
ID50‚ mean infectious dose, A61.9
Ignition temperatures of fuels, F28.2
IGUs. See Insulating glazing units (IGUs)
Illuminance, F37.31
Indoor airflow, A59.1
Indoor air quality (IAQ). (See also Air quality)
Indoor environmental modeling, F13
Indoor environmental quality (IEQ), kitchens, A33.20. (See also Air quality)
Indoor swimming pools. (See also Natatoriums)
Induction
Industrial applications
Industrial environments, A15, A32; A33
Industrial exhaust gas cleaning, S29. (See also Air cleaners)
Industrial hygiene, F10.3
Infiltration. (See also Air leakage)
Infrared applications
In-room terminal systems
Instruments, F14. (See also specific instruments or applications)
Insulating glazing units (IGUs), F15.5
Insulation, thermal
Integrated building design (IBD), A60.1
Integrated project delivery (IPD), A60.1
Integrated project delivery and building design,
Intercoolers, ammonia refrigeration systems, R2.12
Internal heat gains, F19.13
Jacketing, insulation, R10.7
Jails, A10.4
Joule-Thomson cycle, R47.6
Judges’ chambers, A10.5
Juice, R38.1
Jury facilities, A10.5
Justice facilities, A10
Juvenile detention facilities, A9.1. (See also Family courts)
K-12 schools, A8.3
Kelvin’s equation, F25.11
Kirchoff’s law, F4.12
Kitchens, A34
Kleemenko cycle, R47.13
Krypton, recovery, R47.18
Laboratories, A17
Laboratory information management systems (LIMS), A10.8
Lakes, heat transfer, A35.37
Laminar flow
Large eddy simulation (LES), turbulence modeling, F13.3; F24.13
Laser Doppler anemometers (LDA), F37.17
Laser Doppler velocimeters (LDV), F37.17
Latent energy change materials, S50.2
Laundries
LCR. See Load collector ratio (LCR)
LD50‚ mean lethal dose, A61.9
LDA. See Laser Doppler anemometers (LDA)
LDV. See Laser Doppler velocimeters (LDV)
LE. See Life expectancy (LE) rating
Leakage
Leakage function, relationship, F16.15
Leak detection of refrigerants, F29.9
Legionella pneumophila, A50.15; F10.7
Legionnaires’ disease. See Legionella pneumophila
LES. See Large eddy simulation (LES)
Lewis relation, F6.9; F9.4
Libraries. See Museums, galleries, archives, and libraries
Life expectancy (LE) rating, film, A23.3
Lighting
Light measurement, F37.31
LIMS. See Laboratory information management systems (LIMS)
Linde cycle, R47.6
Liquefied natural gas (LNG), S8.6
Liquefied petroleum gas (LPG), F28.5
Liquid overfeed (recirculation) systems, R4
Lithium bromide/water, F30.71
Lithium chloride, S24.2
LNG. See Liquefied natural gas (LNG)
Load calculations
Load collector ratio (LCR), A36.22
Local exhaust. See Exhaust
Loss coefficients
Louvers, F15.33
Low-temperature water (LTW) system, S13.1
LPG. See Liquefied petroleum gas (LPG)
LTW. See Low-temperature water (LTW) system
Lubricants, R6.1; R12. (See also Lubrication; Oil)
Lubrication, R12
Mach number, S38.32
Maintenance. (See also Operation and maintenance)
Makeup air units, S28.8
Malls, A2.7
Manometers, differential pressure readout, A39.25
Manufactured homes, A1.9
Masonry, insulation, F26.7. (See also Building envelopes)
Mass transfer, F6
Mass-transit systems
McLeod gages, F37.13
Mean infectious dose (ID50), A61.9
Mean lethal dose (LD50), A61.9
Mean radiant temperature (MRT), A54.1
Mean temperature difference, F4.22
Measurement, F37. (See also Instruments)
Meat, R30
Mechanical equipment room, central
Mechanical traps, steam systems, S11.8
Medium-temperature water (MTW) system, S13.1
Megatall buildings, A4.1
Meshes, for computational fluid dynamics, F13.4
Metabolic rate, F9.6
Metals and alloys, low-temperature, R48.6
Microbial growth, R22.4
Microbial volatile organic chemicals (MVOCs), F10.8
Microbiology of foods, R22.1
Microphones, F37.29
Mines, A30
Modeling. (See also Data-driven modeling; Energy, modeling)
Model predictive control (MPC), A65.6
Moist air
Moisture (See also Humidity)
Mold, A64.1; F25.16
Mold-resistant gypsum board, A64.7
Molecular sieves, R18.10; R41.9; R47.13; S24.5. (See also Zeolites)
Montreal Protocol, F29.1
Morgues, A9.1; R16.3
Motors, S45
Movie theaters, A5.3
MPC (model predictive control), A65.6
MRT. See Mean radiant temperature (MRT)
Multifamily residences, A1.8
Multiple-use complexes
Multisplit unitary equipment, S48.1
Multizone airflow modeling, F13.14
Museums, galleries, archives, and libraries
MVOCs. See Microbial volatile organic compounds (MVOCs)
Natatoriums. (See also Swimming pools)
Natural gas, F28.5
Navier-Stokes equations, F13.2
NC curves. See Noise criterion (NC) curves
Net positive suction head (NPSH), A35.31; R2.9; S44.10
Network airflow models, F19.25
Neutral pressure level (NPL), A4.1
Night setback, recovery, A43.44
Nitrogen
Noise, F8.13. (See also Sound)
Noise criterion (NC) curves, F8.16
Noncondensable gases
Normalized mean bias error (NMBE), F19.33
NPL. See Neutral pressure level (NPL)
NPSH. See Net positive suction head (NPSH)
NTU. See Number of transfer units (NTU)
Nuclear facilities, A29
Number of transfer units (NTU)
Nursing facilities, A9.17
Nuts, storage, R42.7
Odors, F12
ODP. See Ozone depletion potential (ODP)
Office buildings
Oil, fuel, F28.7
Oil. (See also Lubricants)
Olf unit, F12.6
One-pipe systems
Operating costs, A38.4
Operation and maintenance, A39. (See also Maintenance)
Optimization, A43.4
Outdoor air, free cooling (See also Ventilation)
Outpatient health care facilities, A9.16
Owning costs, A38.1
Oxygen
Ozone
Ozone depletion potential (ODP), F29.5
PACE. (See Property assessment for clean energy S50.2
Packaged terminal air conditioners (PTACs), S49.5
Packaged terminal heat pumps (PTHPs), S49.5
PAH. See Polycyclic aromatic hydrocarbons (PAHs)
Paint, and moisture problems, F25.16
Panel heating and cooling, S6. (See also Radiant heating and cooling)
Paper
Paper products facilities, A27
Parallel compressor systems, R15.14
Particulate matter, indoor air quality (IAQ), F10.5
Passive heating, F19.27
Pasteurization, R33.2
Peak dew point, A64.10
Peanuts, drying, A26.8
PEC systems. See Personal environmental control (PEC) systems
PEL. See Permissible exposure limits (PEL)
Performance contracting, A42.2
Performance monitoring, A48.6
Permafrost stabilization, R45.4
Permeability
Permeance
Permissible exposure limits (PELs), F10.5
Personal environmental control (PEC) systems, F9.26
Pharmaceutical manufacturing cleanrooms, A19.11
Pharmacies, A9.13
Phase-change materials, thermal storage in, S50.16, 27
Photographic materials, A23
Photovoltaic (PV) systems, S36.18. (See also Solar energy)
Physical properties of materials, F33
Physiological principles, humans. (See also Comfort)
Pigs. See Swine
Pipes. (See also Piping)
Piping. (See also Pipes)
Pitot tubes, A39.2; F37.17
Places of assembly, A5
Planes. See Aircraft
Plank’s equation, R20.7
Plant environments, A25.10
Plenums
PMV. See Predicted mean vote (PMV)
Police stations, A10.1
Pollutant transport modeling. See Contami- nants, indoor, concentration prediction
Pollution
Polycyclic aromatic hydrocarbons (PAHs), F10.6
Polydimethylsiloxane, F31.12
Ponds, spray, S40.6
Pope cell, F37.12
Positive building pressure, A64.11
Positive positioners, F7.8
Potatoes
Poultry. (See also Animal environments)
Power grid, A63.9
Power-law airflow model, F13.14
Power plants, A28
PPD. See Predicted percent dissatisfied (PPD)
Prandtl number, F4.17
Precooling
Predicted mean vote (PMV), F37.32
Predicted percent dissatisfied (PPD), F9.18
Preschools, A8.1
Pressure
Pressure drop. (See also Darcy-Weisbach equation)
Primary-air systems, S5.10
Printing plants, A21
Prisons, A10.4
Produce
Product load, R15.6
Propane
Property assessment for clean energy (PACE), A38.9
Propylene glycol, hydronic systems, S13.23
Psychrometers, F1.9
Psychrometrics, F1
PTACs. See Packaged terminal air condition- ers (PTACs)
PTHPs. See Packaged terminal heat pumps (PTHPs)
Public buildings. See Commercial and public buildings; Places of assembly
Pumps, F19.18
Purge units, centrifugal chillers, S43.11
PV systems. See Photovoltaic (PV) systems; Solar energy
Radiant heating and cooling, A55; S6.1; S15; S33.4. (See also Panel heating and cooling)
Radiant time series (RTS) method, F18.2, 22
Radiation
Radiators, S36.1, 5
Radioactive gases, contaminants, F11.21
Radiometers, A54.7
Radiosity method, F19.26
Radon, F10.16, 22
Rail cars, R25. (See also Cargo containers)
Railroad tunnels, ventilation
Rain, and building envelopes, F25.4
RANS. See Reynolds-Averaged Navier-Stokes (RANS) equation
Rapid-transit systems. See Mass-transit systems
Rayleigh number, F4.20
Ray tracing method, F19.27
RC curves. See Room criterion (RC) curves
Receivers
Recycling refrigerants, R9.3
Refrigerant/absorbent pairs, F2.15
Refrigerant control devices, R11
Refrigerants, F29.1
Refrigerant transfer units (RTU), liquid chillers, S43.11
Refrigerated facilities, R23
Refrigeration, F1.16. (See also Absorption; Adsorption)
Refrigeration oils, R12. (See also Lubricants)
Refrigerators
Regulators. (See also Valves)
Relative humidity, F1.8
Residential health care facilities, A9.17
Residential systems, A1
Resistance, thermal, F4; F25; F26. (See also R-values)
Resistance temperature devices (RTDs), F7.9; F37.6
Resistivity, thermal, F25.1
Resource utilization factor (RUF), F34.2
Respiration of fruits and vegetables, R19.17
Restaurants
Retail facilities, A2
Retrofit performance monitoring, A42.4
Retrofitting refrigerant systems, contaminant control, R7.9
Reynolds-averaged Navier-Stokes (RANS) equation, F13.3; F24.13
Reynolds number, F3.3
Rice, drying, A26.9
RMS. See Root mean square (RMS)
Road tunnels, A16.3
Roofs, U-factors, F27.2
Room air distribution, A58; S20.1
Room criterion (RC) curves, F8.16
Root mean square (RMS), F37.1
RTDs. See Resistance temperature devices (RTDs)
RTS. See Radiant time series (RTS)
RTU. See Refrigerant transfer units (RTU)
RUF. See Resource utilization factor (RUF)
Rusting, of building components, F25.16
R-values, F23; F25; F26. (See also Resistance, thermal)
Safety
Sanitation
Savings-to-investment ratio (SIR), A38.12
Savings-to-investment-ratio (SIR), A38.12
Scale
Schneider system, R23.7
Schools
Seasonal energy efficiency ratio (SEER)
Security. See Chemical, biological, radio- logical, and explosive (CBRE) incidents
Seeds, storage, A26.11
SEER. See Seasonal energy efficiency ratio (SEER)
Seismic restraint, A49.53; A56.1
Semivolatile organic compounds (SVOCs), F10.4, 12; F11.15
Sensors
Separators, lubricant, R11.23
Service water heating, A51
SES. See Subway environment simulation (SES) program
Set points, A65.1
Shading
Ships, A13
Shooting ranges, indoor, A10.8
Short-tube restrictors, R11.31
Silica gel, S24.1, 4, 6, 12
Single-duct systems, all-air, S4.11
SIR. See Savings-to-investment ratio (SIR)
Skating rinks, R44.1
Skylights, and solar heat gain, F15.21
Slab heating, A52
Slab-on-grade foundations, A45.11
SLR. See Solar-load ratio (SLR)
Smart building systems, A63.1
Smart grid, A63.9, 11
Smoke control, A54
Snow-melting systems, A52
Snubbers, seismic, A56.8
Sodium chloride brines, F31.1
Soft drinks, R39.10
Software, A65.7
Soils. (See also Earth)
Solar energy, A36; S37.1 (See also Solar heat gain; Solar radiation)
Solar heat gain, F15.14; F18.16
Solar-load ratio (SLR), A36.22
Solar-optical glazing, F15.14
Solar radiation, F14.8; F15.14
Solid fuel
Solvent drying, constant-moisture, A31.7
Soot, F28.20
Sorbents, F32.1
Sorption isotherm, F25.10; F26.20
Sound, F8. (See also Noise)
Soybeans, drying, A26.7
Specific heat
Split-flux method, F19.26
Spot cooling
Spot heating, A54.4
Stack effect
Stadiums, A5.4
Stairwells
Standard atmosphere, U.S., F1.1
Standards, A66. (See also Codes)
Static air mixers, S4.8
Static electricity and humidity, S22.2
Steam
Steam systems, S11
Steam traps, S11.7
Stefan-Boltzmann equation, F4.2, 12
Stevens’ law, F12.3
Stirling cycle, R47.14
Stokers, S31.17
Storage
Stoves, heating, S34.5
Stratification
Stroboscopes, F37.28
Subcoolers
Subway environment simulation (SES) program, A16.3
Subway systems. (See also Mass-transit systems)
Suction risers, R2.24
Sulfur content, fuel oils, F28.9
Superconductivity, diamagnetism, R47.5
Supermarkets. See Retail facilities, supermarkets
Supertall buildings, A4.1
Supervisory control, A43
Supply air outlets, S20.2. (See also Air outlets)
Surface effect. See Coanda effect
Surface transportation
Surface water heat pump (SWHP), A35.3
Sustainability, F16.1; F35.1; S48.2
SVFs. See Synthetic vitreous fibers (SVFs)
SVOCs. See Semivolatile organic compounds (SVOCs)
SWHP. See Surface water heat pump (SWHP)
Swimming pools. (See also Natatoriums)
Swine, recommended environment, A25.7
Symbols, F38
Synthetic vitreous fibers (SVFs), F10.6
TABS. See Thermally activated building systems (TABS)
Tachometers, F37.28
Tall buildings, A4
Tanks, secondary coolant systems, R13.2
TDD. See Tubular daylighting devices
Telecomunication facilities, air-conditioning systems, A20.1
Temperature
Temperature-controlled transport, R25.1
Temperature index, S22.3
Terminal units. [See also Air terminal units (ATUs)], A48.13, F19.16; S20.7
Terminology, of refrigeration, R50
Terrorism. See Chemical, biological, radio- logical, and explosive (CBRE) incidents
TES. See Thermal energy storage (TES)
Testing
Testing, adjusting, and balancing. (See also Balancing)
TETD/TA. See Total equivalent temperature differential method with time averaging (TETD/TA)
TEWI. See Total equivalent warning impact (TEWI)
Textile processing plants, A22
TFM. See Transfer function method (TFM)
Theaters, A5.3
Thermal bridges, F25.8
Thermal comfort. See Comfort
Thermal displacement ventilation (TDV), F19.17
Thermal emittance, F25.2
Thermal energy storage (TES), S8.6; S50
Thermally activated building systems (TABS), A43.3, 33
Thermal-network method, F19.11
Thermal properties, F26.1
Thermal resistivity, F25.1
Thermal storage. See Thermal energy storage (TES) S50
Thermal transmission data, F26
Thermal zones, F19.14
Thermistors, R11.4
Thermodynamics, F2.1
Thermometers, F37.5
Thermopile, F7.4; F37.9; R45.4
Thermosiphons
Thermostats
Three-dimensional (3D) printers, F11.18
Three-pipe distribution, S5.6
Tobacco smoke
Tollbooths
Total equivalent temperature differential method with time averaging (TETD/TA), F18.57
Total equivalent warming impact (TEWI), F29.5
Trailers and trucks, refrigerated, R25. (See also Cargo containers)
Transducers, F7.10, 13
Transfer function method (TFM); F18.57; F19.3
Transmittance, thermal, F25.2
Transmitters, F7.9, 10
Transpiration, R19.19
Transportation centers
Transport properties of refrigerants, F30
Traps
Trucks, refrigerated, R25. (See also Cargo containers)
Tubular daylighting devices (TDDs), F15.30
Tuning automatic control systems, F7.19
Tunnels, vehicular, A16.1
Turbines, S7
Turbochargers, heat recovery, S7.34
Turbulence modeling, F13.3
Turbulent flow, fluids, F3.3
Turndown ratio, design capacity, S13.4
Two-node model, for thermal comfort, F9.19
Two-pipe systems, S5.5; S13.20
U.S. Marshal spaces, A10.6
U-factor
Ultralow-penetration air (ULPA) filters, S29.6; S30.3
Ultraviolet (UV) lamp systems, S17
Ultraviolet air and surface treatment, A62
Ultraviolet germicidal irradiation (UVGI), A60.1; S17.1. [See also Ultraviolet (UV) lamp systems]
Ultraviolet germicidal irradiation (UVGI), A62.1; S17.1. [See also Ultraviolet (UV) lamp systems]
Uncertainty analysis
Underfloor air distribution (UFAD) systems, A4.6; A58.14; F19.17
Unitary systems, S48
Unit heaters. See Heaters
Units and conversions, F39
Unit ventilators, S28.1
Utility interface, electric, S7.43
Utility rates, A63.11
UV. See Ultraviolet (UV) lamp systems
UVGI. See Ultraviolet germicidal irradiation (UVGI)
Vacuum cooling, of fruits and vegetables, R28.9
Validation, of airflow modeling, F13.9, 10, 17
Valves. (See also Regulators)
Vaporization systems, S8.6
Vapor pressure, F27.8; F33.2
Vapor retarders, jackets, F23.12
Variable-air-volume (VAV) systems
Variable-frequency drives, S45.14
Variable refrigerant flow (VRF), S18.1; S48.1, 14
Variable-speed drives. See Variable-frequency drives S50
VAV. See Variable-air-volume (VAV) systems
Vegetables, R37
Vehicles
Vena contracta, F3.4
Vending machines, R16.5
Ventilation, F16
Ventilators
Venting
Verification, of airflow modeling, F13.9, 10, 17
Vessels, ammonia refrigeration systems, R2.11
Vibration, F8.17
Viral pathogens, F10.9
Virgin rock temperature (VRT), and heat release rate, A30.3
Viscosity, F3.1
Volatile organic compounds (VOCs), F10.11
Voltage, A56.1
Volume ratio, compressors
VRF. See Variable refrigerant flow (VRF)
VRT. See Virgin rock temperature (VRT)
Walls
Warehouses, A3.8
Water
Water heaters
Water/lithium bromide absorption
Water-source heat pump (WSHP), S2.4; S48.11
Water systems, S13
Water treatment, A50
Water use and management (See Energy and water use and management)
Water vapor control, A45.6
Water vapor permeance/permeability, F26.12, 17, 18
Water vapor retarders, F26.6
Water wells, A35.30
Weather data, F14
Weatherization, F16.18
Welding sheet metal, S19.12
Wet-bulb globe temperature (WBGT), heat stress, A32.5
Wheels, rotary enthalpy, S26.9
Whirlpools and spas
Wien’s displacement law, F4.12
Wind. (See also Climatic design information; Weather data)
Wind chill index, F9.23
Windows. (See also Fenestration)
Wind restraint design, A56.15
Wineries
Wireless sensors, A63.7
Wood construction, and moisture, F25.10
Wood products facilities, A27.1
Wood pulp, A27.2
Wood stoves, S34.5
WSHP. See Water-source heat pump (WSHP)
Xenon, R47.18
Zeolites, R18.10; R41.9; R47.13; S24.5. (See also Molecular sieves)

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Licensed for single user. © 2020 ASHRAE, Inc.

2020 ASHRAE HANDBOOK

Heating, Ventilating, and Air-Conditioning SYSTEMS AND EQUIPMENT SI Edition

ASHRAE, 1791 Tullie Circle, N.E., Atlanta, GA 30329 www.ashrae.org

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© 2020 ASHRAE. All rights reserved. DEDICATED TO THE ADVANCEMENT OF

Licensed for single user. © 2020 ASHRAE, Inc.

THE PROFESSION AND ITS ALLIED INDUSTRIES

No part of this publication may be reproduced without permission in writing from ASHRAE, except by a reviewer who may quote brief passages or reproduce illustrations in a review with appropriate credit; nor may any part of this book be reproduced, stored in a retrieval system, or transmitted in any way or by any means—electronic, photocopying, recording, or other—without permission in writing from ASHRAE. Requests for permission should be submitted at www.ashrae.org/permissions. Volunteer members of ASHRAE Technical Committees and others compiled the information in this handbook, and it is generally reviewed and updated every four years. Comments, criticisms, and suggestions regarding the subject matter are invited. Any errors or omissions in the data should be brought to the attention of the Editor. Additions and corrections to Handbook volumes in print will be published in the Handbook published the year following their verification and, as soon as verified, on the ASHRAE Internet website. DISCLAIMER ASHRAE has compiled this publication with care, but ASHRAE has not investigated, and ASHRAE expressly disclaims any duty to investigate, any product, service, process, procedure, design, or the like that may be described herein. The appearance of any technical data or editorial material in this publication does not constitute endorsement, warranty, or guaranty by ASHRAE of any product, service, process, procedure, design, or the like. ASHRAE does not warrant that the information in this publication is free of errors. The entire risk of the use of any information in this publication is assumed by the user. ISBN 978-1-947192-53-9 ISSN 1930-7705

The paper for this book is both acid- and elemental-chlorine-free and was manufactured with pulp obtained from sources using sustainable forestry practices.

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CONTENTS Contributors ASHRAE Technical Committees, Task Groups, and Technical Resource Groups ASHRAE Research: Improving the Quality of Life Preface AIR-CONDITIONING AND HEATING SYSTEMS

Licensed for single user. © 2020 ASHRAE, Inc.

Chapter

1. 2. 3. 4. 5. 6. 7. 8. 9. 10. 11. 12. 13. 14. 15. 16. 17. 18.

HVAC System Analysis and Selection (TC 9.1, Large Building Air-Conditioning Systems) Decentralized Cooling and Heating (TC 9.1) Central Cooling and Heating Plants (TC 9.1) Air Handling and Distribution (TC 9.1) In-Room Terminal Systems (TC 9.1) Radiant Heating and Cooling (TC 6.5, Radiant Heating and Cooling) Combined Heat and Power Systems (TC 1.10, Cogeneration Systems) Combustion Turbine Inlet Cooling (TC 1.10) Applied Heat Pump and Heat Recovery Systems (TC 6.8, Geothermal Heat Pump and Energy Recovery Applications) Small Forced-Air Heating and Cooling Systems (TC 6.3, Central Forced Air Heating and Cooling Systems) Steam Systems (TC 6.1, Hydronic and Steam Equipment and Systems) District Heating and Cooling (TC 6.2, District Energy) Hydronic Heating and Cooling (TC 6.1) Condenser Water Systems (TC 6.1) Medium- and High-Temperature Water Heating (TC 6.1) Infrared Radiant Heating (TC 6.5) Ultraviolet Lamp Systems (TC 2.9, Ultraviolet Air and Surface Treatment) Variable Refrigerant Flow [TC 8.7, Variable Refrigerant Flow (VRF)]

AIR-HANDLING EQUIPMENT AND COMPONENTS Chapter

19. 20. 21. 22. 23. 24. 25. 26. 27. 28. 29. 30.

Duct Construction (TC 5.2, Duct Design) Room Air Distribution Equipment (TC 5.3, Room Air Distribution) Fans (TC 5.1, Fans) Humidifiers (TC 5.11, Humidifying Equipment) Air-Cooling and Dehumidifying Coils (TC 8.4, Air-to-Refrigerant Heat Transfer Equipment) Desiccant Dehumidification and Pressure-Drying Equipment (TC 8.12, Desiccant Dehumidification Equipment and Components) Mechanical Dehumidifiers and Related Components (TC 8.10, Mechanical Dehumidification Equipment and Heat Pipes) Air-to-Air Energy Recovery Equipment (TC 5.5, Air-to-Air Energy Recovery) Air-Heating Coils (TC 8.4) Unit Ventilators, Unit Heaters, and Makeup Air Units (TC 6.1 and TC 5.8, Industrial Ventilation) Air Cleaners for Particulate Contaminants (TC 2.4, Particulate Air Contaminants and Particulate Contaminant Removal Equipment) Industrial Gas Cleaning and Air Pollution Control [TC 5.4, Industrial Process Air Cleaning (Air Pollution Control)]

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HEATING EQUIPMENT AND COMPONENTS Chapter

31. 32. 33. 34. 35. 36. 37.

Automatic Fuel-Burning Systems (TC 6.10, Fuels and Combustion) Boilers (TC 6.1) Furnaces (TC 6.3) Residential In-Space Heating Equipment (TC 6.5) Chimney, Vent, and Fireplace Systems (TC 6.10) Hydronic Heat-Distributing Units and Radiators (TC 6.1) Solar Energy Equipment (TC 6.7, Solar Energy Utilization)

COOLING EQUIPMENT AND COMPONENTS

Licensed for single user. © 2020 ASHRAE, Inc.

Chapter

38. Compressors (TC 8.1, Positive Displacement Compressors, and TC 8.2, Centrifugal Machines) 39. Condensers (TC 8.4, TC 8.5, Liquid-to-Refrigerant Heat Exchangers, and TC 8.6, Cooling Towers and Evaporative Condensers) 40. Cooling Towers (TC 8.6) 41. Evaporative Air-Cooling Equipment (TC 5.7, Evaporative Cooling) 42. Liquid Coolers (TC 8.5) 43. Liquid-Chilling Systems (TC 8.1 and TC 8.2)

GENERAL COMPONENTS 44. Centrifugal Pumps (TC 6.1) 45. Motors, Motor Controls, and Variable-Frequency Drives (TC 1.11, Electric Motors and Motor Control) 46. Valves (TC 6.1) 47. Heat Exchangers (TC 6.1)

PACKAGED, UNITARY, AND SPLIT-SYSTEM EQUIPMENT Chapter

48. Unitary Air Conditioners and Heat Pumps (TC 8.11, Unitary and Room Air Conditioners and Heat Pumps) 49. Room Air Conditioners and Packaged Terminal Air Conditioners (TC 8.11)

GENERAL 50. Thermal Storage (TC 6.9, Thermal Storage) 51. Dedicated Outdoor Air Systems (TC 8.10 52. Codes and Standards

Additions and Corrections Index Composite index to the 2017 Fundamentals, 2018 Refrigeration, 2019 HVAC Applications, and 2020 HVAC Systems and Equipment volumes

Comment Pages

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CONTRIBUTORS In addition to the Technical Committees, the following individuals contributed significantly to this volume. The appropriate chapter numbers follow each contributor’s name. Howard McKew (1, 5) BuildingSmartSoftware, LLC

Mick Schwedler (13, 14) The Trane Company

Joseph Brooks (21) AMCA International

Steven Nicklas (2) Gene Strehlow (2)

Forrest B. Fencl (17) UV Resources

Patrick Chinoda (21) Revcor, Inc.

Stephen W. Duda (3) Ross & Baruzzini, Inc.

Jaak Geboers (17) Philips Lighting BV

Armin Hauer (21, 45) ebm-papst, Inc.

R. Dan Leath (3) Murphy Company

Stephen B. Martin, Jr. (17) Centers for Disease Control/National Institute for Occupational Safety and Health

Zhiping Wang (21) Morrison Products, Inc.

Licensed for single user. © 2020 ASHRAE, Inc.

Rachel Romero (4) National Renewable Energy Laboratory Lynn Werman (4)

Dean A. Saputa (17) UV Resources

Dove Feng (6) Taylor Engineering

Richard L. Vincent (17) Icahn School of Medicine at Mount Sinai

Ryan MacGillivray (6) Daniels Wingerak Engineering Ltd. Paul Raftery (6) University of California, Berkeley Peter Simmonds (6) Building and System Analytics Gearoid Foley (7) Integrated CHP Systems

David L. Witham (17) UltraViolet Devices, Inc. William Artis (18) Daikin Applied N.Y. Brian Bogdan (18) LG Electronics Yvette Daniel (18) Mitsubishi Electric USA, Inc.

Lucas B. Hyman (7, 12, 51)

Paul Doppel (18) Mitsubishi Electric USA, Inc.

Birol I. Kilkis (7) Baskent University

Dermot McMorrow (18) Mitsubishi Electric Canada

John S. Andrepont (8, 51) The Cool Solutions Company

Douglas Tucker (18) Mitsubishi Electric USA, Inc.

Dharam Punwani (7, 8) Avalon Consulting, Inc.

Herman Behls (19)

Chris Gray (9) Southern Company

Ralph Koerber (19) ATCO Rubber Products

Gary Phetteplace (9, 12) GWA Research

Craig Wray (19, 21)

Gary Berlin (22) Humidity Consulting LLC Sukru Erisgen (22) DriSteem Corporation Nicholas Lea (22) Nortec Humidity Ltd. William Fox (23, 27, 39) Ingersoll Rand/Trane Steve Brickley (24) Munters Corporation Phillip Farese (24) Advantix Systems Michael Sherber (24) PPL SavageALERT Richard Wolcott (24) Ralph Kittler (25) Seresco Technologies, Inc. Alois Malik (25) Dectron International, Inc. Harry Milliken (25) Desert-Aire Corporation Prakash Dhamshala (26) University of Tennessee Gursaran D. Mathur (26, 41) Calsonic Kansei North America

Charles Gaston (10, 33) The Pennsylvania State University

Kevin Cash (20) Trox

Louis Starr (10) Northwest Energy Efficiency Alliance

Jose Palma (20) Titus

Ramez Afify (11) E4P Consulting Engineering PLLC

Curtis Peters (20) Nailor Industries

Jason Atkisson (11, 32) Affiliated Engineers, Inc.

David Pich (20) Titus

Rex Scare (11) Armstrong International, Inc.

Jack Stegall (20) Energistics

Ken Mead (28) National Institute of Occupational Safety and Health

Steve Tredinnick (12, 13, 14) Burns & McDonnell, Inc.

Michael Brendel (21) Lau Industries

Eric Brodsky (29) Research Products

Tricia Bruenn (28) Belimo Americas Scott Fisher (28, 36, 48) State Farm John McKernan (28) U.S. Environmental Protection Agency

ThisfileislicensedtoOsamaKhayata([email protected]).CopyrightASHRAE2020.

Tom Justice (29) ZENE, LLC Filtration

Ray Good (38) Danfoss Turbocor Compressors, Inc.

Carolyn (Gemma) Kerr (29)

Rick Heiden (38) The Trane Company

Phil Maybee (29) The Filter Man Ltd.

Ken Fonstad (45) ABB, Inc. Paul Lin (45) Regal Beloit

Justin Kauffman (38, 43) Johnson Controls

Larry Brand (31, 35) Gas Technology Institute

Alexander D. Leyderman (38, 43) Triumph Thermal Management SystemsMD

Mehdi M. Doura (31, 35) Lochinvar LLC Jennifer Guerrero-Ferreira (31, 35) Bekaert Corporation Tom Neill (31, 35) Mestek Inc.

Michael Perevozchikov (38) Emerson Climate Technologies Lorenzo Cremaschi (39) Oklahoma State University

Robert Walker (47) Belimo Aircontrols (USA), Inc. Theodore E. Duffy (49, 50) Johnson Supply

Mike Scofield (41) Conservation Mechanical Systems

Cory Weiss (31, 35) Field Controls LLC Diane Jakobs (33) Rheem Manufacturing Company George Yaeger (33) Sears Holdings Corporation Constantinos A. Balaras (37) National Observatory of Athens Elena G. Dascalaki (37) National Observatory of Athens Svein Morner (37) Sustainable Engineering Group LLC Khalid Nagidi (37) Energy Management Consulting Group, LLC

Marcelo Acosta (47) Armstrong Fluid Technology

Steve Taylor (47) Taylor Engineering, LLC

Sankar Padhmanabhan (39) Danfoss LLC

Paul Sohler (31, 35) Crown Boiler Co

Tom Lowery (45) Schneider Electric

Eric Rosenberg (47) Grumman/Butkus Associates

Satheesh Kulankara (39) Johnson Controls, Inc.

Bill Roy (31, 35) Timco Rubber Licensed for single user. © 2020 ASHRAE, Inc.

Greg Towsley (44)

Kevin Mercer (49) Carrier Corporation

Satyam Bendapudi (42) Carrier Corporation

Ray Rite (49, 50) Ingersoll Rand

Laurant Abbas (43) Arkema

Don Schuster (49) Ingersoll Rand

Fred Betz (38, 43) PEDCO E & A Services, Inc.

Craig Messmer (50) Unico

Hermann Renz (38) Bitzer International Niels Bidstrup (44) Grundfos Holding A/S

Geoff Bares (51) CB&I

Larry Konopacz (44) Xylem - Applied Water Systems

Henry Becker (51) H-O-H Water Technology

David Lee (44) Armstrong Fluid Technology

Paul Steffes (51) Steffes Corporation

ASHRAE HANDBOOK COMMITTEE Suzanne LeViseur, Chair 2020 HVAC Systems and Equipment Volume Subcommittee: Michael P. Patton, Chair Caroline C. Calloway

Nicolas Lemire

Prakash R. Dhamshala

Florentino Roson Rodriguez

ASHRAE HANDBOOK STAFF Mark S. Owen, Publisher Director of Publications and Education Heather E. Kennedy, Editor Hayden Spiess, Editorial Assistant Nancy F. Thysell, Typographer/Page Designer David Soltis, Group Manager, and Jayne E. Jackson, Publication Traffic Administrator Publishing Services

ThisfileislicensedtoOsamaKhayata([email protected]).CopyrightASHRAE2020.

Steven C. Sill

ASHRAE TECHNICAL COMMITTEES, TASK GROUPS, AND TECHNICAL RESOURCE GROUPS

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SECTION 1.0—FUNDAMENTALS AND GENERAL 1.1 Thermodynamics and Psychrometrics 1.2 Instruments and Measurements 1.3 Heat Transfer and Fluid Flow 1.4 Control Theory and Application 1.5 Computer Applications 1.6 Terminology 1.7 Business, Management & General Legal Education 1.8 Mechanical Systems Insulation 1.9 Electrical Systems 1.10 Cogeneration Systems 1.11 Electric Motors and Motor Control 1.12 Moisture Management in Buildings 1.13 Optimization SECTION 2.0—ENVIRONMENTAL QUALITY 2.1 Physiology and Human Environment 2.2 Plant and Animal Environment 2.3 Gaseous Air Contaminants and Gas Contaminant Removal Equipment 2.4 Particulate Air Contaminants and Particulate Contaminant Removal Equipment 2.5 Global Climate Change 2.6 Sound and Vibration 2.7 Seismic and Wind Resistant Design 2.8 Building Environmental Impacts and Sustainability 2.9 Ultraviolet Air and Surface Treatment TG2 Heating Ventilation and Air-Conditioning Security (HVAC) SECTION 3.0—MATERIALS AND PROCESSES 3.1 Refrigerants and Secondary Coolants 3.2 Refrigerant System Chemistry 3.3 Refrigerant Contaminant Control 3.4 Lubrication 3.6 Water Treatment 3.8 Refrigerant Containment SECTION 4.0—LOAD CALCULATIONS AND ENERGY REQUIREMENTS 4.1 Load Calculation Data and Procedures 4.2 Climatic Information 4.3 Ventilation Requirements and Infiltration 4.4 Building Materials and Building Envelope Performance 4.5 Fenestration 4.7 Energy Calculations 4.10 Indoor Environmental Modeling TRG4 Indoor Air Quality Procedure Development SECTION 5.0—VENTILATION AND AIR DISTRIBUTION 5.1 Fans 5.2 Duct Design 5.3 Room Air Distribution 5.4 Industrial Process Air Cleaning (Air Pollution Control) 5.5 Air-to-Air Energy Recovery 5.6 Control of Fire and Smoke 5.7 Evaporative Cooling 5.8 Industrial Ventilation 5.9 Enclosed Vehicular Facilities 5.10 Kitchen Ventilation 5.11 Humidifying Equipment SECTION 6.0—HEATING EQUIPMENT, HEATING AND COOLING SYSTEMS AND APPLICATIONS 6.1 Hydronic and Steam Equipment and Systems 6.2 District Energy 6.3 Central Forced Air Heating and Cooling Systems 6.5 Radiant Heating and Cooling 6.6 Service Water Heating Systems 6.7 Solar Energy Utilization 6.8 Geothermal Heat Pump and Energy Recovery Applications

6.9 6.10

Thermal Storage Fuels and Combustion

SECTION 7.0—BUILDING PERFORMANCE 7.1 Integrated Building Design 7.2 HVAC&R Construction & Design Build Technologies 7.3 Operation and Maintenance Management 7.4 Exergy Analysis for Sustainable Buildings (EXER) 7.5 Smart Building Systems 7.6 Building Energy Performance 7.7 Testing and Balancing 7.8 Owning and Operating Costs 7.9 Building Commissioning SECTION 8.0—AIR-CONDITIONING AND REFRIGERATION SYSTEM COMPONENTS 8.1 Positive Displacement Compressors 8.2 Centrifugal Machines 8.3 Absorption and Heat Operated Machines 8.4 Air-to-Refrigerant Heat Transfer Equipment 8.5 Liquid-to-Refrigerant Heat Exchangers 8.6 Cooling Towers and Evaporative Condensers 8.7 Variable Refrigerant Flow (VRF) 8.8 Refrigerant System Controls and Accessories 8.9 Residential Refrigerators and Food Freezers 8.10 Mechanical Dehumidification Equipment and Heat Pipes 8.11 Unitary and Room Air Conditioners and Heat Pumps 8.12 Desiccant Dehumidification Equipment and Components SECTION 9.0—BUILDING APPLICATIONS 9.1 Large Building Air-Conditioning Systems 9.2 Industrial Air Conditioning 9.3 Transportation Air Conditioning 9.4 Justice Facilities 9.6 Healthcare Facilities 9.7 Educational Facilities 9.8 Large Building Air-Conditioning Applications 9.9 Mission Critical Facilities, Data Centers, Technology Spaces and Electronic Equipment 9.10 Laboratory Systems 9.11 Clean Spaces 9.12 Tall Buildings SECTION 10.0—REFRIGERATION SYSTEMS 10.1 Custom Engineered Refrigeration Systems 10.2 Automatic Icemaking Plants and Skating Rinks 10.3 Refrigerant Piping, Controls, and Accessories 10.5 Refrigerated Processing and Storage 10.6 Transport Refrigeration 10.7 Commercial Food and Beverage Refrigeration Equipment 10.8 Refrigeration Load Calculations SECTION MTG—MULTIDISCIPLINARY TASK GROUPS MTG.ASEC Avoided Sources Energy Consumption Due to Waste Heat Recovery and Heat Pump Technologies MTG.BD Building Dampness MTG.BIM Building Information Modeling MTG.ET Energy Targets MTG.HCDG Hot Climate Design Guide MTG.IAST Impact of ASHRAE Standards and Technology on Energy Savings/Performance MTG.ISPAQE Indoor Swimming Pool Air Quality and Evaporation MTG.LowGWP Lower Global Warming Potential Alternative Refrigerants MTG.O&MEE Operations and Maintenance Activities That Impact Energy Efficiency MTG.OBB Occupant Behavior in Buildings

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ASHRAE Research: Improving the Quality of Life ASHRAE is the world’s foremost technical society in the fields of heating, ventilation, air conditioning, and refrigeration. Its members worldwide are individuals who share ideas, identify needs, support research, and write the industry’s standards for testing and practice. The result is that engineers are better able to keep indoor environments safe and productive while protecting and preserving the outdoors for generations to come. One of the ways that ASHRAE supports its members’ and industry’s need for information is through ASHRAE Research. Thousands of individuals and companies support ASHRAE Research annually, enabling ASHRAE to report new data about material

properties and building physics and to promote the application of innovative technologies. Chapters in the ASHRAE Handbook are updated through the experience of members of ASHRAE Technical Committees and through results of ASHRAE Research reported at ASHRAE conferences and published in ASHRAE special publications, ASHRAE Transactions, and ASHRAE’s journal of archival research, Science and Technology for the Built Environment. For information about ASHRAE Research or to become a member, contact ASHRAE, 1791 Tullie Circle N.E., Atlanta, GA 30329; telephone: 404-636-8400; www.ashrae.org.

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Preface The 2020 ASHRAE Handbook—HVAC Systems and Equipment฀ discusses various systems and the equipment (components or฀ assemblies) they comprise, and describes features and differences.฀ This information helps system designers and operators in selecting฀ and using equipment. ASHRAE Technical Committees in each subject area have reviewed all chapters and revised them as needed for฀ current technology and practice. Some of the volume’s revisions and additions are as follows: • Chapter 7, Combined Heat and Power Systems, has a new section on economic evaluation and includes an update on EU Directive 2004/8/EC. • Chapter 9, Applied Heat Pump and Heat Recovery Systems, has new content on waste heat recovery, district applications, and industrial process heat pumps. • Chapter 12, District Heating and Cooling, has new content from ASHRAE research project RP-1267 (the new District Heating Guide and District Cooling Guide). • Chapter 18, Variable Refrigerant Flow, has new sections on modeling and system commissioning, and an updated system design example. • Chapter 19, Duct Construction, has extensive revisions on system leakage and air dispersion systems. • Chapter 20, Room Air Distribution Equipment, has updates for current technology, with new information on specialized components and air curtains. • Chapter 21, Fans, has new sections on series fan operation and field performance testing plus added content on fan and motor efficiency grades and parallel multiple-fan operation. • Chapter 24, Desiccant Dehumidification and Pressure-Drying Equipment, has expanded content on applications, air filters, and liquid strainers, plus recommendations from ASHRAE research project RP-1339 on rating equipment at altitude. • Chapter 25, Mechanical Dehumidifiers and Related Components, has new content on psychrometrics, outdoor air, controls, and industrial dehumidifiers. • Chapter 26, Air-to-Air Energy Recovery Equipment, has new information on heat pipes and desiccant and heat wheel systems. • Chapter 28, Unit Ventilators, Unit Heaters, and Makeup Air Units, has revisions on standards, controls, and fan selection for makeup air units. • Chapter 29, Air Cleaners for Particulate Contaminants, has updates on standards and performance testing. • Chapter 31, Automatic Fuel-Burning Systems, has added content on pneumatically and electronically linked gas/air ratio burner systems.

• Chapter 33, Furnaces, has updates for current technology and efficiency requirements. • Chapter 37, Solar Energy Equipment, has new data on worldwide solar technology use, plus an expanded section on photovoltaic equipment. • Chapter 38, Compressors, has revisions on general theory; screw and scroll compressors; and bearings, including oil-free technologies. • Chapter 44, Centrifugal Pumps, has new content on vertical, inline, split-coupled pumps; hydronic system pump selection; and differential pressure control. • Chapter 45, Motors, Motor Controls, and Variable-Frequency Drives, has new content on standards, bearing currents, and permanent-magnet motors. • Chapter 47, Valves, has new content on control valve sizing; electronic actuators; and ball, butterfly, flow-limiting, and pressureindependent control valves. • Chapter 49, Unitary Air Conditioners and Heat Pumps, has a new map of U.S. regional appliance efficiency standards. • Chapter 50, Room Air Conditioners and Packaged Terminal Air Conditioners, has updates for efficiency standards. • Chapter 52, Dedicated Outdoor Air Systems, is a completely new chapter focusing on DOAS equipment. This volume is published, as a bound print volume and in electronic format as a PDF download and online, in two editions: one using inch-pound (I-P) units of measurement, the other using the International System of Units (SI). Corrections to the 2017, 2018, and 2019 Handbook volumes can be found on the ASHRAE website at www.ashrae.org and in the Additions and Corrections section of this volume. Corrections for this volume will be listed in subsequent volumes and on the ASHRAE website. Reader comments are enthusiastically invited. To suggest improvements for a chapter, please comment using the form on the ASHRAE website or, using the cutout page(s) at the end of this volume’s index, write to Handbook Editor, ASHRAE, 1791 Tullie Circle N.E., Atlanta, GA 30329, or fax 678-539-2187, or e-mail [email protected].

Heather E. Kennedy Editor

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Related Commercial Resources CHAPTER 1

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HVAC SYSTEM ANALYSIS AND SELECTION Selecting a System ...................................................................... 1.1 HVAC Systems and Equipment .................................................. 1.4 Space Requirements ................................................................... 1.6 Air Distribution .......................................................................... 1.8 Pipe Distribution ........................................................................ 1.9

Security and Environmental Health and Safety ......................... 1.9 Automatic Controls and Building Management Systems .............................................................. 1.9 Maintenance Management ....................................................... 1.10 Building System Commissioning .............................................. 1.10

A

engineered options. Occupant comfort (as defined by ASHRAE Standard 55), process heating, space heating, cooling, and ventilation criteria must be considered when selecting the optimum system(s), as well as the following:

N HVAC system maintains desired environmental conditions in a space. In almost every application, many options are available to the design engineer to satisfy a client’s building program and design intent. In the analysis, selection, and implementation of these options, the design engineer should consider the criteria defined here, as well as project-specific parameters to achieve the functional requirements associated with the project design intent. In addition to the design, equipment, and system aspects of the proposed design, the design engineer should consider sustainability as it pertains to responsible energy and environmental design, as well as constructability of the design. HVAC systems are categorized by the method used to produce, deliver, and control heating, ventilating, and air conditioning in the conditioned area. This chapter addresses procedures for selecting an appropriate system for a given application while taking into account pertinent issues associated with designing, building, commissioning, operating, and maintaining the system. It also addresses specific owner requirements and constraints associated with selecting the optimum HVAC system for the application. Chapters 2 to 18 describe specific approaches and systems along with their attributes, based on their heating and cooling medium, the commonly used variations, constructability, commissioning, operation, and maintenance. This chapter is intended as a guide for the design engineer, builder, facility manager, and student needing to know or reference the analysis and selection process that ultimately leads to recommending the optimum system for the job. The approach applies to HVAC equipment conversions, building system upgrades, system retrofits, building renovations and expansion, and new construction for any building: small, medium, large, below grade, at grade, lowrise, and high-rise. This system analysis and selection process (Figure 1) approach helps guide the design engineer in drafting a report that recommends the best system(s) for any building program, regardless of facility type. This chapter’s analysis examines objective, subjective, short-term, and long-term goals along with constraints. Figure 1 also highlights five project delivery methods: performance contracting, design-bid-build, design-build, integrated building design, and construction management.

1.

SELECTING A SYSTEM

The design engineer is responsible for considering various systems and equipment and recommending one or more system options that will meet the project goals and perform per the design intent. It is imperative that the design engineer and owner collaborate to identify and prioritize criteria associated with the design goals. In addition, if the project includes preconstruction services, the designer, owner, and operator should consult with a builder to take advantage of a constructability analysis as well as the consideration of valueThe preparation of this chapter is assigned to TC 9.1, Large Building AirConditioning Systems.

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• • • • • • • • • • • • • • • • • • • • • • • • • • • • • • •

Temperature Humidity Air motion Air/water velocity Water quality and/or reuse Outdoor air quality or purity Indoor air purity or quality Air changes per hour Acoustics and vibration Local climate Mold and mildew prevention Capacities (existing, proposed, and future expansion) Redundancy Spatial requirements (present and future) Environmental health and safety design Security First cost Return-on-investment cost Energy consumption costs Operator labor costs Maintenance costs Serviceability Reliability Flexibility Controllability Replaceability Life-cycle analysis Sustainability of design Seismic protection Filtration and filtration effects Changing codes and standards

Because these factors are interrelated, the owner, design engineer, operator, and builder must consider how these criteria affect each other. The relative importance of factors such as these varies with different owners, and can often change from one project to another for the same owner. For example, typical owner concerns include first cost compared to operating cost, extent and frequency of maintenance and whether that maintenance requires entering the occupied space, expected frequency of system failure, effect of failure, and time required to correct the failure. Each concern has a different priority, depending on the owner’s goals.

Additional Goals In addition to the primary goal of providing the desired environment, the design engineer should be aware of and account for other goals the owner may require. These goals may include the following:

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1.2 • • • • • • • • • •

2020 ASHRAE Handbook—HVAC Systems and Equipment (SI)

Seasonal start-up date Occupant move-in date Operator training Supporting a process, such as operation of computer equipment Promoting a germ-free environment Increasing marketability of rental spaces Increasing net rental income Increasing property saleability Public image of the property Certification of energy use (LEED®, ENERGYSTAR®, etc.)

The owner can only make appropriate value judgments if the design engineer provides complete information on the advantages and disadvantages of each system option. Just as the owner does not usually know the relative advantages and disadvantages of different HVAC systems, the design engineer rarely knows all the owner’s financial and functional goals. Hence, the owner must be proactive and involved in system selection in the conceptual phase of the job. The same can be said for operator participation so that the final design intent is sustainable and can continue to fulfill design intent. All owners should request and/or require the design team to provide building and HVAC security (as defined in Chapter 61 of the 2019 ASHRAE Handbook—HVAC Applications). This security, in addition to environmental health and safety criteria, should address outside-the-building influences (e.g., smoke, toxic fumes, flood, etc.) as well as influences inside the building. System selection pertaining to HVAC security may require another design team member, a security consultant, and/or the owner’s own security group.

Equipment and System Constraints

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Once the goal criteria and additional goal options are verified, equipment and system constraints must be assessed and documented as part of the system analysis and selection recommendation report. These constraints may include the following: • Performance limitations (e.g., temperature, humidity, space pressure, etc.) • Code updates and/or new codes • Available capacity including equipment size, as well as ductwork and/or pipe size • Available space and access in and/or out with new or old equipment • Available utility source • Energy budget (e.g., conceptual Btu/h·ft2 per year{W/m2 per year}), code-driven or targeted • Equipment efficiency • Operator knowledge and capabilities • Existing building occupants • Building architecture

Fig. 1 Process Flow Diagram (Courtesy RDK Engineers)

Few projects allow detailed quantitative evaluation of all alternatives. Common sense, historical data, and subjective experience can be used to narrow choices to one or two potential systems. Heating and air-conditioning loads often contribute to constraints, narrowing the choice to systems that fit in available space and are compatible with building architecture. Chapters 17 and 18 of the 2017 ASHRAE Handbook—Fundamentals describe methods to determine the size and characteristics of heating and air-conditioning loads. By establishing the capacity requirement, equipment size can be determined, and the choice may be narrowed to those systems that work well on projects within the required size range. Loads vary over time based on occupied and unoccupied periods, changes in weather, type of occupancy, activities, internal loads, and solar exposure. Each space with a different use and/or exposure may require its own control zone to maintain space comfort. Some areas with special requirements (e.g., dehumidification requirements) may need individual systems. The extent of zoning, degree of control required in each zone, and space required for individual zones also narrow system choices. No matter how efficiently a particular system operates or how economical it is to install, it can only be considered if it (1) maintains the desired building space environment within an acceptable tolerance under expected conditions and occupant activities and (2) physically fits into, on, or adjacent to the building without causing objectionable occupancy conditions. Cooling, heating, and humidity control are often the basis of sizing HVAC components and subsystems, but ventilation requirements can also significantly impact system sizing. For example, if

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HVAC System Analysis and Selection large quantities of outdoor air are required for ventilation or to replace air exhausted from the building, the design engineer may only need to consider systems that transport and effectively condition those large outdoor air volumes, thus eliminating many of the other HVAC systems (e.g., unit ventilators, through-the-wall window units) from system consideration. The cooling, heating, humidification, and dehumidification delivery performance compromises may be minor for one application in a given climate, but may be unacceptable in another that has more stringent requirements. HVAC systems and associated distribution systems often occupy a significant amount of space. Major components may also require special support from the structure. The size and appearance of terminal devices (e.g., grilles, registers, diffusers, fan-coil units, radiant panels, chilled beams) affect architectural design because they are visible in the occupied space, which constrains the design. Sustainable energy consumption can be compromised and longterm project success lost if building operators are not trained to efficiently and effectively operate and maintain the building systems. For projects in which the design engineer used some form of energy software simulation, the resultant data should be passed on to the building owner so that goals and expectations can be measured and benchmarked against actual system performance. Even though the HVAC designer’s work may be complete after system commissioning and turnover to the owner, continuous acceptable performance is expected and will only occur if the owner understands the background and systems selected and their anticipated performance. See ASHRAE Guideline 0 and to ASHRAE’s Building Energy Quotient (bEQ) program (www.buildingeq.com). System operability should be considered in system selection. Constructing a highly sophisticated, complex HVAC system in a building where maintenance personnel lack the required skills for maintenance can be a recipe for disaster at worst, and at best may require the use of costly outside maintenance and service contractors to achieve successful system operation. The design engineer should closely coordinate the system constraints with the rest of the design team, as well as the owner, to overcome design obstacles associated with the HVAC systems under consideration for the project.

Constructability Constraints The design engineer must take into account HVAC system constructability issues before the project reaches the construction document phase. Some of these constraints may significantly affect the success of the design and cannot be overlooked in the conceptual and design development phases. Some issues and concerns associated with constructability are • Existing conditions (e.g., floor load, access into and through a building) • Rigging equipment into and out of a building to a designated area • Demolition and impact on adjacent space and existing systems in operation • Maintaining existing building occupancy, use of the building, and system operation • Ability to phase HVAC system installation • Temporary HVAC • Equipment availability (e.g., delivery lead time) • Construction schedule • Construction budget It can be very advantageous to involve a builder into the early stages of an HVAC design to bring their prospective and experience to the project system analysis and selection process. Whether a construction manager, design-builder, or integrated project delivery leader, each brings years of experience to the topic of construct-ability that can influence the final selection of HVAC system(s).

1.3 Construction budget constraints can also influence the choice of HVAC systems. Based on historical data, some systems may not be economically viable within the budget limitations of an owner’s building program. In addition, annual maintenance and operating budget (utilities, labor, and materials) can be limiting although more often than not, annual operating budgets do not get included in the total construction “soft costs” (e.g., furnishing, general conditions). This can be particularly important for building owners who will retain the building for a substantial number of years.The building’s estimator can assist in the overall project cost by introducing value-engineered solutions that may involve (1) cost-driven performance, which may provide a better solution for lower first cost; (2) a more sustainable solution over the life of the equipment; or (3) best value based on a reasonable return on investment.

Narrowing the Choices The following chapters in this volume present information to help the design engineer narrow the choices of HVAC systems: • Chapter 2 focuses on a distributed approach to HVAC. • Chapter 3 provides guidance for large equipment centrally located in or adjacent to a building. • Chapters 4 to 18 address the numerous types of air-conditioning and heating systems. Each chapter summarizes positive and negative features of various systems. Comparing the criteria, other factors and constraints, and their relative importance usually identifies one or two systems that best satisfy project goals. In making choices, notes should be kept on all systems considered and the reasons for eliminating those that are unacceptable. Each selection may require combining a primary system with a secondary (or distribution) system. The primary system converts energy derived from fuel or electricity to produce a heating and/or cooling medium. The secondary system delivers heating, ventilation, cooling, humidification, and/or dehumidification to the occupied space. The systems are independent to a great extent, so several secondary systems may work with a particular primary system. In some cases, however, only one secondary system may be suitable for a particular primary system. Once subjective analysis has identified one or more HVAC systems (sometimes only one choice remains), detailed quantitative evaluations must be made. All systems considered should provide satisfactory performance to meet the owner’s essential goals. The design engineer should provide the owner with specific data on each system to make an informed choice. Consult the following chapters to help narrow the choices: • Chapter 10 of the 2017 ASHRAE Handbook—Fundamentals covers physiological principles, comfort, and health. • Chapter 19 of the 2017 ASHRAE Handbook—Fundamentals covers methods for estimating annual energy costs. • Chapters 37 to 44 of the 2019 ASHRAE Handbook—HVAC Applications address important topics pertinent to the long-term success of building programs. Other documents and guidelines that should be consulted are ASHRAE standards and guidelines; local, state, and federal guidelines; and special agency requirements (e.g., U.S. General Services Administration [GSA], Food and Drug Administration [FDA], Joint Commission on Accreditation of Healthcare Organizations [JCAHO], Facility Guidelines Institute [FGI], Leadership in Energy and Environmental Design [LEED®]). Additional sources of detailed information are listed in the Bibliography and/or available in the ASHRAE Bookstore (www.ashrae.org/bookstore).

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1.4

2020 ASHRAE Handbook—HVAC Systems and Equipment (SI) Table 1

Sample HVAC System Analysis and Selection Matrix (0 to 10 Score)

Goal: Furnish and install an HVAC system that provides moderate space temperature control with minimum humidity control at an operating budget of 220 kW/m2 per year Categories

System #1 System #2 System #3

Remarks

1. Criteria for selection: • 25.6°C space temperature with 1.7 K control during occupied cycle, with 40% rh and  rh control during cooling. • 20°C space temperature with , with 20% rh and 5% rh control during heating season. • First cost • Equipment life cycle 2. Important factors: • First-class office space stature • Individual tenant utility metering 3. Other goals: • Engineered smoke control system • ASHRAE Standard 62.1 ventilation rates • Direct digital control building automation

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4. System constraints: • No equipment on first floor • No equipment on ground adjacent to building 5. Energy use as predicted by use of an industry-acceptable computerized energy model 6. Other constraints: • No perimeter finned-tube radiation or other type of in-room equipment TOTAL SCORE

Selection Report As the last step, the design engineer should prepare a system analysis and selection report in a format that incorporates the following: • • • • • • • • • • • • • •

Introduction Building program primary HVAC goals Criteria for selection Important factors, including advantages and disadvantages System integration with other building systems Other HVAC goals Security and environmental health and safety criteria Building envelope Project timeline (design, equipment delivery, commissioning, training, and construction) Basis of design HVAC system analysis and selection matrix System narratives Budget costs (first cost, operating cost, and energy cost) Final recommendation(s)

A brief outline of each of the final selections should be provided along with overall budget and/or cost per square metre. In addition, HVAC systems deemed inappropriate should be noted as having been considered but not found applicable to meet the owner’s primary HVAC goal. The report should include an HVAC system selection matrix that identifies the one or two suggested HVAC system selections (primary and secondary, when applicable), system constraints, and other constraints and considerations. In completing this matrix assessment, the design engineer should identify the owner’s input to the analysis. This input can also be applied as weighted multipliers, because not all criteria carry the same weighted value. Many grading methods are available to complete an analytical matrix analysis. Probably the simplest is to rate each item excellent, very good, good, fair, or poor. A numerical rating system such

as 0 to 10, with 10 equal to excellent and 0 equal to poor or not applicable, can provide a quantitative result. The HVAC system with the highest numerical value then becomes the recommended system to accomplish the goal. The system selection report should include a summary followed by a more detailed account of the HVAC system analysis and system selection. This summary should highlight key points and findings that led to the recommendation(s). The analysis should refer to the system selection matrix (such as in Table 1) and the reasons for scoring. With each HVAC system considered, the design engineer should note the criteria associated with each selection. Issues such as close-tolerance temperature and humidity control may eliminate some HVAC systems from consideration. System constraints, noted with each analysis, may eliminate potential HVAC systems. Advantages and disadvantages of each system should be noted with the scoring from the HVAC system selection matrix. This process should reduce HVAC selection to one or two optimum choices for presentation to the owner. Examples of similar installations for other owners can be included with this report to support the final recommendation. Identifying a third party for an endorsement allows the owner to inquire about the success of other HVAC installations.

2.

HVAC SYSTEMS AND EQUIPMENT

Many built, expanded, and/or renovated buildings may be ideally suited for decentralized HVAC systems, with equipment located in, throughout, adjacent to, or on top of the building. The alternative to this decentralized approach is to use primary equipment located in a central plant (either inside or outside the building) with water and/or air required for HVAC needs distributed from this plant.

Decentralized System Characteristics The various types of decentralized systems are described in Chapter 2. The common element is that the required cooling is

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HVAC System Analysis and Selection

1.5

distributed throughout the building, generally with direct-expansion cooling, electric, or gas-fired heating of air systems, and on occasion, oil-fired warm-air heating units. Temperature, Humidity, and Space Pressure Requirements. A decentralized system may be able to fulfill any or all of these design parameters, but typically not as efficiently or as accurately as a central system. Capacity Requirements. A decentralized system usually requires each piece of equipment to be sized for zone peak capacity, unless the systems are variable-volume. Depending on equipment type and location, decentralized systems do not benefit as much from equipment sizing diversity as centralized systems do. Redundancy. A decentralized system may not have the benefit of back-up or standby equipment. This limitation may need review. Facility Management. A decentralized system can allow the building manager to maximize performance using good business/ facility management techniques in operating and maintaining the HVAC equipment and systems. With some decentralized equipment (e.g., rooftop HVAC units), operation and maintenance may be performed annually via a service contract with a mechanical contracting company. Spatial Requirements. A decentralized system may or may not require equipment rooms. Because of space restrictions imposed on the design engineer or architect, equipment may be located on the roof and/or the ground adjacent to the building. Depending on system components, additional space may be required in the building for chillers and boilers. Likewise, a decentralized system may or may not require duct and pipe shafts throughout the building. First Cost. A decentralized system probably has the best firstcost benefit. This feature can be enhanced by phasing in the purchase of decentralized equipment as needed (i.e., buying equipment as the building is being leased/occupied). Operating Cost. A decentralized system can save operating cost by strategically starting and stopping multiple pieces of equipment. When comparing energy consumption based on peak energy draw, decentralized equipment may not be as attractive as larger, more energy-efficient centralized equipment. Maintenance Cost. A decentralized system can save maintenance cost when equipment is conveniently located and equipment size and associated components (e.g., filters) are standardized. When equipment is located outdoors, maintenance may be difficult during bad weather. Reliability. A decentralized system usually has reliable equipment, although the estimated equipment service life may be less than that of centralized equipment. Decentralized system equipment may, however, require maintenance in the occupied space. Flexibility. A decentralized system may be very flexible because it may be placed in numerous locations. Level of Control. Decentralized systems often use direct refrigerant expansion (DX) for cooling, and on/off or staged heat. This step control results in greater variation in space temperature and humidity, where close control is not desired or necessary. As a caution, oversizing DX or stepped cooling can allow high indoor humidity levels and mold or mildew problems. Noise and Vibration. Decentralized systems often locate noisy machinery close to building occupants. Constructability. Decentralized systems frequently consist of multiple and similar-in-size equipment that makes standardization a construction feature, as well as purchasing units in large quantities.

cogeneration central equipment/systems, or condenser water systems used with water-source heat pumps. Among the largest centralized systems are HVAC plants serving groups of large buildings. These plants improve diversity and generally operate more efficiently, and with lower maintenance costs, than individual central plants. Economic considerations of larger centralized systems require extensive analysis. The utility analysis may consider multiple fuels and may also include gas and steam turbinedriven equipment. Multiple types of primary equipment using multiple fuels and types of HVAC-generating equipment (e.g., centrifugal and absorption chillers) may be combined in one plant. Temperature, Humidity, and Space Pressure Requirements. A central system may be able to fulfill any or all of these design parameters, typically with greater precision and efficiency than a decentralized system. Capacity Requirements. A central system usually allows the design engineer to consider HVAC diversity factors that significantly reduce installed equipment capacity. As a result, this offers some attractive first-cost and operating-cost benefits. Redundancy. A central system can accommodate standby equipment that decentralized configurations may have trouble accommodating because of cost. Facility Management. A central system usually allows the building manager to maximize performance using good business/ facility management techniques in operating and maintaining the HVAC equipment and systems. Operation and maintenance are usually handled by on-site staff, whether employees of the building owner or outsourced building management placed on site. Spatial Requirements. The equipment room for a central system is normally located outside the conditioned area: in a basement, penthouse, service area, or adjacent to or remote from the building. A disadvantage of this approach may be the additional cost to furnish and install secondary equipment for the air and/or water distribution. Other considerations include access requirements and physical constraints that exist throughout the building to the installation of the secondary distribution network of ducts and/ or pipes and for equipment replacement. First Cost. Even with HVAC diversity, a central system may not be less costly than decentralized HVAC systems. Historically, central system equipment has a longer equipment service life to compensate for this shortcoming. Thus, a life-cycle cost analysis is very important when evaluating central versus decentralized systems. Operating Cost. A central system usually has the advantage of larger, more energy-efficient primary equipment compared to decentralized system equipment. In addition, the availability of multiple pieces of HVAC equipment allows staging of this equipment operation to match building loads while maximizing operational efficiency. Maintenance Cost. The equipment room for a central system provides the benefit of being able to maintain HVAC equipment away from occupants in an appropriate service work environment. Access to occupant workspace is not required, thus eliminating disruption to the space environment, product, or process. Because of the typically larger capacity of central equipment, there are usually fewer pieces of HVAC equipment to service. Reliability. Centralized system equipment generally has a longer service life. Flexibility. Flexibility can be a benefit when selecting equipment that provides an alternative or back-up source of HVAC.

Centralized System Characteristics

Air Distribution Systems

These systems are characterized by central refrigeration equipment/systems with chilled-water distribution and central boiler equipment/system with hot-water distribution. Other central systems include low-, medium-, and high-pressure steam plants and

The various air distribution systems, including dedicated outdoor air systems (DOAS), are detailed in Chapter 4. Any of the preceding system types discussed within this chapter can be used in conjunction with DOAS.

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Level of Control. Centralized air systems generally use chilled water for cooling, and steam or hot water for heating. This usually allows for close control of space temperature and humidity where desired or necessary. Sound and Vibration. Centralized air system placement should consider the associated equipment noise and vibration and locate machinery sufficiently remote from building occupants or noisesensitive processes. Constructability. Centralized air systems usually require more coordinated installation than decentralized systems. When reusing an existing central air fan(s) and associated ductwork, it is important to verify construction class remains the same or is reduced (e.g., increasing supply air fan capacity in a project renovation may increase the fan construction from a Class 1 to a Class 2 construction). The same can be said when increasing airflow within an existing duct distribution system (e.g., ductwork may require reinforcement and/or duct seams sealed airtight due to the increase in duct static and velocity pressures).

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Primary Equipment The type of decentralized and centralized equipment selected for buildings depends on a well-organized HVAC analysis and selection report. The choice of primary equipment and components depends on factors presented in the selection report (see the section on Selecting a System). Primary HVAC equipment includes refrigeration equipment; heating equipment; and air, water, and steam delivery equipment. Many HVAC designs recover internal heat from lights, people, and equipment to reduce the size of the heating plant. In buildings with core areas that require cooling while perimeter areas require heating, one of several heat reclaim systems can heat the perimeter to save energy. Sustainable design is also important when considering recovery and reuse of materials and energy. Chapter 9 describes heat pumps and some heat recovery arrangements, Chapter 37 describes solar energy equipment, and Chapter 26 introduces air-to-air energy recovery. In the 2019 ASHRAE Handbook— HVAC Applications, Chapter 37 covers energy management and Chapter 42 covers building energy monitoring. Chapter 35 of the 2017 ASHRAE Handbook—Fundamentals provides information on sustainable design, and Chapters 10 and 16 of that volume address indoor air quality, ventilation, and infiltration in the context of assessing optimum HVAC application options. The search for energy savings has extended to cogeneration or total energy (combined heat and power [CHP]) systems, in which on-site power generation is added to the HVAC project. The economic viability of this function is determined by the difference between gas and electric rates and by the ratio of electricity to heating demands for the project. In these systems, waste heat from generators can be transferred to the HVAC systems (e.g., to drive turbines of centrifugal compressors, serve an absorption chiller, provide heating or process steam). Chapter 7 covers cogeneration or total energy systems. Alternative fuel sources, such as waste heat boilers, are now included in fuel evaluation and selection for HVAC applications. Thermal storage is another cost-saving concept, which provides the possibility of off-peak generation of chilled water or ice. Thermal storage can also be used for storing hot water for heating. Many electric utilities impose severe charges for peak summer power use or offer incentives for off-peak use. Storage capacity installed to level the summer load may also be available for use in winter, thus making heat reclaim a viable option. Chapter 50 has more information on thermal storage. With ice storage, colder supply air can be provided than that available from a conventional chilled-water system. This colder air allows use of smaller fans and ducts, which reduces first cost and (in some locations) operating cost. Additional pipe and duct

insulation is often required, however, contributing to a higher first cost. These life-cycle savings can offset the first cost for storage provisions and the energy cost required to make ice. Similarly, thermal storage of hot water can be used for heating.

Refrigeration Equipment Chapters 2 and 3 summarize the primary types of refrigeration equipment for HVAC systems. When chilled water is supplied from a central plant, as on university campuses and in downtown areas of large cities or in large buildings, the utility service provider should be contacted during system analysis and selection to determine availability, cost, and the specific requirements of the service.

Heating Equipment Steam boilers and heating-water boilers are the primary means of heating a space using a centralized system, as well as some decentralized systems. These boilers may be (1) used both for comfort and process heating; (2) manufactured to produce high- or low-pressure steam; and (3) fired with coal, oil, electricity, gas, and even some waste materials. Low-pressure boilers are rated for a working pressure of 100 kPa for steam, and 1100 kPa for water, with a temperature limit of 120°C. Packaged boilers, with all components and controls assembled at the factory as a unit, are available. Electrode or resistance electric boilers that generate either steam or hot water are also available. Chapter 32 has further information on boilers, and Chapter 27 details air-heating coils. Where steam, hot water, or chilled water is supplied from a central plant, as on university campuses and in downtown areas of large cities, the utility provider should be contacted during project system analysis and selection to determine availability, cost, and specific requirements of the service. When primary heating equipment is selected, the fuels considered must ensure maximum efficiency. Chapter 31 discusses design, selection, and operation of the burners for different types of primary heating equipment. Chapter 28 of the 2017 ASHRAE Handbook—Fundamentals describes types of fuel, fuel properties, and proper combustion factors.

Air Delivery Equipment Primary air delivery equipment for HVAC systems is classified as packaged, manufactured and custom-manufactured, or fieldfabricated (built-up). Most air delivery equipment for large systems uses centrifugal or axial fans; however, plug or plenum fans are often used. Centrifugal fans are frequently used in packaged and manufactured HVAC equipment. One system rising in popularity is a fan array, which uses multiple plug fans on a common plenum wall, thus reducing unit size. Axial fans are more often part of a custom unit or a field-fabricated unit. Both types of fans can be used as industrial process and high-pressure blowers. Chapter 21 describes fans, and Chapters 19 and 20 provide information about air delivery components.

3.

SPACE REQUIREMENTS

In the initial phase of building design, the design engineer seldom has sufficient information to render the optimum HVAC design for the project, and its space requirements are often based on percentage of total area or other experiential rules of thumb. The final design is usually a compromise between the engineer’s recommendations and the architectural considerations that can be accommodated in the building. An integrated project design (IPD) approach, as recommended by the American Institute of Architects (AIA), can address these problems early in the design process; also see Chapter 60 of the 2019 ASHRAE Handbook—HVAC Applications. Other times, the

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HVAC System Analysis and Selection building owner may prefer either a centralized or decentralized system and dictate final design and space requirements. This section discusses some of these requirements.

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Equipment Rooms Total mechanical and electrical space requirements range between 4 and 9% of gross building area, with most buildings in the 6 to 9% range. These ranges include space for HVAC, electrical, plumbing, and fire protection equipment and may also include vertical shaft space for mechanical and electrical distribution through the building. Most equipment rooms should be centrally located to (1) minimize long duct, pipe, and conduit runs and sizes; (2) simplify shaft layouts; and (3) centralize maintenance and operation. With shorter duct and pipe runs, a central location could also reduce pump and fan motor power requirements, which reduces building operating costs. But, for many reasons, not all mechanical and electrical equipment rooms can be centrally located in the building. In any case, equipment should be kept together whenever possible to minimize space requirements, centralize maintenance and operation, and simplify the electrical system. Equipment rooms generally require clear ceiling height ranging from 10 to 18 ft {3 to 5 m}, depending on equipment sizes and the complexity of air and/or water distribution. The main electrical transformer and switchgear rooms should be located as close to the incoming electrical service as practical. If there is an emergency generator, it should be located considering (1) proximity to emergency electrical loads and sources of combustion and cooling air and fuel, (2) ease of properly venting exhaust gases to the outdoors, and (3) provisions for noise control. Noise control is generally a minor issue because equipment only runs during emergencies. Primary Equipment Rooms. The heating equipment room houses the boiler(s) and possibly a boiler feed unit (for steam boilers), chemical treatment equipment, pumps, heat exchangers, pressure-reducing equipment, air compressors, and miscellaneous ancillary equipment. The refrigeration equipment room houses the chiller(s) and possibly chilled-water and condenser water pumps, heat exchangers, air-conditioning equipment, air compressors, and miscellaneous ancillary equipment, and in some cases, direct-expansion compressors. Design of these rooms needs to consider (1) equipment size and weight, (2) installation, maintenance, and replacement, (3) applicable regulations relative to combustion air and ventilation air, and (4) noise and vibration transmission to adjacent spaces. Consult ASHRAE Standard 15 for refrigeration equipment room safety requirements. Note that primary equipment rooms should not be located below ground or even at ground level if the project is to be located in a designated flood zone area, due to the potential of the equipment room being flooded. Many air-conditioned buildings require a cooling tower or other type of heat rejection equipment. If the cooling tower or watercooled condenser is located at ground level, it should be at least 100 ft {30 m} away from the building to reduce tower noise in the building, and to keep discharge air and moisture carryover from fogging the building’s windows and discoloring the building facade or from contaminating outdoor air being introduced into the building. Cooling towers should be kept a similar distance from parking lots to avoid staining car finishes with atomized water treatment chemicals. Chapter 40 has further information on this equipment. It is often economical to locate the heating and/or refrigeration plant in the building, on an intermediate floor, in a roof penthouse, or on the roof when roof space is available; air-cooled refrigeration systems can be considered. Electrical service and structural costs are higher, but these may be offset by reduced costs for piping, pumps and pumping energy, and chimney requirements for fuel-

1.7 fired boilers. Also, initial cost of equipment in a tall building may be less for equipment located on a higher floor (e.g., boilers located in a roof penthouse of a high-rise building) because some operating pressures may be lower. Regulations applicable to both gas and fuel oil systems must be followed. Gas fuel may be more desirable than fuel oil because of the physical constraints on the required fuel oil storage tank, as well as specific environmental and safety concerns related to oil leaks. Also, the cost of an oil leak detection and prevention system may be substantial. Oil pumping presents added design and operating problems, depending on location of the oil tank relative to the oil burner. Energy recovery systems can reduce the size of the heating and/ or refrigeration plant. The possibility of failure or the need to take heat recovery equipment out of operation for service should be considered in the heating plant design, to ensure the ability to heat the building with the heating plant without heat recovery. Wellinsulated buildings and electric and gas utility rate structures may encourage the design engineer to consider energy conservation concepts such as limiting demand, ambient cooling, and thermal storage.

Fan Rooms Fan rooms house HVAC air delivery equipment and may include other miscellaneous equipment. The room must have space for removing the fan(s), shaft(s), coils, and filters. Installation, replacement, and maintenance of this equipment should be considered when locating and arranging the room. Fan rooms in a basement that has an airway for intake of outdoor air present a potential problem. Low air intakes are a security concern, because harmful substances could easily be introduced (see the section on Security). Debris, and in some climate zones snow, may fill the area, resulting in safety, health, and fan performance concerns. Parking areas close to the building’s outdoor air intake may compromise ventilation air quality. Fan rooms located at the perimeter wall can have intake and exhaust louvers at the location of the fan room, subject to coordination with architectural considerations. Interior fan rooms often require intake and exhaust shafts from the roof because of the difficulty (typically caused by limited ceiling space) in ducting intake and exhaust to the perimeter wall. Fan rooms on the second floor and above have easier access for outdoor and exhaust air. Depending on the fan room location, equipment replacement may be easier. The number of fan rooms required depends largely on the total floor area and whether the HVAC system is centralized or decentralized. Buildings with large floor areas may have multiple decentralized fan rooms on each or alternate floors. High-rise buildings may opt for decentralized fan rooms for each floor, or for more centralized service with one fan room serving the lower 10 to 20 floors, one serving the middle floors of the building, and one at the roof serving the top floors. Life safety is a very important factor in HVAC fan room location. Chapter 54 of the 2019 ASHRAE Handbook—HVAC Applications discusses fire and smoke control. State and local codes have additional fire and smoke detection and damper criteria. Building codes that require smoke control and building pressurization should be accounted for in equipment selection.

Horizontal Distribution Most decentralized and central systems rely on horizontal distribution. To accommodate this need, the design engineer needs to take into account the duct and/or pipe distribution criteria for installation in a ceiling space or below a raised floor space. Systems using water distribution usually require the least amount of ceiling or raised floor depth, whereas air distribution systems have the largest demand for horizontal distribution height. Steam systems need

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to accommodate pitch of steam pipe, end of main drip, and condensate return pipe pitch. Another consideration in the horizontal space cavity is accommodating the structural members, light fixtures, rain leaders, cable trays, etc., that can fill up this space.

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Vertical Shafts Buildings over two stories high usually require vertical shafts to consolidate mechanical (e.g., air distribution ducts and pipes), electrical, plumbing, and telecommunication distribution through the facility. Air distribution includes HVAC supply, return, and exhaust air ductwork. If a shaft is used as a return air plenum, close coordination with the architect is necessary to ensure that the shaft is airtight. If the shaft is used to convey outdoor air to decentralized systems, close coordination with the architect is also necessary to ensure that the shaft is constructed to meet mechanical code requirements and to accommodate the anticipated internal pressure. The temperature of air in the shaft must be considered when using the shaft enclosure to convey outdoor air. Pipe distribution includes heating, chilled, and condenser water, and steam supply and condensate return. It may also include fuel oil and gas pumping. Other distribution systems found in vertical shafts or located vertically in the building include electric conduits/bus ducts/closets, telephone cabling/closets, uninterruptible power supply (UPS), plumbing, fire protection piping, pneumatic tubes, and conveyers. Vertical shafts should not be adjacent to stairs, electrical closets, and elevators unless at least two sides are available to allow access to ducts, pipes, and conduits that enter and exit the shaft while allowing maximum headroom at the ceiling. In general, duct shafts with an aspect ratio of 2:1 to 4:1 are easier to develop than large square shafts. The rectangular shape also facilitates transition from the equipment in the fan rooms to the shafts. In addition, significant space is required for motorized and fire/smoke dampers when ducts leave shafts. In multistory buildings, a central vertical distribution system with minimal horizontal branch ducts is desirable. This arrangement (1) is usually less costly; (2) is easier to balance; (3) creates less conflict with pipes, beams, and lights; and (4) enables the architect to design lower floor-to-floor heights. These advantages also apply to vertical water and steam pipe distribution systems. The number of shafts is a function of building size and shape. In larger buildings, it is usually more economical in cost and space to have several small shafts rather than one large shaft. Separate HVAC supply, return, and exhaust air duct shafts may be desirable to reduce the number of duct crossovers. The same applies for steam supply and condensate return pipe shafts because the pipe must be pitched in the direction of flow. When future expansion is a consideration, a pre-agreed percentage of additional shaft space should be considered for inclusion. This, however, affects the building’s first cost because of the additional floor space that must be constructed. The need for access doors into shafts and gratings at various locations throughout the height of the shaft should also be considered.

Rooftop Equipment For buildings three stories or less, system analysis and selection frequently locates HVAC equipment on the roof or another outdoor location, where the equipment is exposed to the weather. Decentralized equipment and systems are sometimes more advantageous than centralized HVAC for smaller buildings, particularly those with multiple tenants with different HVAC needs. Selection of rooftop equipment is usually driven by first cost versus operating cost and/or maximum service life of the equipment. For buildings with larger floor plates, centralized equipment may be advantageous and should be considered.

Equipment Access Properly designed mechanical and electrical equipment rooms must allow for movement of large, heavy equipment in, out, and through the building. If a building is being renovated for another application, floor loading should be assessed to determine whether heavier equipment can be installed on the existing floor. Equipment replacement and maintenance can be very costly if access is not planned properly. Access to rooftop equipment should be by means of, at a minimum, a ship’s ladder and not by a vertical ladder, taking into account a technician carrying boxes of replacement filters. Use caution when accessing equipment on sloped roofs. Because systems vary greatly, it is difficult to estimate space requirements for refrigeration and boiler rooms without making block layouts of the system selected. Block layouts allow the design engineer to develop the most efficient arrangement of the equipment with adequate access and serviceability. Block layouts can also be used in preliminary discussions with the owner and architect. Only then can the design engineer verify the estimates and provide a workable and economical design.

4.

AIR DISTRIBUTION

Ductwork should deliver conditioned air to an area as directly, quietly, and economically as possible. Structural features of the building generally require some compromise and often limit the depth of space available for ducts. Chapter 10 discusses air distribution design for small heating and cooling systems. Chapters 20 and 21 of the 2017 ASHRAE Handbook—Fundamentals discuss space air distribution and duct design. The design engineer must coordinate duct layout with the structure as well as other mechanical, electrical, plumbing, and communication systems. In commercial projects, the design engineer is usually encouraged to reduce floor-to-floor dimensions: with the resulting decrease in available interstitial space, depth for ducts is a major design challenge. In institutional and information technology buildings, higher floor-to-floor heights are required because of the sophisticated, complex mechanical, electrical, and communication distribution systems. Exhaust systems, especially those serving fume exhaust, grease exhaust, dust and/or particle collection, and other process exhaust, require special design considerations. Capture velocity, duct material, and pertinent duct fittings and fabrication are a few of the design parameters necessary for this type of distribution system to function properly, efficiently, and per applicable codes. See Chapters 15 to 33 of the 2019 ASHRAE Handbook—HVAC Applications for additional information on industrial applications.

Air Terminal Units In some instances, such as in low-velocity, all-air systems, air may enter from the supply air ductwork directly into the conditioned space through a grille, register, or diffuser. In medium- and high-velocity air systems, an intermediate device normally controls air volume, reduces air pressure from the duct to the space, or both. Various types of air terminal units are available, including (1) a fan-powered terminal unit, which uses an integral fan to mix ceiling plenum air and primary air from the central or decentralized fan system rather than depending on induction (mixed air terminal unit) delivering to low-pressure ductwork and then to the space); (2) a variable-air-volume (VAV) terminal unit, which varies the amount of air delivered to the space (this air may be delivered to low-pressure ductwork and then to the space, or the terminal may contain an integral air diffuser); or (3) other in-room terminal type (see Chapter 5). Chapter 20 has more information about air terminal units. See Chapters 20 and 21 of the 2017 ASHRAE Handbook—Fundamentals for information on space air diffusion and duct design.

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HVAC System Analysis and Selection Duct Insulation

Pipe Insulation

In new construction and renovation upgrade projects, HVAC supply air ducts should be insulated and sealed in accordance with energy code requirements. ASHRAE Standard 90.1 and Chapter 23 of the 2017 ASHRAE Handbook—Fundamentals have more information about insulation and calculation methods.

In new construction and renovation projects, most HVAC piping must be insulated. ASHRAE Standard 90.1 and Chapter 23 of the 2017 ASHRAE Handbook—Fundamentals have information on insulation and calculation methods. In most applications, watersource heat pump loop piping and condenser water piping may not need to be insulated.

Ceiling and Floor Plenums

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Frequently, the space between the suspended ceiling and the floor slab above it is used as a return air plenum to reduce distribution ductwork. Check regulations before using this approach in new construction or renovation because most codes prohibit combustible material in a ceiling return air plenum. Ceiling plenums and raised floors can also be used for supply air displacement systems to minimize horizontal distribution, along with other features discussed in Chapter 4. Some ceiling plenum applications with lay-in panels do not work well where the stack effect of a high-rise building or highrise elevators creates a negative pressure. If the plenum leaks to the low-pressure area, tiles may lift and drop out when the outside door is opened and closed. The air temperature in a return air plenum directly below a roof deck is usually higher by 2 to 3 K during the air-conditioning season than in a ducted return. This can be an advantage to the occupied space below because heat gain to the space is reduced. Conversely, return air plenums directly below a roof deck have lower return air temperatures during the heating season than a ducted return and may require supplemental heat in the plenum. Using a raised floor for air distribution is popular for computer rooms and cleanrooms, and is now being used in other HVAC applications. Underfloor air displacement (UFAD) systems in office buildings use the raised floor as a supply air plenum, which could reduce overall first cost of construction and ongoing improvement costs for occupants. Making the underfloor supply air plenum as airtight as possible is critical to the success of this type of plenum. This UFAD system improves air circulation to the occupied area of the space. See Chapter 20 of the 2019 ASHRAE Handbook—HVAC Applications and Chapter 20 of the 2017 ASHRAE Handbook—Fundamentals for more information on displacement ventilation and underfloor air distribution.

5.

PIPE DISTRIBUTION

Piping should deliver refrigerant, heating water, chilled water, condenser water, fuel oil, gas, steam, and condensate drainage and return to and from HVAC equipment as directly, quietly, and economically as possible. Structural features of the building generally require mechanical and electrical coordination to accommodate P-traps, pipe pitch-draining of low points in the system, and venting of high points. When assessing application of pipe distribution to air distribution, the floor-to-floor height requirement can influence the pipe system: it requires less ceiling space to install pipe when compared to air ductwork. An alternative to horizontal piping is vertical pipe distribution, which may further reduce floor-tofloor height criteria. See Chapter 22 of the 2017 ASHRAE Handbook—Fundamentals for pipe sizing.

Pipe Systems HVAC piping systems can be divided into two parts: (1) piping in the central plant equipment room and (2) piping required to deliver refrigerant, heating water, chilled water, condenser water, fuel oil, gas, steam, and condensate drainage and return to and from decentralized HVAC and process equipment throughout the building. Chapters 11 to 15 discuss piping for various heating and cooling systems. Chapters 1 to 4 of the 2018 ASHRAE Handbook— Refrigeration discuss refrigerant system design and practices.

6.

SECURITY AND ENVIRONMENTAL HEALTH AND SAFETY

Since September 11, 2001, much attention has been given to protecting buildings against terrorist attacks via their HVAC systems. The first consideration should be to perform a risk assessment of the particular facility, which may be based on usage, size, population, and/or significance. The risk assessment is a subjective judgment by the building owner (and sometimes by a registered/ certified security professional) of whether the building is at low, medium, or high risk. An example of low-to-medium risk buildings may be suburban office buildings, shopping malls, hospitals, educational institutions, or major office buildings. High-risk buildings may include major government buildings. The level of protection designed into these buildings may include enhanced particulate filtration, gaseous-phase filtration, and various control schemes to allow purging of the facility using either the HVAC system or an independent dedicated system. Enhanced particulate filtration for air-handling systems to the level of MERV 14 to 16 filters not only tends to reduce circulation of dangerous substances (e.g., anthrax), but also provides better indoor air quality (IAQ). Gaseous-phase filtration can remove harmful substances such as sarin and other gaseous threats. For buildings of all risk levels, consideration should be taken to include proper location of outdoor air intakes to account for other security concerns, such as toxic fumes from a local accident (e.g., punctured tank from a nearby railroad car or tanker truck). The extent to which the HVAC system designer should use these measures could depend on the recommendations of an experienced security and environmental health and safety consultant or agency. In any building, consideration should be given to protecting outdoor air intakes against insertion of dangerous materials by locating the intakes on the roof or substantially above grade level. Separate systems for mailrooms, loading docks, and other similar spaces should be considered so that any dangerous material received cannot be spread throughout the building from these types of vulnerable spaces. Emergency ventilation systems for these types of spaces should be designed so that upon detection of suspicious material, toxic fumes, or chemical spill, these spaces can be quickly purged and maintained at a negative pressure. A more extensive discussion of this topic can be found in ASHRAE’s Guideline 29. Note, however, that fan energy required to move air through these high efficiency filters is very significant.

7.

AUTOMATIC CONTROLS AND BUILDING MANAGEMENT SYSTEMS

Basic HVAC system controls are primarily direct digital controls (DDC), but there are also electric, electronic, and pneumatic controls (found largely as part of existing building HVAC systems).Depending on the application, the design engineer may recommend a simple and basic system strategy as a cost-effective solution to an owner’s heating, ventilation, and cooling needs. Chapter 48 of the 2019 ASHRAE Handbook—HVAC Applications and Chapter 7 of the 2017 ASHRAE Handbook—Fundamentals discuss automatic control in more detail. The next level of HVAC system management is DDC wireless systems. Many DDC installations will use either pneumatic or

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electric control damper and valve actuators. This automatic control enhancement may include energy monitoring and energy management software. Controls may also be accessible by the building manager using a modem to a remote computer at an off-site location or via a smartphone or tablet computer with Internet connection to the control system. Building size has little to no effect on modern computerized controls: programmable controls can be furnished on the smallest HVAC equipment for the smallest projects. Chapter 43 of the 2019 ASHRAE Handbook—HVAC Applications covers building operating dynamics. Automatic controls often come prepackaged and prewired on the HVAC equipment, creating a design challenge to have the equipment communicate with a third party building automation system (BAS). Current HVAC controls and their capabilities need to be compatible with other new and existing automatic controls. Chapter 41 of the 2019 ASHRAE Handbook—HVAC Applications discusses computer applications, and ASHRAE Standard 135 discusses interfacing building automation systems. Furthermore, compatibility with BACnet® and/or LonWorks® provides an additional level of compatibility between equipment made by numerous manufacturers. Using computers and proper software, the design engineer and building manager can provide complete facility management, as well as remote facility management. A comprehensive building management system (BMS) may include HVAC system control, energy management, operation and maintenance management, medical gas system monitoring, fire alarm system, security system, lighting control, and other reporting and trending software. This system may also be integrated and accessible from the owner’s computer network and the Internet. The automatic control system can be as simple as a time clock to start and stop equipment, or as sophisticated as computerized building automation serving a decentralized HVAC system, multiple building systems, central plant system, and/or a large campus. With a focus on energy management, the building management system can be an important business tool in achieving sustainable facility management that begins with using the system selection matrix.

8.

MAINTENANCE MANAGEMENT

Whereas BMS focus on operation of HVAC, electrical, plumbing, and other systems, maintenance management systems focus on maintaining assets, which include mechanical and electrical systems along with the building structure. A rule of thumb is that only about 20% of the cost of the building is in the first cost, with the other 80% being operation, maintenance, and rejuvenation of the building and building systems over the life cycle. When considering the optimum HVAC selection and recommendation at the start of a project, a maintenance management system should be considered for HVAC systems with an estimated long useful service life. Another maintenance management business tool is a computerized maintenance management software (CMMS) system. The CMMS system can include an equipment database, parts and material inventory, project management software, labor records, etc., pertinent to sustainable management of the building over its life. CMMS also can integrate computer-aided drawing (CAD), building information modeling (BIM), digital photography and audio/video systems, equipment run-time monitoring and trending, and other proactive facility management systems. In scoring the HVAC system selection matrix, consideration should also be given to the potential for interface of the BMS with the CMMS. In addition, onboard equipment diagnostics and remote troubleshooting capabilities as part of considering the optimum HVAC selection. Planning in the design phase and the early compilation of record documents (e.g., computer-aided drawing and electronic text files, checklists, digital photos taken during construction) are also integral to successful building management and maintenance. The

use of bar coding equipment, valves, etc. can link them to the CMMS via an Internet accessed scanner.

9.

BUILDING SYSTEM COMMISSIONING

When compiling data to complete the HVAC system selection matrix to analytically determine the optimum HVAC system for the project, a design engineer should begin to produce the design intent document/basis of design that identifies the project goals. This process is the beginning of building system commissioning and should be an integral part of the project documentation. As design progresses and the contract documents take shape, the commissioning process continues to be built into what eventually becomes the final commissioning report, approximately one year after the construction phase has been completed and the warranty phase comes to an end. For more information, see Chapter 44 in the 2019 ASHRAE Handbook—HVAC Applications or ASHRAE Guideline 1.1. Building commissioning contributes to successful sustainable HVAC design by incorporating the system training requirements necessary for building management staff to efficiently take over ownership, operation, and maintenance of the HVAC systems over the installation’s useful service life. Building system commissioning is often contracted directly by an owner and is required by many standards to achieve peak building system performance. Review in the design phase of a project should consider both commissioning and air and water balancing, which should continue through the construction and warranty phases. For additional information, refer to ASHRAE Guideline 0. With building certification programs (e.g., ASHRAE’s bEQ, LEED®), commissioning is a prerequisite because of the importance of ensuring that high-performance energy and environmental designs are long-term successes.

REFERENCES ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal.ashrae .org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org/bookstore. ASHRAE. 2013. The commissioning process. Guideline 0-2013. ASHRAE. 2007. HVAC&R technical requirements for the commissioning process. Guideline 1.1-2007. ASHRAE. 2009. Guideline for the risk management of public health and safety in buildings. Guideline 29-2009. ASHRAE. 2013. Safety standard for refrigeration systems. ANSI/ASHRAE Standard 15-2013. ASHRAE. 2013. Thermal environmental conditions for human occupancy. ANSI/ASHRAE Standard 55-2013. ASHRAE. 2013. Ventilation for acceptable indoor air quality. ANSI/ ASHRAE Standard 62.1-2013. ASHRAE. 2013. Energy standard for buildings except low-rise residential buildings. ANSI/ASHRAE Standard 90.1-2013. ASHRAE. 2015. Energy efficiency in existing building. ANSI/ASHRAE Standard 100-2015. ASHRAE. 2012. BACnet®—A data communication protocol for building automation and control networks. ANSI/ASHRAE Standard 135-2012. ASHRAE. 2013. Climate data for building design standards. ANSI/ ASHRAE Standard 169-2013. ASHRAE. 2012. Standard practice for inspection and maintenance of commercial building. ANSI/ASHRAE/ACCA Standard 180-2012. ASHRAE. 2014. Standard for the design of high-performance green buildings. ANSI/ASHRAE/USGBC Standard 189.1-2014. ASHRAE. 2013. Commissioning process for buildings and systems. ANSI/ ASHRAE Standard 202-2013.

BIBLIOGRAPHY ASHRAE. 2016. Ventilation and acceptable indoor air quality in residential buildings. ANSI/ASHRAE Standard 62.2-2016. ASHRAE. 2018. Energy efficient design of low-rise residential buildings. ANSI/ASHRAE Standard 90.2-2018.

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DECENTRALIZED COOLING AND HEATING System Characteristics.............................................................. S2.1 Design Considerations.............................................................. S2.2 Window-Mounted and Through-theWall Room HVAC Units........................................................ S2.3 Water-Source Heat Pump Systems............................................ S2.4 Multiple-Unit Systems............................................................... S2.5 Residential and Light Commercial Split Systems ..................... S2.7

Commercial Self-Contained (Floor-by-Floor) Systems ............S2.7 Commercial Outdoor Packaged Systems ..................................S2.9 Single-Zone VAV Systems .......................................................S2.10 Automatic Controls and Building Management Systems ........S2.11 Maintenance Management ......................................................S2.11 Building System Commissioning .............................................S2.11

F

parameters, such as the sensible heat ratio at a given room condition or the airflow per kilowatt of refrigeration capacity. Components are matched and assembled to achieve specific performance objectives. These limitations make manufacture of low-cost, quality-controlled, factory-tested products practical. For a particular kind and capacity of unit, performance characteristics vary among manufacturers. All characteristics should be carefully assessed to ensure that the equipment performs as needed for the application. Several trade associations have developed standards by which manufacturers may test and rate their equipment. See Chapters 48 and 49 for more specific information on pertinent industry standards and on decentralized cooling and heating equipment used in multiple-packaged unitary systems. Large commercial/industrial-grade packaged equipment can be custom-designed by the factory to meet specific design conditions and job requirements. This equipment carries a higher first cost and is not readily available in smaller sizes. Self-contained units, often called rooftop units, can use multiple compressors to control refrigeration capacity. For variable-airvolume (VAV) systems, compressors are turned on or off or unloaded to maintain discharge air temperature. Variable-speed compressors can be factory integrated for close control. As zone demand decreases, the temperature of air leaving the unit can often be reset upward so that a minimum ventilation rate is maintained. Multiple packaged-unit systems for perimeter spaces are frequently combined with a central all-air or floor-by-floor system. These combinations can provide better humidity control, air purity, and ventilation than packaged units alone. Air-handling systems may also serve interior building spaces that cannot be conditioned by wall or window-mounted units. Water-source heat pump systems often combine packaged units (heat pumps) with a central piping system for heat rejection and heat gain. These systems require heat rejection equipment (ground source or cooling tower) and heat source (ground source or boiler) provided separately from the packaged heat pump. Heating can be included in a packaged unit. Gas heat exchangers or electric heat coils can be provided. Heat can be turned on or off in stages to meet zone demands. In some applications, heat from a centralized source, like a boiler system, is combined with decentralized packaged units, and steam or hot-water coils can be factory mounted in packaged units for connection to local piping systems. For supplementary data on air-side design of decentralized systems, see Chapter 4.

OR MOST small to mid-sized installations, decentralized cooling and heating is usually preferable to a centralized system (see Chapter 3). Frequently classified as packaged unit systems (although many are far from being a single packaged unit), decentralized systems can be found in almost all classes of buildings. They are especially suitable for smaller projects with no central plant, where low initial cost and simplified installation are important. These systems are installed in office buildings, shopping centers, manufacturing plants, schools, health care facilities, hotels, motels, apartments, nursing homes, and other multiple-occupancy dwellings. They are also suited to air conditioning existing buildings with limited life or income potential. Applications also include facilities requiring specialized high performance levels, such as computer rooms and research laboratories. Although some of the equipment addressed here can be applied as a single unit, this chapter covers applying multiple units to form a complete heating and air-conditioning system for a building and the distribution associated with some of these systems. For guidance on HVAC system selection, see Chapter 1.

1.

SYSTEM CHARACTERISTICS

Decentralized systems can be one or more individual HVAC units, each with an integral refrigeration cycle, heating source, and direct or indirect outdoor air ventilation. Components are factorydesigned and assembled into a package that includes fans, filters, heating source, cooling coil, refrigerant compressor(s), controls, and condenser. Equipment is manufactured in various configurations to meet a wide range of applications. Examples of decentralized HVAC equipment include the following: • • • • • • • • • • •

Window air conditioners Through-the-wall room HVAC units Air-cooled heat pump systems Water-cooled heat pump systems Multiple-unit systems Residential and light commercial split systems Self-contained (floor-by-floor) systems Outdoor package systems Packaged, special-procedure units (e.g., for computer rooms) Single-zone variable-air-volume systems Variable-refrigerant-flow systems

For details on window air conditioners and through-the-wall units, see Chapter 49. For variable-refrigerant-flow systems, see Chapter 18; the other examples listed here are discussed further in Chapter 48. (Multiple-unit systems are also covered in Chapter 4.) Commercial-grade unitary equipment packages are available only in preestablished increments of capacity with set performance The preparation of this chapter is assigned to TC 9.1, Large Building AirConditioning Systems.

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Advantages • Heating and cooling can be provided at all times, independent of the mode of operation of other building spaces. • Manufacturer-matched components have certified ratings and performance data. • Assembly by a manufacturer helps ensure better quality control and reliability.

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• Manufacturer instructions and multiple-unit arrangements simplify installation through repetition of tasks. • Only one zone of temperature control is affected if equipment malfunctions. • The equipment is readily available. • Multiple vendors manufacture similar equipment, providing multiple sources. • One manufacturer is responsible for the final equipment package. • For improved energy control, equipment serving vacant spaces can be turned off locally or from a central point, without affecting occupied spaces. • System operation is simple. Trained operators are not usually required. • Less mechanical and electrical room space is required than with central systems. • Initial cost is usually low. • Equipment can be installed to condition one space at a time as a building is completed, remodeled, or as individual areas are occupied, with favorable initial investment. • Energy can be metered directly to each tenant. • Air- or water-side economizers may be applicable based on the unit capacity, depending on type of decentralized system used. • High-efficiency units with advanced controls are available from several manufacturers. Disadvantages • Performance options may be limited because airflow, cooling coil size, and condenser size are often fixed. Variable-speed fan and compressor control may be available. • Larger total building installed cooling capacity is usually required because diversity factors used for moving cooling needs do not apply to multiple dedicated packages. • Temperature and humidity control may be less stable, especially with mechanical cooling at very low loads. • Standard commercial units are not generally suited for large percentages of outdoor air or for close humidity control. Custom or special-purpose equipment, such as packaged units for computer rooms, hot-gas reheat for dehumidification, or large custom units, may be required. • Energy use is usually greater than for central systems if efficiency of the unitary equipment is less than that of the combined central system components. • Low-cost cooling by economizers may not be available or practical. • Air distribution design may be limited by fan capacity available. • Operating sound levels can be high, and noise-producing machinery is often closer to building occupants than with central systems. • Ventilation capabilities may be limited by equipment selection and design. • Equipment can make the building less aesthetically pleasing. • Air filtration options may be limited. • Discharge temperature varies because of on/off or step control. • Condensate drain is required with each air-conditioning unit. • Maintenance may be difficult or costly because of multiple pieces of equipment and their location.

2.

DESIGN CONSIDERATIONS

Rating classifications and typical sizes for equipment addressed in this chapter can be found in Chapters 48 and 49, which also address available components, equipment selection, distribution piping, and ductwork. Selection of a decentralized system should follow guidance provided in Chapter 1. The design engineer can use the HVAC system analysis selection matrix to analytically assess and select the optimum decentralized system for the project. Combined with the

design criteria in Chapters 48 and 49, the basis of design can be documented. Unlike centralized cooling and heating equipment, capacity diversity is limited with decentralized equipment, because each piece of equipment must be sized for peak capacity. For indoor decentralized refrigeration systems, each refrigeration system must be treated independently with regard to compliance with ASHRAE Standard 15 requirements; this may complicate individual mechanical room exhaust and air exchange. Noise from this type of equipment may be objectionable and should be checked to ensure it meets sound level requirements. Chapter 49 of the 2019 ASHRAE Handbook—HVAC Applications has more information on HVAC-related sound and vibration concerns. Decentralized packaged equipment requires design consideration because the building structure must support the units: a typical full compressorized package unit weighs more than a similarly sized air-handling-only system. When packaged units are located outdoors, designers must consider weather conditions and accommodate rain, snow, or high winds. Outdoor packaged units can conveniently be provided with integrated outdoor air intakes or air economizers. Gas-fired furnaces on outdoor package units can be vented directly from the unit. When packaged units are located indoors, the designer must consider a separate source for ventilation air and/or economizer. Gas heat might be limited by combustion air source and available flue pathways. When split-system packaged units are used, design limitations for refrigerant piping and distances must be considered. Design must consider noise generated by packaged unit’s compressors, fans, or both. Units mounted on roofs should not be located above sound-sensitive spaces, such as conference rooms or sleeping areas. Units inside the building might require enclosure in a mechanical room or closets with sound-absorbing wall construction. Decentralized units require electric and/or gas to each location. Designers must consider building type and installation costs for the utilities in addition to the HVAC system cost.

Air-Side Economizer With some decentralized systems, an air-side economizer is an option, if not an energy code requirement (check state code for criteria). The air-side economizer uses cool outdoor air to either assist mechanical cooling or, if the outdoor air is cool enough, provide total cooling. It requires a mixing box designed to allow 100% of the supply air to be drawn from the outdoors. See Chapter 43 in the 2019 ASHRAE Handbook—HVAC Applications for additional information on control of economizer cooling. Decentralized systems can be selected with a field-installed accessory that includes an outdoor air damper, relief damper, return air damper, filter, actuator, and linkage. Controls are usually a factory-installed option. Dampers must be the low-leakage type. Self-contained units usually do not include return air fans. A barometric relief, fan-powered relief fan, or return/exhaust fan may be provided as an air-side economizer. The relief fan is off and discharge/exhaust dampers are closed when the air-side economizer is inactive. Advantages • Substantially reduces compressor, cooling tower, and condenser water pump energy requirements, generally saving more energy than a water-side economizer. • Has a lower air-side pressure drop than a water-side economizer. • Reduces tower makeup water and related water treatment. • May improve indoor air quality by providing large amounts of outdoor air during mild weather.

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Decentralized Cooling and Heating Disadvantages • In systems with larger return air static pressure requirements, return or exhaust fans are needed to properly exhaust building air and take in outdoor air. • If the unit’s leaving air temperature is also reset up during the airside economizer cycle, humidity control problems may occur and the fan may use more energy, unless a return air humidity sensor resets the supply air discharge temperature. • Humidification may be required during winter. • More and/or larger air intake louvers, ducts, or shafts may be required for indoor package units.

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Water-Side Economizer The water-side economizer is another option for reducing energy use. ASHRAE Standard 90.1 addresses its application, as do some state energy codes. The water-side economizer consists of a water coil in a self-contained unit upstream of the direct-expansion cooling coil. All economizer control valves, piping between economizer coil and condenser, and economizer control wiring can be factory installed. This applies typically for indoor packaged compressor units using a water loop for heat rejection. The water-side economizer uses the low cooling tower or evaporative condenser water temperature to either (1) precool entering air, (2) assist mechanical cooling, or (3) provide total system cooling if the cooling water is cold enough. If the economizer is unable to maintain the air-handling unit’s supply air or zone set point, factory-mounted controls integrate economizer and compressor operation to meet cooling requirements. For constant condenser water flow control using a economizer energy recovery coil and the unit condenser, two control valves are factory-wired for complementary control, with one valve driven open while the other is driven closed. This keeps water flow through the condenser relatively constant. In variable-flow control, condenser water flow varies during unit operation. The valve in bypass/energy recovery loop is an on/off valve and is closed when the economizer is enabled. Water flow through the economizer coil is modulated by its automatic control valve, allowing variable cooling water flow as cooling load increases (valve opens) and reduced flow on a decrease in cooling demand. If the economizer is unable to satisfy the cooling requirements, factory-mounted controls integrate economizer and compressor operation. In this operating mode, the economizer valve is fully open. When the self-contained unit is not in cooling mode, both valves are closed. Reducing or eliminating cooling water flow reduces pumping energy. Advantages • Compressor energy is reduced by precooling entering air. Often, building load can be completely satisfied with an entering condenser water temperature of less than 13°C. Because the wet-bulb temperature is always less than or equal to the dry-bulb temperature, a lower discharge air temperature is often available. • Building humidification does not affect indoor humidity by introducing outdoor air. • No external wall penetration is required for exhaust or outdoor air ducts. • Controls are less complex than for air-side economizers, because they are often inside the packaged unit. • The coil can be mechanically cleaned. • More net usable floor area is available because large outdoor and relief air ducts are unnecessary. Disadvantages • Cooling tower water treatment cost is greater. • Cooling tower water usage is greater due to blowdown requirements.

2.3 • Air-side pressure drop may increase with the constant added resistance of an economizer coil in the air stream. • Condenser water pump pressure may increase slightly. • The cooling tower must be designed for winter operation (see Chapter 40). • The increased operation (including in winter) required of the cooling tower may reduce its life.

3.

WINDOW-MOUNTED AND THROUGH-THEWALL ROOM HVAC UNITS

Window air conditioners (air-cooled room conditioners) and through-the-wall room air conditioners with supplemental heating are designed to cool or heat individual room spaces. Window units are used where low initial cost, quick installation, and other operating or performance criteria outweigh the advantages of more sophisticated systems. Room units are also available in through-thewall sleeve mountings. Sleeve-installed units are popular in low-cost apartments, motels, and homes. Ventilation can be through operable windows or limited outdoor air ventilation introduced through the self-contained room HVAC unit. These units are described in more detail in Chapter 49. Window units may be used as auxiliaries to a central heating or cooling system or to condition selected spaces when the central system is shut down. These units usually serve only part of the spaces conditioned by the basic system. Both the basic system and window units should be sized to cool the space adequately without the other operating. A through-the-wall air-cooled room air conditioner is designed to cool or heat individual room spaces. Design and manufacturing parameters vary widely. Specifications range from appliance grade through heavy-duty commercial grade, the latter known as packaged terminal air conditioners (PTACs) or packaged terminal heat pumps (PTHPs) (AHRI Standard 310/380). With proper maintenance, manufacturers project an equipment life of 10 to 15 years for these units. More sophisticated through-the-wall units are available for school applications to replace heat-only unit ventilators with heating and cooling unit ventilators. These applications may have steam or hot-water coils incorporated. Typically, packaged control systems provide control of outdoor ventilation, economizer, or compressor operation as required. Advantages • Lowest first cost HVAC system. • Installation of in-room unit is simple. It usually only requires an opening in the wall or displacement of a window to mount the unit, and connection to electrical power. • Equipment is often available from stock. • Generally, the system is well-suited to spaces requiring many zones of individual temperature control. • Designers can specify electric, hydronic, or steam heat or use an air-to-air heat pump design. • Service of in-room equipment can be quickly restored by replacing a defective chassis. • Ductwork or air distribution systems are not required. Disadvantages • Equipment life may be less than for large central equipment, typically 10 to 15 years, and units are built to appliance standards, rather than building equipment standards. • Energy use may be relatively high. • Direct access to outdoor air is needed for condenser heat rejection; thus, these units cannot be used for interior rooms. • The louver and wall box must stop wind-driven rain from collecting in the wall box and leaking into the building.

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• The wall box should drain to the outdoor, which may cause dripping on walls, balconies, or sidewalks. • Temperature control is usually two-position, which causes swings in room temperature. • Ventilation capabilities are fixed by equipment design. • Economizer is limited. • Humidification, when required, must be provided by separate equipment. • Noise and vibration levels vary considerably and are not generally suitable for sound-critical applications. • Routine maintenance of a large number of units is required to maintain capacity. Condenser and cooling coils must be cleaned, and filters must be changed regularly. • Discharge air location limited to unit location.

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Design Considerations A through-the-wall or window-mounted air-conditioning unit incorporates a complete air-cooled refrigeration and air-handling system in an individual package. Each room is an individual occupantcontrolled zone. Cooled or warmed air is discharged in response to thermostatic control to meet room requirements (see the discussion on controls following in this section). Each PTAC or PTHP has a self-contained, air-cooled directexpansion or heat pump cooling system; a heating system (electric, hot water, steam, and/or a heat pump cycle); and controls. See Figure 3 in Chapter 49 for unit configuration. A through-the-wall air conditioner or heat pump system is installed in buildings requiring many temperature control zones such as office buildings, motels and hotels, apartments and dormitories, schools and other education buildings, and areas of nursing homes or hospitals where air recirculation is allowed. These units can be used for renovation of existing buildings, because existing heating systems can still be used. The equipment can be used in both low- and high-rise buildings. In buildings where a stack effect is present, use should be limited to areas that have dependable ventilation and a tight wall of separation between the interior and exterior. Room air conditioners are often used in parts of buildings primarily conditioned by other systems, especially where spaces to be conditioned are (1) physically isolated from the rest of the building and (2) occupied on a different time schedule (e.g., clergy offices in a church, ticket offices in a theater). Ventilation air through each terminal may be inadequate in many situations, particularly in high-rise structures because of the stack effect. Chapter 16 of the 2017 ASHRAE Handbook—Fundamentals explains combined wind and stack effects. Electrically operated outdoor air dampers, which close automatically when the equipment is stopped, can reduce heat losses in winter. Refrigeration Equipment. Room air conditioners are generally supplied with hermetic reciprocating or scroll compressors. Capillary tubes are used in place of expansion valves in most units. Some room air conditioners have only one motor to drive both the evaporator and condenser fans. The unit circulates air through the condenser coil whenever the evaporator fan is running, even during the heating season. Annual energy consumption of a unit with a single motor is generally higher than one with a separate motor, even when the energy efficiency ratio (EER) or the coefficient of performance (COP) is the same for both. Year-round, continuous flow of air through the condenser increases dirt accumulation on the coil and other components, which increases maintenance costs and reduces equipment life. Because through-the-wall conditioners are seldom installed with drains, they require a positive and reliable means of condensate disposal. Conditioners are available that spray condensate in a fine mist over the condenser coil. These units dispose of more condensate than can be developed without any drip, splash, or spray. In heat

pumps, provision must be made for disposal of condensate generated from the outdoor coil during defrost. Many air-cooled room conditioners experience evaporator icing and become ineffective when outdoor temperatures fall below about 18°C. Units that ice at a lower outdoor temperature may be required to handle the added load created by high lighting levels and high solar radiation found in contemporary buildings. Heating Equipment. The air-to-air heat pump cycle described in Chapter 48 is available in through-the-wall room air conditioners. Application considerations are similar to conventional units without the heat pump cycle, which is used for space heating when the outdoor temperature is above 2 to 5°C. The prime advantage of the heat pump cycle is that it reduces annual energy consumption for heating. Savings in heat energy over conventional electric heating ranges from 10 to 60%, depending on the climate. In some applications, when steam or hot water heat is used, designers should consider piping isolation to allow removal of the units’ chassis for service and repair. Electric resistance elements supply heating below this level and during defrost cycles. Controls. All controls for through-the-wall air conditioners are included as a part of the conditioner. The following control configurations are available: • Thermostat control is either unit-mounted or remote wallmounted. • Motel and hotel guest room controls allow starting and stopping units from a central point. • Occupied/unoccupied controls (for occupancies of less than 24 h) start and stop the equipment at preset times with a time clock. Conditioners operate normally with the unit thermostat until the preset cutoff time. After this point, each conditioner has its own reset control, which allows the occupant of the conditioned space to reset the conditioner for either cooling or heating, as required. • Master/slave control is used when multiple conditioners are operated by the same thermostat. • Emergency standby control allows a conditioner to operate during an emergency, such as a power failure, so that the roomside blowers can operate to provide heating. Units must be specially wired to allow operation on emergency generator circuits. • BACnet™ integration is generally available for multiple installations to improve automation (see ASHRAE Standard 135). A building automation system (BAS) allows more sophisticated unit control by time-of-day scheduling, optimal start/stop, duty cycling, demand limiting, custom programming, etc. This control can keep units operating at peak efficiency by alerting the operator to conditions that could cause substandard performance. When several units are used in a single space, controls should be interlocked to prevent simultaneous heating and cooling. In commercial applications (e.g., motels), centrally operated switches can de-energize units in unoccupied rooms.

4.

WATER-SOURCE HEAT PUMP SYSTEMS

Water-source heat pump systems use multiple cooling/heating units distributed throughout the building. For more in-depth discussion of water-source heat pumps, see Chapters 9 and 48. Outdoor air ventilation requires either direct or indirect supply air from an additional air-handling system. Designed to cool and heat individual rooms or multiple spaces grouped together by zone, water-source heat pumps may be installed along the perimeter with a combination of horizontal and vertical condenser water piping distribution, or stacked vertically with condenser water piping also stacked vertically to minimize

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equipment space effects on the rooms they serve. They can also be ceiling mounted or concealed above the ceiling with duct distribution to the area served. Water-source heat pump systems can transfer energy within the system by using the common water loop. Units on sunny exposures in cooling mode add heat to the water loop, from which units in heating modes can draw heat. Water-source heat pump systems also require further decentralized equipment that includes a source of heat rejection such as a cooling tower or groundwater installation. A water distribution piping system with pumps is required. A supplemental heating source such as a boiler may be required, depending on the installation’s location (e.g., in colder winter climates). Data on condenser water systems and the necessary heat rejection equipment can be found in Chapters 14 and 40. Advantages • Water-source heat pump systems use renewable energy when designed with a ground source. • Unit installation is simple for both vertical and horizontal installation and for concealed locations. • Units are available in multiple electrical configurations. • Water-source heat pumps can be installed with a lift to rig horizontal units at or above a ceiling. • Generally, the system is well suited to spaces requiring many zones of individual temperature control. • Units are available in multiple sizes, typically in 1.8 kW increments, to meet any need. • Designers can specify energy recovery for ventilation air separately. • Units can be individually metered easily. • Units are easily relocated if space is remodeled. Disadvantages • Outdoor air introduction to the building must be separately designed. • Condensate drain piping installation and routine maintenance can be a problem, particularly for units installed above ceilings. • Economizer systems typically do not apply. • Humidification, when required, must be provided by separate equipment. • Discharge heating air temperatures are lower compared to gas heat. • Noise and vibration levels vary considerably and are not generally suitable for sound-critical applications. • Routine maintenance within occupied space is required to maintain capacity.

2.5 5.

MULTIPLE-UNIT SYSTEMS

Multiple-unit systems generally use under 70 kW, single-zone unitary HVAC equipment with a unit for each zone (Figure 1). These units typically have factory-charged, indoor compressor refrigeration systems. Some use remote air-cooled condensers with field-provided refrigeration piping. Equipment can be packaged air cooled, and wall louvers can be used to reject condenser heat to the outdoors. Zoning is determined by (1) cooling and heating loads, (2) occupancy, (3) flexibility requirements, (4) appearance, and (5) equipment and duct space availability. Multiple-unit systems are popular for low-rise office buildings, manufacturing plants, shopping centers, department stores, and apartment buildings. Unitary selfcontained units are excellent for renovation. The system configuration can be horizontal distribution of equipment and associated ductwork and piping, or vertical distribution of equipment and piping with horizontal distribution of ductwork. Outdoor air ventilation requires either direct or indirect supply air from an additional air-handling system. Heating media may be steam or hot water piped to the individual units or electric heat at the unit and/or at individual air terminals that provide multiple-space temperature-controlled zones. Heat pumps are available in some equipment. Cooling media may be refrigerant/ direct expansion (DX) air, or condenser water piped to individual units. A typical system features zone-by-zone equipment with a central cooling tower and boiler, although electric heat and DX refrigerant cooling may be used. When these units are cooled by DX coils connected to a variable-refrigerant-flow (VRF) system, the system can take on more “centralized” characteristics; see Chapter 18. Supply air is typically constant volume; outdoor ventilation is typically a fixed minimum. Some designs can include variable volume based on bypass volume control. Usually, multiple units do not come with a return air fan; the supply fan must overcome return air and supply air duct static resistance. Air-side economizers are rare; a water-side economizer is sometimes more practical where condenser water is used, and may be required by state code. Multiple-unit systems require a localized equipment room where one or more units can be installed. This arrangement takes up floor space, but allows equipment maintenance to occur out of the building’s occupied areas, minimizing interruptions to occupants. Advantages • The system can be applied to both large or small buildings (building height).

Design Considerations The location for the central heat rejection and/or heat gain equipment requires consideration. If a cooling tower is used, designers may consider separating the indoor water loop from the outdoor water loop with a heat exchanger. Pump systems can be variable flow when units are provided with automatic flow valves integrated to open with compressor operation. Units are usually furnished with communicating digital controls. Units are typically provided with occupied/unoccupied sequence. These units can be used for renovation of existing buildings where limited ceiling space would prevent other types of HVAC systems (e.g., all-air systems) from being installed. The equipment can be used in both low- and high-rise buildings; both applications require some form of outdoor ventilation to serve the occupants. See Chapters 9 and 48 for data on refrigeration cycle, heating cycle, automatic controls, and other information on design and operation and maintenance.

Fig. 1 Multiple-Unit Systems: Single-Zone Unitary HVAC Equipment for Interior and Packaged Terminal Air Conditioners (PTACs) for Perimeter (Courtesy RDK Engineers)

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• Installation is simple. Equipment is readily available in sizes that allow easy handling. • Installation of outdoor heat pumps is simple with rigging onto concrete pad at grade level or on the roof. • Equipment and components can be standardized. • Relocation of units to other spaces or buildings is practical, if necessary. • Energy efficiency can be quite good, particularly where climate or building use results in a balance of heating and cooling zones. • Units are available with complete, self-contained control systems that include variable-volume control, night setback, and morning warm-up. • Equipment is out of the occupied space, making the system quieter than water-source heat pumps. • Easy access to equipment facilitates routine maintenance. • Failure of one unit affects only a limited area. • System is repetitive and simple, facilitating operator training. • Tenant utility metering is easier than with other systems.

Economizer Cycle. When outdoor temperature allows, energy use can be reduced by cooling with outdoor air in lieu of mechanical refrigeration. Units must be located close to an outdoor wall or outdoor air duct shafts. Where this is not possible, it may be practical to add a water-side economizer cooling coil, with cold water obtained by sending condenser water through a winterized cooling tower. Chapter 40 has further details. Economizer cycle may be required by state code. Acoustics and Vibration. Because these units are typically located near occupied space, they can affect acoustics. The designer must study both the airflow breakout path and the unit’s radiated sound power when coordinating selection of wall and ceiling construction surrounding the unit. Locating units over noncritical work spaces such as restrooms or storage areas around the equipment room helps reduce noise in occupied space. Chapter 49 of the 2019

Disadvantages • Access to outdoor air must be provided at each location. • Fans may have limited static pressure ratings. • Air filtration options are limited. • Discharge air temperature varies with on/off or step control. • Humidification can be impractical on a unit-by-unit basis and may need to be provided by a separate system. • Integral air-cooled condensing units for some direct-expansion cooling installations should be located outdoors within a limited distance. • Multiple units and equipment closets or rooms may occupy rentable floor space. • Multiple pieces of equipment may increase maintenance requirements. • Redundant equipment or easy replacement may not be practical.

Design Considerations Unitary systems can be used throughout a building or to supplement perimeter packaged terminal units (Figures 1 and 2). Because core areas frequently have little or no heat loss, unitary equipment with air- or water-cooled condensers can be applied. The equipment can be used in both low- and high-rise buildings; both applications require some form of outdoor ventilation to serve the occupants. Typical application may be an interior work area, computer room, or other space requiring continual cooling. Special-purpose unitary equipment is frequently used to cool, dehumidify, humidify, and reheat to maintain close control of space temperature and humidity in computer areas (Figure 2). For more information, see Chapters 19 and 20 of the 2019 ASHRAE Handbook—HVAC Applications, as well as Chapters 48 and 49 of this volume. In the multiple-unit system shown in Figure 3, one unit may be used to precondition outdoor air for a group of units. This all-outdoor-air unit prevents hot, humid air from entering the conditioned space during periods of light load. The outdoor unit should have sufficient capacity to cool the required ventilation air from outside design conditions to interior design dew point. Zone units are then sized to handle only the internal load for their particular area. Units are typically under 70 kW of cooling and are typically constant-volume units. VAV distribution may be accomplished on these units with a bypass damper that allows excess supply air to bypass to the return air duct. The bypass damper ensures constant airflow across the direct-expansion cooling coil to avoid coil freeze-up caused by low airflow. The damper is usually controlled by supply duct pressure. Variable-frequency drives can also be used. Controls. Units are usually furnished with individual electric controls, but can be enhanced to a more comprehensive building management system.

Fig. 2

Vertical Self-Contained Unit (Courtesy RDK Engineers)

Fig. 3 Dedicated Outdoor-Air-Conditioning Unit (Courtesy RDK Engineers)

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6.

RESIDENTIAL AND LIGHT COMMERCIAL SPLIT SYSTEMS

These systems distribute cooling and heating equipment throughout the building. A split system consists of an indoor unit with air distribution and temperature control, with either a remote air-cooled condenser or remote air-cooled condensing unit. These units are commonly used in single-story or low-rise buildings, and in residential applications where condenser water is not readily available. Commercial split systems are well suited to small projects with variable occupancy schedules. Indoor equipment is generally installed in service areas adjacent to the conditioned space. When a single unit is required, the indoor unit and its related ductwork constitute a central air system, as described in Chapter 4. Typical components of a split-system air conditioner include an indoor unit with evaporator coils, heating coils, filters, valves, and a condensing unit with the compressors and condenser coils. The configuration can be horizontal distribution of equipment and associated ductwork and piping, or vertical distribution of equipment and piping with horizontal distribution of ductwork. These applications share some of the advantages of multiple-unit systems, but may only have one system installation per project. Outdoor air ventilation requires either direct or indirect supply from another source (e.g., operable window). Heating is usually electric or gas, but may be steam, hot water, or possibly oil-fired at the unit. Cooling is usually by direct expansion. Heat pump operation is readily accommodated. Supply air may be constant- or variable-volume bypass damper control; outdoor ventilation is usually a fixed minimum or barometric relief economizer cycle. A supplemental exhaust fan may be required to complete the design. Air-cooled heat pumps located on roofs or adjacent to buildings are another type of package equipment with most of the features noted here, with the additional benefit of supply air distribution and equipment outside the occupied space. This improved ductwork arrangement makes equipment accessible for servicing out the occupied space, unlike in-room units. See Chapter 48 for additional design, operating, and constructability discussion. Advantages • Low first cost. • Unitary split-system units allow air-handling equipment to be placed close to the heating and cooling load, which allows ample air distribution to the conditioned space with minimum ductwork and fan power. • Tenant utility metering is easy. • Heat rejection through a remote air-cooled condenser allows the final heat rejector (and its associated noise) to be remote from the air-conditioned space. • A floor-by-floor arrangement can reduce fan power consumption because air handlers are located close to the conditioned space. • Equipment is generally located in the building interior near elevators and other service areas and does not interfere with the building perimeter. Disadvantages • Multiple units and equipment closet may occupy rentable floor space. • The proximity of the air handler to the conditioned space requires special attention to unit inlet and outlet airflow and to building acoustics around the unit. • Multiple pieces of equipment may increase maintenance requirements.

2.7 • Ducting ventilation air to the unit and removing condensate from the cooling coil should be considered. • Refrigeration piping to outdoors has limited length (typically about 30 m). • A unit that uses an air-side economizer must be located near an outer wall or outdoor air shaft. Split-system units do not generally include return air fans. • A separate method of handling and controlling relief air may be required. • Filter options and special features may be limited. • Discharge temperature varies because of on/off or step control. • Gas-heat furnace needs access for flue venting to roof or outdoors.

Design Considerations Characteristics that favor split systems are their low first cost, simplicity of installation, and simplicity of training required for operation. Servicing is also relatively inexpensive. The modest space requirements of split-system equipment make it excellent for renovations or for spot cooling a single zone. Control is usually one- or two-step cooling and one- or two-step or modulating heat. VAV operation is possible with a supply air bypass. Some commercial units can modulate airflow, with additional cooling modulation using hot-gas bypass. Commercial split-system units are available as constant-volume equipment for use in atriums, public areas, and industrial applications. Basic temperature controls include a room-mounted or returnair-mounted thermostat that cycles the compressor(s) as needed. Upgrades include fan modulation and VAV control. When applied with VAV terminals, commercial split systems provide excellent comfort and individual zone control.

7.

COMMERCIAL SELF-CONTAINED (FLOORBY-FLOOR) SYSTEMS

Commercial self-contained (floor-by-floor) systems are a type of multiple-unit, decentralized cooling and heating system. Equipment is usually configured vertically, but may be horizontal. Supply air distribution may be a discharge air plenum, raised-floor supply air plenum (air displacement), or horizontal duct distribution, installed on a floor-by-floor basis. Outdoor air ventilation requires either direct or indirect supply air from an additional air-handling system. Typical components include compressors, water-cooled condensers, evaporator coils, economizer coils, heating coils, filters, valves, and controls (Figure 4). To complete the system, a building needs cooling towers and condenser water pumps. See Chapter 40 for more information on cooling towers. Advantages • This equipment integrates refrigeration, heating, air handling, and controls into a factory package, thus eliminating many field integration problems. • Units are well suited for office environments with variable occupancy schedules. • Floor-by-floor arrangement can reduce fan power consumption. • Large vertical duct shafts and fire dampers are eliminated. • Electrical wiring, condenser water piping, and condensate removal are centrally located. • Equipment is generally located in the building interior near elevators and other service areas, and does not interfere with the building perimeter. • Integral water-side economizer coils and controls are available, which allow interior equipment location and eliminate large outdoor air and exhaust ducts and relief fans. • An acoustical discharge plenum is available, which allows lower fan power and lower sound power levels.

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Disadvantages • Units must be located near an outdoor wall or outdoor air shaft to incorporate an air-side economizer. • A separate relief air system and controls must be incorporated if an air-side economizer is used. • Close proximity to building occupants requires careful analysis of space acoustics. • Filter options may be limited. • Discharge temperature varies because of on/off or step control.

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Design Considerations Commercial self-contained units have criteria similar to those for multiple and light commercial units, and can serve either VAV or constant-volume systems. These units contain one or two fans inside the cabinet. The fans are commonly configured in a draw-through arrangement. The size and diversity of the zones served often dictate which system is optimal. For comfort applications, self-contained VAV units coupled with terminal boxes or fan-powered terminal boxes are popular for their energy savings, individual zone control, and acoustic benefits. Constant-volume self-contained units have low installation cost and are often used in noncomfort or industrial airconditioning applications or in single-zone comfort applications. Unit airflow is reduced in response to terminal boxes closing. Common methods used to modulate airflow delivered by the fan to match system requirements include fan speed control and multiplespeed fan motors. Appropriate outdoor air and exhaust fans and dampers work in conjunction with the self-contained unit. Their operation must be coordinated with unit operation to maintain design air exchange and building pressurization. Refrigeration Equipment. Commercial self-contained units usually feature reciprocating or scroll compressors, although screw compressors are available in some equipment. Thermostatic or electronic expansion valves are used. Condensers are water cooled and usually reject heat to a common condenser-water system serving multiple units. A separate cooling tower or other final heat rejection device is required. Self-contained units may control capacity with multiple compressors. For VAV systems, compressors are turned on or off or

Fig. 4 Commercial Self-Contained Unit with Discharge Plenum

unloaded to maintain discharge air temperature. Hot-gas bypass is often incorporated to provide additional capacity control. As system airflow decreases, the temperature of air leaving the unit is often reset upward so that a minimum ventilation rate can be maintained. Resetting the discharge air temperature limits the unit’s demand, thus saving energy. However, increased air temperature and volume increase fan energy. Heating Equipment. In many applications, heating is done by perimeter radiation, with heating installed in the terminal boxes or other such systems when floor-by-floor units are used. If heating is incorporated in these units (e.g., preheat or morning warm-up), it is usually provided by hot-water coils or electric resistance heat, but could be by a gas- or oil-fired heat exchanger. Controls. Self-contained units typically have built-in capacity controls for refrigeration, economizers, and fans. Although units under 50 kW of cooling tend to have basic on/off/automatic controls, many larger systems have sophisticated microprocessor controls that monitor and take action based on local or remote programming. These controls provide for stand-alone operation, or they can be tied to a building automation system (BAS). A BAS allows more sophisticated unit control by time-of-day scheduling, optimal start/stop, duty cycling, demand limiting, custom programming, etc. This control can keep units operating at peak efficiency by alerting the operator to conditions that could cause substandard performance. The unit’s control panel can sequence the modulating valves and dampers of an economizer. A water-side economizer is located upstream of the evaporator coil, and when condenser water temperature is lower than entering air temperature to the unit, water flow is directed through the economizer coil to either partially or fully meet building load. If the coil alone cannot meet design requirements, but the entering condenser water temperature remains cool enough to provide some useful precooling, the control panel can keep the economizer coil active as stages of compressors are activated. When entering condenser water exceeds entering air temperature to the unit, the coil is valved off, and water is circulated through the unit’s condensers only. Typically, in an air-side economizer, an enthalpy or dry-bulb temperature switch energizes the unit to bring in outdoor air as the first stage of cooling. An outdoor air damper modulates flow to meet design temperature, and when outdoor air can no longer provide sufficient cooling, compressors are energized. A temperature input to the control panel, either from a discharge air sensor or a zone sensor, provides information for integrated economizer and compressor control. Supply air temperature reset is commonly applied to VAV systems. In addition to capacity controls, units have safety features for the refrigerant-side, air-side, and electrical systems. Refrigeration protection controls typically consist of high and low refrigerant pressure sensors and temperature sensors wired into a common control panel. The controller then cycles compressors on and off or activates hot-gas bypass to meet system requirements. Constant-volume units typically have high-pressure cut-out controls, which protect the unit and ductwork from high static pressure. VAV units typically have some type of static pressure probe inserted in the discharge duct downstream of the unit. As terminal boxes close, the control modulates airflow to meet the set point, which is determined by calculating the static pressure required to deliver design airflow to the zone farthest from the unit. Acoustics and Vibration. Because self-contained units are typically located near occupied space, their performance can significantly affect occupant comfort. Units of less than 50 kW of cooling are often placed inside a closet, with a discharge grille penetrating the common wall to the occupied space. Larger units have their own equipment room and duct system. Common sound paths to consider include the following:

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• Fan inlet and compressor sound radiates through the unit casing to enter the space through the separating wall. • Fan discharge sound is airborne through the supply duct and enters the space through duct breakout and diffusers. • Airborne fan inlet sound enters the space through the return air ducts, or ceiling plenum if unducted. Air is often discharged from the self-contained unit with a plenum on top of the unit. This plenum facilitates multiple duct discharges that reduce the amount of airflow over a single occupied space adjacent to the equipment room (see Figure 4). Reducing airflow in one direction reduces the sound that breaks out from the discharge duct. About a metre of internally lined round duct immediately off the discharge plenum may significantly reduce noise levels in adjacent areas. In addition to the airflow breakout path, the system designer must study unit-radiated sound power when determining equipment room wall and door construction. A unit’s air-side inlet typically has the highest radiated sound. The inlet space and return air ducts should be located away from the critical area to reduce the effect of this sound path. Selecting a fan that operates near its peak efficiency point helps design quiet systems. Fans are typically dominant in the first three octave bands, and selections at high static pressures or near the fan’s surge region should be avoided. Units may be isolated from the structure with neoprene pads or spring isolators. Manufacturers often isolate the fan and compressors internally, which generally reduces external isolation requirements.

8.

COMMERCIAL OUTDOOR PACKAGED SYSTEMS

Commercial outdoor packaged systems are available in sizes from 7 kW to over 700 kW. There are multiple choices of rooftop and/or horizontal grade-mounted configurations to meet building needs. Heating is usually electric or gas but may be steam, hot water, or possibly oil-fired at the unit. Cooling is usually by direct expansion, but could be chilled water. Supply air distribution can be for multiple floors by vertical duct shafts or horizontal duct distribution installed on a floor-by-floor basis. Outdoor air ventilation can be provided by barometric relief, fan-powered relief, or return air/exhaust air fan. Equipment is generally mounted on the roof (rooftop units [RTUs]), but can also be mounted at grade level. RTUs are designed as central-station equipment for single-zone, multizone, and VAV applications. Systems are available in several levels of design sophistication, from simple factory-standard light commercial packaged equipment, to double-wall commercial packaged equipment with upgraded features, up to fully customized industrial-quality packages. Often, factory-standard commercial rooftop unit(s) are satisfactory for small and medium-sized office buildings. On large projects and highly demanding systems, the additional cost of a custom packaged unit can be justified by life-cycle cost analyses. Custom systems offer great flexibility and can be configured to satisfy almost any requirement. Special features such as heat recovery, service vestibules, boilers, chillers, and space for other mechanical equipment can be designed into the unit. For additional information, see Chapter 48. Advantages • Equipment location allows easy service access without maintenance staff entering or disturbing occupied space. • Construction costs are offset toward the end of the project because the unit can be one of the last items installed.

2.9 • Installation is simplified and field labor costs are reduced because most components are assembled and tested in a controlled factory environment. • A single source has responsibility for design and operation of all major mechanical systems in the building. • Valuable building space for mechanical equipment is conserved. • It is suitable for floor-by-floor control in low-rise office buildings. • Outdoor air is readily available for ventilation and economy cycle use. • Combustion air intake and flue gas exhaust are facilitated if natural gas heat is used. • Upgraded design features, such as high-efficiency filtration or heat recovery devices, are available from some manufacturers. Disadvantages • Maintaining or servicing outdoor units is sometimes difficult, especially in inclement weather. • With all rooftop equipment, safe access to the equipment is a concern. Even slightly sloped roofs are a potential hazard. • Frequent removal of panels for access may destroy the unit’s weatherproofing, causing electrical component failure, rusting, and water leakage. • Rooftop unit design must be coordinated with structural design because it may represent a significant building structural load. • In cold climates, provision must be made to keep snow from blocking air intakes and access doors, and the potential for freezing of hydronic heating or steam humidification components must be considered. • Casing corrosion is a potential problem. Many manufacturers prevent rusting with galvanized or vinyl coatings and other protective measures. • Outdoor installation can reduce equipment life. • Depending on building construction, sound levels and transmitted vibration may be excessive. • Architectural considerations may limit allowable locations or require special screening to minimize visual effect.

Design Considerations Centering the rooftop unit over the conditioned space reduces fan power, ducting, and wiring. Avoid installation directly above spaces where noise and vibration level is critical. All outdoor ductwork should be insulated, if not already required by associated energy codes. In addition, ductwork should be (1) sealed to prevent condensation in insulation during the heating season and (2) weatherproofed to keep it from getting wet. Use multiple single-zone, not multizone, units where feasible to simplify installation and improve energy consumption. For large areas (e.g., manufacturing plants, warehouses, gymnasiums), single-zone units are less expensive and provide protection against total system failure. Use units with return air fans whenever return air static pressure loss exceeds 125 Pa or the unit introduces a large percentage of outdoor air via an economizer. Units are also available with relief fans for use with an economizer in lieu of continuously running a return fan. Relief fans can be initiated by static pressure control. In a rooftop application, the air handler is outdoors and needs to be weatherproofed against rain, snow, and, in some areas, sand. In coastal environments, enclosure materials’ resistance (e.g., to salt spray) must also be considered. In cold climates, fuel oil does not atomize and must be warmed to burn properly. Hot-water or steam heating coils and piping must be protected against freezing. In some areas, enclosures are needed to maintain units effectively during inclement weather. A permanent safe access to the roof, as well as a roof walkway to protect against roof damage, are essential.

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2020 ASHRAE Handbook—HVAC Systems and Equipment (SI)

Rooftop units are generally mounted using (1) integral frames or (2) light steel structures. Integral support frames are designed by the manufacturer to connect to the base of the unit. Separate openings for supply and return ducts are not required. The completed installation must adequately drain condensed water. Light steel structures allow the unit to be installed above the roof using separate, flashed duct openings. Any condensed water can be drained through the roof drains. Accessories such as economizers, special filters, and humidifiers are available. Factory-installed and wired economizer packages are also available. Other options offered are return and exhaust fans, variable-volume controls with hot-gas bypass or other form of coil frost protection, smoke and fire detectors, portable external service enclosures, special filters, and microprocessor-based controls with various control options, including BACnet™ integration. For projects with custom-designed equipment, it may be desirable to require additional witnessed factory testing to ensure performance and quality of the final product. Refrigeration Equipment. Large systems incorporate reciprocating, screw, or scroll compressors. Chapter 38 has information about compressors and Chapters 43 and 48 discuss refrigeration equipment, including the general size ranges of available equipment. Air-cooled or evaporative condensers are built integral to the equipment. Air-cooled condensers pass outdoor air over a dry coil to condense the refrigerant. This results in a higher condensing temperature and, thus, a larger power input at peak conditions. However, this peak time may be relatively short over 24 h. The air-cooled condenser is popular in small reciprocating systems because of its low maintenance requirements. Evaporative condensers pass air over coils sprayed with water, using adiabatic saturation to lower the condensing temperature. As with the cooling tower, freeze prevention and close control of water treatment are required for successful operation. The lower power consumption of the refrigeration system and the much smaller footprint from using an evaporative versus air-cooled condenser are gained at the expense of the cost of water used and increased maintenance costs. Heating Equipment. Natural-gas, propane, oil, electricity, hotwater, steam, and refrigerant gas heating options are available. These are normally incorporated directly into the air-handling sections. Custom equipment can also be designed with a separate prepiped boiler and circulating system. Controls. Multiple lower-capacity outdoor units are usually single-zone, constant-volume, or potentially VAV if units are larger capacity. Zoning for temperature control determines the number of units; each zone has a unit. Zones are determined by the cooling and heating loads for the space served, occupancy, allowable roof loads, flexibility requirements, appearance, duct size limitations, and equipment size availability. These units can also serve core areas of buildings, with perimeter spaces served by PTACs. Most operating and safety controls are provided by the equipment manufacturer. Although remote monitoring panels are optional, they are recommended to allow operating personnel to monitor performance. Acoustics and Vibration. Most unitary equipment is available with limited separate vibration isolation of rotating equipment. Custom equipment is available with several (optional) degrees of internal vibration isolation. Isolation of the entire unit casing is rarely required; however, use care when mounting on light structures. If external isolation is required, it should be coordinated with the unit manufacturer to ensure proper separation of internal versus external isolation deflection. Outdoor noise from unitary equipment should be reduced to a minimum. Evaluate sound power levels at all property lines. Indoorradiated noise from the unit’s fans, compressors, and condensers

travels directly through the roof into occupied space below. Mitigation usually involves adding mass, such as two layers of gypsum board, inside the roof curb beneath the unit. Airborne duct discharge noise, primarily from the fans themselves, can be attenuated by silencers in the supply and return air ducts or by acoustically lined ductwork.

9.

SINGLE-ZONE VAV SYSTEMS

These systems distribute cooling and heating to a single zone by supplying constant-temperature airflow at varying volumes. A variable-speed fan controls the quantity of air provided to the space by modulating the fan speed based on space load. The compressor modulates based on the temperature of the supply air leaving the unit to determine the amount of refrigerant flow needed to maintain the supply air set point. Heating is usually electric with silicon controlled rectifier (SCR) control or modulating gas-fired stainless steel heat exchangers. Advantages • Equipment location allows easy service access without maintenance staff entering or disturbing occupied space. • Installation is simple and equipment is readily available in sizes for easy handling. • Installation is simplified and field labor costs are reduced because most components are assembled and tested in a controlled factory environment. • Units are available with complete VAV controls including night setback and morning warm-up. • Valuable building space for mechanical equipment is conserved. • It is suitable for floor-by-floor control in low-rise office buildings. • Outdoor air is readily available for ventilation and economizer cycle use. • Combustion air intake and flue gas exhaust are facilitated if natural gas heat is used. • Tenant utility metering is easier. • Failure of one systems affects a limited area. Disadvantages • Maintaining or servicing outdoor units is sometimes difficult, especially in inclement weather. • Air filtration options are limited. • With all rooftop equipment, safe access to the equipment is a concern. Even slightly sloped roofs are a potential hazard. • Frequent removal of panels for access may destroy the unit’s weatherproofing, causing electrical component failure, rusting, and water leakage. • Rooftop unit design must be coordinated with structural design because it may represent a significant building structural load. • In cold climates, provision must be made to keep snow from blocking air intakes and access doors, and the potential for freezing of hydronic heating or steam humidification components must be considered. • Casing corrosion is a potential problem. Many manufacturers prevent rusting with galvanized or vinyl coatings and other protective measures. • Outdoor installation can reduce equipment life. • Humidification options are impractical on a unit by unit basis. • Depending on building construction, sound levels and transmitted vibration may be excessive. • Architectural considerations may limit allowable locations or require special screening to minimize visual effect. • Multiple units might be required.

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Decentralized Cooling and Heating Design Considerations Design considerations for single-zone VAV are the same as those in the section on Commercial Outdoor Packaged Systems, with the following exceptions:

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• Potential for freezing in cold climates need not be addressed • VAV systems may require separate supply and return duct openings • The discussion of accessories and custom-designed equipment does not apply to VAV systems Factory-installed and wired economizer packages are also available as well as microprocessor-based controls with various control options, including BACnet™ integration. Refrigeration Equipment. Chapter 38 has information about compressors and Chapters 43 and 48 discuss refrigeration equipment, including the general size ranges of available equipment. Air-cooled condensers are built integral to the equipment. Air-cooled condensers pass outdoor air over a dry coil to condense the refrigerant. This results in a higher condensing temperature and, thus, a larger power input at peak conditions. However, this peak time may be relatively short over 24 h. The aircooled condenser is popular in small reciprocating systems because of its low maintenance requirements. Heating Equipment. Natural-gas, propane, electricity, and sometimes hot-water, steam, and refrigerant gas heating options are available. These are normally incorporated directly into the airhandling sections. Controls. Zoning for temperature control determines the number of units; each zone has a unit. Zones are determined by the cooling and heating loads for the space served, occupancy, allowable roof loads, flexibility requirements, appearance, duct size limitations, and equipment size availability. Most operating and safety controls are provided by the equipment manufacturer. Although remote monitoring panels are optional, they are recommended to allow operating personnel to monitor performance. Acoustics and Vibration. Most unitary equipment is available with limited separate vibration isolation of rotating equipment. Isolation of the entire unit casing is rarely required; however, use care when mounting on light structures. If external isolation is required, it should be coordinated with the unit manufacturer to ensure proper separation of internal versus external isolation deflection. Outdoor noise from unitary equipment should be reduced to a minimum. Evaluate sound power levels at all property lines. Indoorradiated noise from the unit’s fans, compressors, and condensers travels directly through the roof into occupied space below. Mitigation usually involves adding mass, such as two layers of gypsum board, inside the roof curb beneath the unit. Airborne duct discharge noise, primarily from the fans themselves, can be attenuated by silencers in the supply and return air ducts or by acoustically lined ductwork.

10.

AUTOMATIC CONTROLS AND BUILDING MANAGEMENT SYSTEMS

A building management system can be an important tool in achieving sustainable facility energy management. Basic HVAC system controls are electric or electronic, and usually are prepackaged and prewired with equipment at the factory. Controls may also be accessible by the building manager using a remote off-site computer. The next level of HVAC system management is to integrate the manufacturer’s control package with the building management system. If the project is an addition or major renovation, prepackaged controls and their capabilities need to be compatible with existing automated

2.11 controls. Chapter 41 of the 2019 ASHRAE Handbook—HVAC Applications discusses computer applications, and ANSI/ASHRAE Standard 135 discusses interfacing building automation systems.

11.

MAINTENANCE MANAGEMENT

Because they are simpler and more standardized than centralized systems, decentralized systems can often be maintained by less technically trained personnel. Maintenance management for many packaged equipment systems can be specified with a service contract from a local service contracting firm. Frequently, small to midlevel construction projects do not have qualified maintenance technicians on site once the job is turned over to a building owner, and service contracts can be a viable option. These simpler decentralized systems allow competitive solicitation of bids for annual maintenance to local companies.

12.

BUILDING SYSTEM COMMISSIONING

Commissioning a building system that has an independent control system to be integrated with individual packaged control systems requires that both control contractors participate in the process. Before the commissioning functional performance demonstrations to the client, it is important to obtain the control contractors’ individual point checkout sheets, program logic, and list of points that require confirmation with another trade (e.g., fire alarm system installer). Frequently, decentralized systems are installed in phases, requiring multiple commissioning efforts based on the construction schedule and owner occupancy. This applies to new construction and expansion of existing installations. During the warranty phase, decentralized system performance should be measured, benchmarked, and course-corrected to ensure the design intent can be achieved. If an energy analysis study is performed as part of the comparison between decentralized and centralized concepts, or lifecycle comparison of the study is part of a Leadership in Energy and Environmental Design (LEED®) project, the resulting month-tomonth energy data should be a good electronic benchmark for actual energy consumption using the measurement and verification plan implementation. Ongoing commissioning or periodic recommissioning further ensures that design intent is met, and that cooling and heating are reliably delivered. Retro- or recommissioning should be considered whenever the facility is expanded or an additional connection made to the existing systems, to ensure the original design intent is met. The initial testing, adjusting, and balancing (TAB) also contributes to sustainable operation and maintenance. The TAB process should be repeated periodically to ensure levels are maintained. When completing TAB and commissioning, consider posting laminated system flow diagrams at or adjacent to cooling and heating equipment indicating operating instructions, TAB performance, commissioning functional performance tests, and emergency shutoff procedures. These documents also should be filed electronically in the building manager’s computer server for quick reference. Original basis of design and design criteria should be posted as a constant reminder of design intent, and to be readily available in case troubleshooting, expansion, or modernization is needed. As with all HVAC applications, to be a sustainable design success, building commissioning should include the system training requirements necessary for building management staff to efficiently take ownership and operate and maintain the HVAC systems over the useful service life of the installation. Commissioning should continue up through the final commissioning report, approximately one year after the construction phase has been completed and the warranty phase comes to an end. For further details on commissioning, see Chapter 44 of the 2019 ASHRAE Handbook—HVAC Applications.

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2020 ASHRAE Handbook—HVAC Systems and Equipment (SI) BIBLIOGRAPHY

ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal.ashrae .org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org/bookstore.

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AHAM. 2015. Room air conditioners. ANSI/AHAM Standard RAC-12015. Association of Home Appliance Manufacturers, Chicago. AHRI. 2017. Packaged terminal air conditioners and heat pumps. Standard 310/380-2017. Air-Conditioning, Heating, and Refrigeration Institute, Arlington, VA.

ASHRAE. 2006. Single zone air handlers and unitary equipment. Ch. 6 in Fundamentals of HVAC systems. ASHRAE. 2007. Air-conditioning system design manual. ASHRAE. 2019. Safety standard for refrigeration systems. ANSI/ASHRAE Standard 15-2019. ASHRAE. 2016. Energy standard for buildings except low-rise residential buildings. ANSI/ASHRAE Standard 90.1-2016. ASHRAE. 2016. BACnet®—A data communication protocol for building automation and control networks. ANSI/ASHRAE Standard 135-2016.

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Related Commercial Resources CHAPTER 3

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CENTRAL COOLING AND HEATING PLANTS System Characteristics................................................................ 3.1 Design Considerations................................................................ 3.2 Equipment ................................................................................... 3.5 Distribution Systems ................................................................... 3.7 Sound, Vibration, Seismic, and Wind Considerations ................ 3.8

Space Considerations.................................................................. 3.8 Automatic Controls and Building Management Systems .......... 3.10 Maintenance Management Systems .......................................... 3.10 Building System Commissioning ............................................... 3.11 System Replacements and Expansions ...................................... 3.11

C

also typically includes cooling towers and pumps, if it is a watercooled plant. See Chapter 1 for information on selecting a central cooling or heating plant. The plant and equipment can be located as part of the facility, or in remote stand-alone plants. Also, different combinations of centralized and decentralized systems (e.g., a central cooling plant and decentralized heating and ventilating systems) can be used. Chillers and boilers are available in different sizes, capacities, and configurations to serve a variety of building applications. Operating a few pieces of primary equipment (often with back-up equipment) gives central plants different benefits from decentralized systems (see Chapter 2). Multiple types of equipment and fuel sources can be combined in one plant, but typically only in large plants. The heating and cooling energy may be a combination of electricity, natural gas, oil, coal, solar, geothermal, waste heat, etc. This energy is converted into chilled water, hot water, or steam that is distributed through the facility for air conditioning, heating, and processes. The operating, maintenance, and first costs of all these options should be discussed with the owner before final selection. When combining heating generation systems, it is important to note the presence of direct-firing combustion systems and chilled-water production systems using refrigerants, because ASHRAE Standard 15 requires most refrigerants to be isolated from combustion equipment for safety. A central plant can be customized without sacrificing the standardization, flexibility, and performance required to support the primary cooling and heating equipment by carefully selecting ancillary equipment, automatic control, and facility management. Plant design can vary widely based on building use, life-cycle costs, operating economies, and the need to maintain reliable building HVAC, process, and electrical systems. These systems can require more extensive engineering, equipment, and financial analysis than do decentralized systems. In large buildings with interior areas that require cooling at the same time perimeter areas require heating, one of several types of centralized heat reclaim units can meet both these requirements efficiently. Chapter 9 describes these combinations, and Chapters 11 to 15 provide design details for central plant systems. Using recovered energy for reheat at the zone level is a common use for lower-grade heat. It can significantly reduce the total energy use of buildings that commonly have simultaneous heating and cooling loads (e.g., hospitals, large hotels). Using recovered energy for reheat is one of the allowed exceptions in ASHRAE Standard 90.1 for simultaneous heating and cooling. Central plants can be designed to accommodate both occupied/ unoccupied and constant, year-round operation. Maintenance can be performed with traditional one-shift operating crews, but may require 24 h coverage. Higher-pressure steam boiler plants (usually greater than100 kPa [gage]) or combined cogeneration and steam heating plants require multiple-operator, 24 h shift coverage.The need for full-time operating personnel or certified operating engineers is typically defined by the authorities having jurisdiction (AHJs).

ENTRAL cooling and/or heating plants generate cooling and/or heating in one location for distribution to multiple locations in one building or an entire campus. Central cooling and heating systems are used in almost all types of buildings, but most commonly in very large buildings and complexes where there is a high density of energy use. This chapter covers plants with cooling and/or heating (referred to as central plants). They are especially suited to applications where maximizing equipment service life and using energy and operational workforce efficiently are important. Good candidates for a central plant have significant loads and a designated mechanical space. Appendix G of ASHRAE Standard 90.1 shows that central plants are typical in buildings more than five floors high or greater than 14 000 m2. Smaller buildings may use them for efficiency and control reasons, but the cost per square foot increases as compared to more typical systems. Building or facility types that commonly use central plants include • Campus environments with distribution to several buildings (described further in Chapter 12). • High-rise facilities • Large office buildings (typically over 14 000 m2 • Large public assembly facilities, entertainment complexes, stadiums, arenas, and convention and exhibition centers • Urban centers (e.g., city centers/districts) • Shopping malls • Large condominiums, hotels, and apartment complexes • Educational facilities • Hospitals and other health care facilities • Industrial facilities (e.g., pharmaceutical, manufacturing) • Large museums and similar institutions • Locations where waste heat is readily available (result of power generation or industrial processes) • Larger systems where higher efficiency offsets the potentially higher first cost of a chilled-water system This chapter addresses design alternatives that should be considered for central cooling and heating plants. Distribution system options and equipment are discussed when they relate to the central equipment, but more information on distribution systems can be found in Chapters 11 to 15.

1.

SYSTEM CHARACTERISTICS

Central systems are characterized by large chilling and heating equipment located in one facility or multiple installations interconnected to operate as one. Equipment configuration and ancillary equipment vary significantly, depending on the facility’s use. Typically, central plants include water-chilling equipment, pumps, and water system specialty items. Boilers are often in a separate mechanical room to provide proper separation from the refrigeration equipment, as required by ASHRAE Standard 15. The central plant The preparation of this chapter is assigned to TC 9.1, Large Building AirConditioning Systems.

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Advantages • Simultaneous primary cooling and heating can be provided at all times, independent of the operation mode of equipment and systems beyond the central plant. • Using larger but fewer pieces of equipment generally reduces the facility’s overall operation and maintenance costs. It also allows wider operating ranges and more flexible operating sequences. • A centralized location minimizes restrictions on servicing accessibility. • Energy-efficient design strategies, energy recovery, thermal storage, and energy management can be simpler and more cost effective to implement. • Multiple energy sources can be used in central plants, providing flexibility and leverage when purchasing fuel and in the event of shortages of one or more fuel sources. • Standardizing equipment can be beneficial for redundancy and stocking replacement parts. However, strategically selecting different-sized equipment for a central plant can provide better part-load capability and efficiency. • Standby capabilities (capacity/redundancy) and back-up fuel sources can be more easily added to equipment and plant when planned in advance. • Equipment operation can be staged to match load profile and can allow select pieces of equipment to be taken offline for maintenance. • A central plant and its distribution can be economically expanded to accommodate future growth (e.g., adding new equipment to the plant or new buildings to the service group). • Load diversity can substantially reduce the total installed equipment capacity requirement. • Submetering secondary distribution can allow individual billing of cooling and heating uses outside the central plant. • Major vibration and sound-producing equipment can be grouped away from occupied spaces, making acoustic and vibration controls simpler. Acoustical treatment can be applied in a single location instead of many separate locations. • Issues such as cooling tower plume and plant emissions are centralized, allowing a more economic and aesthetically acceptable solution.

Disadvantages • New or replacement equipment of the required capacity may not be readily available, resulting in long lead time for production and delivery. • Equipment may be more complicated than decentralized equipment, and thus require more knowledgeable equipment operators. • A central location within or adjacent to the building(s) served is needed. • Additional equipment room height may be needed for the larger central plant equipment. • Depending on the fuel source, large underground or surface storage tanks may be required on site. If coal is used, space for storage bunker(s) will be needed. • Access may be needed for large deliveries of fuel (oil, propane, wood/biomass, or coal). • Fossil-fuel-fired heating plants require a chimney or flue and possibly special emission treatment, permits, and/or monitoring. • Multiple equipment manufacturers are required when combining primary and ancillary equipment in a common facility. • System control logic may be complex. • First costs can be higher compared to alternatives with rooftop units (RTUs), water-source heat pumps (WSHPs), self-contained equipment, and other systems. • Special permitting may be required.

• Safety requirements are increased. • A large pipe distribution system may be necessary (which may actually be an advantage for some applications).

2.

DESIGN CONSIDERATIONS

Cooling and Heating Loads Design cooling and heating loads are determined by considering individual and simultaneous loads. That is, the simultaneous peaks or instantaneous load for all areas served or a block of buildings served by the HVAC and/or process load is less than the sum of the individual peak cooling and heating loads (e.g., buildings do not receive peak solar load on the east and west exposures at the same time). The difference between the sum of the space design loads and system peak load, called the diversity factor, can be as little as 5% less than the sum of individual loads (e.g., 95% diversity factor) or can represent a more significant reduction of the plant load (e.g., 45% diversity factor), as is possible in multiple-building applications. Computerized load calculation programs can be used to model different schedules (occupancy, lighting, etc.) with peak thermal loads, to determine appropriate diversity factors. The peak central plant load is typically based on a diversity factor, reducing the total installed equipment capacity serving larger building cooling and heating loads. The design engineer should evaluate the full point-of-use load requirements of each facility served by the central system. It is important to review applicable codes and requirements of the local authority having jurisdiction (AHJ), which may limit the use of diversity factors. Opportunities for improving energy efficiency include • Using multiple chillers or boilers can provide better part-load operation, with a wider range of operation without cycling off and on. Using correctly sized equipment is imperative to accurately provide the most flexible and economical sequencing of equipment. Modeling programs or spreadsheets are an excellent way to compare different staging sequences. Controls can handle a greater number of sequences, but controls for special staging will have a higher first cost. Installing more units to deliver the same capacity that can be delivered by a single piece of equipment typically increases the cost and space required. This needs to be checked against the staging and redundancy requirements that multiple units can provide. If multiple units are to be used, the designer must verify that the expected load profiles fit the full load capabilities of one, two, or more of the units installed. • Rightsizing equipment allows it to operate closer to peak efficiency, as opposed to incorrectly sized units that are forced to operate at low loads for extended periods of time. It also saves on the cost of installing extra capacity. • For central chiller plants, consider using variable-frequency drives (VFDs) on chillers. Multiple VFD installations on chillers allow more flexibility in control and energy efficiency of chiller plant operation. Remember that part-load efficiency increases when the required lift decreases. This occurs when the condensing temperature is lowered (cooler condensing fluid) or the discharge chilled-water temperature is raised. • Staging cooling towers with the chillers can increase efficiency and reliability by using as much cooling tower surface and fans as possible. Coordinating equipment flow (maximum and minimum at various stages) characteristics is now required by ASHRAE Standard 90.1 (as of 2013). • Using exchangers to transfer free or waste energy from one source to another can provide significant savings. This can also allow mechanical equipment to unload or turn off completely on nonpeak days. Examples include recovering condenser water for reheat, recovering energy from a process load to provide free heat, or preheating domestic or laundry water.

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Central Cooling and Heating Plants Discrete loads (e.g., server rooms) may be best served by an independent system, depending on the load profile of the central plant; a small, independent system, designed for the discrete load, can be the most cost-effective approach to handle the 24/7 server room load while allowing the primary HVAC system to shut off for a significant number of unoccupied hours. Central plants may not be able to reliably serve a small, discrete load. For example, an 8 MW chiller plant serving a large office facility could be selected with two 4 MW chillers. An independent computer room server has a year-round 80 kW load. Operating one 4 MW chiller, plus towers and pumps, at less than 500 kW reliably would be extremely inefficient and possibly detrimental. Serving the discrete load independently allows the designer flexibility to evaluate the chiller plant independent of individual subsystems that may have very different schedules from the main plant. Peak cooling load time is affected by outdoor ventilation, outdoor dry- and wet-bulb temperatures, hours of occupancy, interior equipment heat gain, and relative amounts of north, east, south, and west exposures. For typical office or classroom buildings with a balanced distribution of solar exposures, the peak usually occurs on a midsummer afternoon when the west solar load and outdoor dry-bulb temperature are at or near concurrent maximums. The peak cooling period can shift if ventilation rates and internal load profiles change significantly, as in a restaurant or high-rise residential building. Building occupancy diversity can significantly affect the overall diversity factor. For example, in a system serving a high school, the peak cooling period for a classroom is different from that for the administration offices. Planning load profiles at academic facilities requires special consideration. Unlike office and residential applications, educational facilities typically have peak cooling loads during late summer and early fall. Peak heating load has less opportunity to accommodate a diversity factor, because equipment is most likely to be selected based on the sum of individual heating loads when the spaces are not being used. This load may often occur when the building must be warmed after an unoccupied weekend setback period. Peak demand might occur during unoccupied periods when outdoor temperatures are coldest and there is little internal heat gain, or during occupied times if a significant amount of outdoor air must be preconditioned. To accommodate part-load conditions and energy efficiency, variable flow may be the best economical choice. It is important for the designer to evaluate plant operation and system use. Boiler heating loads can typically be handled with variable flow by using primary/secondary pumping. Some boilers (e.g., condensing) can also be used in variable primary pumping arrangements. Chapter 32 discusses this in greater detail.

3.3 System Flow Design The flow configuration of a central system is based on use and application. Primary variable flow uses variable flow through the primary equipment (chiller or heating device) and directly pumps the water to the loads. Variable primary flow typically uses twoway automatic control valves at terminal equipment and variablefrequency drive (VFD) pumping (Figure 1). An older version of variable flow uses distribution pressure control with a bypass valve (Figure 2). This option keeps full flow in the chiller or boiler, but sacrifices the savings of the VFD on the pumps. Both methods control system pressure, usually at the hydraulically most remote point (last control valve and terminal unit) in the water system. The design engineer should work with the equipment manufacturer to ensure minimum equipment flow rates are maintained throughout the building load profile, and to determine whether any additional components or ancillary equipment is required. Chillers are typically selected with higher water velocities (0.03 to 0.04 m/ s) and water pressure drops at full load in the evaporator to provide improved turndown on the water flow while maintaining the required minimum equipment flow. The chilled-water pumps are not dedicated to specific chillers/boilers, but run off of a common header to allow the chillers to operate over a wide range of flows (Figure 1). The valve in the bypass line is closed until system flow drops to the minimum flow of the operating chiller/boiler(s). Flow transitions when staging a chiller off and on are the most critical. Check each chiller’s minimum and maximum flow to ensure that the pumps can maintain the required flow rate range from minimum to 100% load. Multiple chillers or those with higher turndown ratios make this easier. Avoid variable primary systems where there are poor controls or limited operator training. Constant primary/variable secondary flow hydraulically decouples the primary production system (chilled- or heatingwater), which is commonly constant flow, from the secondary dis-

Fig. 1 Primary Variable-Flow System

Security The designer needs to evaluate the security requirements as part of the design, commissioning, and maintenance plans for the building. The team should evaluate the building for security risks in terms of intentional (terrorist) acts, accidental events, and natural disasters. Central plants can make it easier to isolate parts of the HVAC system, but components outside the mechanical room or outside the building (e.g., air intakes, power and controls connections) may be of concern. Special sequences of operation can provide ventilation and isolation during a security event. HVAC and power redundancy, increased filtration, and sensors are typically included in the security plan. Chapter 61 of the 2019 ASHRAE Handbook—Applications has more detail on developing design, plans, and planned responses to these types of events. Depending on the building function, location, and occupants, the security plan may change. It is very important to remember that early planning and the use of commissioning can reduce the risk of harm to occupants.

Fig. 2 Primary (Limited) Variable-Flow System Using Distribution Pressure Control (Courtesy RDK Engineers)

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Fig. 3 Primary/Secondary Pumping Chilled-Water System (Courtesy RDK Engineers)

tribution. A variable-flow secondary piping system distributes the chilled or heating medium to the point of use (Figures 3 and 4). This is probably the most common pumping flow arrangement in use, and allows for easy redundancy with all primary components and easy isolation of units as they cycle off or need to be serviced. Primary/secondary pumping maintains constant design flow in each chiller (or boiler) during operation. The secondary (load) pumps vary their flow rates to match system demand by maintaining a set pressure in the loop. A decoupler line between the primary and secondary loops allows the two sets of pumps to operate at different flow rates. The constant-flow primary helps because many chillers and boilers require full flow or are limited in the amount of flow reduction they can handle. The primary pumps are sized for the chiller flows and pressure drops associated with the central plant. The secondary pumps are selected for the building flows and their pressure drops, plus the secondary distribution system, which typically exceeds that of the primary loop. This allows for saving on the largest part of the pumping energy by varying the flows on the building loads. Improved chiller controls and lower VFD costs have made VFDs on primary pumps a flexible and economical way to tune chiller flow without trimming pump impellers or using a throttling valve. Condenser pumps generally need less head capability than the secondary pumps because the towers are typically close. VFDs can be used on these pumps, but they are typically smaller pumps. The designer should be careful when reducing flow on the towers, which can increase the return tower water temperature and chiller energy use. Also, varying flow across the cooling tower can greatly affect tower performance. In Figure 3, multiple tower cells have been selected to match the number of chillers. Note that running 50% of the water through both cells with each fan at 50% speed uses less energy than one tower at 100% flow and fan speed. The designer should confirm that the towers can handle reduced flow (fairly common with newer towers).

Fig. 4 Primary/Secondary Pumping Hot-Water System (Courtesy RDK Engineers)

When using either primary/secondary or primary variable-flow designs, the engineer should understand the design differences between the use of two- and three-way modulating valves (see Chapter 46). Variable-flow designs modulate chilled- or heating-water flow through the distribution loop. As terminal units satisfy demand, a two-way valve modulates toward closed, causing the system pressure to increase above the design point. To compensate, a VFD reduces pump speed, thereby reducing flow and pressure to the control point. Pump speed is usually controlled by a differential pressure sensor. Other recent technologies sense valve positions throughout the system and optimize flow accordingly. As the pressure differential increases (the result of valves closing), the VFD reduces speed to maintain the pressure set point. As system

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Central Cooling and Heating Plants

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demand requires increased flow, the control valves modulate open, reducing system differential pressure. The reduction in differential pressure causes the VFD to increase pump speed to maintain the pressure set point and meet system flow demand. Care should be taken during design to ensure minimum flow is achieved in the secondary loop under part-load operation throughout the system. When selecting the distribution system pump, ensure distribution flow does not exceed production flow, which could cause a temperature drop in the secondary loop if flow exceeds demand, leading to low T syndrome (see, e.g., Taylor [2002]). The engineer should evaluate operational conditions at part load to minimize the potential for this to occur With primary variable-flow, a single set of pumps modulates the primary chilled- or heating-water flow, using a VFD to control pump speed. Primary variable-flow systems use modulating twoway control valves across each heat transfer device (as in the primary/secondary system). Chillers and boilers have minimum water flow rates that must be maintained to ensure proper operation. A bypass or other method to maintain minimum flows must be installed if the minimum system loads can be less than the minimum equipment limits. An older design, which has been used for many years, is a straight constant-volume primary system. Hydronic pumps distribute water through the energy equipment and to the point of use. Both pumping energy and distribution flow are constant. This design is best suited to stable process loads, or may be seen in older systems that were installed before VFDs were common. Constant flow generally requires three-way valves to maintain the minimum required flow rate through generating equipment and not over pump the coils and loads. These systems are usually much more expensive to operate and, therefore, less attractive in central HVAC systems. Take care with this type of design, because many current energy codes (including ASHRAE Standard 90.1) do not allow using pump motors without variable-frequency drives.

(CHP) system or other energy-using system that would normally reject heat to the atmosphere. Chapters 38 and 43 discuss refrigeration equipment, including the size range of available equipment. Compressorized chillers feature scroll, helical rotary (screw), and centrifugal compressors, which may be driven by electric motors; natural gas-, diesel-, or oil-fired internal combustion engines; combustion turbines; or steam turbines. Compressor selection typically depends on capacity, first-cost budget limitations, and life-cycle cost. Compressors are a part of the chiller that also includes the evaporator, condenser, safety and operating controls, and possibly a VFD to control the compressor driver. Reciprocating and helical rotary compressor units can be field assembled and include air- or water-cooled (evaporative) condensers designed for remote installation. Centrifugal compressors are usually included in larger packaged chillers, though they can be field erected on very-large-capacity chiller systems. Chillers require heat rejection equipment that can be unit mounted or remote, as well as air or water cooled. Chapter 38 has more detailed information about compressors. Absorption chillers may be single or double effect. These terms refer to the number of times the solution is distilled, with double effect typically being higher capacity and more efficient, but using more energy than single effect. Like centrifugal chillers, absorption chillers require cooling towers to reject heat. Absorption chillers use a lithium bromide/water cycle in which the lithium bromide/ water solution is the refrigerant. They are generally available in the following configurations: (1) natural gas direct fired, (2) indirect generated by low pressure steam or hot water, (3) indirect generated by high-pressure steam or high-temperature hot water, and (4) indirect generated by hot exhaust gas. Chapter 18 of the 2018 ASHRAE Handbook—Refrigeration discusses absorption air-conditioning and refrigeration equipment in more detail.

Energy Recovery and Thermal Storage

Ancillary equipment for central cooling plants consists primarily of heat-rejection equipment (air-cooled condensers, evaporative condensers, and cooling towers), pumps (primary, secondary, and tertiary), and heat exchangers (water-to-water). For more detailed information, see Chapters 26, 39, 40, and 42 to 47. As part of a packaged air-cooled chiller, air-cooled condensers pass outdoor air over a dry coil to condense the refrigerant. This can result in a higher condensing temperature than a water-cooled condenser would provide, and thus a higher power input at peak condition (though peak time may be relatively short over a 24 h period). Newer designs for air-cooled chillers can offer larger condenser surfaces to provide efficiencies closer to those of watercooled units without the requirement for a cooling tower. Aircooled condensers are popular with scroll and screw compressors, which can handle a wide range of ambient air temperatures. They generally have lower maintenance requirements, but incur a penalty for using the dry-bulb temperature as opposed to the wet-bulb temperature used by water-cooled equipment. Air-cooled equipment may be appropriate in high-wet-bulb environments, because cooling tower supply water temperature is a function of the wetbulb temperature. They are also more common in smaller systems and in cooler climates that require more freeze protection and have lower peak temperatures. Evaporative condensers pass outdoor air over coils sprayed with water, thus taking advantage of the heat that is absorbed in the evaporation process to lower the condensing temperature. Freeze prevention and close control of water treatment are required for successful operation. The lower power consumption of the refrigeration system and much smaller footprint of the evaporative condenser are gained at the expense of the cost of water and water treatment used and increased maintenance cost.

Energy recovery and thermal storage strategies can be applied to a central cooling and heating plant. See ASHRAE Standard 90.1 for systems and conditions requiring energy recovery. Water-to-water energy recovery systems are common and readily available. Thermal storage using water or ice storage can be adapted to central plants. See Chapter 26 in this volume and Chapters 35 to 37 and 42 in the 2019 ASHRAE Handbook—HVAC Applications for more information on energy-related opportunities. Thermal energy storage (TES) systems may offer strategies for both part-load and peak-load energy reduction. Base-loading plant operation for a more flat-line energy consumption profile of a large plant, for example, may help reduce energy costs. Thermal energy storage of chilled water, ice, or heating water offers a medium for capacity redundancy, with potential reductions in both heating and cooling infrastructure equipment sizing.

3.

EQUIPMENT

Primary Refrigeration Equipment Chillers are the major cooling equipment in central plants. They chill water or other low-temperature fluid, which is then pumped throughout the buildings served. Chillers vary in type and application, and fall into two major categories: (1) vapor-compression refrigeration (compressorized) chillers, and (2) absorption-cycle chillers. The chiller plant may have either, or a combination of, these machine types. Cooling towers, air-cooled condensers, evaporative condensers, or some combination are also needed to reject heat from this equipment. Energy for the prime driver of cooling equipment may also come from waste heat from a combined heat and power

Ancillary Refrigeration Equipment

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Cooling towers provide heat rejection by passing outdoor air through an open condenser return water spray to achieve adiabatic cooling performance. Either natural or mechanical-draft cooling towers or spray ponds can be used; the mechanical-draft tower (forced draft, induced draft, or ejector) can be designed for most conditions because it does not depend on wind. Cooling tower types and sizes range from packaged units to field-erected towers with multiple cells in unlimited sizes. Location of cooling towers should consider issues such as reingestion or short-circuiting of discharge heat rejection air, tower plume, proximity to outdoor air intakes, and the effect of drift on adjacent roadways, buildings, and parking lots. As with evaporative condensers, freeze protection may be required. Water treatment is necessary for successful operation to prevent scaling, biological growth (e.g., Legionella), and to prevent corrosion. More detail on cooling towers can be found in Chapter 14. Makeup water subtraction meters should be evaluated for both evaporative condensers and cooling towers by the design engineer or owner. In many areas, sewage costs are part of the overall water bill, and most domestic water supplied to a typical facility goes into the sewage system. However, evaporated condenser water is not drained to sewer, and if a utility-grade meter is used to measure the difference between makeup and blowdown water quantity, potential savings can be available to the owner for water not discharged into the sewer. Metering should be evaluated with chilled-water plants, especially those in operation year-round. Consult with the water supplier to identify compliance requirements. Consider piping cooling tower water blowdown and drainage (e.g., to remove dissolved solids or allow maintenance when installing subtraction meters) to the storm drainage system, but note that environmental effects of the water chemistry must be evaluated before taking this step. Where chemical treatment conditions do not meet environmental outfall requirements, drainage to the sanitary system may still be required unless prefiltration can be incorporated. Water pumps move both chilled and condenser water to and from the refrigeration equipment and associated ancillary equipment. Different pumping arrangements can provide operational savings associated with variable speed pumps, but ways to address application issues such as equipment staging or low T syndrome should be considered. See Chapter 44 for additional information on centrifugal pumps, and Chapters 12 to 14 for system design. Pumps should be headered to allow them to operate with any (as shown in Figures 1 to 3). Dedicating pumps to a chiller boiler or tower cell reduces redundancy and limits flow options to equipment. Heat exchangers provide both operational and energy recovery cost-saving opportunities for central cooling plants. Operational opportunities often involve heat transfer between building systems that must be kept separated because of different pressures, media, cleanliness, etc. For example, heat exchangers can thermally link a low-pressure, low-rise building system with a high-pressure, highrise building system. Heat exchangers can also transfer heat between chemically treated or open, contaminated water systems and closed, clean water systems (e.g., between a central cooling plant chilled-water system and a highly purified, process cooling water system, or potentially dirty pond water and a closed-loop condenser water system). To conserve energy, water-to-water heat exchangers can provide water-side economizer opportunities. When outdoor conditions allow, condenser water can cool chilled water through a heat exchanger, using the cooling tower and pumps rather than a compressor. This approach should be considered when year-round chilled water is needed to satisfy a process load, or when an airside economizer is not possible because of design requirements or limitations (e.g., space humidity requirements, lack of space for

the required ductwork).Water-side economizers can also save on humidification costs in cold, dry climates by reducing the percentage of cold, dry outdoor air delivery to the building from air-side economizers. Plant Controls. Direct digital control (DDC) systems should be considered for control accuracy and reliability. Temperature, flow, and energy use are best measured and controlled with modern DDC technology. Electric actuation is most common. Pneumatic actuation might be considered where torque or rapid response of powered actuation is required; medium-pressure air supply (typically 200 to 400 kPa [gage]) is best. Consider using programmable logic controllers (PLCs) for plants where future expansion/growth is a possibility and changes to DDC controls would be cost prohibitive. Smaller plants can use the onboard chiller control panel, providing a simplified control system at a lower installed cost. Codes and Standards. Specific code requirements and standards apply when designing central cooling plants. For cooling equipment, refer to ASHRAE Standard 15; Chapter 51 of this volume provides a comprehensive list of codes and standards associated with cooling plant design, installation, and operation. ASHRAE Standard 90.1, the International Energy Conservation Code® (IECC®; ICC [2018]), or local energy standards typically govern minimum equipment efficiencies, acceptable control sequences, economizer use, and other details. Follow manufacturers’ recommendations and federal, state, and local codes and standards.

Primary Heating Equipment Boilers are the major heating equipment used in central heating plants. They vary in type and application, and include combined heat and power (CHP) equipment and waste heat boilers. Chapter 32 discusses boilers in detail, including the size ranges of typical equipment. A boiler adds heat to the working medium, which is then distributed throughout the building(s) and/or campus. The working medium may be either water or steam, which can further be classified by temperature and pressure range. Steam, often used to transport energy over long distances, is converted to low-temperature hot water in a heat exchanger near the point of use. Although steam is an acceptable medium for heat transfer, low-temperature hot water is the most common and a more uniform and more easily controlled medium for providing heating and process heat (e.g., heating water to 93°C for heating and up to 60 to 71°C for domestic hot water). Elevated steam pressures and high-temperature hot-water boilers are also used. Both hot-water and steam boilers have the same type of construction criteria, based on operating temperature and pressure. A boiler may be purchased as a package that includes the burner, fire chamber, heat exchanger section, flue gas passage, fuel train, and necessary safety and operating controls. Cast-iron and water-side boilers can be field-assembled, but fire-tube, scotch marine, and waste-heat boilers are usually packaged units. Energy Sources. The energy used by a boiler (i.e., its fuel source) may be electricity, natural gas or propane, oil, coal, or combustible waste material, though natural gas and fossil-fuel oil (No. 2, 4, or 6 grade) are most common, either alone or in combination. Selecting a fuel source requires a detailed analysis of energy prices and availability (e.g., natural gas and primary electrical power). The availability of fuel oil or coal delivery is affected by road access and storage for deliveries. Energy for heating may also come from waste heat from a CHP system. Plant production efficiency is improved when waste heat can be reclaimed and transferred to a heat source. A heat recovery generator can be used to convert the heat byproduct of electric generation to steam or hot water to meet a facility’s heating needs

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Central Cooling and Heating Plants when the thermal load meets or exceeds the heat rejection capacity of the electric generation. This can help with cost-effective plant operation. Additionally, attention to the design of an efficient and maintainable condensate return system for a central steam distribution plant is important. A well-designed condensate return system increases overall efficiencies and reduces both chemical and makeup water use. Codes and Standards. Specific code requirements and standards apply when designing central heating boiler plants. It is important to note that for some applications (e.g., high-pressure boilers) and in some locations, continuous attendance by licensed operators is required for the area of the heating surface. Operating cost considerations should be included in determining such applications. Numerous codes, standards, and manufacturers’ recommendations need to be followed. Refer to Chapter 51 for a comprehensive list of codes and standards associated with this equipment, plant design, installation, and operation, and consult local or state fuel gas code requirements.

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Ancillary Heating Equipment Steam Plants. Boiler feed equipment, including the receiver(s) and associated pump(s), serve as a reservoir for condensate and makeup water waiting to be used by the steam boiler. The boiler feed pump provides system condensate and water makeup back to the boiler on an as-needed basis, at the delivery pressure of the steam supply. Deaerators help eliminate oxygen and carbon dioxide from the feed water or steam to reduce the corrosive and detrimental effects of those contaminants (see Chapter 50 of the 2019 ASHRAE Handbook—HVAC Applications for more information). Chemicals can be fed using several methods or a combination of methods, depending on the chemical(s) used (e.g., chelants, amines, oxygen scavengers). Surge tanks are also applicable to steam boiler plants to accommodate large quantities of condensate water return and are primarily used where there is a rapid demand for steam (e.g., morning start-up of a central heating plant). See Chapter 11 for more information on steam systems. Hot-Water Plants. Ancillary equipment associated with central hot-water heating plants consists primarily of pumps and possibly heat exchanger(s). Water pumps move hot-water supply and return to and from the boiler equipment and associated ancillary equipment. See Chapter 44 for additional information on centrifugal pumps, as well as Chapters 11 to 13 and Chapter 15 for system design. Heat Exchangers. Heat exchangers offer operational and energy recovery opportunities for central heating plants. Operational opportunities for heat exchangers involve combining steam system heating delivery capabilities with hot-water heating system supply and control capabilities. Air-to-water and water-to-water heat exchangers provide opportunities for economizing and heat recovery in a central heating plant (e.g., flue gas exhaust heat recovery and boiler blowdown heat recovery). For more detailed information on ancillary equipment, see Chapters 31 to 33 and Chapter 35.

4.

DISTRIBUTION SYSTEMS

The major piping in a central cooling plant can include, but is not limited to, chilled-water, condenser water, city water, natural gas, fuel oil, refrigerant, vent, and drainage systems. For a central steam heating plant, the major piping includes steam supply, condensate return, pumped condensate, boiler feed, city water (and/or softened or otherwise treated water), natural gas, fuel oil, vent, and

3.7 drainage systems. For a central hot-water heating plant, it includes hot-water supply and return, city water, natural gas, fuel oil, vent, and drainage systems. In the 2017 ASHRAE Handbook—Fundamentals, see Chapter 22 for information on sizing pipes, and Chapter 38 for identification, color-coding, abbreviations, and symbols for piping systems. Design selection of cooling and heating temperature set points (supply and return water) can affect first and operating costs. Water systems with a large temperature difference between supply and return water have lower flow requirements and can allow smaller pipe sizing and smaller valves, fittings, and insulation, which can lower installation cost. However, these savings may be offset by the larger coils and heat exchangers at the point of use needed to accomplish the required heat transfer. A similar design strategy can be achieved by reviewing the steam pressure differential. Determining the optimum cooling and heating water supply and return temperatures requires design consideration of equipment performance, particularly the energy required to produce the supply water temperature. Although end users set water temperatures, the colder the cooling water, the more energy is needed by the chiller, and conversely, the warmer the return water, the less energy is consumed by the chiller. Similar issues affect hot-water supply temperature and steam operating pressure. Supply water reset may be used when peak capacity is not needed, potentially reducing energy consumption. This reduces chiller power input, but those savings may be offset by increased pump power input due to higher water flows required to achieve the needed capacity. Also, raising chilled-water temperatures diminishes the latent removal capacity of cooling and dehumidification equipment, and may adversely affect control of interior humidity. Energy implications for the whole system must be considered. The design engineer should consider using higher temperature differences between supply and return to reduce the pump energy required by the distribution system, for both heating and chilledwater systems. Additionally, central plant production systems (e.g., boilers, steam-to-heating-water converters, chillers) operate more efficiently with higher return water temperatures. During conceptual planning and design, or as early in design as possible, the engineer should evaluate the type(s) of facilities (existing and planned future) to which the central system will deliver service. When using variable-flow distribution with a constant design temperature differential between the supply and return medium, it is critical to avoid low T syndrome (see, e.g., Taylor [2002]). The engineer should attempt to ensure this condition is minimized or avoided, or at minimum, perform due diligence to make the owner aware of the potential shortfalls where this is allowed to occur. With existing constant-flow/variable-temperature systems, take measures to avoid losing the conceptual design strategy of a variable-flow/constant-temperature split. Examples of connection strategies that are less costly than full renovation of a connected facility are (1) return recirculation control, to maintain a design temperature difference across a facility connection, or (2) separation of primary distribution supply from the secondary facility connection by a plate-and-frame heat exchanger with a control valve on the primary return controlled by the secondary supply to the facility. If reduction of temperature differential is allowed, increased flow of the medium through the distribution system will be required to accomplish the same capacity. This increase in flow increases pump power and chiller plant energy, compromising the available capacity and operation of the system. Consider hydraulically modeling the cooling and heating media (chilled water, heating water, steam, domestic water, natural gas, etc.). With an emphasis on centralizing the source of cooling and heating, a performance template can be created for large plants by computerized profiling of cooling and heating delivery. Hydraulic

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modeling provides the economic benefits of predesigning the built-out completed system before installing the initial phase of the project. The model also helps troubleshoot existing systems, select pumps, project energy usage, and develop operation strategy. Energy conservation and management can best be achieved with computerized design and facility management resources to simulate delivery and then monitor and measure the actual distribution performance.

5.

SOUND, VIBRATION, SEISMIC, AND WIND CONSIDERATIONS

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Sound and Vibration Proper space planning is key to sound and vibration control in central plant design. For example, central plants are frequently located at or below grade. This provides a very stable platform for vibration isolation and greatly reduces the likelihood of vibration being transmitted into the occupied structure, where it can be regenerated as sound. Also, locating the central ventilation louvers or other openings away from sound- and vibration-sensitive areas greatly reduces the potential for problems in those areas. As a guide, maintain free area around louvers as specified in ASHRAE Standard 15. Vibration and sound transmitted both into the space served by the plant and to neighboring buildings and areas should be considered in determining how much acoustical treatment is appropriate for the design, especially if the plant is located near sensitive spaces such as conference rooms, sleeping quarters, or residences. See the section on Space Considerations for further discussion of space planning. Acoustics must also be considered for equipment outside the central plant. For example, roof-mounted cooling tower fans sometimes transmit significant vibration to the building structure and generate ambient sound. Many communities limit machinery sound pressure levels at the property line, which affects the design and placement of equipment. See Chapter 49 of the 2019 ASHRAE Handbook—HVAC Applications for more detailed information on sound and vibration.

Seismic and Wind Issues Depending on code requirements and the facility’s location with respect to seismic fault lines, seismic bracing may be required for the central plant equipment and distribution systems. For instance, a hospital located in a seismically active area must be able to remain open and operational to treat casualties in the aftermath of an earthquake. Additionally, outdoor equipment in some areas needs appropriate bracing for wind loads. This is particularly apparent in critical operations facilities in areas susceptible to hurricanes or tornadoes. Most building codes require that measures such as anchors and bracing be applied to the HVAC system. Refer to the local authority responsible for requirements, and to Chapter 56 of the 2019 ASHRAE Handbook—HVAC Applications for design guidance.

6.

SPACE CONSIDERATIONS

In the very early phases of building design, architects, owners, and space planners often ask the engineer to estimate how much building space will be needed for mechanical equipment. The type of mechanical system selected, building configuration, and other variables govern the space required, and many experienced engineers have developed rules of thumb to estimate the building space needed. Although few buildings are identical in design and concept, some basic criteria apply to most buildings and help approximate final space allocation requirements. These space

requirements are often expressed as a percentage of the total building floor area; the combined mechanical and electrical space requirement of most buildings is 6 to 9% of total building area. Space for chillers, pumps, and towers should not only include installation footprints but should also account for adequate clearance to perform routine and major maintenance. Generally, 1.2 m service clearance (or the equipment manufacturer’s minimum required clearance, whichever is greater) around equipment for operator maintenance and service is sufficient. For chillers and boilers, one end or side of the equipment should be provided with free space, to allow for tube pull clearance. In many cases, designers provide service bay roll-up doors or removable ventilation louvers (if winter conditions do not cause freeze damage issues) to allow tube access. Overhead service height is also required, especially where chillers are installed. Provision for 6 m ceilings in a central plant is not uncommon, to accommodate piping, component removal, and service clearance dimensions. Plant designs incorporating steam supply from a separate central boiler plant may use steam-to-heating-water converters and the heating distribution equipment (e.g., pumps) along with the chilled-water production equipment and distribution infrastructure. Where a boiler installation and associated heating distribution equipment and appurtenances are required, the plant’s physical size increases to account for the type of boiler and required exhaust emissions treatment. Generally, for central heating plants, steam or heating water and chilled-water production systems must be separated during design, which may further increase the overall footprint of the plant. (Refer to ASHRAE Standard 15 for limitations.) The arrangement and strategic location of the mechanical spaces during planning affects the percentage of space required. For example, the relationship between outdoor air intakes and loading docks, exhaust, and other contaminating sources should be considered during architectural planning. The final mechanical room size, orientation, and location are established after discussion with the architect and owner. The design engineer should keep the architect, owner, and facility engineer informed, whenever possible, about the HVAC analysis and system selection. Space criteria should satisfy both the architect and the owner or owner’s representative, though this often requires some compromise. The design engineer should strive to understand the owner’s needs and desires and the architect’s vision for the building, while fully explaining the advantages, disadvantages, risks, and rewards of various options for mechanical and electrical room size, orientation, and location. All systems should be coordinated during the space-planning stage to safely and effectively operate and maintain the central cooling and heating plant. In addition, the mechanical engineer sometimes must represent other engineering disciplines in central plant space planning. If so, it is important for the engineer to understand the basics of electrical and plumbing required for central plant equipment. The main electrical transformer and switchgear rooms should be located as close to the incoming electrical service as practical. The main electric transformers and switchgear for the plant and the mechanical equipment switchgear panels should be in separate rooms that only authorized electricians can enter. If there is an emergency generator, it should be located considering (1) proximity to emergency electrical loads, (2) sources of combustion and cooling air, (3) fuel sources, (4) ease of properly venting exhaust gases outside, and (5) provisions for sound control. The main plumbing equipment usually contains gas and domestic water meters, the domestic hot-water system, the fire protection system, and elements such as compressed air, special gases, and vacuum, sewage ejector, and sump pumps. Some water and gas utilities require a remote outdoor meter location. The heating and air-conditioning equipment room houses the (1) boiler, pressure-reducing station, or both; (2) refrigeration

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Central Cooling and Heating Plants machines, including chilled-water and condensing-water pumps; (3) converters for furnishing hot or cold water for heating and/or air conditioning; (4) control air compressors, if any; (5) vacuum and condensate pumps; and (6) miscellaneous equipment. For both chillers and boilers, especially in a centralized application, full access is needed on all sides (including overhead) for extended operation, maintenance, annual inspections of tubes, and tube replacement and repair. Where appropriate, the designer should include overhead structural elements to allow safe rigging of equipment. Ideally, large central facilities should include overhead cranes or gantries. It is critical to consult local codes and ASHRAE Standard 15 for special equipment room requirements. Many jurisdictions require monitoring, alarms, evacuation procedures, separating refrigeration and fuel-fired equipment, and high rates of purge ventilation. A proper operating environment for equipment and those maintaining it must also be provided. This may involve heating the space for freeze protection and/or cooling the space to prevent overheating motors and controls. HVAC equipment serving the central plant itself may be housed in its own equipment room and serve the chiller, boiler room, and adjacent rooms (e.g., switchgear room, office space, generator room, pump room, machine shop). Where refrigeration equipment is installed, follow design requirements in ASHRAE Standard 15.

Location of Central Plant and Equipment Although large central plants are most often located at or below grade, it is often economical to locate the refrigeration plant at the top of the building, on the roof, or on intermediate floors. In floodprone areas, an above-grade installation can provide necessary protection against submersion of critical equipment. For central plants, on-grade should be the first choice, followed by below grade. The designer must always ensure access for maintenance and component or total replacement. Locating major equipment in roof penthouses may pose significant access problems for repair and replacement. For single high-rise central plants, intermediatefloor locations that are closer to the load may allow pumping equipment to operate at a lower pressure. A life-cycle cost analysis (LCCA) should include differences in plant location to identify the most attractive plant sustainability options during the planning stage. The LCCA should consider equipment maintenance access, repair, and replacement during the life of the plant, as determined in the owner’s criteria. If not identified, an engineer should suggest to the owner or client that maintenance criteria be included as a component in developing the LCCA (see the section on Maintenance Management Systems). Electrical service and structural costs are greater when intermediate-floor instead of ground-level locations are used, but may be offset by reduced energy consumption and condenser and chilled-water piping costs. The boiler plant may also be placed on the roof, eliminating the need for a chimney through the building. Benefits of locating the air-cooled or evaporative condenser and/or cooling tower on the ground versus the roof should be evaluated. Personnel safety, security, ambient sound, and contamination from hazardous water vapors are some of the considerations that help determine final equipment location. Also, structural requirements (e.g., steel to support roof-mounted equipment, or a concrete pad and structural steel needed to locate equipment at or near grade) require evaluation. When locating a cooling tower at or near grade, the net positive suction head on the pump suction and overflow of condenser water out of the cooling tower sump should be studied if the tower is below, at, or slightly above the level of the condenser water pump or chiller. Numerous variables should be considered when determining the optimum location of a central cooling or heating plant. When locating the plant, consider the following:

3.9 • Operating mass of the equipment and its effect on structural costs • Vibration from primary and ancillary equipment and its effect on adjacent spaces in any direction • Sound levels from primary and ancillary equipment and their effect on adjacent spaces in any direction • Location of electrical utilities for the central plant room, including primary electric service and associated switchgear and motor control center, as well as electrical transformer location and its entrance into the building • Location of city water and fire pump room (it may be desirable to consolidate these systems near the central plant room) • Accessibility into the area and clearances around equipment for employee access, equipment and material delivery, and major equipment replacement, repair, scheduled teardown, and rigging • Location of cooling, refrigerant relief piping, heating, vents, and boiler flue and stack distribution out of the central plant and into the building, along with the flow path of possible vented hazardous chemical, steam, or combustion exhaust products • Need for shafts to provide vertical distribution of cooling and heating services in the building • Future expansion plans of the central plant (e.g., oversizing the central plant now for adding more primary equipment later, based on master planning of the facility) • Architectural effect on the site • Location of boiler chimney/flue • Loading dock for materials and supplies • Roadway and parking considerations • Storage of fuel • Electrical transformer location • Underground and/or overhead utility and central cooling and heating system distribution around the central plant and to the building(s) • Wind effect on cooling tower plume or other volatile discharges such as boiler emissions.

Central Plant Security Security around a central plant must be considered in design, and should include standby electrical power generation and production of central cooling and heating for critical applications in times of outages or crisis. Restricted access and proper location of exposed intakes and vents must be designed into the central plant layout to protect the facility from attack and protect people from injury. Use care in locating exposed equipment, vents, and intakes, especially at ground level. Above ground and at least 3 to 4.5 m from access to intake face is preferred for the location of intakes. When this is not possible, fencing around exposed equipment, such as cooling towers and central plant intakes, should be kept locked at all times to prevent unauthorized access. Ensure that fencing is open to airflow so it does not adversely affect equipment performance. Air intakes should be located above street level if possible, and vents should be directed so they cannot discharge directly on passing pedestrians or into an air intake of the same or an adjacent facility.

7.

AUTOMATIC CONTROLS AND BUILDING MANAGEMENT SYSTEMS

One advantage of central cooling and heating plants is easier implementation of building automation because the major and ancillary equipment is consolidated in one location. Computerized automatic controls can significantly affect system performance. A facility management system to monitor system points and overall system performance should be considered for any large, complex air-conditioning system. This allows a single operator to monitor performance at many points in a building and make adjustments to increase occupant comfort and to free maintenance staff for other

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duties. Chapter 48 of the 2019 ASHRAE Handbook—HVAC Applications describes design and application of controls. Software to consider when designing, managing, and improving central plant performance should include the following:

The coefficient of performance (COP) for the entire chilledwater plant can be monitored and allows the plant operator to determine the overall operating efficiency of a plant. Central plant COP can be expressed in the following terms:

• Automatic controls that can interface with other control software (e.g., equipment manufacturers’ unit-mounted controls) • Energy management system (EMS) control • Hydraulic modeling, as well as metering and monitoring of distribution systems • Using VFDs on equipment (e.g., chillers, pumps, cooling tower fans) to improve system control and to control energy to consume only the energy required to meet the design parameters (e.g., temperature, flow, pressure). • Computer-aided facility management (CAFM) for integrating other software (e.g., record drawings, operation and maintenance manuals, asset database) • Computerized maintenance management software (CMMS) • Automation from other trades (e.g., fire alarm, life safety, medical gases, etc.) • Regulatory functions (e.g., refrigerant management, federal, state and local agencies, etc.) • Use of three-dimensional/building information modeling (BIM) in plant design that can be incorporated into the operational and maintenance scheme when the plant is completed and placed in operation

• Annual heating or refrigeration per unit of building area (kilowatts per square metre per year) • Energy used per unit of refrigeration (megajoules per kilowatt) • Annual power required per unit of refrigeration per unit of building area (kilowatts per kilowatt of refrigeration per square metre per year)

Automatic controls for central cooling and/or heating plants may include standard equipment manufacturer’s control logic along with optional, enhanced energy-efficiency control logic. These specialized control systems can be based on different architectures such as distributed controls, programmable logic controllers, or microprocessor-based systems. Beyond standard control technology, the following control points and strategies may be needed for primary equipment, ancillary equipment, and the overall system: • Discharge temperatures and/or pressures to verify component operation • Return and supply temperatures to monitor and control T • Head pressure for refrigerant and/or water to ensure compressors and pumps are operating correctly • Stack temperature to ensure boilers are operating properly • Carbon monoxide and/or carbon dioxide levels (can indicate over- or underventilation) • Differential pressures (indicate whether flow or pressure drops are changing) • Flow rates of supply and return water and air, ventilation, etc. (all indicate energy use) • Peak and hourly refrigeration output • Peak and hourly heat energy output • Peak and hourly steam output • Flow rate of fuel(s) and electric demand by type of use • Temperature and/or pressure set points and status • Night setback status and set points • Economizer cycle status and set points • Variable flow through equipment and/or system control • Variable-frequency drive speeds, set points, and override status • Thermal storage control set points, charge mode, and percent capacity • Heat recovery cycle temperatures and performance • Occupancy status and override settings • Local weather conditions, including dry-bulb temperature, dew point, and forecast • History of these data points and building performance See Chapter 43 of the 2019 ASHRAE Handbook—HVAC Applications for more information on control strategies and optimization.

Instrumentation It is very important to measure and track performance of HVAC equipment to ensure that it is performing as intended and not drifting over time. Making it easier for personnel to review and trend the data is a key step in continuous commissioning. Dashboards and trending graphs can provide immediate visual representations of energy use and efficiency to highlight problems and items that need to be addressed. Whole-building and occupancy measurements are useful for scheduling and usage issues (some of the most common and wasteful problems). Additional breakdowns of energy use by type and zone (chiller, boiler, fans, lights, etc.) allows better tracking and conservation over time. Some codes require separate power meters and trending by type of use. All instrument operations where cooling or heating output are measured should have instrumentation calibration that is traceable to the National Institute of Standards and Technology (NIST). The importance of local gages and indicating devices, with or without a facility management system, should not be overlooked. All equipment must have adequate pressure gages, thermometers, flow meters, balancing devices, and dampers for effective performance, monitoring, and commissioning (see also ASHRAE Guideline 22). In addition, capped thermometer wells, gage cocks, capped duct openings, and volume dampers should be installed at strategic points for system balancing. Chapter 39 of the 2019 ASHRAE Handbook—HVAC Applications indicates the locations and types of fittings required. Chapter 37 of the 2017 ASHRAE Handbook—Fundamentals has more information on measurement and instruments.

8.

MAINTENANCE MANAGEMENT SYSTEMS

A review with the end user (owner) should be done, to understand the owner’s requirements for operation (e.g., if an owner has an in-house staff, more frequent access may be required, which may affect the extent to which a designer incorporates access). Reviewing ASHRAE Guideline 4 and Standard 15 with the owner can provide insight to the development of a maintenance plan. If maintenance is outsourced, extensive access may not be as high a priority. In some cases, regulatory and code access may be the only determining factors. Operations and maintenance considerations include the following: • Accessibility around equipment, as well as above and below when applicable, with at least minimum clearances provided per manufacturers’ recommendations and applicable codes (typical space required for regular service is often larger than the minimums, which eases the burden on maintenance staff) • Clearances for equipment removal • Minimizing tripping hazards (e.g., drain piping extending along the floor) • Adequate headroom to avoid injuries • Trenching in floor, if necessary • Cable trays, if applicable • Adequate lighting levels

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Central Cooling and Heating Plants • • • • • • •

Task lighting, when needed Eyewash stations for safety Exterior access for outdoor air supply and for exhaust Storage of mechanical and electrical parts and materials Documentation storage and administrative support rooms Proper drainage for system maintenance Outlets for service maintenance utilities (e.g., water, electricity) in locations reasonably accessible to equipment operators • Adequate lines of sight to view thermometers, pressure gages, etc. • Structural steel elements for major maintenance rigging of equipment

Typical operator maintenance functions include cleaning of condensers, evaporators, and boiler tubes. Cooling tower catwalk safety railing and ladders should be provided to comply with U.S. Occupational Safety and Health Administration (OSHA) requirements.

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9.

BUILDING SYSTEM COMMISSIONING

Because a central plant consumes a major portion of the annual energy operating budget, building system commissioning is imperative for new construction and expansion of existing installations. During the warranty phase, central-plant performance should be measured, benchmarked, and course-corrected to ensure design intent is achieved. If an energy analysis study is performed as part of the comparison between decentralized and centralized concepts, the resulting month-to-month energy data should be a good electronic document to benchmark actual energy consumption using the measurement and verification plan implementation. Ongoing commissioning or periodic recommissioning further ensures that the design intent is met, and that cooling and heating are reliably delivered after the warranty period. Many central systems are designed and built with the intent to connect to facilities in a phased program, which also requires commissioning. Consider retro- or recommissioning whenever the plant is expanded or an additional connection made to the existing systems, to ensure the original design intent is met. It is very important to remember that all the heating and cooling is used outside of the mechanical room. Chiller and boiler commissioning will be severely limited if the engineer does not look at the coils, fans, valves, and other components that may be in other locations. Dirty coils and bypass valves can waste more energy than chiller and boiler maintenance can save. Initial testing, adjusting, and balancing (TAB) also contributes to sustainable operation and maintenance. The TAB process should be repeated periodically to ensure original design levels are maintained. When completing TAB and commissioning, consider posting laminated system flow diagrams at or adjacent to the central cooling and heating equipment, and include operating instructions, TAB performance, commissioning functional performance test results, and emergency shutoff procedures. These documents also should be filed electronically in the central plant computer server for quick reference. Original basis of design and design criteria should be posted as a constant reminder of design intent, and to be readily available in case troubleshooting, expansion, or modernization is needed. As with all HVAC applications, for maintainable, long-term operational success, building commissioning should include the system training requirements necessary for building management staff to efficiently take ownership and operate and maintain the HVAC systems over the useful service life of the installation.

3.11 10.

SYSTEM REPLACEMENTS AND EXPANSIONS

More equipment is sold as replacements than for new construction. During replacements, the designer/specifier has more information available than during design (e.g., from observation and possibly operational trend logs). Typically, the difficulty is in finding the building’s original design details. It is usually necessary to survey the building to determine what is installed and whether it is the same as what is represented in the original design. The engineer also must determine if the project will be merely a replacement of existing equipment, a system expansion, or even a reduction of equipment capacity (if the original equipment was oversized). A like-for-like replacement has the advantage of simplicity, and may also not require an update of the mechanical system to current code requirements. For a straightforward replacement, the engineer still must verify that the replacement equipment will work within the capacity, electrical, and dimensional limitations, and that it is compatible with existing the control system and other equipment. Replacements often offer the chance to correct shortcomings of the original design, as well, such as • Resizing when original equipment was over- or undersized (e.g., where equipment cannot handle facility loads, or where cooling systems cannot properly dehumidify the space) • Adjusting to better fit building loads that have changed (e.g., outdated lighting has been replaced with higher-efficiency LEDs) or that were inappropriately calculated (e.g., the building needs more [or less] cooling capacity) • Improving water and airflow rates Other equipment (e.g., coils, towers, fans) that affect the system performance may also need replacement. A common issue with central plants is that loads on the plant operate on different schedules. These nonsimultaneous, varied loads can introduce a low load demand on the system. If the intent is to expand or reduce the capacity of the plant, the engineer must verify the existing equipment’s operational capabilities before deciding what equipment to replace.

REFERENCES ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal.ashrae .org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org/bookstore. ASHRAE. 2008. Preparation of operating and maintenance documentation for building systems. Guideline 4-2008. ASHRAE 2012. Instrumentation for monitoring central chilled-water plant efficiency. Guideline 22-2012. ASHRAE. 2016. Safety standard for refrigeration systems. ANSI/ASHRAE Standard 15-2016. ASHRAE. 2016. Energy standard for buildings except low-rise residential buildings. ANSI/ASHRAE/IES Standard 90.1-2016. ICC. 2018. International Energy Conservation Code®. International Code Council, Washington, D.C. Taylor, S.T. 2002. Degrading chilled water plant Delta-T: Causes and mitigation. ASHRAE Transactions 108(1). Paper AC-02-6-1.

BIBLIOGRAPHY ASHRAE. 2013. District cooling guide. ASHRAE. 1998. Fundamentals of water system design. Avery, G. 1998. Controlling chillers in variable flow systems. ASHRAE Journal 40(2):42-45. Avery, G. 2001. Improving the efficiency of chilled water plants. ASHRAE Journal 43(5):14-18. Duda, S. 2017. Engineer's notebook: Chillers & boilers in the same room: A cautionary tale. ASHRAE Journal 59(9):60-66.

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2020 ASHRAE Handbook—HVAC Systems and Equipment (SI) Seymore, M., 2014. Simplified chiller sequencing for a primary/secondary variable chilled water flow system. ASHRAE Journal 56(10). Severini, S.C. 2004. Making them work: Primary-secondary chilled water systems. ASHRAE Journal 46(7). Taylor, S.T. 2013. Engineer’s notebook: Tips to reduce chilled water plant costs. ASHRAE Journal 55(10). Taylor, S.T. 2014. Engineer’s notebook: How to design and control waterside economizers. ASHRAE Journal 56(6). Taylor, S. 2017. Fundamentals of design and control of central chilled-water plants. ASHRAE.

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Groenke, S., and M. Schwedler. 2002. Series-series counterflow for central chilled water plants. ASHRAE Journal 44(6):23-29. Hegberg, M. 2015. Fundamentals of water system design I-P, 2nd ed. ASHRAE Peterson, K. 2014. Engineer’s notebook: Improving performance of large chilled water plants. ASHRAE Journal 56(1):52-57. Peterson, K. 2016. Engineer's notebook: Open cooling tower design considerations. ASHRAE Journal 56(2):52-57. Rodriguez, R., and A. Pearson. 2016. Lessons learned from central heating plant upgrade ASHRAE Journal 58(11):88-91. Schwedler, M., 2017. Using low-load chillers to improve system efficiency, ASHRAE Journal 59(2).

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Related Commercial Resources CHAPTER 4

AIR HANDLING AND DISTRIBUTION AIR-HANDLING UNITS............................................................. 4.4 Air-Handling Unit Psychrometric Processes.............................. 4.5 Air-Handling Unit Components.................................................. 4.7 Air Distribution......................................................................... 4.11 AIR-HANDLING SYSTEMS...................................................... 4.12 Single-Duct Systems.................................................................. 4.12 Dual-Duct Systems.................................................................... 4.13

Multizone Systems ..................................................................... 4.14 Special Systems ......................................................................... 4.15 Air Terminal Units .................................................................... 4.17 Air Distribution System Controls .............................................. 4.18 Automatic Controls and Building Management Systems .......... 4.19 Maintenance Management System ............................................ 4.19 Building System Commissioning ............................................... 4.20

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V

ERY early in the design of a new or retrofit building project, the HVAC design engineer must analyze and ultimately select the basic systems, as discussed in Chapter 1, and determine whether production of primary heating and cooling should be decentralized (see Chapter 2) or central (see Chapter 3). This chapter covers the options, processes, available equipment, and challenges of all-air systems; for all-water, air-and-water, and local terminal systems, see Chapter 5. For additional system selection tools, refer to the HVAC System Analysis and Selection Matrix in ASHRAE Handbook Online (handbook.ashrae.org). Building air systems can be designed to provide ventilation air as well as complete sensible and latent cooling, preheating, dehumidification, and humidification capacity in air supplied by the system. No additional ventilation, cooling, or humidification is then required at the zone, except for certain industrial and hospital systems. Heating may be accomplished by the same airstream, either in the central system or at a particular zone. In some applications, heating is accomplished by a separate heat source. The term zone implies the provision of, or the need for, separate thermostatic control, whereas the term room implies a partitioned area that may or may not require separate control. The basic all-air system concept is to supply air to the room at conditions such that the sensible and latent heat gains in the space, when absorbed by supply air flowing through the space, bring the air to the desired room conditions. Because heat gains in the space vary with time, a mechanism to vary the energy removed from the space by the supply air is necessary. There are two such basic mechanisms: (1) vary the amount of supply air delivered to the space by varying the flow rate or supplying air intermittently; or (2) vary the temperature of air delivered to the space, either by modulating the temperature or conditioning the air intermittently. Both of these basic mechanisms must also accommodate variable ventilation flow rates. All-air systems may be adapted to many applications for comfort or process work. They are used in buildings of all sizes that require individual control of multiple zones, such as office buildings, schools and universities, laboratories, hospitals, stores, hotels, and even ships. All-air systems are also used virtually exclusively in special applications for close control of temperature, humidity, ventilation, space pressure, and/or air quality classification (e.g., ISO 14644-1 Class 3 space), including cleanrooms, computer rooms, hospital operating rooms, research and development facilities, and many industrial/manufacturing facilities.

Advantages of All-Air Systems • Operation and maintenance of major equipment can be performed in an unoccupied area (e.g., a central mechanical room). It also maximizes choices of filtration equipment, vibration and The preparation of this chapter is assigned to TC 9.1, Large Building AirConditioning Systems.

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noise control, humidification and dehumidification options, and selection of high-quality and durable equipment, including enhanced filtration for returning and outdoor airstreams. Piping, electrical equipment, wiring, filters, and vibration- and noise-producing equipment are away from the conditioned area, minimizing (1) disruption for service needs and (2) potential harm to occupants, furnishings, and processes. These systems offer the greatest potential for using outdoor air for economizer cooling instead of mechanical refrigeration. Seasonal changeover is simple and adapts readily to automatic control. A wide choice of zoning, flexibility, and humidity control under all operating conditions is possible. Simultaneous heating of one zone and cooling of another zone during off-season periods is available. Air-to-air and other energy recovery may be readily incorporated. Designs are flexible for optimum air distribution, draft control, and adaptability to varying local requirements. The systems are well-suited to applications requiring unusual exhaust, ventilation, or makeup air quantities (negative or positive pressurization, etc.). All-air systems adapt well to winter humidification and dehumidification for high latent loads. All-air systems take advantage of load diversity. In other words, a central air-handling unit serving multiple zones needs to be sized only for the peak coincident load, not the sum of the peak loads of each individual zone. In buildings with significant fenestration loads, diversity can be significant, because the sun cannot shine on all sides of a building simultaneously. By increasing the air change rate and using high-quality controls, these systems can maintain the closest operating condition of ±0.15 K dry bulb and ±0.5% rh. Some systems can maintain essentially constant space conditions ideal for museums. Removal and disposal of cold condensate from cooling coils, and capture and return of steam condensate from heating coils, is generally simpler and more practical in an all-air system. Central air-handling equipment life expectancies are longer and operation and maintenance costs are less than for many terminal systems.

Disadvantages of All-Air Systems • Ducts installed in ceiling plenums require additional duct clearance, sometimes reducing ceiling height and/or increasing building height. In retrofits, these clearances may not be available. • Larger floor plans may be necessary to allow adequate space for vertical shafts (if required for air distribution). In a retrofit application, shafts may be impractical. Shafts, as a general rule, consume 1 to 2% of multiple-story building gross areas; for high-rise buildings, over 2% can be expected.

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2020 ASHRAE Handbook—HVAC Systems and Equipment (SI)

• Transport energy used by the fans to distribute air and overcome duct and equipment static resistance is a larger part of the total building’s HVAC energy use than in other systems. • In commercial buildings, air-handling equipment rooms represent nonrentable or non-revenue-generating spaces. • Accessibility to terminal devices, duct-balancing dampers, etc., requires close cooperation between architectural, mechanical, and structural designers; therefore, accessible ceilings are recommended. • Air balancing, particularly on large systems, can be cumbersome. • Permanent heating is not always available sufficiently early to provide temporary heat during construction. • Mechanical failure of a central air-handling component, such as a fan or a cooling-coil control valve, affects all zones served by that unit. Energy conservation measures implemented on any building system also may have an effect.

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Heating and Cooling Calculations Basic calculations for airflow, temperatures, relative humidity, loads, and psychrometrics are covered in Chapters 1, 17, and 18 of the 2017 ASHRAE Handbook—Fundamentals. System selection should be related to the need, as indicated by the load characteristics. The designer should understand the operation of system components, their relationship to the psychrometric chart, and their interaction under various operating conditions and system configurations. The design engineer must properly determine an air-handling system’s required supply air temperature and volume; outdoor ventilation air requirements; desired space pressures; heating and cooling coil capacities; humidification and dehumidification capacities; return, relief, and exhaust air volume requirements; and required pressure capabilities of the fan(s). The HVAC designer should work closely with the architect to optimize the building envelope design. Close cooperation of all parties during an integrated design process can result in reduced building loads, which allows the use of smaller mechanical systems.

Zoning Exterior zones are affected by weather conditions (e.g., wind, temperature, sun) and, depending on the geographic area and season, may require both heating and cooling at different times. The system must respond to these variations. The need for separate perimeter zone heating is determined by the following: • Severity of heating load (i.e., geographic location) • Nature and orientation of building envelope • Effects of downdraft at windows and radiant effect of cold glass surfaces (i.e., type of glass, area, height, U-factor) • Type of occupancy (i.e., sedentary versus transient). • Operating costs (i.e., in buildings such as offices and schools that are unoccupied for considerable periods, fan operating cost can be reduced by heating with perimeter radiation during unoccupied periods rather than operating the main or local unit supply fans.) Separate perimeter heating can operate with any all-air system. However, its greatest application has been in conjunction with variable-air-volume (VAV) systems for cooling-only service. Careful design must minimize simultaneous heating and cooling. See the section on Variable Air Volume for further details. Interior spaces have relatively constant conditions because they are isolated from external influences. Cooling loads in interior zones may vary with changes in the operation of equipment and appliances in the space and changes in occupancy, but usually interior spaces require cooling throughout the year. A VAV system has limited energy advantages for interior spaces, but it does provide simple temperature control. Interior spaces with a roof exposure, however, may require treatment similar to perimeter spaces that require heat.

Space Heating Although steam is an acceptable medium for central system preheat or reheat coils, low-temperature hot water provides a simple and more uniform means of perimeter and general space heating. Individual automatic control of each terminal provides the ideal space comfort. A control system that varies water temperature inversely with the change in outdoor temperature provides water temperatures that produce acceptable results in most applications. For best results, the most satisfactory ratio can be set after installation is completed and actual operating conditions are ascertained. Multiple perimeter spaces on one exposure served by a central system may be heated by supplying warm air from the central system. Areas with heat gain from lights and occupants and no heat loss require cooling in winter, as well as in summer. In some systems, very little heating of return and outdoor air is required when the space is occupied. Local codes dictate the amount of outdoor air required (see ASHRAE Standard 62.1 for recommended outdoor air ventilation). For example, with return air at 24°C and outdoor air at –18°C, the temperature of a 25% outdoor/75% return air mixture would be 12°C, which is close to the temperature of air supplied to cool such a space in summer. In this instance, a preheat coil installed in the minimum outdoor airstream to warm outdoor air can produce overheating, unless it is sized so that it does not heat the air above 2 to 7°C. Assuming good mixing, a preheat coil in the mixed airstream prevents this problem. The outdoor air damper should be kept closed until room temperatures are reached during warm-up. Low-leakage dampers should be specified and may be required by some model codes. A return air thermostat can terminate warm-up. When a central air-handling unit supplies both perimeter and interior spaces, supply air must be cool to handle interior zones. Additional control is needed to heat perimeter spaces properly. Reheating the air is the simplest solution, but is often restricted under energy codes. An acceptable solution is to vary the volume of air to the perimeter and to combine it with a terminal heating coil or a separate perimeter heating system, either baseboard, overhead air heating, or a fan-powered terminal unit with supplemental heat. The perimeter heating should be individually controlled and integrated with the cooling control. Lowering the supply water temperature when less heat is required generally improves temperature control. For further information, refer to Chapter 13 in this volume and Chapter 48 of the 2019 ASHRAE Handbook—HVAC Applications.

Air Temperature Versus Air Quantity Designers have considerable flexibility in selecting supply air temperature and corresponding air quantity within the limitations of the procedures for determining heating and cooling loads. The difference between supply air temperature and desired room temperature is often referred to as the T of the all-air system. The relationship between T and air volume is approximately linear and inverse: doubling the T results in halving of the air volume. ASHRAE Standard 55 addresses the effect of these variables on comfort. The traditional all-air system is typically designed to deliver approximately 13°C supply air, for a conventional building with a desired indoor temperature of approximately 24°C. That supply air temperature is commonplace because the air is low enough in absolute moisture to result in reasonable space relative humidity in conventional buildings with modest latent heat loads. However, lower supply air temperatures may be required in spaces with high latent loads, such as gymnasiums or laundries, and higher supply air temperatures can be applied selectively with caution. Obviously, not all buildings are conventional or typical, and designers are expected not to rely on these conventions unquestioningly. (Trends toward lower lighting levels and lower internal room loads from energy-efficient

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Air Handling and Distribution lighting mean that room sensible heat is usually lower compared to latent heat.) Commercially available load calculation software programs, when applied correctly, help the designer find the optimum supply air temperature for each application. In cold-air systems, the supply air temperature is designed significantly lower than 13°C (perhaps as low as 7°C) in an effort to reduce the size of ducts and fans. In establishing supply air temperature, the initial cost of lower airflow and low air temperature (smaller fan and duct systems) must be calculated against potential problems of distribution, condensation, air movement, and decreased removal of odors and gaseous or particulate contaminants (Duda 2016). Terminal devices that use low-temperature air can reduce the air distribution cost. These devices mix room and primary air to maintain reasonable air movement in the occupied space, or a dual-fan dual-duct (DFDD) system that mixes air before it enters the room can be used. Because the amount of outdoor air needed is the same for any system, the percentage in low-temperature systems is high, requiring special care in design to avoid freezing preheat or cooling coils. Duda (2016) gives the following pros and cons of using colder supply air.

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Advantages. • A greater ΔT between supply air temperature (SAT) and desired room temperature on the air side means lower airflow to each room can achieve the same cooling effect, thereby reducing fan energy. • Colder SAT can reduce supply air duct size, air-handling unit physical size, and (perhaps) ceiling space, saving building first cost. The reduction in duct and/or ceiling cavity size may be especially advantageous when adding ductwork to an older building with smaller floor-to-floor heights constructed before air conditioning became commonplace. • Greater moisture removal associated with colder SAT results in lower indoor relative humidity, making it possible for the room temperature set point to be a degree or two higher without compromising occupant comfort. • Cold SAT may be complementary to systems using ice-based or low-temperature liquid-based thermal storage. • Lower fan airflow associated with colder SAT may lead to a quieter HVAC system. Disadvantages. • Cold SAT condenses more moisture at the cooling coil, meaning that a greater latent cooling load is forced onto the coil and cooling system when dehumidifying outdoor air. This can be a significant energy penalty because each kilogram of additional water vapor condensed represents an additional 2256 kJ. More latent heat removal means not only more chiller/compressor energy, but also more chilled- and condenser-water pumping energy, cooling tower fan energy (where applicable), and a higher first cost for the central cooling plant. • Colder SAT reduces the number of hours per year that a full airside economy cycle is in effect. For example, if the SAT is 13°C and it is 13°C outdoors, the system can use 100% outdoor air without any mechanical cooling. However, if the SAT is 9°C and it is outdoors, some mechanical cooling remains necessary. • Colder SAT increases reheat energy because it causes a lightly loaded room to overcool faster and need reheat sooner, once the VAV box reaches its minimum allowable airflow. • Colder SAT may require colder chilled water, depending on how aggressive the cold air target is. Chillers making colder chilled water generally use more energy on a kW/kW (coefficient of performance [COP]) basis. • Colder supply air ducts are more prone to condensation or sweating on bare sheet metal surfaces (leading to risk of mold growth)

4.3 if they are not flawlessly insulated. This is of concern particularly in humid climates, on startup after setback for a night or weekend, or where there are high internal moisture loads (e.g., kitchens). If using colder SAT in these cases, specifications regarding duct insulation must be well written and thoroughly enforced. Similarly, air-handling units (AHUs) producing colder supply air may be more expensive than conventional AHUs if specifications for thermal breaks or no through-metal are needed to avoid condensation, dripping, and puddling in unconditioned or less-conditioned mechanical rooms. • Diffuser selection can be more challenging and diffuser choices more limited, to avoid dumping cold air and to ensure a high air diffusion performance index (ADPI) in the space. • Using cold SAT with a direct-expansion (DX) system has a greater chance of ice build-up on the DX evaporator coil due to reduced airflow per unit of refrigeration, and more moisture being condensed on the coil surface.

Space Pressure Many special applications, such as isolation rooms, research labs, and cleanrooms, require constant-volume supply and exhaust air to the space to ensure space pressure control. Some of these applications allow the designer to reduce air volume during unoccupied periods while still maintaining space pressure control. Allair systems are generally the only systems able to combine space pressure control with temperature, humidity, and/or air filtration control.

Other Considerations All-air systems operate by maintaining a temperature differential between the supply air and the space. Any load that affects this differential and the associated airflow must be calculated and considered, including the following: • All fans (supply, return, and supplemental) add heat. All of the fan shaft power eventually converts to heat in the system, either initially as fan losses or downstream as duct friction losses. Motor inefficiencies are an added load if that motor is in the airstream. Whether the fan is upstream of the cooling coil (blow-through) or downstream (draw-though) affects how this load must be accounted for (Duda 2018a). The effect of these gains can be considerable, particularly in process applications. Heat gain in medium-pressure systems is about 1 K per kilopascal static pressure. • The supply duct may gain or lose heat from the surroundings. Most energy codes require that the supply duct be insulated, which is usually good practice regardless of code requirements. Uninsulated supply ducts delivering cool air are subject to condensation formation, leading to building water damage and potential mold growth, depending on the dew-point temperature of surrounding air. • Controlling humidity in a space can affect the air quantity and become the controlling factor in selecting supply airflow rate. All-air systems provide only limited humidity control, particularly at part loads, so if humidity is critical, extra care must be taken in design.

First, Operating, and Maintenance Costs As with all systems, the initial cost of an air-handling system varies widely (even for identical systems), depending on location, condition of the local economy, and contractor preference. For example, a dual-duct system is more expensive because it requires up to twice the amount of material for ducts as that of a comparable single-duct system. Systems requiring extensive use of terminal units are also comparatively expensive. The operating cost depends on the system selected, the designer’s skill in selecting and correctly

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2020 ASHRAE Handbook—HVAC Systems and Equipment (SI)

Fig. 1 Typical Air-Handling Unit Configurations

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sizing components, efficiency of the duct design, and effect of building design and type on the operation. Because an all-air system separates the air-handling equipment from occupied space, maintenance on major components in a central location is more economical. Also, central air-handling equipment requires less maintenance than a similar total capacity of multiple small packaged units. The many terminal units used in an all-air system do, however, require periodic maintenance. Because these units (including reheat coils) are usually installed throughout a facility, maintenance costs and scheduling for these devices must be considered.

Energy in Air Handling The engineer’s early involvement in the design of any facility can considerably affect the building’s energy consumption. Careful design minimizes system energy costs. In practice, however, a system might be selected based on a low first cost or to perform a particular task. VAV systems are generally more energy efficient than constantair-volume systems. Savings from a VAV system come from the savings in fan power and because the system does not overheat or overcool spaces, and reheat is minimized. Attention to fan static pressure settings and static pressure reset with VAV air terminal unit (ATU) unloading is needed to minimize energy use. The air distribution system for an all-air system consists of two major subsystems: (1) air-handling units that generate conditioned air under sufficient positive pressure to circulate it to and from the conditioned space, and (2) a distribution system that only carries air from the air-handling unit to the space being conditioned. The air distribution subsystem often includes means to control the amount or temperature of air delivered to each space.

1. AIR-HANDLING UNITS The basic air-handling system is an all-air, single-zone HVAC system consisting of an air-handling unit and an air distribution system. The air-handling unit may be designed to supply constant or variable air volume for low-, medium-, or high-velocity air distribution. Normally, the equipment is located outside the conditioned area in a basement, penthouse, or service area. The equipment can be adjacent to the primary heating and refrigeration equipment or at considerable distance, with refrigerant, chilled water, hot water, or steam circulated to it for energy transfer.

Figure 1 shows a typical draw-through central system that supplies conditioned air to a single zone or to multiple zones. A blowthrough configuration may also be used if space or other conditions dictate. The quantity and quality of supplied air are fixed by space requirements and determined as indicated in Chapters 17 and 18 of the 2017 ASHRAE Handbook—Fundamentals. Air gains and loses heat by contacting heat transfer surfaces and by mixing with air of another condition. Some of this mixing is intentional, as at the outdoor air intake; other mixing results from the physical characteristics of a particular component, as when untreated air passes through a coil without contacting the fins (bypass factor). All treated and untreated air must be well mixed for maximum performance of heat transfer surfaces and for uniform temperatures in the airstream. Stratified, parallel paths of treated and untreated air must be avoided, particularly in the vertical plane of systems using double-inlet or multiple-wheel fans. Because these fans may not completely mix the air, different temperatures can occur in branches coming from opposite sides of the supply duct. Poor mixing in units with high outdoor percentages can also result in freezing temperatures at the coils. Single or dual plenum fans can improve mixing conditions as well as reduce space requirements between upstream or downstream unit coils.

Primary Equipment Cooling. Either central station or localized equipment, depending on the application, can provide cooling. Most large systems with multiple central air-handling units use a central refrigeration plant. Small, individual air-handling equipment can (1) be supplied with chilled water from central chillers, (2) use direct-expansion cooling with a central condensing (cooling tower) system, or (3) be air cooled and totally self-contained. The decision to provide a central plant or local equipment is based on factors similar to those for air-handling equipment, and is further addressed in Chapters 1 to 3. Heating. The same criteria described for cooling are usually used to determine whether a central heating plant or a local one is desirable. Usually, a central, fuel-fired plant is more desirable for heating large facilities. In facilities with low heating loads, electric heating is a viable option and is often economical, particularly where care has been taken to design energy-efficient systems and buildings. Another option is a local indirect fuel-fired furnace in an air-handling unit.

Air-Handling Equipment Packaged air-handling equipment is commercially available in many sizes, capacities, and configurations using any desired method of cooling, heating, humidification, filtration, etc. These systems can be suitable for small and large buildings. In large systems (over 25 m3/s), air-handling equipment is usually custom-designed and fabricated to suit a particular application. Air-handling units may be either centrally or remotely located. Air-handling units (AHUs) can be one of the more complicated pieces of equipment to specify or order, because a vast array of choices are available, and because there is no single-number identifier (e.g., a “200 kW unit” or a “20 m3/s unit”) that adequately describes the desired product. Regardless of size or type, the designer must properly determine an air-handling unit’s required supply air temperature and volume; outdoor air requirements; desired space pressures; heating and cooling coil capacities; humidification and dehumidification capacities; return, relief, and exhaust air volume requirements; filtration; and required pressure capabilities of the fan(s). Typically, these parameters and more must be specified or scheduled by the design engineer before an installer or equipment supplier can provide an AHU.

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Air Handling and Distribution Central Mechanical Equipment Rooms (MERs) The type of facility and other factors help determine where the air-handling equipment is located. Central fan rooms are more common in large buildings, where maintenance is isolated from the conditioned space. The advent of fan arrays and increased cost competitiveness in the custom AHU market have led to very large air-handling units (mega-AHUs) designed for over 50,000 L/s (Taylor 2018). Reasons a design engineer may consider a central air-handling unit or bank of central air-handling units include the following: • • • • •

Fewer total pieces of equipment to maintain. Maintenance is concentrated at one location. Filtration is easily enhanced. Energy recovery opportunities may be more practical. Compared to floor-by-floor AHUs, very large AHUs with fan arrays are less expensive, require less space, have lower maintenance costs, and generally are more energy efficient. • Vibration and noise control, seismic bracing, outdoor air intakes, economizers, filtration, humidification, and similar auxiliary factors may be handled more simply when equipment is centralized. Licensed for single user. © 2020 ASHRAE, Inc.

Decentralized MERs Many buildings locate air-handling equipment at each floor, or at other logical subdivisions of a facility. Reasons a design engineer may consider multiple distributed air-handling unit locations include the following:

4.5 • Direct expansion (refrigerant) takes advantage of the latent heat of the refrigerant fluid, and cools as shown in the psychrometric diagram in Figure 2. • Chilled-water (fluid-filled) coils use temperature differences between the fluid and air to exchange energy by the same process as in Figure 2 (see the section on Dehumidification). • Direct spray of water in the airstream (Figure 3), an adiabatic process, uses the latent heat of evaporation of water to reduce drybulb temperature while increasing moisture content. Both sensible and latent cooling are also possible by using chilled water. A conventional evaporative cooler uses the adiabatic process by spraying or dripping recirculated water onto a filter pad (see the section on Humidification). • The wetted duct or supersaturated system is a variation of direct spray. In this system, tiny droplets of free moisture are carried by the air into the conditioned space, where they evaporate and provide additional cooling, reducing the amount of air needed for cooling the space (Figure 4). • Indirect evaporation adiabatically cools outdoor or exhaust air from the conditioned space by spraying water, then passes that cooled air through one side of a heat exchanger. Air to be supplied to the space is cooled by passing through the other side of the heat exchanger. Chapter 41 has further information on this method of cooling. Chapter 1 of the 2017 ASHRAE Handbook—Fundamentals details the psychrometric processes of these methods.

• Reduced size of ducts reduces space required for distribution ductwork and shafts. • Reduced equipment size as a result of decentralized systems allows use of less expensive packaged equipment and reduces the necessary sophistication of training for operating and maintenance personnel. • For facilities with varied occupancy, multiple decentralized airhandling units can be set back or turned off in unoccupied areas. • Failure of an air-handling unit affects only the part of the building served by that one unit. • Having greater quantity of smaller AHUs is well suited to buildings with multiple tenants who may have varying start times, or if rogue zones cause duct static pressure and supply air temperature reset strategies to be ineffective.

Fans Both packaged and built-up air-handling units can use any type of fan. Centrifugal fans may be forward-curved, backward-inclined, or airfoil, and single-width/single-inlet (SWSI) or double-width/ double-inlet (DWDI). Many packaged air-handling units feature a single fan, but packaged and custom air-handling units can use multiple DWDI centrifugal fans on a common shaft with a single drive motor. SWSI centrifugal plenum (plug) fans without a scroll are common on larger packaged air-handling units to make them more compact. Also, direct-drive fans are common with variable-frequency drives, thus avoiding belt maintenance and replacement. Fan selection should be based on efficiency and sound power level throughout the anticipated range of operation, as well as on the ability of the fan to provide the required flow at the anticipated static pressure. Chapter 21 further discusses fans and fan selection.

1.1

Fig. 2 Direct-Expansion or Chilled-Water Cooling and Dehumidification

AIR-HANDLING UNIT PSYCHROMETRIC PROCESSES

Cooling The basic methods used for cooling and dehumidification include the following:

Fig. 3 Direct Spray of Water in Airstream Cooling

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2020 ASHRAE Handbook—HVAC Systems and Equipment (SI)

Heating The basic methods used for heating include the following: • Steam uses the latent heat of the fluid. • Hot-water (fluid-filled) coils use temperature differences between the warm fluid and the cooler air. • Electric heat also uses the temperature difference between the heating coil and the air to exchange energy. • Direct or indirect gas- or oil-fired heat exchangers can also be used to add sensible heat to the airstream. The effect on the airstream for each of these processes is the same and is shown in Figure 5. For basic equations, see Chapter 1 of the 2017 ASHRAE Handbook—Fundamentals.

Dehumidification

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Humidification Methods used to humidify air include the following: • Direct spray of recirculated water into the airstream (air washer) reduces the dry-bulb temperature while maintaining an almost constant wet bulb, in an adiabatic process (see Figure 3, path 1 to 3). The air may also be cooled and dehumidified, or heated and humidified, by changing the spray water temperature. In one variation, the surface area of water exposed to the air is increased by spraying water onto a cooling/heating coil. The coil surface temperature determines leaving air conditions. Another method is to spray or distribute water over a porous medium, such as those in evaporative coolers. This method requires careful

Fig. 4

monitoring of the water condition to keep biological contaminants from the airstream (Figure 6). • Compressed air that forces water through a nozzle into the airstream is essentially a constant wet-bulb (adiabatic) process. The water must be treated to keep particles from entering the airstream and contaminating or coating equipment and furnishings. Many types of nozzles are available. • Steam injection is a constant-dry-bulb process (Figure 7). As the steam injected becomes superheated, the leaving dry-bulb temperature increases. If live steam is injected into the airstream, the boiler water treatment chemical must be nontoxic to occupants and nondamaging to building interior and furnishings.

Moisture condenses on a cooling coil when its surface temperature is below the air’s dew point, thus reducing the humidity of the air. Similarly, air will also be dehumidified if a fluid with a temperature below the airstream dew point is sprayed into the airstream (see the section on Air Washers in Chapter 41). The process is identical to that shown in Figure 2, except that the moisture condensed from the airstream condenses on, and dissolves in, the spray droplets instead of on the solid coil surface. Chemical dehumidification involves either passing air over a solid desiccant or spraying the air with a solution of desiccant and water. Both of these processes add heat, often called the latent heat of wetting, to the air being dehumidified. Usually about 465 kJ/kg of moisture is removed (Figure 8). These systems should be

Supersaturated Evaporative Cooling Fig. 6 Direct Spray Humidification

Fig. 5 Steam, Hot-Water, and Electric Heating, and Direct and Indirect Gas- and Oil-Fired Heat Exchangers

Fig. 7 Steam Injection Humidification

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Air Handling and Distribution

4.7 outdoor air if the return path has a significant pressure drop (greater than approximately 75 Pa). It provides a positive return and exhaust from the conditioned area, particularly when mixing dampers allow cooling with outdoor air in intermediate seasons and winter. The return air fan ensures that the proper volume of air returns from the conditioned space. It prevents excess building pressure when economizer cycles introduce more than the minimum quantity of outdoor air, and reduces the static pressure against which the supply fan must work. The return air fan should handle a slightly smaller amount of air to account for fixed exhaust systems, such as toilet exhaust, and to ensure a slight positive pressure in the conditioned space. Chapter 48 of the 2019 ASHRAE Handbook—HVAC Applications provides design details; also, see ASHRAE Guideline 16 and Taylor (2014).

Relief Air Fan Fig. 8 Chemical Dehumidification reviewed with the user to ensure that the space is not contaminated. Chapter 24 has more information on this topic. Licensed for single user. © 2020 ASHRAE, Inc.

Air Mixing or Blending Adiabatic mixing of two or more airstreams (e.g., outdoor and return air) into a common airstream can be shown on a psychrometric chart with reasonable accuracy (see Chapter 1 of the 2017 ASHRAE Handbook—Fundamentals).

1.2

AIR-HANDLING UNIT COMPONENTS

The following sections describe many commonly available airhandling unit components. Not all of these components will necessarily be used in any one system. To determine the system’s air-handling requirement, the designer must consider the function and physical characteristics of the space to be conditioned, and the air volume and thermal exchange capacities required. Then, the various components may be selected and arranged by considering the fundamental requirements of the central system. Figure 1 shows two basic arrangements of air-handling unit components for a single-zone, all-air central system suitable for yearround air conditioning. These arrangements allow close control of temperature and humidity. Most of the components in Figure 1 are available from many manufacturers completely assembled or in subassembled sections that can be bolted together in the field. When selecting central system components, specific design parameters must be evaluated to balance cost, controllability, operating expense, maintenance, noise, and space. The sizing and selection of primary air-handling units substantially affect the results obtained in the conditioned space. The equipment must be adequate, accessible for easy maintenance, and not overly complex in its arrangement and control to provide the required conditions. Further, the designer should consider economics in component selection. Both initial and operating costs affect design decisions. For example, the designer should not arbitrarily design for a 2.5 m/s face velocity, which has been common for selecting cooling coils and other components. Filter and coil selection at 1.5 to 2 m/s, with resultant lower pressure loss, could produce a substantial payback on constant-volume systems (Peterson 2014). Chapter 38 of the 2019 ASHRAE Handbook—HVAC Applications has further information on energy and life-cycle costs.

In many situations, a relief (or exhaust) air fan may be used instead of a return fan. A relief air fan relieves ventilation air introduced during air economizer operation and operates only when this control cycle is in effect. When a relief air fan is used, the supply fan must be designed for the total supply and return pressure losses in the system. During economizer mode, the relief fan must be controlled to ensure a slight positive pressure in the conditioned space, as with the return air fan system. The section on Economizers describes the required control for relief air fans.

Automatic Dampers The section on Mixing Plenums discusses conditions that must be considered when choosing, sizing, and locating automatic dampers for this critical mixing process. These dampers throttle the air with parallel- or opposed-blade rotation. These two forms of dampers have different airflow throttling characteristics (see Chapter 7 of the 2017 ASHRAE Handbook—Fundamentals). Pressure relationships between various sections of this mixing process must be considered to ensure that automatic dampers are properly sized for wide-open and modulating pressure drops. See ASHRAE Guideline 16 for additional detail.

Relief Openings Relief openings in large buildings should be constructed similarly to outdoor air intakes, but may require motorized or self-acting backdraft dampers to prevent high wind pressure or stack action from causing airflow to reverse when the automatic dampers are open. Pressure loss through relief openings should be 25 Pa or less. Lowleakage dampers, such as those for outdoor intakes, prevent rattling and minimize leakage. The relief air opening should be located so that air does not short-circuit to the outdoor air intake. Damper relief openings may also be used in conjunction with area pressure control.

Return Air Dampers Negative pressure in the outdoor air intake plenum is a function of the resistance or static pressure loss through the outdoor air louvers, damper, and duct. Positive pressure in the relief air plenum is likewise a function of the static pressure loss through the relief damper, the relief duct between the plenum and outdoors, and the relief louver. The pressure drop through the return air damper must accommodate the pressure difference between the positive-pressure relief air plenum and the negative-pressure outdoor air plenum. Proper sizing of this damper facilitates better control and mixing. An additional manual damper may be required for proper air balancing.

Return Air Fan

Outdoor Air Intakes

A return air fan is optional on small systems, but is essential for proper operation of air economizer systems for free cooling from

Resistance through outdoor air intakes varies widely, depending on construction. Frequently, architectural considerations dictate the

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type and style of louver. The designer should ensure that the louvers selected offer minimum pressure loss, preferably not more than 25 Pa. High-efficiency, low-pressure louvers that effectively limit carryover of rain are available. Flashing installed at the outer wall and weep holes or a floor drain will carry away rain and melted snow entering the intake. Cold regions may require a snow baffle to direct fine snow to a low-velocity area below the dampers. Outdoor air dampers should be low-leakage types with special gasketed edges and endseals. A separate damper section and damper operator are strongly recommended for ensuring minimum ventilation. The maximum outdoor air damper controls the air needed for economizer cycles. Carefully consider the location of intake and exhaust louvers; in some jurisdictions, location is governed by codes. Louvers must be separated enough to avoid short-circuiting air. Furthermore, intake louvers should not be near a potential source of contaminated air, such as a boiler stack or hood exhaust from a kitchen, laundry, or laboratory. Relief air should also not interfere with other systems. If heat recovery devices are used, intake and exhaust airstreams may need to be run in parallel, such as through air-to-air heat exchangers. A past complaint was a lack of outdoor air, especially in VAV systems where outdoor air quantities were established for peak loads and were then reduced in proportion to the air supplied during periods of reduced load. A simple airflow measurement control scheme added to the outdoor air damper eliminates this problem and keeps the amount of outdoor air constant, regardless of VAV system operation. However, the need to preheat outdoor air must be considered.

Economizers An air-side economizer uses outdoor air to reduce refrigeration requirements. Whereas a logic circuit maintains a fixed minimum of ventilation outdoor air in all weather, the air-side economizer takes advantage of cool outdoor air either to assist mechanical cooling or, if outdoor air is cool enough, to provide total system cooling. When weather allows, temperature controls systems can modulate outdoor and return air in the correct proportion to produce desired supply air temperatures without mechanical heating or cooling. For example, if 13°C supply air is necessary for space cooling and it is 13°C outdoors, direct digital controls (DDC) will position dampers to use 100% outdoor air without mechanical heating or cooling. If the return air is 24°C and it is 2°C outdoors, a 50/50 blend of return and outdoor air will produce 13°C supply air, again without mechanical heating or cooling. Designers should evaluate drying of the conditioned space when using economizers; exceptionally dry conditions may require mechanical humidification, which may reduce the attractiveness of air-side economizers. To exhaust the extra outdoor air brought in by the economizer, a method of variable-volume relief must be provided. The relief volume may be controlled by modulating the relief air dampers in response to building space pressure. Another common approach is opening the relief/exhaust and outdoor air intake dampers simultaneously, although this alone does not address space pressurization. A powered relief or return/relief fan may also be used. The relief system is off and relief dampers are closed when the air-side economizer is inactive. Intake dampers must be low leakage. Advantages and disadvantages of air-side economizers are discussed in Chapter 2.

Mixing Plenums Mixing plenums provide space for airstreams with different properties to mix as they are introduced into a common section of ductwork or air-handling unit, allowing the system to operate as intended. If the airstreams are not sufficiently mixed, the resulting

stratification adversely affects system performance. Some problems associated with stratification are nuisance low-temperature safety cutouts, frozen cooling coils, excess energy use by the preheat coil, inadequate outdoor air, control hunting, and poor outdoor air distribution throughout occupied spaces. A common example of a mixing plenum is the air-handling unit mixing box, in which outdoor and recirculated airstreams are combined. In air-handling units, mixing boxes typically have one inlet, with control dampers, for each airstream. There are no performance standards for mixing boxes or mixing plenums. Thus, it is difficult to know whether a particular mixing box design will provide sufficient mixing. In the absence of performance data, many rules of thumb have been developed to increase the mixing provided by mixing boxes. It is important to note that few supporting data exist; the following suggestions are based largely on common-sense solutions and anecdotal evidence: • The minimum outdoor air damper should be located as close as possible to the return air damper. • An outdoor air damper sized for 7.5 m/s gives good control. • Low-leakage outdoor air dampers minimize leakage during system shutdown. • A higher velocity through the return air damper facilitates air balance and may increase mixing. • Parallel-blade dampers may aid mixing. Positioning the dampers so that the return and outdoor airstreams are deflected toward each other may increase mixing. • Placing the outdoor air damper above the return air damper increases mixing by density differences: denser, cold outdoor air mixes as it drops through the warm, less dense return air. • Mixing dampers should be placed across the full width of the unit, even if the location of the return duct makes it more convenient to return air through the side. Return air entering through the side of an air-handling unit can pass through one side of a double-inlet fan while outdoor air passes through the other side. This same situation can exist whenever two parallel fans are used in an airhandling unit receiving two different airstreams. Wherever there are two fans and two airstreams, an air mixer should be used. • Field-built baffles may be used to create additional turbulence and to enhance mixing. Unfortunately, the mixing effectiveness and pressure drop of field-built solutions are unknown. If stratification is anticipated in a system, then special mixing equipment that has been tested by the manufacturer (see the section on Static Air Mixers) should be specified and used in the airhandling system.

Static Air Mixers Static air mixers are designed to enhance mixing in the mixing plenum to reduce or eliminate problems associated with stratification. These devices have no moving parts and create turbulence in the airstream, which increases mixing. They are usually mounted between the mixing box and the heating or cooling coil; the space required depends on the amount of mixing that is required. Typical pressure loss for these devices is 25 to 75 Pa. There are no performance standards for air mixers. Thus, manufacturers of air mixers and air-handling units should demonstrate that their devices provide adequate mixing.

Filter Section A system’s overall performance depends heavily on the filter. Unless the filter is regularly maintained, system resistance increases and airflow diminishes. Accessibility for replacement is an important consideration in filter arrangement and location. In smaller air-handling units, filters are often placed in a slide-out rack for side-access replacement. In larger units and built-up systems with internal or front-loading access, there should be at least

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Air Handling and Distribution 1 m between the upstream face of the filter bank and any obstruction. Other requirements for filters can be found in Chapter 29 and in ASHRAE Standard 52.2. Good mixing of outdoor and return air is also necessary for good filter performance. A poorly placed outdoor air duct or a bad duct connection to the mixing plenum can cause uneven filter loading and poor distribution of air through the coil section. Particulate filters are rated according to ASHRAE Standard 52.2’s minimum efficiency rating value (MERV) system, a numeric ranking from 1 (least) to 20 (highest). A particulate filter bank of at least MERV 6 should be placed upstream of the first coil, to maintain coil cleanliness. Depending on the spaces served, many applications demand higher-efficiency filters. Some studies suggest filters up to MERV 14 can pay for themselves in reduced coil maintenance and better heat transfer effectiveness. Where higher-MERV filters are used, many designers specify a lower-MERV prefilter as an inexpensive sacrificial filter to capture bulk particulate and extend the life of the more expensive final filter. Filter bank(s) location may be governed by codes. For example, many prevailing health care codes mandate a prefilter upstream of all fans, coils, and humidifiers, plus a final filter bank downstream of all fans, coils, and humidifiers. Designers are not limited to particulate filters. Electronic air cleaners and gaseous-phase (e.g., activated carbon) filters are available for added protection. For example, ASHRAE Standard 62.1 requires use of gaseous-phase filters for certain, usually urban, regions where outdoor air quality has been measured to exceed threshold values for ozone or other gaseous contaminants. Odor control using activated carbon or potassium permanganate as a filter medium is also available. Chapters 11 and 12 of the 2017 ASHRAE Handbook—Fundamentals have more information on odor control.

Preheat Coil Preheat coils are heating coils placed upstream of a cooling coil; they can use steam, hot water, or electric resistance as a medium. Some air-handling units do not require a preheat coil at all, particularly if the percentage of outdoor air is low and if building heating is provided elsewhere (e.g., perimeter baseboard). Where used, a preheat coil should have wide fin spacing, be accessible for easy cleaning, and be protected by filters. If a preheat coil is located in the minimum outdoor airstream rather than in the mixed airstream as shown in Figure 11, it should not heat the air to an exit temperature above 2 to 7°C; preferably, it should become inoperative at outdoor temperatures above 7°C. For use with steam, inner distributing tube or integral face-and-bypass coils are preferable. Hot-water preheat coils should be piped for counterflow so that the coldest air contacts the warmest part of the coil surface first. Consider a constant-flow recirculating pump if the local climate and anticipated percentage of outdoor air may result in freezing conditions at a hot-water preheat coil. Chapter 27 provides more detailed information on heating coils.

Cooling Coil Sensible and latent heat are removed from the air by the cooling coils. The cooling medium can be either chilled water or refrigerant, in which case the refrigerant coil serves as the evaporator in a vaporcompression refrigeration cycle. The psychrometrics of cooling and dehumidification were described previously. In all finned coils, some air passes through without contacting the fins or tubes. The amount of this bypass can vary from 30% for a four-row coil at 3.5 m/s, to less than 2% for an eight-row coil at 1.5 m/s. The dew point of the air mixture leaving a four-row coil might satisfy a comfort installation with 25% or less outdoor air (10% for humid climates), a small internal latent load, and sensible temperature control only. For close control of room conditions for precision work, a deeper coil may be required. Chapter 23 provides more information on cooling coils and their selection.

4.9 Coil freezing can be a serious problem with chilled-water coils. Full-flow circulation of chilled water during freezing weather, or even reduced flow with a small recirculating pump, minimizes coil freezing and eliminates stratification. Further, continuous full-flow circulation can provide a source of off-season chilled water in airand-water systems. Antifreeze solutions or complete coil draining also prevent coil freezing. However, because it is difficult (if not impossible) to drain most cooling coils completely, caution should be used if this option is considered. Another design consideration is the drain pan. ASHRAE Standard 62.1 calls for drain pans to be sloped to a drain, to avoid holding standing water in the air-handling unit. Because of the constant presence of moisture in the cooling coil drain pan and nearby casing, many designers require stainless steel construction in that portion of the air-handling unit.

Reheat Coil Reheat coils are heating coils placed downstream of a cooling coil. Reheat systems are strongly discouraged, unless recovered energy is used (see ASHRAE Standard 90.1). Positive humidity control is required to provide comfort conditions for most occupancies. Either reheat or desiccant is usually required to dehumidify outdoor air. Reheating is necessary for laboratory, health care, or similar applications where temperature and relative humidity must be controlled accurately. Heating coils located in the reheat position, as shown in Figure 11, are frequently used for warm-up, although a coil in the preheat position is preferable. Hot-water coils provide a very controllable source of reheat energy. Inner-distributing-tube coils are preferable for steam applications. Electric coils may also be used. See Chapter 27 for more information. Reheat is highly inefficient. In health care occupancies, many spaces have total air turnover requirements, which, in concert with the apparatus dew-point temperature required for humidity control in the space, often results in reheat to avoid space overcooling. Coilbypass air handling avoids most of the required reheat through a strategy described by Nall (2017). Properly controlled to maintain the required operating parameters, this strategy can significantly reduce required cooling and heating capacity, and thus first cost, as well as save significant annual energy costs. However, close attention is needed to the override strategies to prevent supply air temperature optimization algorithms from compromising the required space operating conditions.

Humidifiers Humidifiers may be installed as part of the air-handling unit, or in terminals at the point of use, or both. Where close humidity control of selected spaces is required, the entire supply airstream may be humidified to a low humidity level in the air-handling unit. Terminal humidifiers in the supply ducts serving selected spaces bring humidity up to the required levels. For comfort installations not requiring close control, moisture can be added to the air by mechanical atomizers or point-of-use electric or ultrasonic humidifiers. Steam grid humidifiers with dew-point control usually are used for accurate humidity control. Air to a laboratory or other space that requires close humidity control must be reheated after leaving a cooling coil before moisture can be added. Humidifying equipment capacity should not exceed the expected peak load by more than 10%. If humidity is controlled from the room or return air, a limiting humidistat and fan interlock may be needed in the supply duct. This tends to minimize condensation and mold or mildew growth in the ductwork. Humidifiers add some sensible heat that should be accounted for in the psychrometric evaluation. See Chapter 22 for additional information. An important question for air-handling unit specifiers is where to place the humidification grid. Moisture cannot be successfully added to cold air, so placement is typically downstream of a preheat

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coil. For general building humidification, one satisfactory location is between a preheat coil and cooling coil. Another consideration is absorption distance (i.e., distance required for steam to be absorbed into the airstream). This can vary from 450 mm to 1.5 m and must be allowed for in the layout and dimensioning of the air-handling unit.

Dehumidifiers For most routine applications in dry or limited-cooling climates, such as offices, residences, and schools, the air-handling unit’s cooling coil provides adequate dehumidification. For humid climates, separate dehumidifiers or some form of reheat or desiccant is usually necessary. Where a specialty application requires additional moisture removal, desiccant dehumidifiers are an available accessory. Dust can be a problem with solid desiccants, and lithium contamination is a concern with spray equipment. Chapter 23 discusses dehumidification by cooling coils, and Chapter 24 discusses desiccant dehumidifiers.

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Energy Recovery Devices Energy recovery devices are in greater demand as outdoor air percentage increases along with increasing HVAC efficiency requirements. ASHRAE Standard 90.1 requires energy recovery devices for air-handling units except for systems with low outdoor air percentages or very low overall capacity. They are used extensively in research and development facilities and in hospitals and laboratories with high outdoor air requirements. Many types are available, and the type of facility usually determines which is most suitable. Choices include heat pipes, runaround loops, fixed-plate energy exchangers (both sensible only and enthalpic), and rotary wheel energy exchangers. See Chapter 26 for details. Select these devices with care, and minimize the differential pressure between airstreams. Most manufacturers of factory-packaged air-handling units offer optional energy recovery modules for both small and large unit applications. Many countries with extreme climates provide heat exchangers on outdoor/relief air, even for private homes. This trend is now appearing in both modest and large commercial buildings worldwide. Under certain circumstances, heat recovery devices can save energy and reduce the required capacity of primary cooling and heating plants by 20% or more. When applying energy recovery, it is important to have a way to bypass or disable the energy recovery in mild weather. Under some outdoor conditions, an energy recovery device can actually increase energy use. For example, if an air-handling unit is set to provide 13°C supply air, the exhaust is 24°C, and the outdoor temperature is 13°C, an energy recovery device will warm the supply air (via the 24°C exhaust) and then require mechanical cooling to return it to 13°C. An energy recovery device is an excellent choice during very hot or very cold weather, but not during mild weather. A rotating enthalpy wheel can be turned off when necessary. For a fixed-plate heat exchanger, an outdoor-air bypass around the heat exchanger is needed, with motorized dampers and control logic to use the recovery device only when advantageous to do so. Another important consideration is whether to size AHU heating and/or cooling coils as if the energy recovery device has failed and is no longer recovering energy, especially for rotary wheel energy recovery devices whose motor or drive could possibly fail, with potentially significant negative consequences on a very hot/humid or very cold day. Doing so allows such a failure to be ridden through without loss of building temperature control. However, there is some additional pressure loss of the upsized heating and/or cooling coil for the 99% (or more) of the time that the energy recovery device is in good working order. The engineer, in consultation with the building users, should apply judgment based on the critical nature of the building (e.g., hospital versus office), consequences of a failure (e.g., loss of rental income), etc. In colder climates, for

example, many designers size the heating coil (but not the cooling coil) as if the energy recovery device has failed, on the assumption that occupants and property need to be protected from the cold more so than from heat. Many designers do not apply this safety consideration when applying static energy recovery devices with no moving parts, reasoning that the likelihood of failure is extremely low. For more information on technical considerations beyond this chapter’s scope, see Chapter 26.

Sound Control Devices Where noise control is important, air-handling units can be specified with a noise control section, ranging from a plenum lined with acoustic duct liner to a full bank of duct silencers. This option is available in the smallest to largest units. Sound attenuation can be designed into the discharge (supply) end of the air-handling unit to reduce ductborne fan noise. Remember to consider ductborne noise traveling backward down the return or outdoor air paths in a noisesensitive application, and use a sound attenuation module if necessary at the inlet end of an air-handling unit. See Chapter 49 of the 2019 ASHRAE Handbook—HVAC Applications for details.

Supply Air Fan Axial-flow, centrifugal, or plenum (plug) fans may be chosen as supply air fans for straight-through flow applications. In factory-fabricated units, more than one centrifugal fan may be tied to the same shaft. If headroom allows, use a single-inlet fan when air enters at right angles to the flow of air through the equipment. This allows direct airflow from the fan wheel into the supply duct without abrupt change in direction and loss of efficiency. It also allows a more gradual transition from the fan to the duct and increases static regain in the velocity pressure conversion. To minimize inlet losses, the distance between casing walls and fan inlet should be at least the diameter of the fan wheel. With a single-inlet fan, the length of the transition section should be at least half the width or height of the casing, whichever is longer. If fans blow through the equipment, analyze air distribution through the downstream components, and use baffles to ensure uniform air distribution. See Chapter 21 for more information. Two placements of the supply fan section are common. A supply fan placed downstream of the cooling coil is known as a drawthrough arrangement, because air is drawn, or induced, across the cooling coil. Similarly, a supply fan placed upstream of the cooling coil is called the blow-through position. Either arrangement is possible in both small and large air-handling units, and in factorypackaged and custom field-erected units. A draw-through system (illustrated in Figure 1) draws air across the coils. A draw-through system usually provides a more even air distribution over all parts of the coil. However, some fan heat is added to the airstream after the air has crossed the cooling coil and must be taken into account when calculating the desired supply air temperature. Depressing the cooling coil’s leaving air temperature an equivalent amount to account for fan heat, may accidentally add additional latent load to the system because depressing the coil leaving temperature if in an already saturated condition strips additional moisture from the airstream. In other words, a draw-through arrangement has a higher dehumidification (latent cooling) effect for a given leaving air temperature. If a typical cooling coil drops the wet-bulb temperature 6 K and the motor heat rise is 2 K, then the latent cooling is 6 – 2 = 4 K. This can help with applications that need higher dehumidification levels. A blow-through system (illustrated in Figure 1) requires some caution on the part of the designer, because the blast effect of the supply fan outlet can concentrate a high percentage of the total air over a small percentage of the downstream coil surfaces. Air diffusers or diverters may be required. Consequently, blow-through air-handling units may tend to be longer overall than comparable draw-through

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Air Handling and Distribution units. Now that plenum fans and plenum fan arrays have become more common, their air discharge pattern tends to be much more uniform, somewhat mitigating this concern. This arrangement offers the advantage of placing the fan before the cooling coil, raising the temperature of air before it enters the cooling coil and thereby allowing the cooling coil to remove fan heat from the system. The added heat is entirely sensible and therefore typically less than the sensibleplus-latent penalty discussed for draw-through. The blow-through arrangement is mandatory where natural-gas-fired heat exchangers are used for heating. In either arrangement, consider the system effect of the fan arrangement in the unit. Refer to AMCA Standard 210/ASHRAE Standard 51 for details.

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Miscellaneous Components Vibration and sound isolation equipment is required for many central fan installations. Standard mountings of fiberglass, ribbed rubber, neoprene mounts, and springs are available for most fans and prefabricated units. The designer must account for seismic restraint requirements for the seismic zone in which the project is located (see Chapter 56 of the 2019 ASHRAE Handbook—HVAC Applications). In some applications, fans may require concrete inertia blocks in addition to spring mountings. Steel springs require sound-absorbing material inserted between the springs and the foundation. Horizontal discharge fans operating at a high static pressure frequently require thrust arrestors. Ductwork connections to fans should be made with fireproof fiber cloth sleeves having considerable slack, but without offset between the fan outlet and rigid duct. Misalignment between the duct and fan outlet can cause turbulence, generate noise, and reduce system efficiency. Electrical and piping connections to vibration-isolated equipment should be made with flexible conduit and flexible connections. Equipment noise transmitted through ductwork can be reduced by sound-absorbing units, acoustical lining, and other means of attenuation. Sound transmitted through the return and relief ducts should not be overlooked. Acoustical lining sufficient to adequately attenuate objectionable system noise or locally generated noise should be considered. Chapter 49 of the 2019 ASHRAE Handbook—HVAC Applications and Chapter 8 of the 2017 ASHRAE Hand-book—Fundamentals have further information on sound and vibration control. Noise control, both in occupied spaces and outdoors near intake or relief louvers, must be considered. Some local ordinances may limit external noise produced by these devices.

1.3

AIR DISTRIBUTION

Once air-handling system and air-handling equipment have been selected, air must be distributed to the zone(s) served. Ductwork should deliver conditioned air to each zone as directly, quietly, and economically as possible. Air distribution ductwork and terminal devices must be compatible; otherwise, the system will either fail to operate effectively or incur high first, operating, and maintenance costs. Unit connections must be given close attention; see AMCA Standard 210/ASHRAE Standard 51 for details.

Ductwork Design Chapter 21 of the 2017 ASHRAE Handbook—Fundamentals describes ductwork design in detail and gives several methods of sizing duct systems, including static regain and equal friction. Duct sizing is often performed manually for simple systems, but commercially available duct-sizing software programs are often used for larger and complex systems. It is imperative that the designer coordinate duct design with architectural and structural design. Structural features of the building generally require some compromise and often limit depth. In commercially developed projects, great effort is made to reduce floor-to-floor dimensions. In architecturally

4.11 significant buildings, high ceilings, barrel-vault ceilings, rotundas and domes, ceiling coves, and other architectural details can place obstacles in the path of ductwork. The resultant decrease in available interstitial space left for ductwork can be a major design challenge. Layout of ductwork in these buildings requires experience, skill, and patience on the part of the designer. Considerations. Duct systems can be designed for high or low velocity. A high-velocity system has smaller ducts, which save space but require higher pressures and may result in more noise. In some low-velocity systems, medium or high fan pressures may be required for balancing or to overcome high pressure drops from terminal devices. In any variable-flow system, changing operating conditions can cause airflow in the ducts to differ from design flow. Thus, varying airflow in the supply duct must be carefully analyzed to ensure that the system performs efficiently at all loads. This precaution is particularly needed with high-velocity air. Return air ducts are usually sized by the equal friction method. ASHRAE is currently developing a ductwork design manual, scheduled for publication in late 2016. In many applications, the space between a suspended ceiling and the floor slab or roof above it is used as a return air plenum, so that return air is collected at a central point. Governing codes should be consulted before using this approach in new design, because most codes prohibit combustible material in a ceiling space used as a return air plenum. For example, the National Electrical Code® Handbook (NFPA 2017) requires that either conduit or PTFE-insulated wire (often called plenum-rated cable) be installed in a return air plenum. In addition, regulations often require that return air plenums be divided into smaller areas by firewalls and that fire dampers be installed in ducts, which increases first cost. In research and some industrial facilities, return ducting must be installed to avoid contamination and growth of biological contaminants in the ceiling space. Lobby ceilings with lay-in panels may not work well as return plenums where negative pressure from high-rise elevators or stack effects of high-rise buildings may occur. If the plenum leaks to the low-pressure area, the tiles may lift and drop out when the outer door is opened and closed. Return plenums directly below a roof deck have substantially greater return air temperature increases or decreases than a duct return. Corridors should not be used for return air because they spread smoke and other contaminants. Although most codes ban returning air through corridors, the method is still used in many older facilities. All ductwork should be sealed. Energy waste because of leaks in the ductwork and terminal devices can be considerable. Unsealed ductwork in many commercial buildings can have leakage of 20% or more. Air systems are classified in two categories: • Single-duct systems contain the main heating and cooling coils in a series-flow air path. A common duct distribution system at a common air temperature feeds all terminal apparatus. Capacity can be controlled by varying air temperature or volume. • Dual-duct systems contain the main heating and cooling coils in parallel-flow or series/parallel-flow air paths with either (1) a separate cold- and warm-air duct distribution system that blends air at the terminal apparatus (dual-duct systems), or (2) a separate supply air duct to each zone with the supply air blended at the main unit with mixing dampers (multizone). Dual-duct systems generally vary the supply air temperature by mixing two airstreams of different temperatures, but they can also vary the volume of supply air in some applications. These categories are further divided and described in the following sections.

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2.

AIR-HANDLING SYSTEMS 2.1

SINGLE-DUCT SYSTEMS

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Constant Volume While maintaining constant airflow, single-duct constantvolume systems change the supply air temperature in response to the space load (Figure 9). Single-Zone Systems. The simplest all-air system is a supply unit serving a single zone. The unit can be installed either in or remote from the space it serves, and may operate with or without distribution ductwork. Ideally, this system responds completely to the space needs, and well-designed control systems maintain temperature and humidity closely and efficiently. Single-zone systems often involve short ductwork with low pressure drop and thus low fan energy, and can be shut down when not required without affecting operation of adjacent areas, offering further energy savings. A return or relief fan may be needed, depending on system capacity and whether 100% outdoor air is used for cooling as part of an economizer cycle. Relief fans can be eliminated if overpressurization can be relieved by other means, such as gravity dampers. Multiple-Zone Reheat Systems. Multiple-zone reheat is a modification of the single-zone system. It provides (1) zone or space control for areas of unequal loading, (2) simultaneous heating or cooling of perimeter areas with different exposures, and (3) close control for process or comfort applications. As the word reheat implies, heat is added as a secondary simultaneous process to either preconditioned (cooled, humidified, etc.) primary air or recirculated

room air. Relatively small low-pressure systems place reheat coils in the ductwork at each zone. More complex designs include highpressure primary distribution ducts to reduce their size and cost, and pressure reduction devices to maintain a constant volume for each reheat zone. The system uses conditioned air from a central unit, generally at a fixed cold-air temperature that is low enough to meet the maximum cooling load. Thus, all supply air is always cooled the maximum amount, regardless of the current load. Heat is added to the airstream in each zone to avoid overcooling that zone, for every zone except the zone experiencing peak cooling demand. The result is very high energy use, so use of this system is restricted by ASHRAE Standard 90.1. However, the supply air temperature from the unit can be varied, with proper control, to reduce the amount of reheat required and associated energy consumption. Care must be taken to avoid high internal humidity when the temperature of air leaving the cooling coil is allowed to rise during cooling. Constant-volume reheat can ensure close control of room humidity and/or space pressure. In cold weather, when a reheat system heats a space with an exterior exposure, the reheat coil must not only replace the heat lost from the space, but also must offset cooling of the supply air (enough cooling to meet the peak load for the space), further increasing energy consumption. If a constant-volume system is oversized, reheat cost becomes excessive. In commercial applications, use of a constant-volume reheat system is generally discouraged in favor of variable-volume or other systems. Constant-volume reheat systems may continue to be applied in hospitals, laboratories, and other critical applications where variable airflow may be detrimental to proper pressure relationships (e.g., for infection control).

Variable Air Volume (VAV)

Fig. 9 Constant-Volume System with Reheat (Courtesy RDK Engineers)

Fig. 10

A VAV system (Figure 10) controls temperature in a space by varying the quantity of supply air rather than varying the supply air temperature. A VAV terminal unit at the zone varies the quantity of supply air to the space. The supply air temperature is held relatively constant. Although supply air temperature can be moderately reset depending on the season, it must always be low enough to meet the cooling load in the most demanding zone and to maintain appropriate relative humidity. VAV systems can be applied to interior or perimeter zones, with common or separate fans, with common or separate air temperature control, and with or without auxiliary heating devices. The greatest energy saving associated with VAV occurs in perimeter zones, where variations in solar load and outdoor temperature allow the supply air quantity to be reduced.

Variable-Air-Volume System with Reheat and Induction and Fan-Powered Devices (Courtesy RDK Engineers)

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Air Handling and Distribution Other measures may also maintain enough air circulation through the room to achieve acceptable humidity levels. The human body is more sensitive to elevated air temperatures when there is little air movement. Minimum air circulation can be maintained during reduced load by (1) raising the supply air temperature of the entire system, which increases space humidity, or supplying reheat on a zone-by-zone basis; (2) providing auxiliary heat in each room independent of the air system; (3) using individual-zone recirculation and blending varying amounts of supply and room air or supply and ceiling plenum air with fan-powered VAV terminal units, or, if design allows, at the air-handling unit; (4) recirculating air with a VAV induction unit; or (5) providing a dedicated recirculation fan to increase airflow. VAV reheat can ensure close room space pressure control with the supply terminal functioning in sync with associated room exhaust. A typical application might be a fume hood VAV exhaust with constant open sash velocity (e.g., 0.4 or 0.5 m/s) or occupied/ unoccupied room hood exhaust (e.g., 0.5 m/s at sash in occupied periods and 0.3 m/s in unoccupied periods). Dual-Conduit. This method is an extension of the single-duct VAV system: one supply duct offsets exterior transmission cooling or heating loads by its terminal unit with or without auxiliary heat, and the other supply air path provides cooling throughout the year. The first airstream (primary air) operates as a constant-volume system, and the air temperature is varied to offset transmission only (i.e., it is warm in winter and cool in summer). Often, however, the primary-air fan is limited to operating only during peak heating and cooling periods to further reduce energy use. When calculating this system’s heating requirements, the cooling effect of secondary air must be included, even though the secondary system operates at minimum flow. The other airstream (secondary air) is cool year-round and varies in volume to match the load from solar heating, lights, power, and occupants. It serves both perimeter and interior spaces. Variable Diffuser. The discharge aperture of this diffuser is reduced to keep discharge velocity relatively constant while reducing conditioned supply airflow. Under these conditions, the diffuser’s induction effect is kept high, cold air mixes in the space, and the room air distribution pattern is more nearly maintained at reduced loads. These devices are of two basic types: one has a flexible bladder that expands to reduce the aperture, and the other has a diffuser plate that moves. Both devices are pressure dependent, which must be considered in duct distribution system design. They are either powered by the system or pneumatically or electrically driven.

2.2

DUAL-DUCT SYSTEMS

A dual-duct system conditions all air in a central apparatus and distributes it to conditioned spaces through two ducts, one carrying cold air and the other carrying warm air. In each conditioned zone, air valve terminals mix warm and cold air in proper proportion to satisfy the space temperature and pressure control. Dual-duct systems may be designed as constant volume or variable air volume; a dual-duct, constant-volume system generally uses more energy than a single-duct VAV system. As with other VAV systems, certain primary-air configurations can cause high relative humidity in the space during the cooling season.

4.13 vidual zones (Figure 11), and (2) only part of the supply air is cooled by the cooling coil (except at peak cooling demand); the rest of the supply is heated by the hot-deck coil during most hours of operation. This uses less heating and cooling energy than the terminal reheat system where all the air is cooled to full cooling capacity for more operating hours, and then all of it is reheated as required to match the space load. Fan energy is constant because airflow is constant. Single Fan without Reheat. This system has no heating coil in the fan unit hot deck and simply pushes a mixture of outdoor and recirculated air through the hot deck. A problem occurs during periods of high outdoor humidity and low internal heat load, causing the space humidity to rise rapidly unless reheat is added. This system has limited use in most modern buildings because most occupants demand more consistent temperature and humidity. A single-fan, no-reheat dual-duct system does not use any extra energy for reheat, but fan energy is constant regardless of space load.

Variable Air Volume Dual-duct VAV systems blend cold and warm air in various volume combinations. These systems may include single-duct VAV terminal units connected to the cold-air duct distribution system for cooling-only interior spaces (see Figures 11 and 12), and the cold duct may serve perimeter spaces in sync with the hot duct. This saves reheat energy for the air for those cooling-only zones because space temperature control is by varying volume, not supply air temperature, which may save some fan energy to the extent that the airflow matches the load. Newer dual-duct air terminals provide two damper operators per air terminal, which allows the unit to function like a single-duct VAV cooling terminal unit (e.g., 250 mm inlet damper) and a single-duct VAV heating terminal unit (e.g., 150 mm inlet damper) in one physical terminal package. This arrangement allows the designer to specify the correct cold-air supply damper as part of the dual-duct terminal, which is usually a large supply air quantity in sync with a smaller hot-air supply damper in the same terminal unit. This provides temperature control by means of dual-duct box mixing at minimum airflow; it also saves both heating and cooling energy and fan energy, because the terminal damper sizes are more appropriate for the design flow compared to older dual-duct terminals, which provided same-size cold and hot dampers and a single damper operator that did not allow space temperature and airflow controllability. Single-Fan, Dual-Duct (SFDD) System. This system (see Figure 11), frequently used as a retrofit to an antiquated dual-duct, single-fan application during the 1980s and 1990s, uses a single supply fan sized for the peak cooling load or the coincident peak of the hot and cold decks. Fan control is from two static-pressure controllers, one located in the hot deck and the other in the cold deck. The duct requiring the highest pressure governs the airflow by signaling the supply fan VFD speed control. An alternative is to add discharge supply air duct damper control to both the cold and hot ducts to vary flow while the supply fan operates up and down its fan curve. Return air fan

Constant Volume Dual-duct, constant-volume systems using a single supply fan were common through the mid-1980s, and were used frequently as an alternative to constant-volume reheat systems. Today, dual-fan applications are standard. There are two types of dual-duct, singlefan application: with reheat, and without. Single Fan with Reheat. There are two major differences between this and a conventional terminal reheat system: (1) reheat is applied at a central point in the fan unit hot deck instead of at indi-

Fig. 11 Single-Fan, Dual-Duct System

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Fig. 12 Dual-Fan, Dual-Duct System

Licensed for single user. © 2020 ASHRAE, Inc.

(Courtesy RDK Engineers)

Fig. 13 Multizone System (Courtesy RDK Engineers)

tracking of discharge supply air must be assessed with this application. Usually, the cold deck is maintained at a fixed temperature, although some central systems allow the temperature to rise during warmer weather to save refrigeration. The hot-deck temperature is often raised during periods of low outdoor temperature and high humidity to increase the flow over the cold deck for dehumidification. Other systems, particularly in laboratories, use a precooling coil to dehumidify the total airstream or just the outdoor air to limit humidity in the space. Return air quantity can be controlled either by flow-measuring devices in the supply and return duct or by staticpressure controls that maintain space static pressure. Duda (2018b) strongly cautions against applying SFDD in moist climates unless all air is directed across a cooling coil (humidity control; see Figure 11). Otherwise, air traveling down the warm-air branch has no opportunity for dehumidification and some nondehumidified air will be delivered to the space, which can lead to lack of thermal comfort or even conditions hospitable to mold growth. If a zone begins to overcool, the warm-air damper must modulate open to prevent overcooling. If the space is thermally neutral or has only a small net cooling load, supply air can be brought by the warmer duct, which has not been dehumidified. So, loss of indoor relative humidity control will be the result. Dual-Fan, Dual-Duct (DFDD) System. Supply air volume of each supply fan is controlled independently by the static pressure in its respective duct (Figure 12). The return fan is controlled based on the sum of the hot and cold fan volumes using flow-measuring stations. Each fan is sized for the anticipated maximum coincident hot or cold volume, not the sum of the instantaneous peaks; that is, each

fan provides only the amount and temperature of heating or cooling air needed, while maintaining constant airflow in each zone. The cold deck can be maintained at a constant temperature either with mechanical refrigeration, when minimum outdoor air is required, or with an air-side economizer, when outdoor air is below the temperature of the cold-deck set point. This operation does not affect the hot deck, which can recover heat from the return air, and the heating coil need operate only when heating requirements cannot be met using return air. Outdoor air can provide ventilation air via the hot duct when the outdoor air is warmer than the return air. However, controls should be used to prohibit introducing excessive amounts of outdoor air beyond the required minimum when that air is more humid than the return air. This system is highly regarded for critical-care hospitals.

2.3

MULTIZONE SYSTEMS

The multizone system (Figure 13) supplies several zones from a single, centrally located air-handling unit. Different zone requirements are met by mixing cold and warm air through zone dampers at the air-handling unit in response to zone thermostats. The mixed, conditioned air is distributed throughout the building by single-zone ducts. The return air is handled conventionally. The multizone system is similar to the dual-duct system and has the same potential problem with high humidity levels. This system can provide a smaller building with the advantages of a dual-duct system, and it uses packaged equipment, which is less expensive. Packaged equipment is usually limited to about 12 zones, although built-up systems can include as many zones as can be physically incorporated in the layout. A multizone system is somewhat

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Air Handling and Distribution

4.15

more energy efficient than a terminal reheat system because not all the air goes through the cooling coil, which reduces the amount of reheat required. But a multizone system uses essentially the same fan energy as terminal reheat because the airflow is constant. There are two common variations on the multizone system. A three-deck multizone system is similar to the standard multizone system, but bypass zone dampers are installed in the air-handling unit parallel with the hot- and cold-deck dampers. In the Texas multizone system, the hot-deck heating coil is removed from the air-handling unit and replaced with an air-resistance plate matching the cooling coil’s pressure drop. Individual heating coils are placed in each perimeter zone duct. These heating coils are usually located in the equipment room for ease of maintenance. This system is common in humid climates where the cold deck often produces 9 to 11°C air for humidity control. Using the air-handling units’ zone dampers to maintain zone conditions, supply air is then mixed with bypass air rather than heated air. Heat is added only if the zone served cannot be maintained by delivering return air alone. These arrangements can save considerable reheat energy.

2.4

SPECIAL SYSTEMS

Underfloor Air Distribution

Primary/Secondary Licensed for single user. © 2020 ASHRAE, Inc.

may humidify or dehumidify it, and filters it, then supplies this treated air to each of its assigned spaces. DOAS can accommodate an exhaust or relief airflow for heat recovery between the outdoor and exhaust or relief airflows. Air volume is sized in response to minimum ventilation standards, such as ASHRAE Standard 62.1, or to meet makeup air demands. Often, the DOAS serves multiple spaces and is designed not necessarily to control space temperature, but to provide thermally neutral air to those spaces. A second, more conventional system serves those same spaces and is intended to control space temperature. The conventional system is responsible for offsetting building envelope and internal loads. In this instance, however, the conventional system has no responsibility for conditioning or delivering outdoor air. A common example may be a large apartment building with individual fan-coil units (the conventional system) in each dwelling unit, plus a common building-wide DOAS to deliver code-required outdoor air to each housing unit for good indoor air quality and to make up bathroom and/or kitchen exhaust. Another application is to provide minimum outdoor air ventilation while controlling space temperature with radiant panels, chilled beams, valance heating and cooling, etc.

Some processes use two interconnected all-air systems (Figure 14). In these situations, space gains are very high and/or a large number of air changes are required (e.g., in sterile or cleanrooms or where close-tolerance conditions are needed for process work). The primary system supplies the conditioned outdoor air requirements for the process to the secondary system. The secondary system provides additional cooling and humidification (and heating, if required) to offset space and fan power gains. Normally, the secondary cooling coil is designed to be dry (i.e., sensible cooling only) to reduce the possibility of bacterial growth, which can create air quality problems. The alternative is to have the primary system supply conditioned outdoor air (e.g., 20 air changes per hour [ach]) to the ceiling return air plenum, where fan-powered HEPA filter units provide the total supply air (e.g., 120 ach) to the occupied space. The total heat gain from the numerous fan-powered motors in the return air plenum must be taken into account.

Dedicated Outdoor Air Similar in some respects to the primary/secondary system, the dedicated outdoor air system (DOAS) decouples air-conditioning of the outdoor air from conditioning of the internal loads. Long popular in hotels and multifamily residential buildings, DOAS is now gaining popularity in commercial buildings and many other applications. The DOAS introduces 100% outdoor air, heats or cools it,

An underfloor air distribution (UFAD) system (Figure 15) uses the open space between a structural floor slab and the underside of a raised-floor system to deliver conditioned air to supply outlets at or near floor level. Floor diffusers make up the large majority of installed UFAD supply outlets, which can provide different levels of individual control over the local thermal environment, depending on diffuser design and location. UFAD systems use variations of the same basic types of equipment at the cooling and heating plants and primary air-handling units as conventional overhead systems do. Variations of UFAD include displacement ventilation and task/ ambient systems. Humidity control is necessary for UFAD air-handling units. In climates requiring dehumidification of outdoor air, a cooling coil leaving temperature of 10 to 13°C is typical. To maintain the higher plenum inlet temperatures of 16 to 18°C airflow required for delivery, mixing return air with outdoor air is necessary. Figure 15 shows a bypass arrangement where the incoming outdoor and a portion of the return air are mixed and dehumidified through the cooling coil, before the remainder of the return air is bypassed around the cooling coil in controlled mixing to maintain the higher plenum supply air temperature. Displacement ventilation can be a variant of UFAD, using a large number of low-volume supply air outlets to create laminar flow. See Chapter 58 of the 2019 ASHRAE Handbook—HVAC Applications for more information.

Fig. 14 Primary/Secondary System (Courtesy RDK Engineers)

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A task/ambient conditioning (TAC) system is defined as any space-conditioning system that allows thermal conditions in small, localized zones (e.g., regularly occupied work locations) to be individually controlled by nearby building occupants while automatically maintaining acceptable environmental conditions in the ambient space of the building (e.g., corridors, open-use space, other areas outdoor of regularly occupied work space). Typically, the occupant can control the perceived temperature of the local environment by adjusting the speed and direction, and in some cases the temperature, of incoming air supply, much like on the dashboard of a car. TAC systems are distinguished from standard UFAD systems by the higher degree of personal comfort control provided by the localized supply outlets. TAC supply outlets use direct-velocity cooling to achieve this level of control, and are therefore most commonly configured as fan-driven (active) jet-type diffusers that are part of the furniture or partitions. Active floor diffusers are also possible. Unlike conventional HVAC design, in which conditioned air is both supplied and exhausted at ceiling level, UFAD removes supply ducts from the ceilings and allows a smaller overhead ceiling cavity. Less stratification and improved ventilation effectiveness may be achieved with UFAD because air flows from a floor outlet to a ceiling inlet, rather than from a ceiling outlet to a ceiling inlet. UFAD has long been popular in computer room and data center applications, and in Europe in conventional office buildings. Close attention to underfloor tightness is needed because, unlike a supply air

duct that is pressure tested for leakage and installed only by the sheet metal contractor, the entire raised-floor structure of the plenum contains elements from multiple trades (e.g., electrical conduits, floor drains, wall partitions) that all influence underfloor structure. Humidity control is needed to prevent condensation on the concrete floor mass. The air also gains and loses heat as the mass of the flooring heats and cools. For more information, see Chapter 20 of the 2017 ASHRAE Handbook—Fundamentals, ASHRAE’s UFAD Guide (ASHRAE 2016), and Chapters 20 and 32 of the 2019 ASHRAE Handbook—HVAC Applications.

Wetted Duct/Supersaturated Some industries spray water into the airstream at central airhandling units in sufficient quantities to create a controlled carryover (Figure 16). This supercools the supply air, normally equivalent to an oversaturation of about 1.4 g per kilogram of air, and allows less air to be distributed for a given space load. These are used where high humidity is desirable, such as in the textile or tobacco industry, and in climates where adiabatic cooling is sufficient.

Compressed-Air and Water Spray This is similar to the wetted duct system, except that the water is atomized with compressed air and provides limited cooling while maintaining a relatively humid environment. Nozzles are sometimes

Fig. 15 Underfloor Air Distribution (Courtesy RDK Engineers)

Fig. 16 Supersaturated/Wetted Coil (Courtesy RDK Engineers)

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Air Handling and Distribution placed inside the conditioned space and independent of the cooling air supply, and can be used for large, open manufacturing facilities where humidity is needed to avoid static electricity in the space. Several nozzle types are available, and the designer should understand their advantages and disadvantages. Depending on the type of nozzle, compressed-air and water spray systems can require large and expensive air compressors. The extra first cost might be offset by energy cost savings, depending on the application.

Low-Temperature Ice storage is often used to reduce peak electrical demand. Lowtemperature systems (where air is supplied at as low as 5°C) are sometimes used. Benefits include smaller central air-handling units and associated fan power, and smaller supply air duct distribution. At air terminals, fan-powered units are frequently used to increase supply air to the occupied space. Attention to detail is important because of the low air temperature and the potential for excessive condensate on supply air duct distribution in the return plenum space. These fan-powered terminals mix return air or room air with cold supply air to increase the air circulation rate in the space. Alternatively, DFDD systems can be used with low-temperature supply air in one duct. Licensed for single user. © 2020 ASHRAE, Inc.

Smoke Control Air-conditioning systems are often used for smoke control during fires. Controlled airflow can provide smoke-free areas for occupant evacuation and fire fighter access. Space pressurization creates a low-pressure area at the smoke source, surrounding it with high-pressure spaces. The air-conditioning system can also be designed to provide makeup air for smoke exhaust systems in atria and other large spaces (Duda 2004). For more information, see Chapter 54 of the 2019 ASHRAE Handbook—HVAC Applications. Klote et al. (2012) also have detailed information on this topic.

2.5

AIR TERMINAL UNITS

Air terminal units (ATUs) are located between the primary air distribution systems and the conditioned space. There are two main types: (1) passive devices, such as supply outlets (registers or diffusers) and return or exhaust inlets (grilles or registers), and (2) active devices (also known as boxes). Passive devices deliver and extract air throughout the conditioned space without occupants sensing a draft and without creating excessive noise. In systems such as low-velocity all-air systems, airflow amounts and direction entering and leaving the conditioned space are usually manually adjusted; see Chapter 20 of the 2017 ASHRAE Handbook—Fundamentals. With active devices, the ATU controls the quantity and/or temperature of the supply air to maintain desired conditions in the space. In medium- and high-velocity air systems, ATUs normally control air volume, reduce duct pressure, or both. Various types are available. A VAV terminal varies the amount of air delivered. Air may be delivered to low-pressure ductwork and then to the space, or the air terminal may contain an integral air diffuser. A fan-powered VAV terminal varies the amount of primary air delivered, but it also uses a fan to mix ceiling plenum or return air with primary supply air before it is delivered to the space. An all-air induction terminal controls the volume of primary air, induces a flow of ceiling plenum or space air, and distributes the mixture through low-velocity ductwork to the space. An air/water induction terminal includes a coil or coils in the induced airstream to condition the return air before it mixes with the primary air and enters the space. For more information on ATUs, see Chapter 5.

Constant-Volume Reheat Constant-volume reheat ATUs are used mainly in terminal reheat systems with medium- to high-velocity ductwork. The unit serves as

4.17 a pressure-reducing valve and constant-volume regulator to maintain a constant supply air quantity to the space, and is generally fitted with an integral reheat coil that controls space temperature. The constant supply air quantity is selected to provide cooling to match the peak load in the space, and the reheat coil is controlled to raise the supply air temperature as necessary to maintain the desired space temperature at reduced loads.

Variable Air Volume VAV terminal units are available in many configurations, all of which control the space temperature by varying the volume of cool supply air from the air-handling unit in response to room temperature feedback. VAV terminal units are fitted with automatic controls that are most commonly pressure independent (i.e., they measure actual supply airflow and control flow in response to room temperature). Pressure-independent units may be fitted with a velocity-limit control that overrides the room temperature signal to limit the measured supply velocity to some selected maximum. Velocity-limit control can be used to prevent excess airflow through units nearest the air-handling unit. Excessive airflow at units close to the airhandling unit can draw so much supply air that more distant units do not get enough air. Throttling without Reheat. Throttling (or pinch-off) without reheat essentially uses an air valve or damper to reduce supply airflow to the space in response to falling space temperature. The unit usually includes some means of sound attenuation to reduce air noise created by the throttling action. It is the simplest and least expensive VAV terminal unit, but is suitable for use only where no heat is required and the zone served is not occupied, allowing the unit to go completely closed at reduced cooling loads. If this type of unit is set up with a minimum position, it will constantly provide cooling to the space, whether the space needs it or not, and can overcool the space. This approach offers the lowest fan energy use, because it minimizes airflow to just the amount required by the cooling load. Throttling with Reheat. This simple VAV system integrates heating at the terminal unit with the same type of air valve. It is applied in interior and exterior zones where full heating and cooling flexibility is required. These terminal units can be set to maintain a predetermined minimum air quantity necessary to (1) offset the heating load, (2) limit maximum humidity, (3) provide reasonable air movement in the space, and (4) provide required ventilation air. The reheat coil is most commonly hot water or electric resistance. Variable air volume with reheat allows airflow to be reduced as the first step in control; heat is then initiated as the second step. Compared to constant-volume reheat, this procedure reduces operating cost appreciably because the amount of primary air to be cooled and secondary air to be heated is reduced. Many types of controls can provide control sequences with more than one minimum airflow. This type of dual-maximum control allows the ATU to go to a lower flow rate that just meets ventilation requirements at the lightest cooling loads, then increase to a higher flow rate when the heating coil is energized (Taylor et al. 2012). A feature can be provided to isolate the availability of reheat during the summer, except in situations where even low airflow would overcool the space and should be avoided or where increased humidity causes discomfort (e.g., in conference rooms when the lights are turned off). Because the reheat coil requires some minimum airflow to deliver heat to the space, and because the reheat coil must absorb all of the cooling capacity of that minimum airflow before it starts to deliver heat to the space, energy use can be significantly higher than with throttling ATUs that go fully closed. Dual-Duct. Dual-duct systems typically feature throttling dualduct VAV air terminal units. These terminal units are very similar to the single-duct VAV ATUs except that two primary air inlets are

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provided. This allows connection of one primary air inlet to a heating duct and the other to a cooling duct. The dual-duct ATU then modulates both air dampers in response to instructions from a thermostat. Dual-duct ATUs are generally available in a constantvolume output, with cooling and heating dampers operating in tandem but inversely such that the sum total of heating plus cooling is always relatively constant. Dual-duct ATUs are also available in a variable-volume output, with only the cooling or heating damper allowed to stroke open at any given time, such that cooling damper must be closed before the heating damper can stroke open. Minimum positions are available on these dampers to meet minimum ventilation airflow requirements even when little or no airflow would otherwise be required. Induction. The VAV induction system uses an ATU to reduce cooling capacity by simultaneously reducing primary air and inducing room or ceiling air (replacing the reheat coil) to maintain a relatively constant room supply volume. This operation is the reverse of the bypass air terminal. The primary-air quantity decreases with load, retaining the savings of VAV, and the air supplied to the space is kept relatively constant to avoid the effect of stagnant air or low air movement. VAV induction units require a higher inlet static pressure, which requires more fan energy, to achieve the velocities necessary for induction. Today, induction units have for the most part been displaced by fan-powered ATUs, which allow reduction of inlet static pressure and, in turn, reduced central air-handling unit fan power. Fan-Powered. Fan-powered systems are available in either parallel or series airflow. In parallel-flow units, the fan is located outside the primary airstream to allow intermittent fan operation. A backdraft damper on the terminal fan prevents conditioned air from escaping into the return air plenum when the terminal fan is off. In series units, the fan is located in the primary airstream and runs continuously when the zone is occupied. These constant-airflow fan ATUs in a common return plenum can help maintain indoor air quality by recirculating air from overventilated zones to zones with greater outdoor air ventilation requirements. Fan-powered systems, both series and parallel, are often selected because they maintain higher air circulation through a room at low loads but still retain the advantages of VAV systems. As the cold primary-air valve modulates from maximum to minimum (or closed), the unit recirculates more plenum air. In a perimeter zone, a hot-water heating coil, electric heater, baseboard heater, or remote radiant heater can be sequenced with the primary-air valve to offset external heat losses. Between heating and cooling operations, the fan only recirculates ceiling air. This allows heat from lights to be used for space heating, for maximum energy saving. During unoccupied periods, the main supply air-handling unit remains off and individual fan-powered heating zone terminals are cycled to maintain required space temperature, thereby reducing operating cost during unoccupied hours. Fans for fan-powered air-handling units operated in series are sized and operated to maintain minimum static pressures at the unit inlet connections. This reduces fan energy for the central airhandling unit, but the small fans in fan-powered units are less efficient than the large air-handling unit fans. As a result, the series fanpowered unit (where small fans operate continuously) may use more fan energy than a throttling unit system. However, the extra fan energy may be more than offset by the reduction in reheat through the recovery of plenum heat and the ability to operate a small fan to deliver heat during unoccupied hours where heat is needed. Because fan-powered ATUs involve an operating fan, they may generate higher sound levels than throttling ATUs. Acoustical ceilings generally are not very effective sound barriers, so extra care should be taken in considering the sound level in critical spaces near fan-powered terminal units. Both parallel and series fan-powered terminal units should be provided with filters. A disadvantage of this type of terminal unit is

the need to periodically change these filters, making them unsuitable for installation above inaccessible ceilings. A large building could contain hundreds of fan-powered terminal units, some of which might be located in inconvenient locations above office furniture or executive offices. Select installation locations carefully for maximum accessibility. The constant (series) fan VAV terminal can accommodate minimum (down to zero) flow at the primary-air inlet while maintaining constant airflow to the space. Both types of fan-powered units and induction terminal units are usually located in the ceiling plenum to recover heat from lights. This sometimes allows these terminals to be used without reheat coils in internal spaces. Perimeter-zone units are sometimes located above the ceiling of an interior zone where heat from lights maintains a higher plenum temperature. Provisions must still be made for morning warm-up and night heating. Also, interior spaces with a roof load must have heat supplied either separately in the ceiling or at the terminal.

Terminal Humidifiers Most projects requiring humidification use steam. This can be centrally generated as part of the heating plant, where potential contamination from water treatment of the steam is more easily handled and therefore of less concern. Where there is a concern, local generators (e.g., electric or gas) that use treated water are used. Compressed-air and water humidifiers are used to some extent, and supersaturated systems are used exclusively for special circumstances, such as industrial processes. Spray-type washers and wetted coils are also more common in industrial facilities. When using water directly, particularly in recirculating systems, the water must be treated to avoid dust accumulation during evaporation and the build-up of bacterial contamination.

Terminal Filters In addition to air-handling unit filters, terminal filters may be used at the supply outlets to protect particular conditioned spaces where an extra-clean environment is desired (e.g., in a hospital’s surgery suite). Chapter 29 discusses this topic in detail.

2.6

AIR DISTRIBUTION SYSTEM CONTROLS

Controls should be automatic and simple for best operating and maintenance efficiency. Operations should follow a natural sequence. Depending on the space need, one controlling thermostat closes a normally open heating valve, opens the outdoor air mixing dampers, or opens the cooling valve. In certain applications, an enthalpy controller, which compares the heat content of outdoor air to that of return air, may override the temperature controller. This control opens the outdoor air damper when conditions reduce the refrigeration load. On smaller systems, a dry-bulb control saves the cost of enthalpy control and approaches these savings when an optimum changeover temperature, above the design dew point, is established. Controls are discussed in more detail in Chapter 48 of the 2019 ASHRAE Handbook—HVAC Applications and in ASHRAE Guideline 36. Air-handling systems, especially variable-air-volume systems, should include means to measure and control the amount of outdoor air being brought in to ensure adequate ventilation for acceptable indoor air quality. Strategies include the following: • • • •

Separate constant-volume 100% outdoor air ventilation systems Outdoor air injection fan Directly measuring the outdoor air flow rate Modulating the return damper to maintain a constant pressure drop across a fixed outdoor air orifice • Airflow-measuring systems that measure both supply and return air volumes and maintain a constant difference between them.

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Air Handling and Distribution

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• CO2- and/or VOC-based demand-controlled ventilation A minimum outdoor air damper with separate motor, selected for a velocity of 7.5 m/s, is preferred to one large outdoor air damper with minimum stops. A separate damper simplifies air balancing. Proper selection of outdoor, relief, and return air dampers is critical for efficient operation. Most dampers are grossly oversized and are, in effect, unable to control. One way to solve this problem is to provide maximum and minimum dampers. A high velocity across a wide-open damper is essential to its providing effective control. A mixed-air temperature control can reduce operating costs and also reduce temperature swings from load variations in the conditioned space. Chapter 48 of the 2019 ASHRAE Handbook—HVAC Applications shows control diagrams for various arrangements of central system equipment. Direct digital control (DDC) is common, and most manufacturers offer either a standard or optional DDC package for equipment, including air-handling units, terminal units, etc. These controls offer considerable flexibility. DDC controls offer the additional advantage of the ability to record actual energy consumption or other operating parameters of various components of the system, which can be useful for optimizing control strategies. Constant-Volume Reheat. This system typically uses two subsystems for control: one controls the discharge air conditions from the air-handling unit, and the other maintains the space conditions by controlling the reheat coil. Variable Air Volume. Air volume can be controlled by ductmounted terminal units serving multiple air outlets in a control zone or by units integral to each supply air outlet. Pressure-independent volume-regulator units control flow in response to the thermostat’s call for heating or cooling. The required flow is maintained regardless of fluctuation of the VAV unit inlet or system pressure. These units can be field- or factory-adjusted for maximum and minimum (or shutoff) air settings. They operate at inlet static pressures as low as 50 Pa. The type of controls available for VAV units varies with the terminal device. Most use electronic controls. Components for both control and regulation are usually contained in the terminal device. As individual-zone VAV terminals begin to modulate closed, static pressure in the supply air duct increases. To save energy, the AHU’s supply air fan speed should be reset based on VAV box damper position (i.e., the duct static pressure set point is reset lower until one zone damper is nearly wide open). The preferred strategy is trim and respond (Taylor 2015), in which pressure set point is reset slowly and regularly until a zone requests that more pressure is required, in which case the controller bumps the set point up a small amount. Dual Duct. Because dual-duct systems are generally more costly to install than single-duct systems, their use is less widespread. DDC, with its ability to maintain set points and flow accurately, can make dual-duct systems worthwhile for certain applications. They should be seriously considered as alternatives to single-duct systems. Personnel. The skill levels of personnel operating and maintaining the air conditioning and controls should be considered. In large research and development or industrial complexes, experienced personnel are available for maintenance. On small and sometimes even large commercial installations, however, office managers are often responsible, so designs must be in accordance with their capabilities. Water System Interface. On large hydronic installations where direct blending is used to maintain (or reset) the secondary-water temperature, the system valves and coils must be accurately sized for proper control. Many designers use variable flow for hydronic as well as air systems, so the design must be compatible with the air system to avoid operating problems. Relief Fans. In many applications, relief or exhaust fans can be started in response to a signal from the economizer control or to a space pressure controller. The main supply fan must be able to handle the return air pressure drop when the relief fan is not running.

4.19 2.7

AUTOMATIC CONTROLS AND BUILDING MANAGEMENT SYSTEMS

Central air-handling units increasingly come with prepackaged and prewired automatic control systems. Controls may also be accessible by the building manager using a modem to an off-site computer. Larger building HVAC systems integrate manufacturers’ control packages with the building management system (BMS). If the project is an addition or major renovation of space, prepackaged controls and their capabilities need to be compatible with existing automatic controls. Chapter 41 of the 2019 ASHRAE Handbook — HVAC Applications discusses computer applications, and ASHRAE Standard 135 discusses interfacing building automation systems. Automatic temperature controls can be important in establishing a simple or complex control system, more so with all-air systems than with air-and-water and all-water systems. Maintaining these controls can be challenging to building management staff. With a focus on energy management and indoor air quality, the building management system can be an important business tool in achieving sustainable facility management success. Rather than having every consulting engineer write custom control sequences of operation for numerous similar air-handling unit systems, those engineers are now encouraged to specify alreadyvetted controls diagrams, control points lists, and sequences of operation found in ASHRAE Guideline 36. This guideline was created to publish and maintain optimized sequences for VAV air handlers and VAV terminal units, including the trim and respond sequence. Controls manufacturers are expected to preprogram those sequences into their controllers and simply use the programming directly, with only minimal project-specific configuration needed.

2.8

MAINTENANCE MANAGEMENT SYSTEM

Maintenance management for central air-handling units involves many component and devices, with a varied lists of tasks (e.g., check belts, lube fittings, replace filters, adjust dampers), and varied times and frequencies, depending on components and devices (e.g., check damper linkage monthly, change filters based on pressure drop). Small installations may be best served by local service contractors, in lieu of in-house personnel; larger installations may be best served with in-house technicians. See Chapter 40 of the 2019 ASHRAE Handbook — HVAC Applications for further discussion. AHU location affects maintenance. Units located outdoors, especially in climates with long, cold winters, may discourage or impede routine or preventative maintenance. Unfavorable weather can impair important unplanned repairs, such as a broken fan belt during a snow storm. For units located on a building roof, acceptable and safe access must be provided. Roof-mounted units must be located at least 3 m away from the roof edge. Alternatively, if the air-handling equipment must be located closer to the roof edge, a 1 m high parapet or guard rail (designed not to pass a 533 mm sphere) must be provided and needs to extend at least 762 mm beyond each end of the equipment. Air-handling units located indoors require adequate access and clearance for maintenance as well. Model codes require that any room containing equipment be provided with an unobstructed passageway from the door to the equipment, measuring no less than 914 mm wide and 2 m high, and a level working space at least 762 by 762 mm must be provided in front of the control side for servicing. Engineers and designers should insist on adequate mechanical room service space and annotate the service aisles clearly in their designs. In addition to the obvious need for maintenance access, a clear, unobstructed path of egress out of the room is necessary in the event of fire or other emergency.

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2020 ASHRAE Handbook—HVAC Systems and Equipment (SI) BUILDING SYSTEM COMMISSIONING

Prepackaged control systems use a different automatic control checkout process than traditional control contractors. When commissioning a building system that integrates an independent (nonGuideline 36) control system with individual packaged control systems, the process can be more cumbersome because both control contractors need to participate. Air and water balancing for each allair system are also important. With the complexity of air systems and numerous modes of operation (e.g., economizer cycle, minimum outdoor air, smoke control mode), it is essential to adjust and balance systems before system commissioning. Ongoing commissioning or recommissioning should be integral to maintaining each central air system. ASHRAE Guideline 0 and Standard 202 should be applied. During the warranty phase, all-air system performance should be measured and benchmarked to ensure continuous system success. Retro- or recommissioning should be considered whenever the facility is expanded or an additional connection made to the existing systems, to ensure the original design intent is met. When completing TAB and commissioning, consider posting laminated system flow diagrams at or adjacent to the air-handling unit, indicating operating instructions, TAB performance, commissioning functional performance tests, and emergency shutoff procedures. These documents also should be filed electronically in the building manager’s computer server for quick reference. Original basis of design and design criteria should be posted as a constant reminder of design intent, and be readily available in case troubleshooting, expansion, or modernization is needed. For the HVAC design to succeed, commissioning should include the system training requirements necessary for building management staff to efficiently take ownership and operate and maintain the HVAC systems over the service life of the installation. Commissioning continues until the final commissioning report, approximately one year after the construction phase has been completed and the warranty phase comes to an end.

ASHRAE. 2016. Ventilation for acceptable indoor air quality. ANSI/ ASHRAE Standard 62.1-2016. ASHRAE. 2016. Energy standard for buildings except low-rise residential buildings. ANSI/ASHRAE Standard 90.1-2016. ASHRAE. 2016. BACnet®—A data communication protocol for building automation and control networks. ANSI/ASHRAE Standard 135-2016. ASHRAE. 2013. Commissioning process for buildings and systems. ANSI/ ASHRAE/IES Standard 202-2013. ASHRAE. 2013. The commissioning process. ASHRAE Guideline 0-2013. ASHRAE. 2010. Selecting outdoor, return, and relief dampers for air-side economizer systems. Guideline 16-2010. ASHRAE. 2018. High-performance sequences of operation for HVAC systems. ASHRAE Guideline 36-2018. ASHRAE. 2016. UFAD O&M guide: A practical guide for operation and maintenance of underfloor air distribution systems. Duda, S.W. 2016. A critical look at cold supply air systems. ASHRAE Journal 58(12):56-59. Duda, S.W. 2018a. Blow-through vs. draw-through: Air handling units. ASHRAE Journal 60(1)48-52. Duda, S.W. 2018b. Pitfalls of single-fan dual-duct systems in humid climates. ASHRAE Journal 60(9):60-66. Duda, S.W. 2004. Atria smoke exhaust: 3 approaches to replacement air delivery. ASHRAE Journal 46(6):21-27. Klote, J.H., J.A.Milke, P.G. Turnbull., and A. Kashef. 2012. Handbook of smoke control engineering. ASHRAE and the Society of Fire Protection Engineers. Nall, D.H. 2017. Coil bypass AHUs: Avoiding reheat in health-care applications. ASHRAE Journal 59(10):54-61. NFPA. 2017. National electrical code® handbook. National Fire Protection Association, Quincy, MA. Peterson, K.P. 2014. Face velocity considerations in air handler selection. ASHRAE Journal 56(5):56-58. Taylor, S.T. 2014. Return fans in VAV systems. ASHRAE Journal 56(10):54-59. Taylor, S.T. 2015. Resetting setpoints using trim & respond logic. ASHRAE Journal 57(11):52-57. Taylor, S.T. 2018. Designing mega-AHUs. ASHRAE Journal 60(4):50-59. Taylor, S.T., J. Stein, G. Paliaga, and H. Cheng. 2012. Dual maximum VAV box control logic. ASHRAE Journal 54(12):16-25.

REFERENCES

BIBLIOGRAPHY

ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal.ashrae .org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org/bookstore.

ASHRAE. 2017. Method of testing general ventilation air-cleaning devices for removal efficiency by particle size. ANSI/ASHRAE Standard 52.22017. Bauman, F.S. 2003. Underfloor air distribution design guide. ASHRAE. Goodfellow, H., and E. Tahti, eds. 2001. Industrial ventilation design guidebook. Academic Press, New York. McKew, H. 1978. Double duct design—A better way. Heating/Piping/AirConditioning, December. Taylor, S.T. 2016. Making UFAD systems work. ASHRAE Journal 58(3):44-53. Warden, D. 2004. Dual fan, dual duct goes to school. ASHRAE Journal 46(5):18-25. Warden, D. 1996. Dual fan dual duct, better performance at lower cost. ASHRAE Journal 38(1)36-41.

AMCA. 2016. Laboratory methods of testing fans for certified aerodynamic performance rating. ANSI/AMCA Standard 210-07/ANSI/ASHRAE Standard 51-16. Air Movement and Control Association, Arlington Heights, IL. ASHRAE. 2017. Method of testing general ventilation air-cleaning devices for removal efficiency by particle size. ANSI/ASHRAE Standard 52.22017. ASHRAE. 2017. Thermal environmental conditions for human occupancy. ANSI/ASHRAE Standard 55-2017.

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Related Commercial Resources CHAPTER 5

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IN-ROOM TERMINAL SYSTEMS System Characteristics................................................................ 5.1 System Components and Configurations .................................... 5.4 Secondary-Water Distribution .................................................... 5.5 Piping Arrangements .................................................................. 5.5 Fan-Coil Unit and Unit Ventilator Systems................................ 5.6 Variable-Refrigerant-Flow (VRF) Units .................................... 5.8 Chilled-Beam Systems................................................................. 5.8 Radiant-Panel Heating Systems.................................................. 5.9 Radiant-Floor Heating Systems.................................................. 5.9 Induction Unit Systems ............................................................. 5.10

Supplemental Heating Units ..................................................... 5.11 Primary-Air Systems ................................................................. 5.11 Performance Under Varying Load............................................ 5.11 Changeover Temperature.......................................................... 5.12 Two-Pipe Systems with Central Ventilation.............................. 5.12 Four-Pipe Systems .................................................................... 5.16 Automatic Controls and Building Management Systems .......... 5.17 Maintenance Management Systems and Building System Commissioning ...................................................................... 5.17

V

primary, air is delivered by a separate ducted system, either to the terminal unit or ducted directly to the space, and should be pretreated to handle the total latent load of the ventilation air, occupancy, and the space, as well as the sensible load of the ventilation air. Terminal units without central ventilation require additional coil capacity to heat or cool and dehumidify the ventilation air required for the end space. Terminal units are commonly small, with minimal coil rows; therefore, providing this additional capacity is often difficult. Care must be taken to minimize the risk of frozen coils in the winter, and to have enough cooling capacity to not only cool, but also dehumidify ventilation air in the summer. Terminal units have very small fans, so ventilation air must be provided to the unit from a central fan-powered source or supplied from a nearby opening in the building skin, thereby limiting the location of terminal units to the exterior wall of the building. Individual outdoor air fans with filters are also an option. Although a single in-room terminal unit can be applied to a single room of a large building, this chapter covers applying multiple inroom terminal units to form a complete air-conditioning system for a building.

ERY early in the design process, the HVAC design engineer must analyze and ultimately select appropriate systems, as discussed in Chapter 1. Next, production of heating and cooling is selected as decentralized (see Chapter 2) or centralized (see Chapter 3). Finally, distribution of heating and cooling to the end-use space can be done by an all-air system (see Chapter 4), or a variety of allwater or air/water systems and local terminals, as discussed in this chapter. One option is using in-room terminal systems to provide heating and/or cooling to individual zones. Terminal units include consoles, fan-coils, blower coils, unit ventilators, chilled beams, and radiant panels. Terminal systems add heat energy or absorb the heat in the conditioned space served. The medium that transfers the heat either from the space to the outdoors or from a heat source to the conditioned spaces may be the same as used with nonterminal systems. Typical uses of in-room terminal unit systems include hotels/motels, apartments and multifamily dwellings, classrooms, and health care facilities. In older office buildings, in-room terminal units were commonly used for perimeter rooms, combined with central air handlers that served the interior spaces. Systems of this type are making a comeback with the introduction of variable-refrigerant-flow (VRF) equipment, combined with dedicated outdoor air systems (DOAS). Historical preservation projects often use in-room terminal units to minimize the space problems due to running ductwork in historical structures.

1.

SYSTEM CHARACTERISTICS

Terminal-unit systems can be designed to provide complete sensible and latent cooling and heating to an end-use space; however, most terminal systems are best used with a central ventilation system providing pretreated air to the space. Heat can be provided by hot water, steam, direct expansion (DX), or an electric heating coil. Cooling can be provided by chilled-water or DX coils. Heat pumps (discussed in Chapter 2) can be used, either with a piped water loop (water-source) or air cooled. In-room terminals usually condition a single space, but some (e.g., a large fan-coil unit) may serve several spaces. In-room terminal systems can allow individual space control of heating or cooling during intermediate seasons; satisfying the heating and cooling needs of various rooms on a single system. A thermostat for each terminal unit provides individual zone temperature control. A terminal unit used with central ventilation provides the cooling or heating necessary to handle only the sensible heat gain or loss caused by the building envelope and occupancy. Ventilation, or The preparation of this chapter is assigned to TC 9.1, Large Building AirConditioning Systems.

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Advantages Advantages of all in-room terminal unit systems include the following: • The delivery system for the space heating and cooling needs (piping versus duct systems) requires less building space (a smaller central fan room, or none, and little duct space) • System has all the benefits of a central cooling and heating plant, but allows local terminals to be shut off in unused areas • Individual room temperature control allows each thermostat to be adjusted for a different temperature • Minimal cross contamination of recirculated air • Because this system can heat with low-temperature water, it is particularly suitable for use with solar or low-temperature/highefficiency boilers or with heat recovery equipment • Failure of a single in-room unit affects only the one room or area, allowing other spaces to continue to operate • Facilities personnel can remove and replace an in-room terminal unit in hours, allowing them to install a spare and have the room back in service quickly; units are comparatively inexpensive, small, and light, so the owner has the option of stocking spare units on the premises • Maintenance procedures generally can be done by nonlicensed HVAC personnel, allowing in-house crews to complete the tasks • Controls for each unit are very simple • Central control systems can be incorporated to control unit operation and space temperatures during unoccupied hours

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In-room terminal unit systems with central ventilation air have these additional advantages: • The central air-handling apparatus used for central ventilation air is smaller than that of an all-air system because less air must be conditioned at that location. • Space can be heated without operating the primary air system, using just the in-room terminal units. Nighttime primary fan operation is avoided in an unoccupied building. Emergency power for heating, if required, is much lower than for most all-air systems. • Dehumidification, filtration, and humidification of ventilation air are performed in a central location remote from conditioned spaces. • They allow using central heat recovery devices such as heat wheels. • Ventilation air is positively supplied and can accommodate constant recommended outdoor air quantities, regardless of the temperature control of the room. • Use of central ventilation air with terminal units in some climates can prevent the negative pressurization problems that occur when occupants turn off in-room units.

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Disadvantages • For many buildings, in-room terminals are limited to perimeter space; separate systems are required for other areas. • In-room terminal unit fans with very little, if any, ductwork can be noisy. Many manufacturers have addressed this and provided units that are acceptable in many situations. • Because of the individual space control, more controls are needed than for many all-air systems. • The system is not appropriate for spaces with high exhaust requirements (e.g., research laboratories) unless supplementary ventilation air is provided. • Central dehumidification eliminates condensation on the secondarywater heat transfer surface (see the section on Secondary-Water Distribution) under maximum design latent load, but abnormal moisture sources (e.g., open windows, cooking, people congregating, a failed primary-air system) can cause annoying or damaging condensation. Therefore, a condensate pan must be provided as for other systems. • Primary-air supply usually is constant with no provision for shutoff. This is a disadvantage in residential applications, where tenants or hotel room guests may prefer to turn off the air conditioning, or where management may desire to do so to reduce operating expense; however, this can help to stabilize pressurization of a building when exhaust fans are centrally controlled or in constant operation. • Low chilled-water temperature and/or deep chilled-water coils are needed at the central ventilation air unit to control space humidity adequately. Low chilled-water temperatures can result in excessive condensation occurring at terminal units, if chilledwater valves are not used to shut off water flow through the terminal unit when the terminal unit fan is off. • Low primary-air temperatures can require heavily insulated ducts; however, using neutral air temperatures minimizes this requirement and prevents overcooling of some spaces. • In-room terminal units without central ventilation air may result in greater infiltration levels due to numerous penetrations of the exterior of the building, which must be sealed adequately to prevent air infiltration and water intrusion. This may be accentuated in winter conditions, when wind pressures and stack effects become more significant. • Adding necessary humidity in the winter is difficult. • Maintenance must be done within the occupied space, which may disrupt space use.

Heating and Cooling Calculations Basic calculations for airflow, temperatures, relative humidity, loads, and psychrometrics are covered in Chapters 1, 17, and 18 of the 2017 ASHRAE Handbook—Fundamentals. Use caution in determining peak load conditions for each space served by a terminal unit. Rather than depending on guidelines for typical lighting, ventilation, infiltration, equipment, and occupancy levels, the designer should try to get realistic values based on the particular owner’s use plans for the facility. If the client has an existing or similar facility, visiting it to understand the actual occupancy hours, concentration of equipment, and occupancy should help avoid unrealistic assumptions. Incorporating effects of planned energysaving features (e.g., daylighting; high-efficiency/low-heat-producing lighting; shading apparatus for privacy, glare, or solar radiant control; fan and outdoor air modulation; full-building energy management control strategies) can prevent oversizing of terminal units and resulting potential loss of humidity control. Determining areas where the normal base load will be a small percentage of the concentrated usage load and understanding the usage schedule for the area can allow equipment selection and control strategies to maximize energy savings and still provide excellent comfort control in both extremes. Energy standards may result in calculations that differ from actual use. For example, in a zone with a terminal unit sized for 2 kW (common for offices), a simple change of four 100 W incandescent light bulbs to compact fluorescents could reduce the space’s heat load by about 18% of the unit’s capacity, or even more with LED. If this is the consistent peak load of the space for future years, the terminal unit is oversized. Unless a central ventilation pretreatment system handles the entire latent load of the space and the outdoor air, humidity control will be lost in the space. In many climate zones, this may not be significant. Integrated building design (IBD) techniques should be used to ensure the building envelope provides adequate energy efficiency, airtightness, and moisture penetration control to allow terminal units to control the indoor environmental conditions without need for excessive moisture control in each space. Close cooperation of all parties during design is necessary to create an overall building design that minimizes the required mechanical systems’ energy consumption while achieving good indoor conditions. For details on IBD, see Chapter 60 of the 2019 ASHRAE Handbook—HVAC Applications. Computer programs generally can model primary ventilation systems as well as secondary in-room terminal systems. Most, however, do not allow the user to assign the room latent load as part of the primary ventilation system capacity requirement. Therefore, the designer needs to either manually determine final capacity requirements for both the primary and in-room units, or use overrides and manual inputs to redistribute the loads within the computer program after initial sensible and latent loads as well as block and peak conditions have been determined. The design refrigeration load is determined by considering the entire portion or block of the building served by the air-and-water or DX system at the same time. Because the load on the secondarywater system depends on the simultaneous demand of all spaces, the sum of the individual room or zone peaks is not considered, unless individual DX systems are used.

Space Heating Some in-room terminal units provide only heating to the end space. Equipment such as cabinet or unit heaters, radiant panels, radiant floors, and finned-tube radiators are designed for heating only. Extreme care must be used with these systems if they are incorporated into a two-pipe changeover piping distribution system, or any other system in which secondary water being piped is not

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In-Room Terminal Systems consistently over 38°C. The heating coils in these units are not designed to handle condensation, and there is no drain pipe in the unit. If cold water is provided to these units, dripping condensation from units, valves, or piping may damage building finishes or saturate the insulation, leading to mold growth. Ball valves tied into the automatic temperature control (ATC) system and/or aquastats should be provided to prevent water at temperatures below space air dew point from reaching heating-only terminal units.

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Central (Primary-Air) Ventilation Systems Generally, the supply air volume from the central apparatus is constant and is called primary or ventilation air to distinguish it from recirculated room air or secondary air. The quantity of primary air supplied to each space is determined by the amount of outdoor air required by codes and other applicable guidelines for ventilation. If in-room terminal units are heating-only units, then the primary-air system must also provide the required sensible cooling capacity at maximum room cooling load, unless cooling is not required. The air may be from outdoors, or may be mixed outdoor and return air. During the cooling season, air is dehumidified sufficiently in the central conditioning unit to maintain required humidity conditions and to prevent condensation on the in-room terminal unit cooling coil. (Both outdoor air and space latent loads should be handled by the central unit.) Centrally supplied air can be supplied at a low enough dew point to absorb moisture generated in the space, but as a minimum should be supplied at a condition such that the room terminal unit has to remove only the space-generated latent load (this is only appropriate with unit ventilators with the capability to handle the latent space loads). In winter, moisture can be added centrally to limit dryness. As the primary air is dehumidified, it is also cooled. The air should be cool enough when delivered to offset part of the room sensible loads without overcooling the spaces. If the necessary amount of cooling (i.e., that needed to dehumidify primary air to handle all of the space and outdoor air latent loads) is likely to overcool the end spaces, the primary air should be adjusted to avoid overcooling. A heating coil may be required as a preheater in areas with freezing weather. Alternatively, heating can occur at the terminal units, where only a minimal number of spaces might be overcooled due to outdoor air conditions. When outdoor air is introduced from a central ventilation system, it may be connected to the inlet plenum of some in-room terminal units (fan-coils, unit ventilators, active chilled beams, or induction units), or introduced directly into the space. If introduced directly, ensure that this air is pretreated, dehumidified, and held at a temperature approximately equal to room temperature so as not to cause occupant discomfort when the space unit is off. Caution should always be used to prevent overcooling or loss of humidity control from ventilation air, which can lead to condensation problems on surfaces as well as discomfort for occupants. In the ideal in-room terminal unit design, the cooling coil is always dry, greatly extending terminal unit life and eliminating odors and the possibility of bacterial growth in the unit in the occupied space, as well as limiting condensation issues in the spaces. In this case, in-room terminals may be replaced by radiant panels (see Chapter 6) or chilled beams and panels, as the primary air controls the space humidity. Therefore, the moisture content of primary air must be low enough to offset the room’s latent heat gain and to maintain a room dew point low enough to preclude condensation on the secondary cooling surface. Even though some systems operate successfully with little or no condensate, a condensate drain is recommended. In systems that shut down during off hours, start-up load may include a considerable dehumidification load, producing moisture to be drained away. In climates with elevated outdoor air dew points during space-cooling

5.3 periods, piped condensate removal systems that can be maintained regularly should always be included.

Central Plant Sizing Central equipment size is based on the block load of the entire building at the time of the building peak load, not on the sum of individual in-room terminal unit peak loads. Cooling load should include appropriate diversity factors for lighting and occupant loads. Heating load is based on maintaining the unoccupied building (where there is not continuous occupancy) at design temperature, plus an additional allowance for pickup capacity if the building temperature is set back at night. For additional information, see Chapter 3. If water supply temperatures or quantities are to be reset at times other than at peak load, the adjusted settings must be adequate for the most heavily loaded space in the building. Analysis of individual room load variations is required. If the side of the building exposed to the sun or interior zone loads requires chilled water in cold weather, consider using condenser water with a water-to-water heat exchanger, a four-pipe system, or other water economizer methods. Varying refrigeration loads require the water chiller to operate satisfactorily under all conditions, but verify accordance with local energy laws for allowable operation.

Building Pressurization As with any HVAC system, the amount of ventilation air required depends on the number of occupants in the space as well as other factors (see ASHRAE Standard 62.1). The rate of airflow per person or per unit area is also usually dictated by state codes, based on activity in the space and contaminant loads. If the amount of ventilation air required is considerable (i.e., 10% or more of a space’s total supply air volume), the designer needs to consider how the excess air will move out of the space and the building. Means of preventing overpressurization may have to be provided, depending on building tightness, amount of exhaust, and other considerations. Additionally, the designer should consider heat recovery for the rejected air.

First, Operating, and Maintenance Costs As with all systems, the initial cost of an in-room terminal system varies widely, depending on location, local economy, and contractor preference (even for identical systems). For example, a through-wall unit ventilator system is less expensive than fan-coil units with a central ventilation system, because it does not require extensive ductwork distribution. The operating cost depends on the system selected, the designer’s skill in selecting and correctly sizing components, and efficiency of the duct and piping systems. A terminal unit design without a central ventilation system is often one of the less expensive systems to install, but in most situations will not operate without some condensation and humidity issues. Because in-room terminal equipment is in each occupied space, maintenance may be more time consuming, depending on the size and use of the facility. The equipment and components are less complex than other equipment. A common method of repairing in-room terminal units is to simply disconnect and replace a nonfunctioning unit, minimizing the time spent in the occupied space. The nonfunctioning unit can then be repaired in a workshop and used as a spare. The number of individual units is much greater than in many systems, therefore increasing the number of control valves. If a building automation system (BAS) is used, the number of control points is increased as well, which raises the first cost and maintenance costs of the controls.

Energy The engineer’s early involvement in design of any facility can considerably reduce the building’s energy consumption; collaboration between the architect and engineer allows optimization of

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energy conservation measures, including the building envelope, as well as selection of an HVAC system that is energy efficient with minimal operational costs. In practice, however, a system might be selected based on low first cost or to perform a particular task. In general, terminal units can save energy if the BAS controls operation of the units and can deenergize them if the space is unoccupied. This adds significant cost to the control system, but can save on operational costs. If a central ventilation system is used, energy recovery in the air handler should be considered to minimize operation costs. The choice of two- or four-pipe piping system and the design of the piping system is another energy-impacting decision. Pumping power, insulation to control heat losses, and the number of yearly changeovers all affect the appropriateness and available energy savings of each type of piping system.

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Life-Cycle Costs Life-cycle costs include the first, operational, maintenance, and replacement costs over a predetermined length of time. With an inroom terminal system, a major portion of the cost of the system is the insulated piping infrastructure needed to provide the primary and secondary equipment with water (or refrigerant). When properly selected, installed, and maintained, these components have a very long expected life. The in-room terminal units are generally of simple design and complexity and can be readily repaired. As a general rule, the less condensate that occurs at the in-room units, the longer their life span. Selection of good-quality central ventilation equipment can result in quite lengthy life spans for these units. Operational concerns that affect costs are discussed throughout this chapter. Consult Chapter 38 of the 2019 ASHRAE Handbook— HVAC Applications for more information on equipment life expectancies and predicting maintenance costs.

2.

SYSTEM COMPONENTS AND CONFIGURATIONS

Components Terminal units have common components; mainly, a fan, coil(s), filter, dampers, and controls (although some units only have coils and controls). Damper. If a terminal unit is providing ventilation air through the envelope of the building, a damper is needed to stop airflow when the room is at unoccupied status. Because in-room terminal units often have a ducted primary- or central ventilation air system, a damper on the primary system duct allows airflow to be balanced. Many less expensive in-room terminal units have manual dampers rather than automatic. Regardless of whether dampers are manually operated or automatic, pressurization problems can occur because of damper position compared with exhaust fan operation and primary air system operation, so the designer needs to carefully coordinate how building pressurization is controlled and maintained. Filtration. Filtration capabilities with in-room terminals are generally minimal. The cabinet and component assembly often provide very little ability to improve the filtration capability, although fan-coil-style terminal units commonly can handle better filtration. Manufacturer-supplied filters are often either washable or throwaway filters. Some manufacturers provide an option for a higher-quality pleated-style filter to be used, but their recommendations for motor selection should be obeyed. Maintenance instructions for washable filters should be carefully followed, to avoid filter impaction and reduced airflow; good filter maintenance improves sanitation and provides full airflow, ensuring full-capacity delivery. In most areas, ASHRAE Standard 62.1 now requires filtration of MERV 8 or higher. Heating and Cooling Coils. Coils in terminal units are usually available in one-, two-, three-, and sometimes four-row coils for cooling, and one- or two-row coils for heating. In units with untreated

outdoor air, selecting coil and fin materials and coatings for longer life expectancy should be carefully considered (see Chapter 23). Only the building envelope and internal space heating and cooling loads need to be handled by in-room terminal units when outdoor air is adequately pretreated by a central system to a neutral air temperature of about 21°C. This pretreatment should reduce the size and cost of the terminal units. All loads must be considered in unit selection when outdoor air is introduced directly through building apertures into the terminal unit, as is sometimes done with unit ventilators. For cold climates, coil freeze protection must be considered. Fan. Terminal units typically are not complex. Most modern units have ECM motors, a variable-speed drive (VSD), or other speed control on the fan. Duct Distribution. Terminal units work best without extensive ductwork. With ducts, static pressure on the fan (instead of the coil capacity) may be the determining factor for sizing the terminal units, because multiple fan selections are not normally available. Automatic Controls. Most terminal units are controlled with a standard electronic thermostat, either provided by the manufacturer or packaged by the automatic temperature controls (ATC) contractor. Thermostats capable of seven-day programming and night setback can improve energy savings where space usage is predictable enough to allow consistent programs. Terminal units can be incorporated into a BAS, but the cost to do so may be prohibitive, depending on the number of terminal units in the building. The potential operational cost savings, especially in facilities where a significant percentage of areas are often unoccupied, should be evaluated against the first-cost consideration. Capacity Control. Terminal unit capacity is usually controlled by coil water or refrigerant flow, fan speed, or both. Water flow can be thermostatically controlled by return air temperature or a wall thermostat and two- or three-way valve. Unit controls may be a selfcontained direct digital microprocessor, line voltage or low-voltage electric, or pneumatic. Fan speed control may be automatic or manual; automatic control is usually on/off, with manual speed selection. Room thermostats are preferred where automatic fan speed control is used. Return air thermostats do not give a reliable index of room temperature when the fan is off or when outdoor air is introduced nearby. Residential fan-coil units may have manual three-speed fan control, with water temperature (both heating and cooling) scheduled based on outdoor temperature. On/off fan control can be poor because (1) alternating shifts in fan noise level are more obvious than the sound of a constantly running fan, and (2) air circulation patterns in the room are noticeably affected. However, during cooling cycles, constant fan operation results in higher relative humidity levels than fan cycling, unless all latent load is handled by a primary central ventilation system. For systems without primary central ventilation, summer room humidity levels tend to be relatively high, particularly if modulating chilled-water control valves are used for room temperature control. Alternatives are two-position control with variable-speed fans (chilled water is either on or off, and airflow is varied to maintain room temperature) and the bypass unit variable chilled-water temperature control (chilled-water flow is constant, and face and bypass dampers are modulated to control room temperature). The designer must be careful to understand the unit’s operating conditions for the vast majority of the operation hours, and how the unit will actually perform at those times. Many manufacturers publish sensible and latent capacity information that is developed from testing at a single set of conditions (generally the AHRI standard condition), then use computer modeling or extrapolation rather than actual testing to determine operation capacities at other conditions. In most situations, the unit’s actual operation for the vast majority of time is at conditions other than the rating condition. Additionally, control choices for water flow or airflow may create conditions different from those used by the manufacturer to determine the published operation

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In-Room Terminal Systems capacities. Because peak conditions are often used to select coils but in most applications only occur a small percentage of the time, oversizing of coils and loss of humidity control become common if these issues are not carefully thought through during design and selection of the equipment.

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Configurations Terminal units are available in many different configurations; however, not all configurations are available for all types of terminal units. The designer should evaluate the air pathways of the specific units being considered, because some allow raw outdoor air to bypass the filter and/or the coils. Additionally, on some unit ventilators with face and bypass, the raw outdoor air is allowed to stratify in the bypass section, again bypassing the coils. Depending on climate conditions and filtration requirements, these configurations may not be appropriate. Low-profile vertical units are available for use under windows with low sills; however, in some cases, the low silhouette is achieved by compromising features such as filter area, motor serviceability, and cabinet style. Floor-to-ceiling, chase-enclosed units are available in which the water and condensate drain risers are part of the factory-furnished unit. Stacking units with integral prefabricated risers directly one above the other can substantially reduce field labor for installation, an important cost factor. These units are used extensively in hotels and other residential buildings. For units serving multiple rooms (single-occupancy suite-type spaces), the supply and return air paths must be isolated from each other to prevent air and sound interchange between rooms. Perimeter-located units give better results in climates or buildings with high heating requirements. Heating is enhanced by underwindow or exterior wall locations. Vertical units with finned risers can operate as convectors with the fans turned off during night setback, and overheating can become an issue. Horizontal overhead units may be fitted with ductwork on the discharge to supply several outlets. A single unit may serve several rooms (e.g., in an apartment house where individual room control is not essential and a common air return is feasible). Units must have larger fan motors designed to handle the higher static pressure resistance of the connected ductwork. Horizontal models conserve floor space and usually cost less, but when located in furred ceilings, they can create problems such as condensate collection and disposal, mixing return air from other rooms, leaky pans damaging ceilings, and difficult access for filter and component removal. In addition, possible condensate leakage may present air quality concerns. Other sections in this chapter discuss specific configurations for the type of unit, and indicate additional chapters in the volume that provide diagrams and additional information.

3.

SECONDARY-WATER DISTRIBUTION

The secondary-water system includes the part of the water distribution system that circulates water to room terminal units when the water has been cooled or heated either by extraction from or heat exchange with another source in the primary circuit. In the primary circuit, water is cooled by flow through a chiller or is heated by a heat input source. Water flow through the in-room terminal unit coil performs secondary cooling or heating when the room air (secondary air) gives up or gains heat. Secondary-water system design differs for two- and four-pipe systems. Secondary-water systems are discussed in Chapter 13.

4.

PIPING ARRANGEMENTS

For terminal units requiring chilled and/or hot water, the piping arrangement determines the performance characteristics, ease of

5.5 operation, and initial cost of the system. Each piping arrangement is briefly discussed here; for further details, see Chapter 13.

Four-Pipe Distribution Four-pipe distribution of secondary water has dedicated supply and return pipes for chilled and hot water. The four-pipe system generally has a high initial cost compared to a two-pipe system but has the best system performance. It provides (1) all-season availability of heating and cooling at each unit, (2) no summer/winter changeover requirement, (3) simpler operation, and (4) hot-water heating that uses any heating fuel, heat recovery, or solar heat. In addition, it can be controlled at the terminal unit to maintain a dead band between heating and cooling so simultaneous heating and cooling cannot occur.

Two-Pipe Distribution Two-Pipe Changeover Without Central Ventilation. In this system, either hot or cold water is supplied through the same piping. The terminal unit has a single coil. The simplest system with the lowest initial cost is the two-pipe changeover with (1) outdoor air introduced through building apertures, (2) manual three-speed fan control, and (3) hot- and cold-water temperatures scheduled by outdoor temperatures. The changeover temperature is set at some predetermined set point. If a thermostat is used to control water flow, it must reverse its action depending on whether hot or cold water is available. The two-pipe system cannot simultaneously provide heating and cooling, which may be required during intermediate seasons when some rooms need cooling and others need heat. This characteristic can be especially troublesome if a single piping zone supplies the entire building, but may be partly overcome by dividing the piping into zones based on solar exposure. Then each zone may be operated to heat or cool, independent of the others. However, one room may still require cooling while another room on the same solar exposure requires heating, particularly if the building is partially shaded by an adjacent building or tree. Another system characteristic is the possible need for frequent changeover from heating to cooling, which complicates operation and increases energy consumption to the degree that it may become impractical. For example, two-pipe changeover system hydraulics must consider the water expansion (and relief) that occurs during cycling from cooling to heating. Caution must be used when this system is applied to spaces with widely varying internal loads, and outdoor air is introduced through the terminal unit instead of through a central ventilation system. Continuous introduction of outdoor air when the load is reduced often results in sporadically unconditioned outdoor air, which can cause high space humidity levels, unless additional separate dehumidification or reheat is used. The outdoor air damper in the unit must be motor-operated so it can be closed during unoccupied periods when minimal conditioning is required. The designer should consider the disadvantages of the two-pipe system carefully; many installations of this type waste energy and have been unsatisfactory in climates where frequent changeover is required, and where interior loads require cooling and exterior spaces simultaneously require heat. Most energy codes are difficult to meet with this system. Two-Pipe Changeover with Partial Electric Strip Heat. This arrangement provides heating in intermediate seasons by using a small electric strip heater in the terminal unit. The unit can handle heating requirements in mild weather, typically down to 4°C, while continuing to circulate chilled water to handle any cooling requirements. When the outdoor temperature drops sufficiently to require heating beyond the electric strip heater capacity, the water system must be changed over to hot water. Most energy codes do not allow electric heating.

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Two-Pipe Nonchangeover with Full Electric Strip Heat. This system may not be recommended for energy conservation, but it may be practical in areas with a small heating requirement or in other situations where life-cycle costs support this choice. Many energy codes will not allow this system unless energy analysis can prove compliance by life-cycle analysis.

Three-Pipe Distribution Three-pipe distribution uses separate hot- and cold-water supply pipes. A common return pipe carries both hot and cold water back to the central plant. The terminal unit control introduces hot or cold water to the common unit coil based on the need for heating or cooling. This type of distribution is not recommended because of its energy inefficiency from constantly reheating and recooling water, and it does not comply with most recognized energy codes. Further, modern chillers and boilers cannot survive with the resultant changeover water temperatures.

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Condenser Water Systems with Heat Pump Terminal Units Condenser water systems are very similar to the two-pipe changeover with partial electric strip heat. The supply and return pipes carry water at more moderate temperatures than those of typical chilled or hot water. The heat pumps use the water as a heat source for heating and as a heat sink in cooling mode, allowing various in-room heat pumps to meet room needs during intermediate seasons. If additional heat is needed in a room, a small electric strip heater in the terminal unit can provide it, though this should be avoided for energy reasons. When the outdoor temperature drops sufficiently to require additional heating capacity, the water system must be changed over to hot water. Chapters 9, 14, and 48 contain additional information about condenser water piping systems and heat pumps.

5.

FAN-COIL UNIT AND UNIT VENTILATOR SYSTEMS

Fan-coil units and unit ventilator systems are similar; their common traits and differences are discussed here. Both (1) can provide cooling as well as heating, (2) normally move air by forced convection through the conditioned space or fan-powered flow, (3) filter circulating air, and (4) may introduce outdoor ventilation air. Fancoils are available in various configurations to fit under windowsills, above furred ceilings, in vertical pilasters built into walls, etc., whereas unit ventilators are available in three main configurations: floor-mounted below a window, horizontal overhead with ducted supply and return, and stacking vertical units. Fan-coils are often used in applications where ventilation requirements are minimal. Unit ventilators are similar, except that unit ventilators are designed to provide up to 100% outdoor air to the space. Basic elements of both fan-coils and unit ventilators are a finned-tube heating/cooling coil, filter, and fan section (Figure 1). Unit ventilators can include a face-and-bypass damper. The fan recirculates air from the space through the coil, which contains either hot or chilled water or refrigerant. The unit may contain an additional electric resistance, steam, or hot-water heating coil, though these may be regulated by energy codes. The electric heater is often sized for fall and spring to avoid changeover problems in two-pipe systems; it may also provide reheat for humidity control. Due to comfort requirements and equipment capabilities, two-pipe systems are seldom used for new construction. A cleanable or replaceable moderate-efficiency filter upstream of the fan helps prevent clogging of the coil with dirt or lint entrained in recirculated air. It also helps protect the motor and fan, and can reduce the level of airborne contaminants in the conditioned space. Further, much higher filter efficiencies are recommended by ASHRAE standards and many energy codes. The fan and motor assembly is

arranged for quick removal for servicing. The units generally are also equipped with an insulated drain pan. Most manufacturers furnish units with cooling performance certified as meeting Air-Conditioning, Heating, and Refrigeration Institute (AHRI) standards. Unit prototypes have been tested and labeled by Underwriters Laboratories (UL) or Engineering Testing Laboratories (ETL), as required by some codes. Requirements for testing and standard rating of room fan-coils with airdelivery capacities of 708 L/s or below are described in AHRI Standard 440 and ASHRAE Standard 79. Requirements for testing and standard rating of room unit ventilators with air delivery capacities of 1416 L/s or below are described in AHRI Standard 840. For the U.S. market, fan-coil units are generally available in nominal sizes of 95, 140, 190, 285, 375, 565, 755, and 945 L/s, and unit ventilators in nominal sizes of 350, 470, 700, and 940 L/s. Both types of units can often be purchased with multispeed, highefficiency fan motors.

Types and Location Floor-mounted units have various ventilation air ductwork connections, including from the back or a ducted collar on the top of the cabinet. Ceiling-mounted and stacking units can be mounted completely exposed, partially exposed in a soffit, fully recessed, or concealed. Ventilation air connections can be made in the back or top of the unit, as well as the side in some cases. For existing building retrofit, it may be easier to install piping, outdoor air, and wiring for a terminal unit system than the large ductwork required for an all-air system. Common fan-coil system applications are hotels, motels, apartment buildings, and small office buildings and retail spaces. Fan-coil systems are used in many hospitals, but they are less desirable because of the low-efficiency filtration and difficulty in maintaining adequate cleanliness in the unit and enclosure. In addition, limits set by the American Institute of Architects’ Guidelines for Design and Construction of Hospital and Health Care Facilities (AIA 2001) do not allow air recirculation in certain types of hospital spaces. Unit ventilator systems are most frequently used in classrooms, which need a high percentage of outdoor air for proper ventilation. Unit ventilators are often located under a window along the perimeter wall. They are available in a two-pipe configuration with changeover, two-pipe with electric heating, four-pipe, or with heating and DX coils for spaces (e.g., computer rooms) that may require year-round cooling. Limited ductwork may be allowed, allowing for higher and often exposed ceiling systems. DX heat pump unit ventilators may be used in areas with wide variations in daily loads, such as deserts and areas near oceans. These units can often be misused as shelving in classroom; books and paperwork may be stacked

Unit ventilators use same/or similar components; see Chapter 28 for unit ventilator configurations.

Fig. 1 Typical Fan-Coil Unit

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In-Room Terminal Systems on top of them, impeding airflow to the space. Also, ventilation air intake louvers can become choked by vegetation if they are not properly maintained. In addition, because the fans are sized to accommodate 100% ventilation air, they are typically noisier than fan-coils, although recent developments have led to new units that are much quieter. For existing building retrofit, it is easiest to replace unit ventilators in kind. If a building did not originally use unit ventilators, installing multiple ventilation air intake louvers to accommodate the unit ventilators may be cost prohibitive. Likewise, installing a different type of system in a building originally fitted with unit ventilators requires bricking up intake louvers and installing exposed ductwork (if there is no ceiling plenum) or creating a ceiling space in which to run ductwork. Unit ventilators are best applied where individual space temperature control with large amounts of ventilation air is needed.

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Ventilation Air Requirements Fan-coil and unit ventilators often receive ventilation air from a penetration in the outer wall or from a central air handler. Units that have outdoor air ducted to them from an aperture in the building envelope are not suitable for tall buildings because constant changes in wind pressure cause variations in the amount of outdoor air admitted. In this situation, ventilation rates can also be affected by stack effect, wind direction, and speed. Also, freeze protection may be required in cold climates, because preheating outdoor air is not possible. Historically, fan-coils have often been used in residential construction because of their simple operation and low first cost, and because residential rooms were often ventilated by opening windows or by outer wall apertures rather than a central system. Operable windows can cause imbalances with a ducted ventilation air system; current standards have moved toward requiring specific ventilation airflow control in residences, therefore minimizing the ability to use noncentralized ventilation systems, even in residential applications. In many cases, ventilation can be controlled by using an outdoor air fan with a high-efficiency filter. Unlike fan-coils, unit ventilators can provide the entire volume of ventilation air that is required in many applications. The heating/ cooling coils in unit ventilators therefore differ considerably from fan-coils. Coils in unit ventilators are much deeper, because the unit ventilator needs to be able to heat, cool, and dehumidify up to 100% ventilation air. Coil selection must be based on the temperature of the entering mixture of primary and recirculated air, and air leaving the coil must satisfy the room’s sensible and latent cooling and heating requirements. If variable occupancy levels regularly occur during system operation hours, such as often occurs with classrooms, sizing a single unit for the fully occupied outdoor air requirement could result in oversized cooling capacity during many hours of operating time. This could result in loss of humidity control. For variable-occupancy applications, demand control ventilation or pretreated outdoor air is recommended.

Selection Some designers select fan-coil units and unit ventilators for nominal cooling at medium speed when a three-speed control switch is provided, to enable quieter operation in the space and add a safety factor (sensible capacity can be increased by operating at high speed). Sound power ratings are available from many manufacturers. If using a horizontal overhead unit with ducted supply and return, fan capacity may be the factor that decides the unit’s size, not the coil’s capacity. Static pressure as little as 75 Pa can significantly affect fan air volume and unit capacity. If the unit is selected to provide full capacity at medium speed, the unit must also be able to handle the full volume of required ventilation air at that airflow, which is not an issue when using a supplemental

5.7 outdoor air fan and filter. If cooling loads vary more than 20% during operation hours, it is highly likely this selection could result in oversized capacity operation and loss of humidity control. Current fan motor speed control options, such as electronically commutated motors (ECMs), and units designed to operate more quietly allow unit selection for the fan’s full-speed total capacity, and minimize the chance of oversizing.

Wiring Fan-coil and unit ventilator blower fans are driven by small motors. Fan-coil motors are generally shaded pole or capacitor start with inherent overload protection. Operating wattage of even the largest sizes rarely exceeds 300 W at high speed. Running current rarely exceeds 2.5 A. Unit ventilator motors are typically 0.4 kW or less. Operating power of even the largest sizes rarely exceeds 400 W at high speed. Almost all motors on units in the United States are 120 V, single-phase, 60 Hz current. In planning the wiring circuit, follow all required codes. The preferred wiring method generally provides separate electrical circuits for fan-coil or unit ventilator units and does not connect them into the lighting circuit, or other power circuits. Separate wiring connections may be needed for condensate pumps.

Condensate Even when outdoor air is pretreated, a condensate removal system must be installed for fan-coil units and unit ventilators. Drain pans should be integral for all units. For floor-mounted units along the perimeter of the building, condensate piping may run from the drain pan to the exterior grade, depending on the authorities having jurisdiction. Where drainage by gravity will not be sufficient, provide condensate pumps. Condensate drain lines should be properly sized and maintained to avoid clogging with dirt and other materials. Where approved and safe, use of plastic piping provides a safer drainage system with reduced clogging. Condensation may occur on the outside of drain piping, which requires that these pipes be insulated. Many building codes have outlawed systems without condensate drain piping because of the potential damage and possibility of mold growth in stagnant water accumulated in the drain pan.

Capacity Control Fan-coil and unit ventilator capacity is usually controlled by coil water or refrigerant flow, fan speed, or a combination of these. In addition, unit ventilators often are available with a face-and-bypass damper, which allows for another form of capacity control. For additional information, see the discussion on capacity control in the section on System Components and Configurations.

Maintenance Fan-coil and unit ventilator systems require more maintenance than central all-air systems, and the work must be done in occupied areas. Units operating at low dew points require regular (multiple times per year) cleaning and flushing of condensate pans and drains to prevent overflow and microbial build-up. Coils should be cleaned at least once a year, and more often when they consistently are removing moisture. The physical restraints of in-room terminal units located high in rooms or concealed in ceilings, soffits, etc., can create challenges for proper coil cleaning. Water valves, controls, and dampers should also be checked yearly for proper calibration, operation, and needed repairs. Filters are small and low- to medium-efficiency, and require frequent changing to maintain air volume. Cleaning frequency varies with the application. Units in apartments, hotels, and hospitals usually require more frequent filter service because of lint. Unit motors may require periodic lubrication. Motor failures are not common, but when they occur, the entire fan can be quickly replaced with minimal interruption in the conditioned space. More

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specialized motors with speed control devices may take several days to get a replacement, so large facilities should stock several spares for quick replacement to avoid significant down time for the unit. Chapters 20 and 28 provide more information on fan-coils and unit ventilators, respectively.

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6.

VARIABLE-REFRIGERANT-FLOW (VRF) UNITS

Technological advances in compressor design and control as well as use of electronically controlled expansion valves have allowed design of DX systems in which a single compressor unit provides refrigerant to multiple in-room terminal units. These types of systems are commonly called variable-refrigerant-flow (VRF) systems. Configuration of in-room units is similar to fan-coils, but uses a DX coil instead of water. In these systems, refrigerant piping to terminal units is also called primary piping, and the compressor unit can often be more than 30 m from the in-room terminal units. To allow heating to occur in one space while cooling occurs in an adjacent space, heat pump/heat recovery condensing units must be used and a set of three refrigerant pipes (suction, liquid, and hot gas) must be piped out into the building to the in-room terminal units, or two pipes to terminal units, which then distribute to three pipes. A reversing valve must be provided for each of the in-room terminal units, to allow them to use either liquid refrigerant to cool, or hot gas to heat. The specifics of long-run refrigerant piping are discussed further in Chapter 1 of the 2018 ASHRAE Handbook—Refrigeration, and most manufacturers of these type of systems have very specific installation instructions that must be followed. As with any system that runs refrigerant piping through occupied areas of a building, ASHRAE Standard 15 must be complied with, as well. VRF systems are discussed in more detail in Chapter 18.

7.

CHILLED-BEAM SYSTEMS

Chilled beams are an evolution of chilled ceiling panels. Reports of energy savings over variable-air-volume (VAV) systems, especially in spaces with high concentrations of sensible loads (e.g., laboratories), have been touted in Europe and Australia. Applications such as health care, data centers, and some office areas may be well suited to chilled-beam systems. Two types of chilled beams, passive and active, are in use (Figure 2). Passive chilled beams consist of a chilled-water coil mounted inside a cabinet. Chilled water is piped to the convective coil at between 14 and 15°C. Passive beams use convection currents to cool the space. As air that has been cooled by the beam’s chilledwater coil falls into the space, warmer air is displaced, rises into the coil, and is cooled. Passive beams can provide approximately 1385 kJ/m and, to ensure proper dehumidification and effective ventilation to the spaces, require a separate ventilation system to provide tempered, dehumidified air. Heat can be provided by finned-tube radiation along the space perimeter. Overcooling must be avoided during cooling seasons, to prevent discomfort, condensation, and microbial growth in spaces. Active chilled beams can provide up to approximately 2750 kJ/m. They operate with induction nozzles that entrain room air and mix it with the primary or ventilation air that is ducted to the beam. Chilled water is piped to the coil at between 13 and 15°C. Primary air should be ducted to the beam at 13°C or lower to provide proper dehumidification. The primary air is then mixed with induced room air at a ratio of 1:2. For example, 24 L/s of primary air at 13°C may be mixed with 48 L/s of recirculated room air, and the active beam would distribute 70 L/s at around 18°C. If the low-temperature primary air alone will overcool spaces during any time of the year, there must be provision for reheat. Active beams can have either a two- or four-pipe distribution system. The two-

pipe system may be cooling only or two-pipe changeover. Active beams can be designed to heat and cool the occupied space, but finned-tube radiation is still commonly used to provide heating in a space that is cooled with active beams. Both active and passive beams are designed to operate dry, without condensate. In some models of active beams, a drain pan may be available if the coil is in a vertical configuration. Horizontal coils in passive beams cannot have drain pans, because the area directly below the coil is needed to allow the air in the convection current to circulate. Chilled beams can be used in various applications; however, they are best used in applications with high sensible loads, such as laboratory spaces with high internal heat gains. See manufacturers’ information for beam cooling capacities at various water temperatures and flow rates. Several manufacturers have design guides available on the Internet. The latest generation of chilled beams is multiservice: they can be either passive or active, and combine building operations such as lighting, security sensors, motion detectors, sprinkler systems, smoke detectors, intercoms, and power or fiber-optic distribution with the chilled beam. Proper implementation requires extensive integrated building design.

Types and Location Passive beams are available in sections up to 3 m long and 460 to 610 mm wide. They can be located above the ceiling with perforated panels below it, mounted into the frame of an acoustical tile ceiling, or mounted in the conditioned space. The perforated panels must have a minimum 50% free area and extend beyond either side of the beam for usually half of the unit’s width, so the convection current is not hindered. Also, care must be taken to not locate passive beams too close to window treatments, which can also hinder air movement around the beam. Active beams are available in sections up to 3 m long and 300 to 610 mm wide. They can be mounted into the frame of an acoustical tile ceiling or in the conditioned space.

Ventilation Air Requirements Passive beams require a separate ventilation system to provide tempered and dehumidified air to the space. The ventilation air should be ducted to low-wall diffusers or in an underfloor distribution system so that the ventilation air does not disturb the convection currents in the conditioned space. Ventilation air can be ducted directly into the active beams. If more ventilation air is needed to meet the space requirements, the volume of air can be split by the active beams and high-induction diffusers. Care must be taken in selecting diffuser locations to coordinate well with the convective currents required by the chilled beams.

Selection Chilled beams are selected based on the calculated heat gain for the space less the cooling effect of the primary ventilation air.

Fig. 2 Passive and Active Chilled-Beam Operation (Courtesy of Trox USA, Inc.)

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In-Room Terminal Systems Wiring Chilled beams only require controls wiring. There is no fan or other electrical equipment to be wired.

Condensate Chilled beams are designed to operate dry, with few exceptions. In some active beams with vertical coils, a drain pan may be installed. However, as a rule, a separate ventilation system should be sized to handle the latent cooling load in the space, and the relative humidity should be closely monitored. If the primary ventilation system fails to properly control the space humidity, condensation may form on the beams and their housings. Dripping of this condensate could damage building materials and contents. Chilled-water valves should have drain pans to contain any normal condensate that occurs, when the space is properly dehumidified. In case of loss of humidity control of the space, unpiped condensate drain pans will not be sufficient to avoid overflows and damage.

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limit furniture placement. Electric radiant panels are available from 250 to 750 W in standard single-phase voltages.

Ventilation Air Requirements Radiant panels provide space heating. Ventilation air must be supplied by a central ventilation unit that can provide tempered, humidity-controlled air to the space.

Selection Radiant panels are selected based on the calculated heat loss for the space.

Wiring Electric radiant panels are available in standard single-phase voltages. Panels are often prewired, including the ground wire, with lead wires housed in flexible metal conduit and connector for junction box mounting.

Capacity Control

Capacity Control Capacity of the chilled beam is controlled by a two-way valve on the chilled-water pipe, which is wired to a room thermostat. There is typically one valve per zone (e.g., office, lecture hall). Beams should be piped directly in a reverse/return piping design. Beams are not typically piped in series.

Maintenance Maintenance on chilled beams requires blowing off the coils on a regular basis. Because coils are located throughout occupied spaces, this requires some coordination with occupants and housekeeping personnel to minimize effects on furniture and space contents.

Other Concerns Consistent air movement with natural convection equipment is difficult to achieve. The designer should consult with the manufacturer to determine the proper spacing of chilled beams based on ceiling height, heat generation sources, occupant locations, and movement patterns. Prevention of cold air ponding on the floor, especially when heat sources vary, can be paramount in maintaining comfort. Occupant comfort is affected by more than temperature and humidity. Noise levels and nonstagnant air are also important. Some spaces may not achieve sufficient air movement or background noise to allow occupant comfort through use of passive chilled beams. Additionally, insufficient filtration of the air may occur without sufficient air movement and filtration devices in the space. When introducing pretreated ventilation air to a space, be careful not to interfere with convection currents while still complying with the requirements of ventilation effectiveness and efficiency.

8.

5.9

RADIANT-PANEL HEATING SYSTEMS

Radiant heating panels can use either hot water or electricity. The panels are manufactured in standard 300 by 300 or 300 by 1220 mm panels that can be mounted into the frame of an acoustical tile ceiling or directly to an exposed ceiling or wall. Radiant panels are designed for all types of applications. They are energy efficient, providing a comfortable heat without lowering the moisture content of the room air the way heated air systems may. Occupants in a space heated by radiant heat are comfortable at lower room temperatures, which frequently reduces operational costs. See Chapter 6 for more information on these systems. Electric panels may be regulated by local energy codes.

Types and Location Radiant panels are typically mounted on the ceiling near perimeter walls in a metal frame. Unlike finned-tube radiation, they do not

Panel capacity is usually controlled by coil water flow, or in the case of electric heat, capacity steps. Most radiant panels are controlled with a wall-mounted thermostat located in the space.

Maintenance Because they have no moving parts, radiant panels require little maintenance. Water flow control valves require periodic verification that they are operating correctly.

9.

RADIANT-FLOOR HEATING SYSTEMS

Radiant-floor heat is best applied under a finished floor that is typically cold to the touch. Radiant-floor heat systems in the past used flexible copper pipe heating loops encased in concrete. Unfortunately, the soldered joints could fail or the concrete’s expansion and contraction or chemical composition could corrode the pipes, causing them to leak. However, new technologies include flexible plastic tubing (often referred to as PEX, or cross-linked polyethylene) to replace the old flexible copper tubing. PEX tubing is also available with an oxygen diffusion barrier, because oxygen entrained in the radiant heat tubing can cause corrosion on the ferrous connectors between the tubing and the manifold system. PEX tubing is also available in longer lengths than the flexible copper, which minimizes buried joints. The tubing is run back to a manifold system, which includes valves to balance and shut down the system and a small circulator pump. Multiple zones can be terminated at the same manifold. Historically, radiant-floor heat was commonly designed for residential applications, when ventilation was provided by operable windows, and cooling was not mandatory. Like most other in-room terminal systems, the required ventilation air must be supplied by a central unit that can provide tempered and humidity-controlled air that allows comfort conditions in the space to be met. Common applications of radiant-floor heat systems include large open buildings, such as airplane hangars, where providing heat at the floor is more cost-effective than heating the entire volume of air in the space. Radiant-floor heat is becoming more common for preschools, elementary schools, exercise spaces, and other locations where children or adults sit or lie on floors. Water in the radiantfloor loop is often around 32°C, depending on the floor finish. This is a lower temperature than forced hot-air systems, and reduces the energy required to heat the building. Buildings that have high ceilings, large windows, or high infiltration rates or that require high air change rates may save energy by using radiant-floor heat. Radiant-floor systems are commonly zoned by room. Each room may require multiple pipe circuits, depending on the room’s area and the manifold’s location. Maximum tubing lengths are determined based on tubing diameter and desired heat output. If tubing is

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installed in a slab on or below grade or over an unconditioned space, insulation should be incorporated to minimize heat losses. If a radiant-floor heat system is installed in a slab over a conditioned space, the radiant effect on that space must be considered as well. See Chapter 6 for more information on radiant panel heating systems.

Types and Location

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Radiant-floor heating is located in the flooring or just below it with a heat-reflecting wrap. If the final floor finish is hardwood flooring, the radiant-floor piping can be installed in plywood tracks with a heat reflector, below the finished floor. Ensure that, as the final flooring is nailed down, the flexible tubing is not punctured. If the final floor is a ceramic tile or other surface requiring a poured concrete, the radiant floor can be laid out in the concrete, if hotwater systems are being used. Electric underfloor heating is also available, using a prefabricated mat that is applied on top of the subfloor and under with ceramic tile. Radiant-floor heating systems can also be mounted below the floor joists, with a heat reflector below the piping. This method is often used in renovations where removing existing flooring is not feasible.

Ventilation Air Requirements Ventilation air must be supplied by a central ventilation unit that can provide tempered, humidity-controlled air to the space.

Selection Spacing between rows of tubing that make up the radiant floor varies, depending on the heat loss of the space. Usually, the entire heat loss of the space is calculated and the tubing spaced accordingly. Another method is to place tubing closer together near the room’s perimeter and increase the spacing in the interior. This method is more time consuming, and the difference is only noticeable in large spaces. Supply water temperature in the tubing is determined based on the flooring materials’ resistance to heat flow; the temperature is higher for carpeting and a pad than for ceramic tile. The tubing is also available in different nominal diameters, the most common being 10 or 13 mm.

Wiring Circulator pumps at the manifolds require power. Additional controls for zone control valves may be selected for either line voltage or low voltage.

Capacity Control Because radiant floors heat the mass of the floor, these systems are typically slow to respond to environmental changes. The circulator pumps start on a call for heat from a thermostat; however, rapid solar gains to a space with many windows could cause the space to overheat. Smart systems have been used to anticipate the needs of the space and overcome the thermal flywheel effects inherent to these types of systems. Smart systems anticipate the need for heating based on outdoor conditions and the conditions experienced in the recent past. They anticipate when set-point temperatures are about to be achieved and reduce heat generation, to minimize overshooting. They also learn patterns of operation that help overcome reasonably consistent daily solar gains.

10.

Induction units are very similar to active chilled beams, although they have mostly been replaced by VAV systems. Only the specific differences of higher-pressure air induction units will be discussed here. Primary air is supplied to an induction unit’s plenum at medium to high pressure. The acoustically treated plenum attenuates part of the noise generated in the unit and duct. High-velocity induction unit nozzles typically generate significant high-frequency noise. A balancing damper adjusts the primary-air quantity within limits. Medium- to high-velocity air flows through the induction nozzles and induces secondary air from the room through the secondary coil. This secondary air is either heated or cooled at the coil, depending on the season, room requirement, or both. Ordinarily, the room coil does no latent cooling, but a drain pan without a piped drain collects condensed moisture from temporary latent loads such as at start-up. This condensed moisture then reevaporates when the temporary latent loads are no longer present. Primary and secondary (induced) air is mixed and discharged to the room. Secondary airflow can cause induction-unit coils to become dirty enough to affect performance. Lint screens are sometimes used to protect these terminals, but require frequent in-room maintenance and reduce unit thermal performance. Induction units are installed in custom enclosures, or in standard cabinets provided by the manufacturer. These enclosures must allow proper flow of secondary air and discharge of mixed air without imposing excessive pressure loss. They must also allow easy servicing. Although induction units are usually installed under a window at a perimeter wall, units designed for overhead installation are also available. During the heating season, the floor-mounted induction unit can function as a convector during off hours, with hot water to the coil and without a primary-air supply. Numerous induction unit configurations are available, including units with low overall height or with larger secondary-coil face areas to suit particular space or load needs. Induction units may be noisier than fan-coil units, especially in frequencies that interfere with speech. On the other hand, white noise from the induction unit enhances acoustical privacy by masking speech from adjacent spaces. In-room terminals operate dry, with an anticipated life of 15 to 25 years. Individual induction units do not contain fans, motors, or compressors. Routine service is generally limited to temperature controls, cleaning lint screens, and infrequently cleaning the induction nozzles. In existing induction systems, conserving energy by raising the chilled-water temperature on central air-handling cooling coils can damage the terminal cooling coil, causing it to be used constantly as a dehumidifier. Unlike fan-coil units, the induction unit is not designed or constructed to handle condensation. Therefore, it is critical that an induction terminal operates dry. Induction units are rarely used in new construction. They consume more energy because of the increased power needed to deliver primary air against the pressure drop in the terminal units, and they generate high-frequency noise from the induction nozzles. In addition, the initial cost for a four-pipe induction system is greater than for most all-air systems. However, induction units are still used for direct replacement renovation; because the architecture was originally designed to accommodate the induction unit, other systems may not be easily installed.

11. Maintenance The circulator pumps, valves, controls, and manifolds are the only components requiring maintenance. Once the tubing is laid out, it should be pressure-tested for leaks; once covered, it is extremely difficult and/or expensive to access.

INDUCTION UNIT SYSTEMS

SUPPLEMENTAL HEATING UNITS

In-room supplemental heating units come in all sizes. Units can have either electric (if allowed by local energy code) or hot-water heat, and sometimes steam; they can be surface-mounted, semirecessed, or recessed in the walls on the floor or horizontally along the ceiling. Baseboard radiation is usually located at the source of the

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Fig. 3 Primary-Air System heat loss, such as under a window or along a perimeter wall, and is usually rated for between 1400 and 2075 kJ/m at 77°C. Other supplemental heating units include unit heaters, wall heaters, and cabinet heaters. All supplemental heating units can be supplied with an integral or separate wall-mounted thermostat. If the heater is located low in the space, an integral thermostat is sufficient most of the time; however, if the unit is mounted horizontally near the ceiling, the thermostat should be wired so that it is located in the space, for accurate space temperature readings. In addition, units may have a summer fan option, which allows the fan to turn on for ventilation. Water flow to space supplemental heaters should be cut off anytime the water temperature to the coil is below 27°C, to avoid condensation and consequent damage or mold growth.

12.

PRIMARY-AIR SYSTEMS

Figure 3 illustrates a primary-air system for in-room terminal systems. The components are described in Chapter 4. Some primary-air systems operate with 100% outdoor air at all times. In climates where moisture content is lower outdoors than indoors, systems using return air may benefit from a provision for operating with 100% outdoor air (economizer cycle) to reduce operating cost during some seasons. In some systems, when the quantity of primary air supplied exceeds the ventilation or exhaust required, excess air is recirculated by a return system common with the interior system. A good-quality filter (MERV = 8 to 11) is desirable in the central air treatment apparatus. If it is necessary to maintain a given humidity level in cold weather, a humidifier can usually be installed. Where humidification is needed, steam humidifiers have been used successfully. Water-spray humidifiers must be operated in conjunction with (1) the preheat coil elevating the temperature of the incoming air or (2) heaters in the spray water circuit. Water-spray humidifiers should be used with caution, however, because of the possible growth of undesirable organisms in untreated water. See Chapter 22 for additional information on humidifiers. The primary-air quantity is fixed, and the leaving primary-air temperature varies inversely with the outdoor temperature to provide the necessary amount of heating or cooling and humidity control. Proper leaving-air temperature must be determined based on climatic conditions and the influence of the ventilation air quantity on the final space temperature and relative humidity. In cooling season, many climates require primary air to be cooled to a point low enough to dehumidify the total system (cooling coil leaving temperature about 10°C or less, and almost completely saturated) and then reheated to be provided at an appropriate temperature and humidity level. During winter, primary air is often preheated and supplied at approximately 10°C to provide cooling. All room terminals in a

5.11 given primary-air preheated zone must be selected to operate satisfactorily with the common primary-air temperature. The supply fan should be selected at a point near maximum efficiency to reduce power consumption, supply air heating, and noise. Sound absorbers may be required at the fan discharge to attenuate fan noise. Reheat coils are required in a two-pipe system. Reheat may not be required for primary-air supply of four-pipe systems. Formerly, many primary-air distribution systems for induction units were designed with 2.0 to 2.5 kPa (gage) static pressure. With energy use restrictions, this is no longer economical. Good duct design and elimination of unnecessary restrictions (e.g., sound traps) can result in primary systems that operate at 1.1 to 1.5 kPa (gage) or even lower. Low-initial-cost, smaller ductwork has to be weighed against operating costs during life-cycle evaluation, to provide a system that best meets the owner’s long-term goals. Primary-air distribution systems serving fan-coil systems can operate at pressures 0.25 to 0.37 kPa or lower. Induction units and active chilled beams require careful selection of the primary-air cooling coil and the induction unit nozzles to achieve an overall medium-pressure primary-air system. Primary-air system distribution that is independently supplied to spaces for use with other in-room terminal systems may be lowvelocity or a combination of low- and medium-velocity systems. See Chapter 21 in the 2017 ASHRAE Handbook—Fundamentals for a discussion of duct design. Variations in pressure between the first and last terminals should be minimized to limit the pressure drop required across balancing dampers. Room sound characteristics vary depending on unit selection, air system design, and equipment manufacturer. Select units by considering the unit manufacturer’s sound power ratings, desired maximum room noise level, and the room’s acoustical characteristics. Limits of sound power level can then be specified to obtain acceptable acoustical performance. See Chapter 8 in the 2017 ASHRAE Handbook—Fundamentals.

13.

PERFORMANCE UNDER VARYING LOAD

Under peak load conditions, the psychrometrics of induction units, chilled beams, unit ventilators, and fan-coil unit systems are essentially identical for two- and four-pipe systems. Primary air mixes with secondary air conditioned by the room coil in an induction unit before delivery to a room. Mixing also occurs in a fan-coil unit with a direct-connected primary-air supply. If primary air is supplied to the space separately, as in fan-coil systems with independent primary-air supplies, the same effect occurs in the space. During cooling, the primary-air system provides part of the sensible capacity and all of the dehumidification. The rest of the sensible capacity is accomplished by the cooling effect of secondary water circulating through the in-room terminal unit cooling coils. In winter, when primary air is provided directly into the in-room terminal unit, it can be provided at a low temperature and humidified if necessary. This may allow cooling of internal spaces solely by the primary air, if quantities are sufficient to meet the full cooling requirements. Room heating where needed is then supplied by the secondary-water system circulating through the in-room terminal unit coils. If the economizer cycle is used, cooling energy can be reduced when the moisture level of the outdoor air allows. For systems where primary air does not enter at the terminal unit, the primary air should enter the room at a temperature approximately neutral with the desired room condition and a relative humidity level that enhances room conditions. In buildings where cooling is required during heating conditions, care must be taken to avoid drafts caused by providing primary air at too low a temperature. If interior spaces require cooling, a four-pipe system should be considered to allow cooling of those areas while heating exterior perimeter zones. During fall and spring months, the primary-air

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2020 ASHRAE Handbook—HVAC Systems and Equipment (SI) The outdoor temperature at which the heat gain to every space can be satisfied by the combination of cold primary air and the transmission loss is called the changeover temperature. Below this temperature, cooling is not required. The following empirical equation approximates the changeover temperature at sea level. It should be fine-tuned after system installation (Carrier 1965): q is + q es – 1.2Q p  t r – t p  tco = tr – ---------------------------------------------------------- q td

(1)

where

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Fig. 4

Solar Radiation Variations with Seasons

temperature may be reduced slightly to provide the limited amount of cooling needed for east and west exposures with low internal heat gains, because solar heat gain is typically reduced during these seasons. In the northern hemisphere, the north exposure is not a significant factor because solar gain is very low; for south, southeast, and southwest exposures, the peak solar heat gain occurs in winter, coincident with a lower outdoor temperature (Figure 4). Reducing the primary-air temperature may not be a viable alternative where primary air is supplied directly to the space, because this air could overcool spaces where solar heat gain or internal heat gain is low. In buildings with large areas of glass, heat transmitted from indoors to the outdoors, coupled with the normal supply of cool primary air, does not balance internal heat and solar gains until an outdoor temperature well below freezing is reached. Double-glazed windows with clear or heat-absorbing glass aggravate this condition, because this type of glass increases the heating effect of the radiant energy that enters during the winter by reducing reverse transmission. Therefore, cooling must be available at lower outdoor temperatures. In buildings with very high internal heat gains from lighting or equipment, the need for cooling from the room coil, as well as from the primary air, can extend well into winter. In any case, the changeover temperature at which the cooling capacity of the secondary-water system is no longer required for a given space is an important calculation. All these factors should be considered when determining the proper changeover temperature.

14.

CHANGEOVER TEMPERATURE

For all systems using a primary-air system for outdoor air, there is an outdoor temperature (balance temperature) at which secondary cooling is no longer required. The system can cool by using outdoor air at lower temperatures. For all-air systems operating with up to 100% outdoor air, mechanical cooling is seldom required at outdoor temperatures below 12°C, unless the dew point of the outdoor air equals or exceeds the desired indoor dew point. An important characteristic of in-room terminal unit systems, however, is that secondary-water cooling may still be needed, even when the outdoor temperature is considerably less than 10°C. This cooling may be provided by the mechanical refrigeration unit or by a thermal economizer cycle. Full-flow circulation of primary air through the cooling coil below 10°C often provides all the necessary cooling while preventing coil freeze-up and reducing the preheat requirement. Alternatively, secondary-water-to-condenser-water heat exchangers function well.

tco = temperature of changeover point, °C tr = room temperature at time of changeover, normally 22°C tp = primary-air temperature at unit after system is changed over, normally 13°C Qp = primary-air quantity, L/s qis = internal sensible heat gain, W qes = external sensible heat gain, W qtd = heat transmission per degree of temperature difference between room and outdoor air

In two-pipe changeover systems, the entire system is usually changed from winter to summer operation at the same time, so the room with the lowest changeover point should be identified. In northern latitudes, this room usually has a south, southeast, or southwest exposure because the solar heat gains on these exposures reach their maximum during winter. If the calculated changeover temperature is below approximately 9°C, an economizer cycle should operate to allow the refrigeration plant to shut down. Although factors controlling the changeover temperature of inroom terminal systems are understood by the design engineer, the basic principles may not be readily apparent to system operators. Therefore, it is important that the concept and calculated changeover temperature are clearly explained in operating instructions given before operating the system. Some increase from the calculated changeover temperature is normal in actual operation. Also, a range or band of changeover temperatures, rather than a single value, is necessary to preclude frequent change in seasonal cycles and to grant some flexibility in operation. The difficulties associated with operator understanding and the need to perform changeover several times a day in many areas have severely limited the acceptability of the two-pipe changeover system.

15.

TWO-PIPE SYSTEMS WITH CENTRAL VENTILATION

Two-pipe systems for in-room terminal systems derive their name from the water-distribution circuit, which consists of one supply and one return pipe. Each unit or conditioned space is supplied with secondary water from this distribution system and with conditioned primary air from a central apparatus. The system design and control of primary-air and secondary-water temperatures must be such that all rooms on the same system (or zone, if applicable) can be satisfied during both heating and cooling seasons. The heating or cooling capacity of any unit at a particular time is the sum of its primary-air output plus its secondary-water output. The secondary-water coil (cooling-heating) in each space is controlled by a space thermostat and can vary from 0 to 100% of coil capacity, as required to maintain space temperature. The secondary water is cold in summer and intermediate seasons and warm in winter. All rooms on the same secondary-water zone must operate satisfactorily with the same water temperature. Figure 5 shows the capacity ranges available from a typical two-pipe system. On a hot summer day, loads from about 25 to 100% of the design space cooling capacity can be satisfied. On a 10°C intermediate-season day, the unit can satisfy a heating requirement

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Fig. 5 Capacity Ranges of In-Room Terminal Operating on Two-Pipe System by closing off the secondary-water coil and using only the output of warm primary air. A lesser heating or net cooling requirement is satisfied by the cold secondary-water coil output, which offsets the warm primary air to obtain cooling. In winter, the unit can provide a small amount of cooling by closing the secondary coil and using only the cold primary air. Smaller cooling loads and all heating requirements are satisfied by using warm secondary water.

Critical Design Elements The most critical design elements of a two-pipe system are the calculation of primary-air quantities and the final adjustment of the primary-air temperature reset schedule. All rooms require a minimum amount of heat and latent capacity from the primary-air supply during the intermediate season. Using the ratio of primary air to transmission per degree (A/T ratio) to maintain a constant relationship between the primary-air quantity and the heating requirements of each space fulfills this need. The A/T ratio determines the primary-air temperature and changeover point, and is fundamental to proper design and operation of a two-pipe system. Transmission per Degree. The relative heating requirement of every space is determined by calculating the transmission heat flow per degree temperature difference between the space temperature and the outdoor temperature (assuming steady-state heat transfer). This is the sum of the (1) glass heat transfer coefficient times the glass areas, (2) wall heat transfer coefficient times the wall area, and (3) roof heat transfer coefficient times the roof area. Air-to-Transmission (A/T) Ratio. The A/T ratio is the ratio of the primary airflow to a given space divided by the transmission per degree of that space: A/T ratio = Primary air/Transmission per degree. Spaces on a common primary-air zone must have approximately the same A/T ratios. The design base A/T ratio establishes the primary-air reheat schedule during intermediate seasons. Spaces with A/T ratios higher than the design base A/T ratio tend to be overcooled during light cooling loads at an outdoor temperature in the 21 to 32°C range, whereas spaces with an A/T ratio lower than design lack sufficient heat during the 5 to 15°C outdoor temperature range when primary air is warm for heating and secondary water is cold for cooling. The minimum primary-air quantity that satisfies the requirements for ventilation, dehumidification, and both summer and winter cooling is used to calculate the minimum A/T ratio for each space. If the system operates with primary-air heating during cold

Fig. 6 Primary-Air Temperature Versus Outdoor Air Temperature weather, the heating capacity can also be the primary-air quantity determinant for two-pipe systems. The design base A/T ratio is the highest A/T ratio obtained, and the primary airflow to each space is increased as required to obtain a uniform A/T ratio in all spaces. An alternative approach is to locate the space with the highest A/T ratio requirement by inspection, establish the design base A/T ratio, and obtain the primary airflow for all other spaces by multiplying this A/T ratio by the transmission per degree of all other spaces. For each A/T ratio, there is a specific relationship between outdoor air temperature and temperature of the primary air that maintains the room at 22°C or more during conditions of minimum room cooling load. Figure 6 illustrates this variation based on an assumed minimum room load equivalent to 5 K times the transmission per degree. A primary-air temperature over 50°C at the unit is seldom used. The reheat schedule should be adjusted for hospital rooms or other applications where a higher minimum room temperature is desired, or where a space has no minimum cooling load. Deviation from the A/T ratio is sometimes permissible. A minimum A/T ratio equal to 0.7 of the maximum A/T is suitable, if the building is of massive construction with considerable heat storage effect (Carrier 1965). The heating performance when using warm primary air becomes less satisfactory than that for systems with a uniform A/T ratio. Therefore, systems designed for A/T ratio deviation should be suitable for changeover to warm secondary water for heating whenever the outdoor temperature falls below 5°C. A/T ratios should be more closely maintained on buildings with large glass areas or with curtain wall construction, or on systems with low changeover temperature.

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Fig. 7 Psychrometric Chart, Two-Pipe System, Off-Season Cooling

Fig. 8 Typical Changeover System Temperature Variation

Changeover Temperature Considerations Transition from summer operations to intermediate-season operation is done by gradually raising the primary-air temperature as the outdoor temperature falls, to keep rooms with small cooling loads from becoming too cold. The secondary water remains cold during both summer and intermediate seasons. Figure 7 illustrates the psychrometrics of summer operation near the changeover temperature. As the outdoor temperature drops further, the changeover temperature is reached. The secondary-water system can then be changed over to provide hot water for heating. If the primary airflow is increased to some spaces to elevate the changeover temperature, the A/T ratio for the reheat zone is affected. Adjustments in primary-air quantities to other spaces on that zone will probably be necessary to establish a reasonably uniform ratio. System changeover can take several hours and usually temporarily upsets room temperatures. Good design, therefore, includes provision for operating the system with either hot or cold secondary water over a range of 8 to 11 K below the changeover point. This range makes it possible to operate with warm air and cold secondary water when the outdoor temperature rises above the daytime changeover temperature. Changeover to hot water is limited to times of extreme or protracted cold weather. Optional hot- or cold-water operation below the changeover point is provided by increasing the primary-air reheater capacity to provide adequate heat at a colder outdoor temperature. Figure 8 shows temperature variation for a system operating with changeover, indicating the relative temperature of the primary air and secondary water throughout the year and the changeover temperature range. The solid arrows show the temperature variation when changing over from the summer to the winter cycle. The open arrows show the variation when going from the winter to the summer cycle.

Nonchangeover Design Consider using nonchangeover systems to simplify operation for buildings with mild winter climates, or for south exposure zones of buildings with a large winter solar load. A nonchangeover system operates on an intermediate-season cycle throughout the heating season, with cold secondary water to the terminal unit coils and with

Fig. 9

Typical Nonchangeover System Variations

warm primary air satisfying all the heating requirements. Typical temperature variation is shown in Figure 9. Spaces may be heated during unoccupied hours by operating the primary-air system with 100% return air. This feature is necessary because nonchangeover design does not usually include the ability to heat the secondary water. In addition, cold secondary water must be available throughout the winter. Primary-air duct insulation and observance of close A/T ratios for all units are essential for proper heating during cold weather.

Zoning A two-pipe system can provide good temperature control most of the time, on all exposures during the heating and cooling seasons. Comfort and operating cost can be improved by zoning in the following ways: • Primary air to allow different A/T ratios on different exposures • Primary air to allow solar compensation of primary-air temperature

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• Both air and water to allow a different changeover temperature for different exposures All spaces on the same primary-air zone must have the same A/T ratio. The minimum A/T ratios often are different for spaces on different solar exposures, thus requiring the primary-air quantities on some exposures to be increased if they are placed on a common zone with other exposures. The primary-air quantity to units serving spaces with less solar exposure can usually be reduced by using separate primary-air zones with different A/T ratios and reheat schedules. Primary-air quantity should never be reduced below minimum ventilation requirements. The peak cooling load for the south exposure occurs during fall or winter when outdoor temperatures are lower. If shading patterns from adjacent buildings or obstructions are not present, primaryair zoning by solar exposure can reduce air quantities and unit coil sizes on the south. Units can be selected for peak capacity with cold primary air instead of reheated primary air. Primary-air zoning and solar compensators save operating cost on all solar exposures by reducing primary-air reheat and secondary-water refrigeration penalty. Separate air and water zoning may save operating cost by allowing spaces with less solar exposure to operate on the winter cycle with warm secondary water at outdoor temperatures as high as 16°C during the heating season. Systems with a common secondarywater zone must operate with cold secondary water to cool heavier solar exposures. Primary airflow can be lower because of separate A/T ratios, resulting in reheat and refrigeration cost savings.

Room Control When room temperature rises, the thermostat must increase the output of the cold secondary coil (in summer) or decrease the output of the warm secondary coil (in winter). Changeover from cold to hot water in the unit coils requires changing the action of the room temperature control system. Room control for nonchangeover systems does not require the changeover action, unless it is required to provide gravity heating during shutdown.

Evaluation Characteristics of two-pipe in-room terminal unit systems include the following: • Usually less expensive to install than four-pipe systems • Less capable of handling widely varying loads or providing a widely varying choice of room temperatures than four-pipe systems • Present operational and control changeover problems, increasing the need for competent operating personnel • More costly to operate than four-pipe systems

Electric Heat for Two-Pipe Systems Electric heat can be supplied with a two-pipe in-room terminal unit system by a central electric boiler and terminal coils, or by individual electric-resistance heating coils in the terminal units. One method uses small electric-resistance terminal heaters for intermediate-season heating and a two-pipe changeover chilledwater/hot-water system. The electric terminal heater heats when outdoor temperatures are above 5°C, so cooling can be kept available with chilled water in the chilled-water/hot-water system. System or zone reheating of primary air is greatly reduced or eliminated entirely. When the outdoor temperature falls below this point, the chilled-water/hot-water system is switched to hot water, providing greater heating capacity. Changeover is limited to a few times per season, and simultaneous heating/cooling capacity is available, except in extremely cold weather, when little, if any, cooling is needed. If electric-resistance terminal heaters are used,

5.15 they should be prevented from operating whenever the secondarywater system is operated with hot water. Another method is to size electric resistance terminal heaters for the peak winter heating load and operate the chilled-water system as a nonchangeover cooling-only system. This avoids the operating problem of chilled-water/hot-water system changeover. In fact, this method functions like a four-pipe system, and, in areas where the electric utility establishes a summer demand charge and has a low unit energy cost for high winter consumption, it may have a lower life-cycle cost than hydronic heating with fossil fuel. A variation, especially appropriate for well-insulated office buildings with induction units where cooling is needed in perimeter offices for almost all occupied hours because of internal heat gain, is to use electric heaters in the terminal unit during occupied hours and to provide heating during unoccupied hours by raising primary-air temperature on an outdoor reset schedule.

16.

FOUR-PIPE SYSTEMS

Four-pipe systems have a chilled-water supply, chilled-water return, hot-water supply, and hot-water return. The terminal unit usually has two independent secondary-water coils: one served by hot water, the other by cold water. During peak cooling and heating, the four-pipe system performs in a manner similar to the two-pipe system, with essentially the same operating characteristics. Between seasons, any unit can be operated at any level from maximum cooling to maximum heating, if both cold and warm water are being circulated, or between these extremes without regard to other units’ operation. In-room terminal units are selected by their peak capacity. The A/T ratio does not apply to four-pipe systems. There is no need to increase primary-air quantities on units with low solar exposure beyond the amount needed for ventilation and to satisfy cooling loads. The available net cooling is not reduced by heating the primary air. The changeover point is still important, though, because cooling spaces on the sunny side of the building may still require secondary-water cooling to supplement the primary air at low outdoor temperatures.

Zoning Zoning primary-air or secondary-water systems is not required with four-pipe systems. All terminal units can heat or cool at all times, as long as both hot and cold secondary pumps are operated and sources of heating and cooling are available.

Room Control The four-pipe terminal usually has two completely separated secondary-water coils: one receiving hot water and the other receiving cold water. The coils are operated in sequence by the same thermostat; they are never operated simultaneously. The unit receives either hot or cold water in varying amounts, or else no flow is present, as shown in Figure 10A. Adjustable, dead-band thermostats further reduce operating cost. Figure 10B illustrates another unit and control configuration. A single secondary-water coil at the unit and three-way valves located at the inlet and outlet admit water from either the hot- or cold-water supply, as required, and divert it to the appropriate return pipe. This arrangement requires a special three-way modulating valve, originally developed for one form of the three-pipe system. It controls the hot or cold water selectively and proportionally, but does not mix the streams. The valve at the coil outlet is a two-position valve open to either the hot or cold water return, as required. Overall, the two-coil arrangement provides a superior four-pipe system. Operation of the induction and fan-coil unit controls is the same year-round.

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Evaluation Compared to the two-pipe system, the four-pipe air-and-water system has the following characteristics: • More flexible and adaptable to widely differing loads, responding quickly to load changes • Simpler to operate • Operates without the summer-winter changeover and primary-air reheat schedule • Efficiency is greater and operating cost is lower, though initial cost is generally higher • Can be designed with no interconnection of hot- and cold-water secondary circuits, and the secondary system can be completely independent of the primary-water piping

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17.

AUTOMATIC CONTROLS AND BUILDING MANAGEMENT SYSTEMS

Chapter 48 of the 2019 ASHRAE Handbook—HVAC Applications and Chapter 7 of the 2017 ASHRAE Handbook—Fundamentals discuss automatic controls. The information and concepts discussed there apply to in-room terminal system equipment and systems, as well. The designer should discuss the complexity, technical expertise, and local resources with the facility’s owner/operator before specifying a overly complex or sophisticated control system, because many facilities using in-room terminal units have limited maintenance staff.

18. MAINTENANCE MANAGEMENT SYSTEMS AND BUILDING SYSTEM COMMISSIONING Chapter 1 discusses both of these topics. In-room terminal systems, like every other system, benefit from consideration and implementation of these practices.

Fig. 10 Fan-Coil Unit Control

ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal.ashrae .org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org/bookstore. AHRI. 2008. Performance rating of room fan-coils. ANSI/AHRI Standard 440-2008. Air-Conditioning, Heating, and Refrigeration Institute, Arlington, VA. AHRI. 1998. Unit ventilators. Standard 840-1998. Air-Conditioning, Heating, and Refrigeration Institute, Arlington, VA. AIA. 2001. Guidelines for design and construction of hospital and health care facilities. American Institute of Architects, Washington, D.C. ASHRAE. 2013. Safety standards for refrigeration systems. ANSI/ ASHRAE Standard 15-2013. ASHRAE. 2013. Ventilation for acceptable indoor air quality. ANSI/ ASHRAE Standard 62.1-2013. ASHRAE. 2002. Method of testing for rating fan coil conditioners. ANSI/ ASHRAE Standard 79-2002. Carrier Air Conditioning Company. 1965. Handbook of air conditioning system design. McGraw-Hill, New York.

BIBLIOGRAPHY ASHRAE. 2016. Methods of testing for rating room air conditioners and packaged terminal air conditioners. ANSI/ASHRAE Standard 16-2016. ASHRAE. 2017. Thermal environmental conditions for human occupancy. ANSI/ASHRAE Standard 55-2017. ASHRAE. 2014. Methods of testing for rating room air conditioners and packaged terminal air conditioner heating capacity. ANSI/ASHRAE Standard 58-1986 (RA 2014). ASHRAE. 2016. Ventilation for acceptable indoor air quality in low-rise residential buildings. ANSI/ASHRAE Standard 62.2-2016. ASHRAE. 2011. Method of testing for rating the performance of air outlets and air inlets. ANSI/ASHRAE Standard 70-2006 (RA 2011). ASHRAE. 2016. Energy standard for buildings except low-rise residential buildings. ANSI/ASHRAE Standard 90.1-2016. ASHRAE. 2018. Energy standard of low-rise residential buildings. ANSI/ ASHRAE Standard 90.2-2018. ASHRAE. 2018. Energy efficiency in existing buildings. ANSI/ ASHRAE Standard 100-2018. ASHRAE. 2013. Methods of testing for room air diffusion. ANSI/ASHRAE Standard 113-2013. ASHRAE. 2012. Methods of testing for rating computer and data processing room unitary air conditioners. ANSI/ASHRAE Standard 127-2012. ASHRAE. 2016. Laboratory methods of testing air terminal units. ANSI/ ASHRAE Standard 130-2016. ASHRAE. 2016. BACnet®: A data communication protocol for building automation and control networks. ANSI/ASHRAE Standard 135-2016. ASHRAE. 2013. Methods of testing for radiant ceiling panels for sensible heating and cooling. ANSI/ASHRAE Standard 138-2013 (RA 2016). ASHRAE. 2013. Climate data for building design standards. ANSI/ ASHRAE Standard 169-2013. ASHRAE. 2018. Standard practice for inspection and maintenance of commercial building HVAC systems. ANSI/ASHRAE/ACCA Standard 1802018. ASHRAE. 2017. Standard for the design of high-performance green buildings. ANSI/ASHRAE/USGCB/IES Standard 189.1-2017. ASHRAE. 2013. Method of test for rating air terminal unit controls. ANSI/ ASHRAE Standard 195-2013. ASHRAE. 2018. Commissioning process for buildings and systems. ANSI/ ASHRAE/IES Standard 202-2018. ASHRAE. 2019. The commissioning process. Guideline 0-2019. ASHRAE. 2009. Guideline for the risk management of public health and safety in buildings. Guideline 29-2009.

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Related Commercial Resources CHAPTER 6

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RADIANT HEATING AND COOLING PRINCIPLES OF RADIANT SYSTEMS ....................................... 1 Heat Transfer ................................................................................ 1 Factors Affecting Heat Transfer ................................................... 6 Panel Design ................................................................................. 9 General Design Considerations.................................................. 11 Hybrid Systems ........................................................................... 14

RADIANT HEATING AND COOLING SYSTEMS ...................... 14 Hydronic Ceiling Panels ............................................................. 14 Embedded Systems with Tubing in Ceilings, Walls, or Floors ... 16 Electrically Heated Radiant Systems .......................................... 18 DESIGN PROCEDURE .............................................................. 22 Controls ....................................................................................... 24

P

radiation, although there is little, if any, change in the ambient air temperature. In buildings, thermal radiation occurs mostly in the infrared (longwave) region because the temperature of the bodies in a building environment is low. An example is the radiation heat exchange between human bodies and a cold window on an exterior wall. Another major type of radiation is shortwave radiation; the source is mainly the sun, although lighting fixtures also emit energy in the shortwave region. In this chapter, discussions about radiation thermal properties and their effects are limited to the longwave region, unless specifically indicated otherwise. ASHRAE research project RP-876 (Lindstrom et al. 1998) concluded that surface roughness and texture have insignificant effects on thermal convection and thermal radiation, respectively. Surface longwave emittance (the ratio of the radiant heat flux emitted by a body to that emitted by a blackbody under the same conditions) for typical indoor surfaces, such as carpets, vinyl texture paint, and plastic, remained between 0.9 and 1.0 for panel surface temperatures of 30 to 55°C. For shortwave (solar) absorptance, other factors, such as color and surface roughness, play a significant role. Thermal comfort, as defined in ASHRAE Standard 55-2013, is “that condition of mind which expresses satisfaction with the thermal environment.” Radiant heating and cooling systems can be used to provide unique approaches to dealing with several factors affecting human thermal comfort. Maintaining correct conditions for human thermal comfort by thermal radiation is possible for even the most severe climatic conditions (Buckley 1989). Panel heating and cooling systems modulate the thermal environment by directly controlling surface temperatures as well as indoor air temperature in an occupied space. With a properly designed system, occupants should not be aware that the environment is being heated or cooled. To provide an acceptable thermal environment to the occupants, the requirements for general and local thermal comfort must be taken into account. Chapter 9 of the 2017 ASHRAE Handbook—Fundamentals and ASHRAE Standard 55 have more information on these requirements. Both the air temperature and the mean radiant temperature (MRT) of a space should be taken into account when assessing occupant thermal comfort. The combined influence of these two temperatures is expressed as the operative temperature. Mean radiant temperature has a strong influence on human thermal comfort. The magnitude of the effect is slightly greater than that of air temperature at the low air velocities typical to most indoor spaces. When the temperature of surfaces comprising the building (particularly outdoor exposed walls with extensive fenestration) deviates excessively from the air temperature, convective systems sometimes have difficulty counteracting the discomfort caused by cold or hot surfaces. This is because the air temperature used for controlling the HVAC system is significantly different from the MRT, and both affect thermal comfort. Most building materials have relatively high surface emittance and, therefore, absorb and reradiate heat from active panels. This yields an MRT that is close to the air temperature under most condi-

ANEL heating and cooling systems use temperature-controlled indoor surfaces on the floor, walls, or ceiling; temperature is maintained by circulating water, air, or electric current through a circuit embedded in or attached to the panel. A temperature-controlled surface is called a radiant panel if 50% or more of the design heat transfer on the temperature-controlled surface takes place by thermal radiation. Panel systems are characterized by controlled surface temperatures below 150°C. Panel systems may be combined either with a central forced-air system of one-zone, constant-temperature, constant-volume design, or with dual-duct, reheat, multizone or variable-volume systems, decentralized convective systems, or inspace fan-coil units. In decoupled systems, the air system provides ventilation air and meets dehumidification needs. In hybrid (or load-sharing) systems, the air system may provide significant additional capacity (e.g., where cooling loads exceed the capabilities of a radiant slab system). This chapter covers temperature-controlled surfaces that are the primary source of sensible heating and cooling in the conditioned space. For snow-melting and freeze-protection applications, see Chapter 52 of the 2019 ASHRAE Handbook—HVAC Applications. Chapter 16 of this volume covers high-temperature panels over 150°C, which may be energized by gas or electricity.

1.

PRINCIPLES OF RADIANT SYSTEMS

Thermal radiation (1) is transmitted at the speed of light, (2) travels in straight lines and can be reflected, (3) elevates the temperature of solid objects by absorption but does not noticeably heat the air through which it travels, and (4) is exchanged continuously between all bodies in a building environment. The rate at which thermal radiation occurs depends on the following factors: • • • •

Temperature of the emitting surface and receiver Emittance of the radiating surface Reflectance, absorptance, and transmittance of the receiver View factor between the emitting and receiver surfaces (viewing angle of the occupant to the thermal radiation source)

One example of heating by thermal radiation is the feeling of warmth when standing in the sun’s rays on a cool, sunny day. Some of the rays come directly from the sun and include the entire electromagnetic spectrum. Other rays are absorbed by or reflected from surrounding objects. This generates secondary rays that are a combination of the wavelength produced by the temperature of the objects and the wavelength of the reflected rays. If a cloud passes in front of the sun, there is an instant sensation of cold. This sensation is caused by the decrease in the amount of heat received from solar The preparation of this chapter is assigned to TC 6.5, Radiant Heating and Cooling.

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tions, and thus heating and cooling panels directly lead to a more controlled, acceptable thermal comfort condition in the space.

1.1

HEAT TRANSFER

Sensible heating or cooling panels transfer heat through temperature-controlled (active) surface(s) to or from an indoor space and its enclosure surfaces by thermal radiation and natural convection.

same plane with the panel is not accounted for by AUST. For example, if only part of the floor is heated, the remainder of the floor is not included in the calculation of AUST, unless it is observed by other panels in the ceiling or wall. The radiation interchange factor for two-surface radiation heat exchange is given by the Hottel equation: 1 Fr = -------------------------------------------------------------------- 1  Ap  1 1 ----------- +  ----- – 1 + ------  ---- – 1 Fp – r   p  A   r 

Heat Transfer by Thermal Radiation The basic equation for a multisurface enclosure with gray, diffuse isothermal surfaces is derived by radiosity formulation methods (see Chapter 4 of the 2017 ASHRAE Handbook—Fundamentals). This equation may be written as qr = Jp –  F pj Jj n

(1)

j=1

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where qr = net heat flux because of thermal radiation on active (heated or cooled) panel surface, W/m2 Jp = total radiosity leaving or reaching panel surface, W/m2 Jj = radiosity from or to another surface in room, W/m2 Fpj = radiation angle factor between panel surface and another surface in room (dimensionless) n = number of surfaces in room other than panel(s)

Equation (1) can be applied to simple and complex enclosures with varying surface temperatures and emittances. The net heat flux by thermal radiation at the panel surfaces can be determined by solving the unknown Jj if the number of surfaces is small. More complex enclosures require computer calculations. Radiation angle factors can be evaluated using Figure 15 in Chapter 4 of the 2017 ASHRAE Handbook—Fundamentals. Fanger (1972) shows room-related angle factors; they may also be developed from algorithms in ASHRAE’s Energy Calculations I (1976). Several methods have been developed to simplify Equation (1) by reducing a multisurface enclosure to a two-surface approximation. In the MRT method, the thermal radiation interchange in an indoor space is modeled by assuming that the surfaces radiate to a fictitious, finite surface that has an emittance and surface temperature that gives about the same heat flux as the real multisurface case (Walton 1980). In addition, angle factors do not need to be determined in evaluating a two-surface enclosure. The MRT equation may be written as qr = Fr Tp4 – Tr4

where

 Fr Tp Tr

= = = =

r

where Fpr = radiation angle factor from panel to fictitious surface (1.0 for flat panel) Ap , Ar = area of panel surface and fictitious surface, respectively p , r = thermal emittance of panel surface and fictitious surface, respectively (dimensionless)

In practice, the thermal emittance p of nonmetallic or painted metal nonreflecting surfaces is about 0.9. When this emittance is used in Equation (4), the radiation exchange factor Fr is about 0.87 for most indoor spaces. Substituting this value in Equation (2), Fr becomes 4.93  10–8. Min et al. (1956) showed that this constant was 5.03  10–8 in their test room. Then the equation for heat flux from thermal radiation for panel heating and cooling becomes approximately qr = 5  10–8[(tp + 273.15) 4 – (AUST + 273.15) 4]

tp = effective panel surface temperature, °C AUST = area-weighted temperature of all indoor surfaces of walls, ceiling, floor, windows, doors, etc. (excluding active panel surfaces), °C

Equation (5) establishes the general sign convention for this chapter, which states that heating by the panel is positive and cooling by the panel is negative. Radiation exchange calculated from Equation (5) is given in Figure 1. The values apply to ceiling, floor, or wall panel output. Radiation removed by a cooling panel for a range of normally encountered temperatures is given in Figure 2.

(2)

The temperature of the fictitious surface is given by an area emittance weighted average of all surfaces other than the panel(s):

 Aj j Tj n

j=p

(3)

j =p

where Aj = area of surfaces other than panels, m2 j = thermal emittance of surfaces other than panel(s) (dimensionless)

When the surface emittances of an enclosure are nearly equal, and surfaces directly exposed to the panel are marginally unheated (uncooled), then Equation (3) becomes the area-weighted average unheated (uncooled) temperature (AUST) of such surfaces exposed to the panels. Therefore, any unheated (uncooled) surface in the

(5)

where

Stefan-Boltzmann constant = 5.67  10–8 W/(m2 ·K4) radiation exchange factor (dimensionless) effective temperature of heating (cooling) panel surface, K temperature of fictitious surface (unheated or uncooled), K

Tr = ---------------------n  Aj j

(4)

Fig. 1 Radiation Heat Flux at Heated Ceiling, Floor, or Wall Panel Surfaces

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Radiant Heating and Cooling In many specific instances where normal multistory commercial construction and fluorescent lighting are used, the indoor air temperature at the 1.5 m level closely approaches the AUST. In structures where the main heat gain is through the walls or where incandescent lighting is used, wall surface temperatures tend to rise considerably above the indoor air temperature. Note that Equation (5) is simplified such that it only characterizes longwave radiation heat transfer between the radiant surface and its enclosure surfaces. It does not include the effects of incident solar (shortwave and longwave) radiation and other internal heat gains on the active surfaces. For cases when solar radiation dominates, special considerations must be taken (Feng et al. 2013a; Liu et al. 2015; Simmonds 2006). Dynamic simulation software that implements the heat balance (HB) method (discussed in Chapter 18 of the 2017 ASHRAE Handbook—Fundamentals) is the most appropriate tool to characterize the performance of radiant surfaces in these cases.

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Heat Transfer by Natural Convection Heat flux from natural convection qc occurs between the indoor air and the temperature-controlled panel surface. Thermal convection coefficients are not easily established. In natural convection, warming or cooling the boundary layer of air at the panel surface generates air motion. In practice, however, many factors, such as the indoor space configuration, interfere with or affect natural convection. Infiltration/exfiltration, occupants’ movement, and mechanical ventilating systems can introduce some forced convection that disturbs the natural convection. Parmelee and Huebscher (1947) included the effect of forced convection on heat transfer occurring on heating or cooling panel surfaces as an increment to be added to the natural-convection coefficient. However, increased heat transfer from forced convection should not be factored into calculations because the increments are unpredictable in pattern and performance, and forced convection does not significantly increase the total heat flux on the active panel surface. Natural-convection heat flux in a panel system is a function of the effective panel surface temperature and the temperature of

6.3 the air layer directly contacting the panel. The most consistent measurements are obtained when the dry-bulb air layer temperature is measured close to the region where the fully developed boundary layer begins, usually 50 to 65 mm from the panel surfaces. Min et al. (1956) determined natural-convection coefficients 1.5 m above the floor in the center of a 3.66 by 7.47 m room (De = 4.91 m). Equations (6) through (11), derived from this research, can be used to calculate heat flux from panels by natural convection. Natural-convection heat flux between an all-heated ceiling surface and indoor air  t p – t a  1.25 qc = 0.20 --------------------------De0.25

(6)

Natural-convection heat flux between a heated floor or cooled ceiling surface and indoor air tp – ta  tp – ta  qc = 2.42 --------------------------------------------0.08 De 0.31

(7)

Natural-convection heat flux between a heated or cooled wall panel surface and indoor air tp – ta  tp – ta  qc = 1.87 --------------------------------------------0.05 H 0.32

(8)

where qc tp ta De H

= = = = =

heat flux from natural convection, W/m2 effective temperature of temperature-controlled panel surface, °C indoor space dry-bulb air temperature, °C equivalent diameter of panel (4  area/perimeter), m height of wall panel, m

Schutrum and Humphreys (1954) measured panel performance in furnished test rooms that did not have uniform panel surface temperatures and found no variation in performance large enough to be significant in heating practice. Schutrum and Vouris (1954) established that the effect of room size was usually insignificant except for very large spaces like hangars and warehouses, for which Equations (6) and (7) should be used. Otherwise, Equations (6), (7), and (8) can be simplified to the following by De = 4.91 m and H = 2.7 m: Natural-convection heat flux between an all-heated ceiling surface and indoor air qc = 0.134(tp – ta)0.25(tp – ta)

(9)

Natural-convection heat flux from a heated ceiling may be augmented by leaving cold strips (unheated ceiling sections), which help initiate natural convection. In this case, Equation (9) may be replaced by Equation (10) (Kollmar and Liese 1957): qc = 0.87(tp – ta)0.25(tp – ta)

(10)

For large spaces such as aircraft hangars, if panels are adjoined, Equation (10) should be adjusted with the multiplier (4.91/De )0.25. Natural-convection heat flux between a heated floor or cooled ceiling surface and indoor air qc = 2.13|tp – ta|0.31(tp – ta)

(11)

Natural-convection heat flux between a heated or cooled wall panel surface and indoor air Fig. 2 Heat Removed by Radiation at Cooled Ceiling or Wall Panel Surface

qc = 1.78|tp – ta|0.32(tp – ta)

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(12)

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2020 ASHRAE Handbook—HVAC Systems and Equipment (SI)

Fig. 3 Natural-Convection Heat Transfer at Floor, Ceiling, and Wall Panel Surfaces There are no confirmed data for floor cooling, but Equation (10) may be used for approximate calculations. Under normal conditions, ta is the dry-bulb indoor air temperature. In floor-heated or ceilingcooled spaces with large proportions of outdoor exposed fenestration, ta may be taken to equal AUST. In cooling, tp is less than ta, so qc is negative. Figure 3 shows heat flux by natural convection at floor, wall, and ceiling heating panels as calculated from Equations (11), (12), (9), and (10), respectively. Figure 4 compares heat removal by natural convection at cooled ceiling panel surfaces, as calculated by Equation (11), with data from Wilkes and Peterson (1938) for specific panel sizes. An additional curve illustrates the effect of forced convection on the latter data. Similar adjustment of the data from Min et al. (1956) is inappropriate, but the effects would be much the same.

Combined Heat Flux (Thermal Radiation and Natural Convection) The combined heat flux on the active panel surface can be determined by adding the thermal-radiation heat flux qr as calculated by Equation (5) (or from Figures 1 and 2) to the natural-convection heat flux qc as calculated from Equations (9), (10), (11), or (12) or from Figure 3 or 4, as appropriate. Equation (5) requires the AUST for the indoor space. In calculating the AUST, the surface temperature of interior walls may be assumed to be the same as the dry-bulb indoor air temperature. The inside surface temperature tw of outdoor exposed walls and outdoor exposed floors or ceilings can be calculated from the following relationship: h(ta – tu) = U(ta – to) (13) or tu = ta – U (14) ---- (ta – to) h where h = natural-convection coefficient of the inside surface of an outdoor exposed wall or ceiling

Fig. 4

Empirical Data for Heat Removal by Ceiling Cooling Panels from Natural Convection

U = overall heat transfer coefficient of wall, ceiling, or floor, W/(m2 ·K) ta = dry-bulb indoor space design air temperature, °C tu = inside surface temperature of outdoor exposed wall, °C to = dry-bulb outdoor design air temperature, °C

From Table 10 in Chapter 26 of the 2017 ASHRAE Handbook— Fundamentals, h = 9.26 W/(m2 ·K) for a horizontal surface with heat flow up h = 9.09 W/(m2 ·K) for a vertical surface (wall) h = 8.29 W/(m2 ·K) for a horizontal surface with heat flow down

Figure 5 is a plot of Equation (13) for a vertical outdoor wall with 21°C dry-bulb indoor air temperature and h = 9.09 W/(m2 · K). For rooms with dry-bulb air temperatures above or below 21°C, the values in Figure 5 can be corrected by the factors plotted in Figure 6. Tests by Schutrum et al. (1953a, 1953b) and simulations by Kalisperis (1985) based on a program developed by Kalisperis and Summers (1985) show that the AUST and indoor air temperature are almost equal, if there is little or no outdoor exposure. Steinman et al. (1989) noted that this may not apply to enclosures with large fenestration or a high percentage of outdoor exposed wall and/or ceiling surface area. These surfaces may have a lower (in heating) or higher (in cooling) AUST, which increases the heat flux from thermal radiation. Figure 7 shows the combined heat flux from thermal radiation and natural convection for cooling, as given in Figures 2 and 4. The data in Figure 7 do not include solar, lighting, occupant, or equipment heat gains.

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In suspended-ceiling panels, heat is transferred from the back of the ceiling panel to the floor slab above (in heating) or vice versa (in cooling). The ceiling panel surface temperature is affected because of heat transfer to or from the panel and the slab by thermal radiation and, to a much smaller extent, by natural convection. The thermalradiation component of the combined heat flux can be approximated using Equation (5) or Figure 1. The natural-convection component can be estimated from Equation (10) or (11) or from Figure 3 or 4. In this case, the temperature difference is that between the top of the ceiling panel and the midspace of the ceiling. The temperature of the ceiling space should be determined by testing, because it varies with different panel systems. However, much of this heat transfer is nullified when insulation is placed over the ceiling panel, which, for perforated metal panels, also provides acoustical absorption. If artificial lighting fixtures are recessed into the suspended ceiling space, radiation from the top of the fixtures raises the overhead slab temperature and heat is transferred to the indoor air by natural convection. This heat is absorbed at the top of the cooled ceiling panels both by thermal radiation, in accordance with Equation (5) or

6.5 Figure 2, and by thermal convection, generally in accordance with Equation (10). The amount the top of the panel absorbs depends on the system. Similarly, panels installed under a roof absorb additional heat, depending on configuration and insulation.

1.2

FACTORS AFFECTING HEAT TRANSFER

Panel Thermal Resistance Any thermal resistance between the indoor space and the active panel surface, as well as between the active panel surface and the hydronic tubing or electric circuitry in the panel, reduces system performance. Thermal resistance to heat transfer may vary considerably among different panels, depending on the type of bond between the tubing (electric cabling) and the panel material. Factors such as corrosion or adhesion defects between lightly touching surfaces and the method of maintaining contact may change the bond with time. The actual thermal resistance of any proposed system should be verified by testing. Specific resistance and performance data, when available, should be obtained from the manufacturer based on performance tests such as those described in ANSI/ ASHRAE Standard 138 or European Standard EN 14240 (CEN 2004). Panel thermal resistances include rt = thermal resistance of tube wall per unit tube spacing in a hydronic system, (m·K)/W rs = thermal resistance between tube (electric cable) and panel body per unit spacing between adjacent tubes (electric cables), (m·K)/W rp = thermal resistance of panel body, (m 2 ·K)/W rc = thermal resistance of active panel surface covers, (m 2 ·K)/W ru = characteristic (combined) panel thermal resistance, (m 2 ·K)/W For a given adjacent tube (electric cable) spacing M, ru = rt M + rs M + rp + rc

(15)

When the tubes (electric cables) are embedded in the slab, rs may be neglected. However, if they are externally attached to the body of the panel, rs may be significant, depending on the quality of bonding. Table 1 gives typical rs values for various ceiling panels. Fig. 5 Relation of Inside Surface Temperature to Overall Heat Transfer Coefficient

Fig. 6 Inside Surface Temperature Correction for Exposed Wall at Dry-Bulb Air Temperatures Other Than 21°C

Fig. 7 Cooled Ceiling Panel Performance in Uniform Environment with No Infiltration and No Internal Heat Sources

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6.6

2020 ASHRAE Handbook—HVAC Systems and Equipment (SI) Table 2 Thermal Conductivity of Typical Tube Material

Table 1 Thermal Resistance of Ceiling Panels

Thermal Conductivity kt , W/(m ·K)

Thermal Resistance Type of Panel

rp , (m2 ·K)/W

rs , (m·K)/W

xp ----kp

0.32

xp ----kp

0.38

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xp ----kp

Material Carbon steel (AISI 1020) Aluminum Copper (drawn) Red brass (85 Cu-15 Zn) Stainless steel (AISI 202) Low-density polyethylene (LDPE) High-density polyethylene (HDPE) Cross-linked polyethylene (VPE or PEX) Textile-reinforced rubber hose (HTRH) Polypropylene block copolymer (PP-C) Polypropylene random copolymer (PP-RC)

0.10

x p – Do / 2 ----------------------kp

0

x p – Do / 2 ----------------------kp

0.12

The value of rp may be calculated if the characteristic panel thickness xp and the thermal conductivity kp of the panel material are known. If the tubes (electric cables) are embedded in the panel, x p – Do / 2 rp = ----------------------kp

(16)

where Do = outside diameter of the tube (electric cable). Hydronic floor heating by a heated slab and gypsum-plaster ceiling heating are typical examples. If the tubes (electric cable) are attached to the panel, xp rp = ----kp

ln  D o /D i  rt = ------------------------2k t

(18)

For an elliptical tube with semimajor and semiminor axes of a and b, respectively, measured at both the outside and inside of the tube,  a o + b o /  a i + b i  kt = ln --------------------------------------------2k t

(19)

In an electric cable, rt = 0. In metal pipes, rt is virtually the fluid-side thermal resistance: 1 rt = -------hDi

If the tube has multiple layers, Equations (18) or (19) should be applied to each individual layer and then summed to calculate the tube’s total thermal resistance. Thermal resistance of capillary tube mats can also be calculated from either Equation (18) or (19). Typically, capillary tubes are circular, 2 mm in internal diameter, and 12 mm apart. Capillary tube mats can be easily applied to existing ceilings in a sand plaster cover layer. If the tube material is nonmetallic, oxygen ingress may be a problem, especially in panel heating. To avoid oxygen corrosion in the heating system, either (1) tubing with an oxygen barrier layer, (2) corrosion-inhibiting additives in the hydronic system, or (3) a heat exchanger separating the panel circuit from the rest of the system should be used. Table 8 in Chapter 4 of the 2017 ASHRAE Handbook—Fundamentals may be used to calculate the forced-convection heat transfer coefficient h. Table 2 gives values of kt for different tube and pipe materials.

Effect of Floor Coverings Active panel surface coverings like carpets and pads on the floor can have a pronounced effect on a panel system’s performance. The added thermal resistance rc reduces the panel surface heat flux. To reestablish the required performance, the water temperature must be increased (decreased in cooling). Thermal resistance of a panel covering is xc rc = ---kc

(17)

Metal ceiling panels (see Table 1) and tubes under subfloor (see Figure 23) are typical examples. Thermal resistance per unit on-center spacing (M = unity) of circular tubes with inside diameter Di and thermal conductivity kt is

(20)

52 237 390 159 17 0.31 0.42 0.38 0.29 0.23 0.24

(21)

where rc = thermal resistance of panel covering, (m2 ·K)/W xc = thickness of each panel covering, m kc = thermal conductivity of each panel cover, W/(m·K)

If the active panel surface has more than one cover, individual rc values should be added. Table 3 gives typical rc values for floor coverings. If there are covered and bare floor panels in the same hydronic system, it may be possible to maintain a sufficiently high water temperature to satisfy the covered panels and balance the system by throttling the flow to the bare slabs. In some instances, however, the increased water temperature required when carpeting is applied over floor panels makes it impossible to balance floor panel systems in which only some rooms have carpeting unless the piping is arranged to permit zoning using more than one water temperature.

Panel Heat Losses or Gains Heat transferred from the upper surface of ceiling panels, the back surface of wall panels, the underside of floor panels, or the exposed perimeter of any panel is considered a panel heat loss (gain

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Radiant Heating and Cooling

6.7 Table 3 Thermal Resistance of Floor Coverings Thermal Resistance rc , (m2 ·K)/W

Description

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Bare concrete, no covering Asphalt tile Rubber tile Light carpet Light carpet with rubber pad Light carpet with light pad Light carpet with heavy pad Heavy carpet Heavy carpet with rubber pad Heavy carpet with light pad Heavy carpet with heavy pad 10 mm hardwood 16 mm wood floor (oak) 13 mm oak parquet and pad Linoleum Marble floor and mudset Rubber pad Prime urethane underlayment, 10 mm 1.5 kg/m3 waffled sponge rubber Bonded urethane, 13 mm

0 0.009 0.009 0.106 0.176 0.247 0.300 0.141 0.211 0.281 0.335 0.095 0.100 0.120 0.021 0.031 0.109 0.284 0.137 0.368

Notes: 1. Carpet pad thickness should not be more than 6 mm 2. Total thermal resistance of carpet is more a function of thickness than of fiber type. 3. Generally, thermal resistance (R-value) is approximately 0.018 times the total carpet thickness in millimetres. 4. Before carpet is installed, verify that the backing is resistant to long periods of continuous heat up to 50°C

T = temperature difference, either tp – to in electric heating or tw – to in hydronic heating, K to = operative temperature, °C m = 2 + rc 2rp

Fig. 8 Downward and Edgewise Heat Loss Coefficient for Concrete Floor Slabs on Grade in cooling). Panel heat losses (gains) are part of the building heat loss (gain) if the heat transfer is between the panel and the outside of the building. If heat transfer is between the panel and another conditioned space, the panel heat loss (gain) is a positive conditioning contribution for that space instead. In either case, the magnitude of panel loss (gain) should be calculated. Panel heat loss (gain) to (from) space outside the conditioned space should be kept to a reasonable amount by insulation. For example, a floor panel may overheat the basement below, and a ceiling panel may cause the temperature of a floor surface above it to be too high for comfort unless it is properly insulated. The heat loss from most panels can be calculated by using the coefficients given in Table 1 in Chapter 26 of the 2017 ASHRAE Handbook—Fundamentals. These coefficients should not be used to determine downward heat loss from panels built on grade because heat flow from them is not uniform (ASHAE 1956, 1957; Sartain and Harris 1956). The heat loss from panels built on grade can be estimated from Figure 8 or from Equations (40) and (41) in Chapter 18 of the 2017 ASHRAE Handbook—Fundamentals.

Panel Performance As with other electric or hydronic terminal units, panel performance can be described by the following equation: q = CT | T | m–1

where q = combined heat flux on panel surface, W/m2 C = characteristic performance coefficient, W/(m2 ·K m)

(22)

C and m for a particular panel may be either experimentally determined or calculated from the design material given in this chapter. In either case, sufficient data or calculation points must be gathered to cover the entire operational design range (ASHRAE Standard 138). Radiant panel product performance data can be obtained from the manufacturers. ASHRAE Standard 138 and European Standards EN 14240 and EN 14037 regulate the testing and rating method for radiant panel cooling and heating performance. European Standards EN 1264 and EN 15377 regulate the testing methods for the embedded radiant heating and cooling systems.

1.3

PANEL DESIGN

Either hydronic or electric circuits control the active panel surface temperature. The required effective surface temperature tp necessary to maintain a combined heat flux q (where q = qr + qc) at steady-state design conditions can be calculated by using applicable heat flux equations for qr and qc , depending on the position of the panel. At a given ta , AUST must be predicted first. Figures 9 and 10 can also be used to find tp when q and AUST are known. The next step is to determine the required average water (brine) temperature tw in a hydronic system. It depends primarily on tp , tube spacing M, and the characteristic panel thermal resistance ru . Figure 9 provides design information for heating and cooling panels, positioned either at the ceiling or on the floor. The combined heat flux for ceiling and floor panels can be read directly from Figure 9. Here qu is the combined heat flux on the floor panel and qd is the combined heat flux on the ceiling panel. For an electric heating system, tw scales correspond to the skin temperature of the cable. The following algorithm (TSI 1994) may also be used to design and analyze panels under steady-state conditions:

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2020 ASHRAE Handbook—HVAC Systems and Equipment (SI)  t p – t a M td  ta + ------------------------- + q(rp + rc + rsM ) 2W + D o

f

(23)

where

m  tp – ta 

 ki xi

12

for tp  ta

(26)

i=1

td = average skin temperature of tubing (electric cable), °C q = combined heat flux on panel surface, W/m2 ta = dry-bulb indoor air design temperature, °C. In floor-heated or ceiling-cooled indoor spaces that have large fenestration, ta may be replaced with AUST. Do = outside diameter of embedded tube or characteristic contact width of attached heating or cooling element with panel (see Table 1), m M = on-center spacing of adjacent tubes (electric cables), m 2W = net spacing between tubing (electric cables), M – Do , m  = fin efficiency, dimensionless

The first two terms in Equation (23) give the maximum (minimum in cooling) value of the panel surface temperature profile; consequently, if tube spacing M is too large, hot strips along the panel surface may occur in heating, or local condensation occur on these cold strips in sensible cooling mode. tanh  f W   = -----------------------fW

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q --------------------------------------n

  1f W

(24)

for f W > 2

(25)

The following equation, which includes transverse heat diffusion in the panel and surface covers, may be used to calculate the fin coefficient f:

where f = fin coefficient m = 2 + rc /2rp n = total number of different material layers, including panel and surface covers xi = characteristic thickness of each material layer i, m ki = thermal conductivity of each layer i, W/(m·K)

For a hydronic system, the required average water (brine) temperature is tw = (q + qb)Mrt + td (27) where qb is the flux of back and perimeter heat losses (positive) in a heated panel or gains (negative) in a cooling panel. This algorithm may be applied to outdoor slab heating systems, provided that the combined heat flux q is calculated according to Chapter 51 in the 2015 ASHRAE Handbook—HVAC Applications, with the following conditions: no snow, no evaporation, q (in panel heating) = qh (radiation and convection heat flux in snow-melting calculations). With a careful approach, outdoor slab cooling systems may be analyzed by incorporating the solar radiation gain, thermal radiation, and forced convection from the sky and ambient air. ISO Standard 18566-2014 provides an in-depth description of the design, installation, and control of radiant panels.

Fig. 9 Design Graph for Sensible Heating and Cooling with Floor and Ceiling Panels

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Radiant Heating and Cooling Special Cases Figure 9 may also be used for panels with tubing not embedded in the panel: • xp = 0 if tubes are externally attached. • In spring-clipped external tubing, Di = 0 and Do is the clip thickness. Warm air and electric heating elements are two design concepts influenced by local factors. The warm-air system has a special cavity construction where air is supplied to a cavity behind or under the panel surface. The air leaves the cavity through a normal diffuser arrangement and is supplied to the indoor space. Generally, these systems are used as floor panels in schools and in floors subject to extreme cold, such as in an overhang. Cold outdoor temperatures and heating medium temperatures must be analyzed with regard to potential damage to the building construction. Electric heating elements embedded in the floor or ceiling construction and unitized electric ceiling panels are used in various applications to provide both full heating and spot heating of the space.

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Examples Residential heating applications usually consist of tubes or electric elements embedded in masonry floors or plaster ceilings. This construction is suitable where loads are stable and building design minimizes solar effects. However, in buildings where fenestration is large and thermal loads change abruptly, the slow response, thermal lag, and override effect of masonry panels are unsatisfactory. Metal ceiling panels with low thermal mass respond quickly to load changes (Berglund et al. 1982). Panels are preferably located in the ceiling because it is exposed to all other indoor surfaces, occupants, and objects in the conditioned indoor space. It is not likely to be covered, as floors are, and higher surface temperatures can be maintained. Figure 10 gives design data for ceiling and wall panels with effective surface temperatures up to 140°C. Example 1. An in-slab, on-grade panel (see Figure 20) will be used for both heating and cooling. M = 300 mm, ru = 0.09 (m2 ·K)/W, and rc/rp is less than 4. ta is 20°C in winter and 4.5°C in summer. AUST is expected to be 1 K less than ta in winter heating and 0.5 K higher than ta in summer cooling. What is the average water temperature and effective floor temperature (1) for winter heating when qu = 130 W/m2, and (2) for summer cooling when –qu = 50 W/m2? Solution: Winter heating To obtain the average water temperature using Figure 9, start on the left axis where qu = 130 W/m2. Proceed right to the intersect ru = 0.09 and then down to the M = 300 mm line. The reading is AUST + 31, which is the solid line value because rc/rp < 4. As stated in the initial problem, AUST = ta – 1 or AUST = 20 – 1 = 19°C. Therefore, the average water temperature would be tw = 31 + 19 = 50°C. To find the effective floor temperature, start at qu = 130 W/m2 in Figure 9 and proceed right to AUST = ta – 1 K. The solid line establishes 12 K as the temperature difference between the panel and the indoor air. Therefore, the effective floor temperature tp = ta + 12 or tp = 20 + 12 = 32°C. Summer cooling Using Figure 9, start at the left axis at –qu = 50 W/m2. Proceed to ru = 0.09, and then up (for cooling) to M = 300 mm, which reads ta – 11 or 24.5 – 11 = 13.5°C average water temperature for cooling. To obtain the effective floor temperature at –qu = 50 W/m2, proceed to AUST – ta = +0.5 K, which reads –5°C. Therefore, the effective floor temperature is 24.5 – 5 = 19.5°C. Example 2. An aluminum extrusion panel, which is 0.127 mm thick with heat element spacing of M = 150 mm, is used in the ceiling for heating. If a ceiling heat flux qd of 1260 W/m2 is required to maintain room

6.9 temperature ta at 20°C, what is the required heating element skin temperature td and effective panel surface temperature tp? Solution: Using Figure 10 enter the left axis heat flux qd at 1260 W/m2. Proceed to the line corresponding to ta = 20°C and then move up to the M = 150 mm line. The ceiling heating element temperature td at the intersection point is 160°C. From the bottom axis of Figure 10, the effective panel surface temperature tp is 129°C.

2.

GENERAL DESIGN CONSIDERATIONS

Principal advantages of panel systems are the following: • Because not only indoor air temperature but also mean radiant temperature can be controlled, total human thermal comfort may be better satisfied. • Because the operative temperature for required human thermal comfort may be maintained by primarily controlling the mean radiant temperature of the conditioned indoor space, dry-bulb air temperature may be lower (in heating) or higher (in cooling), which reduces sensible heating or cooling loads (see Chapter 16 for the definition and calculation of operative and mean radiant temperatures). • Hydronic panel systems may be connected in series, following other hydronic heating or cooling systems (i.e., their return water may be used), increasing exergetic efficiency. • Comfort levels can be better than those of other space-conditioning systems because thermal loads are satisfied directly and air motion in the space corresponds to required ventilation only. • Waste and low-enthalpy energy sources and heat pumps may be directly coupled to panel systems without penalty on equipment sizing and operation. Being able to select from a wide range of moderate operation temperatures ensures optimum design for minimum cost and maximum thermal and exergetic efficiency. • Seasonal thermal distribution efficiency in buildings may be higher than in other hydronic systems. • In terms of simple payback period, ceiling cooling panels and chilled beams have the highest technical energy savings potential (DOE 2002). • Part or all of the building structure may be thermally activated (Meierhans and Olesen 2002). • Space-conditioning equipment is not required at outdoor exposed walls, simplifying wall, floor, and structural systems. • Almost all mechanical equipment may be centrally located, simplifying maintenance and operation. • No space within the conditioned space is required for mechanical equipment. This feature is especially valuable in hospital patient rooms, offices, and other applications where space is at a premium, if maximum cleanliness is essential or legally required. • Significantly reduced need for ductwork allows either lower floor-to-floor heights (reducing façade costs and overall building height) or higher ceilings (an architecturally desirable feature). • Draperies and curtains can be installed at outdoor exposed walls without interfering with the space-conditioning system. • When four-pipe systems are used, cooling and heating can be simultaneous, without central zoning or seasonal changeover. • Supply air requirements usually do not exceed those required for ventilation and humidity control. • Modular panels provide flexibility to meet changes in partitioning. • A 100% outdoor air system may be installed with smaller penalties in refrigeration load because of reduced air quantities. • A common central air system can serve both the interior and perimeter zones. • Wet-surface cooling coils are eliminated from the occupied space, reducing the potential for septic contamination.

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6.10

2020 ASHRAE Handbook—HVAC Systems and Equipment (SI) • In-floor heating creates inhospitable living conditions for house dust mites compared to other heating systems (Sugawara et al. 1996). Disadvantages include the following: • Response time can be slow if controls and/or heating elements are not selected or installed correctly. • Improper selection of panel heating or cooling tube or electrical heating element spacing and/or incorrect sizing of heating/cooling source can cause nonuniform surface temperatures or insufficient sensible heating or cooling capacity. • Radiant systems can satisfy only sensible heating and cooling loads. In a stand-alone radiant cooling system, dehumidification and surface condensation may be of prime concern and a second-

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• Modular panel systems can use the automatic sprinkler system piping (see NFPA Standard 13, Section 3.6). The maximum water temperature must not fuse the heads. • Panel heating and cooling with minimum supply air quantities provide a draft-free environment. • Noise associated with fan-coil or induction units is eliminated. • Peak loads are reduced as a result of thermal energy storage in the panel structure, as well as walls and partitions directly exposed to panels. • Panels can be combined with other space-conditioning systems to decouple several indoor requirements (e.g., humidity control, indoor air quality, air velocity) and optimally satisfy them without compromises.

Fig. 10

Design Graph for Heating with Aluminum Ceiling and Wall Panels

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Radiant Heating and Cooling

Fig. 11

Typical Residential Hybrid HVAC System

ary air-handling system must provide ventilation air and handle latent loads. • Radiant systems in which the tubing is embedded in a building surface or in the structural slab typically have a large exposed concrete surface area. Additional acoustical treatments are often required to reduce room reverberation time to a more desirable level.

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2.1

HYBRID SYSTEMS

In general, any HVAC system relies on a single heat transfer mode as its major heat transfer mechanism. For example, central air conditioning is a forced-convection system, and a high-intensity radiant system operates almost purely with thermal radiation. Additionally, every system has a typical range for satisfactory and economical operation, with its own advantages, limitations, and disadvantages. A single system may not be sufficient to encompass all requirements of a given building in the most efficient and economical way. Under these circumstances, it may be more desirable to decouple several components of indoor space conditioning, and satisfy them by several dedicated systems. A hybrid heating and cooling system is an optimum partnership of multiple, collocated, simultaneously operating heating and cooling systems, each of which is based on one of the primary heat transfer modes (i.e., radiation and convection). In its most practical form, hybrid heating and cooling may consist of a radiant system and a forced-air system. The forced-air component may be a central HVAC system or a hydronic system such as fan-coils, as shown in Figure 11. Here, a panel system is added downstream of the condensing fan-coils, and ventilation is provided by a separate system with substantially reduced duct size. Dehumidification and ventilation problems that may be associated with stand-alone panel cooling systems may be eliminated by hybrid HVAC systems. ASHRAE research project RP-1140 (Scheatzle 2003) successfully demonstrates the use of panel systems for both heating and cooling, in conjunction with a forced-convection system, to economically achieve year-round thermal comfort in a residence using both active and passive performance of the building and a groundsource heat pump. Skylighting, an energy recovery ventilator, and packaged dehumidifiers were also used. Twenty-four-month-long tests indicated that a radiant/convective system can offer substantial cost savings, given proper design and control.

3.

6.11

RADIANT HEATING AND COOLING SYSTEMS

The following are the most common forms of panels applied in radiant heating and cooling systems: • Hydronic ceiling panels • Embedded tubing in ceilings, walls, or floors • Electric ceiling panels for heating only

Fig. 12 Metal Ceiling Panels Attached to Pipe Laterals • Electrically heated ceilings or floors

3.1

HYDRONIC CEILING PANELS

Metal ceiling panels can be integrated into a system that heats and cools. In such a system, a source of dehumidified ventilation air is required in summer, so the system is classed as an air-andwater system. Also, various amounts of forced air are supplied year-round. When metal panels are applied for heating only, a ventilation system may be required, depending on local codes. Ceiling panel systems are an outgrowth of perforated metal, suspended acoustical ceilings. These ceiling panel systems are usually designed into buildings where the suspended acoustical ceiling can be combined with panel heating and cooling. The panels can be designed as small units to fit the building module, which provides extensive flexibility for zoning and control, or the panels can be arranged as large continuous areas for maximum economy. Some ceiling installations require active panels to cover only part of the indoor space and compatible matching acoustical panels for the remaining ceiling area. Three types of metal ceiling systems are available. The first consists of light aluminum panels, usually 300 by 600 mm, attached in the field to 15 mm galvanized pipe coils. Figure 12 shows a metal ceiling panel system that uses 15 mm pipe laterals on 150, 300, or 600 mm centers, hydraulically connected in a sinuous or parallelflow welded system. Aluminum ceiling panels are clipped to these pipe laterals and act as a heating panel when warm water is flowing or as a cooling panel when chilled water is flowing. The second type of panel consists of a copper coil secured to the aluminum face sheet to form a modular panel. Modular panels are available in sizes up to about 910 by 1520 mm and are held in position by various types of ceiling suspension systems, most typically a standard suspended T-bar 600 by 1200 mm exposed grid system. Figure 13 shows metal panels using a copper tube pressed into an aluminum extrusion, although other methods of securing the copper tube have proven equally effective. Metal ceiling panels can be perforated so that the ceiling becomes sound absorbent when acoustical material is installed on the back of the panels. The acoustical blanket is also required for thermal reasons, so that reverse loss or upward flow of heat from the metal ceiling panels is minimized. The third type of panel is an aluminum extrusion face sheet with a copper tube mechanically fastened into a channel housing on the back. Extruded panels can be manufactured in almost any shape and size. Extruded aluminum panels are often used as long, narrow panels at the outside wall and are independent of the ceiling system. Panels 380 or 510 mm wide usually satisfy a typical office building’s heating requirements. Lengths up to 6 m are available. Figure 14

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shows metal panels using a copper tube pressed into an aluminum extrusion. Some manufacturers also use graphite as a panel material. Performance data for extruded aluminum panels vary with the copper tube/aluminum contact and test procedures used. Hydronic ceiling panels have a low thermal resistance and respond quickly to changes in space conditions. Table 1 gives thermal resistance values for various ceiling constructions. Metal ceiling panels can be used with any of the all-air cooling systems described in Chapter 2. Chapters 17 and 18 of the 2017 ASHRAE Handbook—Fundamentals describe how to calculate heating loads. Double-glazing and heavy insulation in outside walls reduce transmission heat losses. As a result, infiltration and reheat have become of greater concern. Additional design considerations include the following: • Perimeter radiant heating panels extending not more than 1 m into the indoor space may operate at higher temperatures, as described in the section on Hydronic Panel Systems. • Hydronic panels operate efficiently at low temperature and are suitable for condenser water heat reclaim systems. • Locate ceiling panels adjacent to the outside wall and as close as possible to the areas of maximum load. The panel area within 1 m of the outside wall should have a heating capacity equal to or greater than 50% of the wall transmission load. • Ceiling system designs based on passing return air through perforated modular panels into the plenum space above the ceiling are not recommended because much of this heat is lost to the return air system in heating mode.

• When selecting heating design temperatures for a ceiling panel surface or average water temperature, the design parameters are as follows: • Excessively high temperatures over the occupied zone will cause the occupant to experience a “hot head effect.” • Temperatures that are too low can result in an oversized, uneconomical panel and a feeling of coolness at the outside wall. • Give ceiling panel location priority. • With normal ceiling heights of 2.4 to 2.8 m, panels less than 1 m wide at the outside wall can be designed for 113°C surface temperature. If panels extend beyond 1 m into the indoor space, the panel surface temperature should be limited to the values as given in Figure 15. • Allow sufficient space above the ceiling for installation and connection of the piping that forms the panel ceiling. • Hydronic ceiling panels provide a fast-response system (Watson et al. 1998). Metal acoustic panels provide heating, cooling, sound absorption, insulation, and unrestricted access to the plenum space. They are easily maintained, can be repainted to look new, and have a life expectancy in excess of 30 years. The system is quiet, comfortable, draft-free, easy to control, and responsive. The system is a basic airand-water system. Metal heating panels, hydronic and electric, are applied to building perimeter spaces for heating in much the same way as finnedtube convectors. They can also be integrated into the ceiling design to provide a narrow band of panel heating around the perimeter of the building. The panel system offers advantages over baseboard or overhead air in appearance, comfort, operating efficiency and cost, maintenance, and product life. ASHRAE Standard 138 discusses testing and rating of ceiling panels.

3.2

EMBEDDED SYSTEMS WITH TUBING IN CEILINGS, WALLS, OR FLOORS

Layout and design of embedded systems for heating and cooling begin early in the job. The type and location chosen influences the design and, conversely, thermal considerations may dictate which system type to use. ISO Standard 11855-2012 provides an in-depth description of the design, installation, and control of embedded Fig. 13 Metal Ceiling Panels Bonded to Copper Tubing

Fig. 14

Extruded Aluminum Panels with Integral Copper Tube

Fig. 15

Permitted Design Ceiling Surface Temperatures at Various Ceiling Heights

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Radiant Heating and Cooling radiant systems. One of the following types of construction is generally used:

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• Pipe or tube is embedded in the lower portion of a concrete slab, generally within 25 mm of its lower surface (Figure 16). If plaster is to be applied to the concrete, the piping may be placed directly on the wood forms. If the slab is to be used without plaster finish, the piping should be installed not less than 20 mm above the undersurface of the slab. The minimum coverage must comply with local building code requirements. • Pipe or tube is embedded in a metal lath and plaster ceiling. If the lath is suspended to form a hung ceiling, the lath and heating coils are securely wired to the supporting members so that the lath is below, but in good contact with, the coils. Plaster is then applied to the metal lath, carefully embedding the coil as shown in Figure 17. • Smaller-diameter copper or thermoplastic tube is attached to the underside of wire or gypsum lath. Plaster is then applied to the lath to embed the tube, as shown in Figure 18. • Other forms of ceiling construction are composition board, wood paneling, etc., with warm-water piping, tube, or channels built into the panel sections. Coils are usually laid in a sinusoidal pattern, although some header or grid-type coils have been used in ceilings. Coils are typically thermoplastic tube, spaced from 100 to 300 mm on centers, depending on the required heat flux, pipe or tube size, and other factors. Where plastering is applied to pipe coils, a standard three-coat gypsum plastering specification is followed, with a minimum of 10 mm of cover below the tubes when they are installed below the lath. Generally, the surface temperature of plaster panels should not

6.13 exceed 50°C, which may be satisfied by limiting the average water temperature to a maximum temperature of 60°C. Insulation should be placed above the coils to reduce back loss, which is the difference between heat supplied to the coil and net useful heat output to the heated indoor space. To protect the plaster installation and to ensure proper air drying, heat must not be applied to the panels for two weeks after all plastering work has been completed. When the system is started for the first time, water supplied to the panels should not be more than 10 K above the prevailing indoor air temperature and not in excess of 32°C. Water should be circulated at this temperature for about two days, and then increased at a rate of about 3 K per day to 60°C. During the air-drying and preliminary warm-up periods, there should be adequate ventilation to carry moisture from the panels. No paint or paper should be applied to the panels before these periods have been completed or while the panels are being operated. After paint and paper have been applied, an additional shorter warm-up period, similar to first-time starting, is also recommended.

Hydronic Wall Panels Although piping embedded in walls is not as widely used as floor and ceiling panels, it can be constructed by any of the methods outlined for ceilings or floors. Its design is similar to other hydronic panels (see Equations [22] to [26]). Equations (5) and (12) give the heat flux at the surface of wall panels.

Hydronic Floor Panels Interest has increased in floor heating with the introduction of nonmetallic tubing and new design, application, and control techniques. Whichever method is used for optimum floor output and comfort, it is important that heat be evenly distributed over the floor. Spacing is generally 100 to 300 mm on centers for the coils. Wide spacing under tile or bare floors can cause uneven surface temperatures. Embedded Tubes or Pipes in Concrete Slab. Thermoplastic, rubber, ferrous, and nonferrous pipes, or composite tubes (e.g., thermoplastic tubes with aluminum sleeves) are used in floor slabs

Fig. 16 Coils in Structural Concrete Slab

Fig. 18 Coils in Plaster Below Lath

Fig. 17 Coils in Plaster Above Lath

Fig. 19

Coils in Floor Slab on Grade

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2020 ASHRAE Handbook—HVAC Systems and Equipment (SI)

Fig. 21 Tube in Subfloor

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Fig. 20 Embedded Tube in Thin Slab that rest on grade. Hydronic coils are constructed as sinusoidalcontinuous coils or arranged as header coils with a spacing of 150 to 450 mm on centers. The coils are generally installed with 40 to 100 mm of cover above them. Insulation is recommended to reduce perimeter and back losses where the outer surface is exposed to the exterior. Figure 19 shows application of hydronic coils in slabs resting on grade. Coils should be embedded completely and should not rest on an interface. Any supports used for positioning heating coils should be nonabsorbent and inorganic. Reinforcing steel, angle iron, pieces of pipe or stone, or concrete mounds can be used. No wood, brick, concrete block, or similar materials should support coils. A waterproofing layer is desirable to protect insulation and piping. Where coils are embedded in structural load-supporting slabs above grade, construction codes may affect their position. Otherwise, the coil piping is installed as described for slabs resting on grade. Embedded systems may fail sometime during their life. Adequate valves and properly labeled drawings help isolate the point of failure. Suspended Floor Tubing or Piping. Piping may be applied on or under suspended wood floors using several construction methods. Piping may be attached to the surface of the floor and embedded in a layer of concrete or gypsum, mounted in or below the subfloor, or attached directly to the underside of the subfloor using metal panels to improve heat transfer from the piping. An alternative method is to install insulation with a reflective surface and leave an air gap of 50 to 100 mm to the subfloor. Whichever method is used for optimum floor output and comfort, it is important that heat be evenly distributed throughout the floor. Tubing generally has a spacing of 100 to 300 mm on centers. Wide spacing under tile or bare floors can cause uneven surface temperatures. Figure 20 depicts construction with piping embedded in concrete or gypsum. The embedding material is generally 25 to 50 mm thick when applied to a wood subfloor. Gypsum products specifically designed for floor heating can generally be installed 25 to 40 mm thick because they are more flexible and crack-resistant than concrete. When concrete is used, it should be of structural quality to reduce cracking caused by movement of the wood frame or shrinkage. The embedding material must provide a hard, flat, smooth surface that can accommodate a variety of floor covers. As shown in Figure 21, tubing may also be installed in the subfloor. The tubing is installed on top of the rafters between the subflooring members. Heat diffusion and surface temperature can

Fig. 22 Tube Under Subfloor be improved uniformly by adding metal heat transfer plates, which spread heat beneath the finished flooring. A third construction option is to attach the tube to the underside of the subfloor with or without metal heat transfer plates. The construction is depicted in Figure 22. Transfer from the hot-water tube to the surface of the floor is the important consideration in all cases. The floor surface temperature affects the actual heat transfer to the space. Any hindrance between the heated water tube and the floor surface reduces system effectiveness. The method that transfers and spreads heat evenly through the subfloor with the least resistance produces the best results.

3.3

ELECTRICALLY HEATED RADIANT SYSTEMS

Several panel systems convert electrical energy to heat, raising the temperature of conditioned indoor surfaces and the indoor air. These systems are classified by the temperature of the heated system. Higher-temperature surfaces require less area to maintain occupant comfort. Surface temperatures are limited by the ability of the materials to maintain their integrity at elevated temperatures. The maximum effective surface temperature of floor panels is limited to what is comfortable to occupants’ feet.

Electric Ceiling Panels Prefabricated Electric Ceiling Panels. These panels are available in sizes 300 to 1800 mm wide by 600 to 3600 mm long by 13 to 50 mm thick. They are constructed with metal, glass, or semirigid fiberglass board or vinyl. Heated surface temperatures range from 40 to 150°C, with corresponding heat fluxes ranging from 270 to 1100 W/m2 for 120 to 480 V services. A panel of gypsum board embedded with insulated resistance wire is also available. It is installed as part of the ceiling or between joists in contact with a ceiling. Heat flux is limited to 240 W/m to maintain

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Radiant Heating and Cooling

6.15

the board’s integrity by keeping the heated surface temperature below 40°C. Nonheating leads are furnished as part of the panel. Some panels can be cut to fit; others must be installed as received. Panels may be either flush or surface-mounted, and in some cases, they are finished as part of the ceiling. Rigid 600 by 1200 mm panels for lay-in ceilings (Figure 23) are about 25 mm thick and have a mass of 2.6 to 11 kg. Typical characteristics of an electric panel are listed in Table 4. Panels may also be (1) surfacemounted on gypsum board and wood ceilings or (2) recessed between ceiling joists. Panels range in size from 1200 mm wide to 2400 mm long. Their maximum power output is 1000 W/m2. Electric Ceiling Panel Systems. These systems are laminated conductive coatings, printed circuits, or etched elements nailed to the bottom of ceiling joists and covered by 13 mm gypsum board. Heat flux is limited to 190 W/m2. In some cases, the heating element can be cut to fit available space. Manufacturers’ instructions specify how to connect the system to the electric supply. Appropriate codes should be followed when placing partitions, lights, and air grilles adjacent to or near electric panels. Electric Cables Embedded in Ceilings. Electric heating cables for embedded or laminated ceiling panels are factory-assembled units furnished in standard lengths of 25 to 550 m. These cable lengths cannot be altered in the field. The cable assemblies are normally rated at 9 W/m and are supplied in capacities from 200 to 5000 W in roughly 200 W increments. Standard cable assemblies are available for 120, 208, and 240 V. Each cable unit is supplied with 2 m nonheating leads for connection at the thermostat or junction box. Electric cables for panel heating have electrically insulated sleeves resistant to medium temperature, water absorption, aging effects, and chemical action with plaster, cement, or ceiling lath material. This insulation is normally made of polyvinyl chloride (PVC), which may have a nylon jacket. The thickness of the insulation layer is usually about 3 mm. For plastered ceiling panels, the heating cable may be stapled to gypsum board, plaster lath, or similar fire-resistant materials with

Fig. 24

rust-resistant staples (Figure 24). With metal lath or other conducting surfaces, a coat of plaster (brown or scratch coat) is applied to completely cover the metal lath or conducting surface before the cable is attached. After the lath is fastened on and the first plaster coat is applied, each cable is tested for continuity of circuit and for insulation resistance of at least 100 k measured to ground. The entire ceiling surface is finished with a cover layer of thermally noninsulating sand plaster about 13 to 19 mm thick, or other approved noninsulating material applied according to manufacturer’s specifications. The plaster is applied parallel to the heating cable rather than across the runs. While new plaster is drying, the

Fig. 23 Electric Heating Panels

Electric Heating for Wet Plaster Ceiling

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2020 ASHRAE Handbook—HVAC Systems and Equipment (SI) Table 4 Characteristics of Typical Electric Panels

Resistor material Relative heat intensity Resistor temperature Envelope temperature (in use) Thermal-radiation-generating ratio a Response time (heat-up) Luminosity (visible light) Thermal shock resistance Vibration resistance Impact resistance Resistance to drafts or windb Mounting position Envelope material Color blindness Flexibility Life expectancy

Graphite or nichrome wire Low, 540 to 1350 W/m2 80 to 180°C 70 to 150°C 0.7 to 0.8 240 to 600 s None Excellent Excellent Excellent Poor Any Steel alloy or aluminum Very good Good—wide range of heat flux, length, and voltage practical Over 10 000 h

a Ratio

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b May

of radiation heat flux to input power density (elements only). be shielded from wind effects by louvers, deep-drawn fixtures, or both.

system should not be energized, and the range and rate of temperature change should be kept low by other heat sources or by ventilation until the plaster is thoroughly cured. Vermiculite or other insulating plaster causes cables to overheat and is contrary to code provisions. For laminated drywall ceiling panels, the heating cable is placed between two layers of gypsum board, plasterboard, or other thermally noninsulating fire-resistant ceiling lath. The cable is stapled directly to the first (or upper) lath, and the two layers are held apart by the thickness of the heating cable. It is essential that the space between the two layers of lath be completely filled with a noninsulating plaster or similar material. This fill holds the cable firmly in place and improves heat transfer between the cable and the finished ceiling. Failure to fill the space between the two layers of plasterboard completely may allow the cable to overheat in the resulting voids and may cause cable failure. The plaster fill should be applied according to manufacturer’s specifications. Electric heating cables are ordinarily installed with a 150 mm nonheating border around the periphery of the ceiling. A 200 mm clearance must be provided between heating cables and the edges of the outlet or junction boxes used for surface-mounted lighting fixtures. A 50 mm clearance must be provided from recessed lighting fixtures, trim, and ventilating or other openings in the ceiling. Heating cables or panels must be installed only in ceiling areas that are not covered by partitions, cabinets, or other obstructions. However, it is permissible for a single run of isolated embedded cable to pass over a partition. The National Electrical Code® (NFPA Standard 70) requires that all general electrical power and lighting wiring be run above the thermal insulation or at least 50 mm above the heated ceiling surface, or that the wiring be derated. In drywall ceilings, the heating cable is always installed with the cable runs parallel to the joist. A 65 mm clearance between adjacent cable runs must be left centered under each joist for nailing. Cable runs that cross over the joist must be kept to a minimum. Where possible, these crossings should be in a straight line at one end of the indoor space For cable having a heat flux of 9 W/m, the minimum permissible spacing is 40 mm between adjacent runs. Some manufacturers recommend a minimum spacing of 50 mm for drywall construction. The spacing between adjacent runs of heating cable can be determined using the following equation: M = 1000An C (28) where M = cable spacing, mm

An = net panel heated area, m2 C = length of cable, m

Net panel area An in Equation (28) is the net ceiling area available after deducting the area covered by the nonheating border, lighting fixtures, cabinets, and other ceiling obstructions. For simplicity, Equation (28) contains a slight safety factor, and small lighting fixtures are usually ignored in determining net ceiling area. Electrical resistance of the electric cable must be adjusted according to its temperature at design conditions (Ritter and Kilkis 1998):  1 +  e  t d – 20   R = R ---------------------------------------- 1 +  o  t d – 20  

(29)

where R = electrical resistance of electric cable at standard temperature (20°C), /m e = thermal coefficient for material resistivity, °C 1 o = thermal expansion coefficient, °C 1 td = surface temperature of electric cable at operating conditions [see Equation (23)], °C

The 65 mm clearance required under each joist for nailing in drywall applications occupies one-fourth of the ceiling area if the joists are 400 mm on centers. Therefore, for drywall construction, the net area An must be multiplied by 0.75. Many installations have a spacing of 40 mm for the first 600 mm from the cold wall. Remaining cable is then spread over the balance of the ceiling.

Electric Wall Heating Cable embedded in walls similar to ceiling construction is used in Europe. Because of possible damage from nails driven for hanging pictures or from building alterations, most U.S. codes prohibit such panels. Some of the prefabricated panels described in the preceding section are also used for wall panel heating.

Electric Floor Heating Electric heating cable assemblies such as those used for ceiling panels are sometimes used for concrete floor heating systems. Because the possibility of cable damage during installation is greater for concrete floor slabs than for ceiling panels, these assemblies must be carefully installed. After the cable has been placed, all unnecessary traffic should be eliminated until the concrete layer has been placed and hardened. Preformed mats are sometimes used for electric floor slab heating systems. These mats usually consist of PVC-insulated heating cable woven into or attached to metallic or glass fiber mesh. Such mats are available as prefabricated assemblies in many sizes from 0.2 to 9 m2 and with heat fluxes from 160 to 270 W/m2. When mats are used with a thermally treated cavity beneath the floor, a heat storage system is provided, which may be controlled for off-peak heating. Mineral-insulated (MI) heating cable is another effective method of slab heating. MI cable is a small-diameter, highly durable, flexible heating cable composed of solid electric-resistance heating wire or wires surrounded by tightly compressed magnesium oxide electrical insulation and enclosed by a metal sheath. MI cable is available in stock assemblies in a variety of standard voltages, heat fluxes (power densities), and lengths. A cable assembly consists of the specified length of heating cable, waterproof hot-cold junctions, 2 m cold sections, UL-approved end fittings, and connection leads. Several standard MI cable constructions are available, such as single conductor, twin conductor, and double cable. Custom-designed MI heating cable assemblies can be ordered for specific installations. Other outer-sleeve materials that are sometimes specified for electric floor heating cable include (1) silicone rubber, (2) lead, and (3) tetrafluoroethylene (Teflon®).

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Radiant Heating and Cooling

6.17

4.

DESIGN PROCEDURE

Panel design requires determining panel area, type, and arrangement as well as the supply water temperature and flow rate. Panel performance is directly related to indoor space conditions. Air-side design also must be established. Heating and cooling loads may be calculated by procedures covered in Chapters 17 and 18 of the 2017 ASHRAE Handbook—Fundamentals. For cooling load calculations, the procedure based on the heat balance (HB) method is recommended, because other simplified methods developed for all-air systems (such as the radiant time series [RTS] method) are not as appropriate for radiant-system applications (Feng et al. 2013b, 2014). The procedure is as follows:

Sensible Cooling

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Fig. 25 Electric Heating Cable in Concrete Slab For a given floor heating cable assembly, the required cable spacing is determined from Equation (24). In general, cable heat flux and spacing should be such that floor panel heat flux is not greater than 160 W/m. Check the latest edition of the National Electrical Code® (NFPA Standard 70) and other applicable codes to obtain information on maximum panel heat flux and other required criteria and parameters. Floor Heating Cable Installation. When PVC-jacketed electric heating cable is used for floor heating, the concrete slab is laid in two pours. The first pour should be at least 75 mm thick and, where practical, should be insulating concrete to reduce downward heat loss. For a proper bond between the layers, the finish slab should be placed within 24 h of the first pour, with a bonding grout applied. The finish layer should be between 40 and 50 mm thick. This top layer must not be insulating concrete. At least 25 mm of perimeter insulation should be installed as shown in Figure 25. The cable is installed on top of the first pour of concrete no closer than 50 mm from adjoining walls and partitions. Methods of fastening the cable to the concrete include the following: • Staple the cable to wood nailing strips fixed in the surface of the rough slab. Daubs of cement, plaster of paris, or tape maintain the predetermined cable spacing. • In light or uncured concrete, staple the cable directly to the slab using hand-operated or powered stapling machines. • Nail special anchor devices to the first slab to hold the cable in position while the top layer is being poured. Preformed mats can be embedded in the concrete in a continuous pour. The mats are positioned in the area between expansion and/or construction joints and electrically connected to a junction box. The slab is poured to within 40 to 50 mm of the finished level. The surface is rough-screeded and the mats placed in position. The final cap is applied immediately. Because the first pour has not set, there is no adhesion problem between the first and second pours, and a monolithic slab results. A variety of contours can be developed by using heater wire attached to glass fiber mats. Allow for circumvention of obstructions in the slab. MI electric heating cable can be installed in concrete slab using either one or two pours. For single-pour applications, the cable is fastened to the top of the reinforcing steel before the pour is started. For two-layer applications, the cable is laid on top of the bottom structural slab and embedded in the finish layer. Proper spacing between adjacent cable runs is maintained by using prepunched copper spacer strips nailed to the lower slab.

1. Determine indoor design dry-bulb temperature, relative humidity, and dew point. 2. Calculate sensible and latent heat gains. 3. Establish minimum supply air quantity for ventilation. 4. Calculate latent cooling available from supply air. 5. Calculate sensible cooling available from supply air. 6. Determine remaining sensible cooling load to be satisfied by panel system. 7. Determine minimum permissible effective cooling panel surface temperature that will not lead to surface condensation at design conditions. 8. Determine AUST. 9. Determine necessary panel area for remaining sensible cooling. 10. Determine average panel cooling water (brine) temperature for given tube spacing, or determine necessary tube spacing if average panel cooling water (brine) temperature is known.

Sensible Heating 1. Designate indoor design dry-bulb temperature for panel heating. 2. Calculate room heat loss. 3. Determine AUST. Use Equation (14) to find surface temperatures of exterior walls and exposed floors and ceilings. Interior walls are assumed to have surface temperatures equal to indoor air temperature. 4. Calculate required effective surface temperature of panel. Refer to Figures 9 and 10 if AUST does not greatly differ from indoor air temperature. Otherwise, use Equations (5), (9), (10), (11), and (12) or refer to Figures 1 and 3. 5. Determine panel area. Refer to Figures 9 and 10 if AUST does not vary greatly from indoor air temperature. 6. Refer to manufacturers’ data for panel surface temperatures higher than those given in Figures 9 and 10. For panels with several covers, average temperature of each cover and effective panel surface temperature must be calculated and compared to temperature-withstanding capacity for continuous operation of every cover material. For this purpose, Equation (23) may be used: add thermal resistances of all cover layers between panel surface and cover layer in question, then multiply by q and add to first two terms in Equation (23). This is tha, approximated temperature of the particular layer at design. 7. Determine tube spacing for a given average water temperature or select electric cable properties or electric mat size. 8. In a hydronic panel system, if tube spacing is known, determine required average water temperature. 9. Design panel arrangement.

Other Steps Common for Sensible Heating and Cooling 1. Check thermal comfort requirements in the following steps (see Chapter 9 of the 2017 ASHRAE Handbook—Fundamentals and NRB [1981]).

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2020 ASHRAE Handbook—HVAC Systems and Equipment (SI)

(a) Determine occupant’s clothing insulation value and metabolic rate (see Tables 4, 7, 8, and 9 in Chapter 9 of the 2017 ASHRAE Handbook—Fundamentals). (b) Determine optimum operative temperature at coldest point in conditioned space (see the Comfort Equations for Radiant Heating section in Chapter 9 of the 2017 ASHRAE Handbook—Fundamentals. Note that the same equations may be adopted for panel cooling). (c) Determine MRT at the coldest point in the conditioned space (see Chapter 16 and Fanger [1972]). Note: If indoor air velocity is less than 0.4 m/s and MRT is less than 50°C, operative temperature may be approximated as the average of MRT and ta. (d) From the definition of operative temperature, establish optimum indoor design air temperature at coldest point in the room. If optimum indoor design air temperature varies greatly from designated design temperature, designate a new temperature. (e) Determine MRT at hottest point in conditioned space. (f) Calculate operative temperature at hottest point in conditioned space. (g) Compare operative temperatures at hottest and coldest points. For light activity and normal clothing, the acceptable operative temperature range is 20 to 24°C (see NRB [1981] and ANSI/ASHRAE Standard 55-2013). If the range is not acceptable, the panel system must be modified. (h) Calculate radiant temperature asymmetry (NRB 1981). Acceptable ranges are less than 5 K for warm ceilings, 15 K for cool ceilings, 10 K for cool walls, and 27 K for warm walls at 10% local discomfort dissatisfaction (ANSI/ASHRAE Standard 55-2013). 2. Determine water flow rate and pressure drop. Refer to manufacturers’ guides for specific products, or use the guidelines in Chapter 22 of the 2017 ASHRAE Handbook—Fundamentals. Chapter 13 of this volume also has information on hydronic heating and cooling systems. 3. The supply and return manifolds must be carefully designed. If there are circuits of unequal coil lengths, the following equation may be used (Hansen 1985; Kilkis 1998) for a circuit i connected to a manifold with n circuits: Qi = (Leq Li)1 rQtot

where

 Li n

Leq =

–1  r

(30)

–r

,m

i=1

Qi Qtot Li r

= = = =

flow rate in circuit i, L/s total flow rate in supply manifold, L/s coil length of hydronic circuit i, m 1.75 for hydronic panels (Siegenthaler 1995)

Application, design, and installation of panel systems have certain requirements and techniques: • As with any hydronic system, look closely at the piping system design. Piping should be designed to ensure that water of the proper temperature and in sufficient quantity is available to every grid or coil at all times. Proper piping and system design should minimize the detrimental effects of oxygen on the system. Reverse-return systems should be considered to minimize balancing problems. • Hydronic panels can be used with two- and four-pipe distribution systems. Figure 26 shows a typical system arrangement. It is common to design for a 10 K temperature drop for heating across a given grid and a 3 K rise for cooling, but larger temperature differentials may be used, if applicable.

Fig. 26 Primary/Secondary Water Distribution System with Mixing Control • Individual panels can be connected for parallel flow using headers, or for sinuous or serpentine flow. To avoid flow irregularities in a header-type grid, the water channel or lateral length should be greater than the header length. If the laterals in a header grid are forced to run in a short direction, using a combination seriesparallel arrangement can solve this problem. Serpentine flow ensures a more even panel surface temperature throughout the heating or cooling zone. • Noise from entrained air, high-velocity or high-pressure-drop devices, or pump and pipe vibrations must be avoided. Water velocities should be high enough to prevent separated air from accumulating and causing air binding. Where possible, avoid automatic air venting devices over ceilings of occupied spaces. • Design piping systems to accept thermal expansion adequately. Do not allow forces from piping expansion to be transmitted to panels. Thermal expansion of ceiling panels must be considered. • In hydronic systems, thermoplastic, rubber tubes, steel, or copper pipes are used widely in ceiling, wall, or floor panel construction. Where coils are embedded in concrete or plaster, no threaded joints should be used for either pipe coils or mains. Steel pipe should be the all-welded type. Copper tubing should be a softdrawn coil. Fittings and connections should be minimized. Bending should be used to change direction. Solder-joint fittings for copper tube should be used with a medium-temperature solder of 95% tin, 5% antimony, or capillary brazing alloys. All piping should be subjected to a hydrostatic test of at least three times the working pressure. Maintain adequate pressure in embedded piping while pouring concrete. • Placing the thermostat on a wall where it can observe both the outdoor exposed wall and the warm panel should be considered. The normal thermostat cover reacts to the warm panel, and thermal radiation from the panel to the cover tends to alter the control point so that the thermostat controls 1 to 2 K lower when the outdoor temperature is a minimum and the panel temperature is a maximum. Experience indicates that panel-heated spaces are more comfortable under these conditions than when the thermostat is located on a back wall. • If throttling valve control is used, either the end of the main should have a fixed bypass, or the last one or two rooms on the mains should have a bypass valve to maintain water flow in the

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Radiant Heating and Cooling

6.19 • Selection of summer design indoor dew point below 10°C generally is not economical. • The most frequently applied method of dehumidification uses cooling coils. If the main cooling coil is six rows or more, the dew point of leaving air will approach the leaving water temperature. The cooling water leaving the dehumidifier can then be used for the panel water circuit. • Several chemical dehumidification methods are available to control latent and sensible loads separately. In one application, cooling tower water is used to remove heat from the chemical drying process, and additional sensible cooling is necessary to cool the dehumidified air to the required system supply air temperature. • When chemical dehumidification is used, hygroscopic chemical dew-point controllers are required at the central apparatus and at various zones to monitor dehumidification. • When cooled ceiling panels are used with a variable-air-volume (VAV) system, the air supply rate should be near maximum volume to ensure adequate dehumidification before the cooling ceiling panels are activated.

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Other factors to consider when using panel systems are

Fig. 27 Split Panel Piping Arrangement for Two-Pipe and Four-Pipe Systems main. Thus, when a throttling valve modulates, there will be a rapid response. • When selecting heating design temperatures for a ceiling panel surface, the design parameters are as follows: • Excessively high temperatures over the occupied zone cause the occupant to experience a “hot head effect.” • Temperatures that are too low can result in an oversized, uneconomical panel and a feeling of coolness at the outside wall. • Locate ceiling panels adjacent to perimeter walls and/or areas of maximum load. • With normal ceiling heights of 2.4 to 2.8 m, panels less than 1 m wide at the outside wall can be designed for 113°C surface temperature. If panels extend beyond 1 m into the indoor space, the panel surface temperature should be limited to the values given in Figure 15. The surface temperature of concrete or plaster panels is limited by construction. • Floor panels are limited to surface temperatures of less than 29°C in occupied spaces for comfort reasons. Subfloor temperature may be limited to the maximum exposure temperature specified by the floor cover manufacturer. • When the panel chilled-water system is started, the circulating water temperature should be maintained at indoor air temperature until the air system is completely balanced, the dehumidification equipment is operating properly, and building relative humidity is at design value. When the panel area for cooling is greater than the area required for heating, a two-panel arrangement (Figure 27) can be used. Panel HC (heating and cooling) is supplied with hot or chilled water yearround. When chilled water is used, the controls activate panel CO (cooling only) mode, and both panels are used for cooling. • To prevent condensation on the cooling panels, the panel water supply temperature should be maintained at least 0.5 K above the indoor design dew-point temperature. This minimum difference is recommended to allow for the normal drift of temperature controls for water and air systems, and also to provide a factor of safety for temporary increase in indoor relative humidity.

• Evaluate the panel system to take full advantage in optimizing the physical building design. • Select recessed lighting fixtures, air diffusers, hung ceilings, and other ceiling devices to provide the maximum ceiling area possible for use as panels. • The air-side design must be able to maintain relative humidity at or below design conditions at all times to eliminate any possibility of condensation on the panels. This becomes more critical if indoor space dry- and wet-bulb temperatures are allowed to drift for energy conservation, or if duty cycling of the fans is used. • Do not place cooling panels in or adjacent to high-humidity areas, such as near a lobby entrance, or a kitchen. • Anticipate thermal expansion of the ceiling and other devices in or adjacent to the ceiling. • The design of operable windows should discourage unauthorized opening.

4.1

CONTROLS

Automatic controls for panel heating may differ from those for convective heating because of the thermal inertial characteristics of the panel and the increase in mean radiant temperature in the space under increasing loads. However, low-mass systems using thin metal panels or thin underlay with low thermal heat capacity may be successfully controlled with conventional control technology using indoor sensors. Many of the control principles for hotwater heating systems described in Chapters 13 and 15 also apply to panel heating. Because panels do not depend on air-side equipment to distribute energy, many control methods have been used successfully; however, a control interface between heating and cooling should be installed to prevent simultaneous heating and cooling. High-thermal-mass panels such as concrete slabs require a control approach different from that for low-mass panels. Because of thermal inertia, significant time is required to bring such massive panels from one operating point to another, say from vacation setback to standard operating conditions. This will result in long periods of discomfort during the delay, then possibly periods of uncomfortable and wasteful overshoot. Careful economic analysis may reveal that nighttime setback is not warranted. Once a slab is at operating conditions, the control strategy should be to supply the slab with heat at the rate that heat is being lost from the space (MacCluer et al. 1989). For hydronic slabs with constant circulator flow rates, this means modulating the tempera-

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2020 ASHRAE Handbook—HVAC Systems and Equipment (SI)

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ture difference between the outgoing and returning water; this is accomplished with mixing valves, fuel modulation, or, for constant thermal power sources, pulse-width modulation (on/off control). For some applications, more advanced control techniques, such as model predictive control, may be used to optimize the balance of thermal comfort and energy consumption (Feng et al. 2015; Oldewurtel et al. 2012). Slabs with embedded electric cables can be controlled by pulse-width modulators such as the common round thermostat with anticipator or its solid-state equivalent. Outdoor reset control, another widely accepted approach, measures the outdoor air temperature, calculates the supply water temperature required for steady operation, and operates a mixing valve or boiler to achieve that supply water temperature. If the heating load of the controlled space is primarily a function of outdoor air temperature, or indoor temperature measurement of the controlled space is impractical, then outdoor reset control alone is an acceptable control strategy. When other factors such as solar or internal gains are also significant, indoor temperature feedback should be added to the outdoor reset. In all panel applications, precautions must be taken to prevent excessive temperatures. A manual boiler bypass or other means of reducing the water temperature may be necessary to prevent new panels from drying out too rapidly.

Sensible Cooling Controls The average water temperature in the hydronic circuit of radiant systems can be controlled either by mixing, by heat exchange, or by using the water leaving the dehumidifier. Other considerations are listed in the section on General Design Considerations. It is imperative to dry out the building space before starting the panel water system, particularly after extended down periods such as weekends. Such delayed starting action should be automated. Radiant cooling systems require the following basic areas of temperature control: (1) exterior zones; (2) areas under exposed roofs, to compensate for transmission and solar loads; and (3) each typical interior zone, to compensate for internal loads. For optimum results, each area of the building with significantly different heating or cooling loads should be independently controlled. Temperature control of the indoor air and panel water supply should not be a function of the outdoor weather. Normal thermostat drift is usually adequate compensation for the slightly lower temperatures desirable during winter weather. This drift should result in an indoor air temperature change of no more than 0.8 K. Control of interior zones is best accomplished by devices that reflect the actual presence of the internal load elements. Frequently, time clocks and current-sensing devices are used on lighting feeders.

Heating Slab Controls In comfort heating, the effective surface temperature of a heated floor slab is held to a maximum of 27 to 29°C. As a result, when the heated slab is the primary heating system, thermostatic controls sensing air temperature should not be used to control the heated slab temperature; instead, the heating system should be wired in series with a slab-sensing thermostat. The remote sensing thermostat in the slab acts as a limit switch to control maximum surface temperatures allowed on the slab. The ambient sensing thermostat controls the comfort level. For supplementary slab heating, as in kindergarten floors, a remote sensing thermostat in the slab is commonly used to tune in the desired comfort level. Indoor-outdoor thermostats are used to vary the floor temperature inversely with the outdoor temperature. If the building heat loss is calculated for an outdoor temperature between 21 to –19°C, and the effective floor temperature range is maintained between 21 to 29°C with a remote sensing thermostat, the ratio of change in outdoor temperature to change in the heated slab temperature is 40:8, or 5:1. This means that a 5 K drop in outdoor temperature requires a 1 K increase in the slab temperature.

An ambient sensing thermostat is used to vary the ratio between outdoor and slab temperatures. A time clock is used to control each heating zone if off-peak slab heating is desirable. In heating systems, lowering nighttime air and surface temperatures can produce less satisfactory results in thermally massive radiant systems such as embedded tubing in concrete floors. These systems cannot respond to a quick increase or decrease in heating demand within the relatively short time required, resulting in a very slow reduction of the space temperature at night and a correspondingly slow pickup in the morning. Light panels, such as plaster or metal ceilings and walls, may respond to changes in demand quickly enough for satisfactory results from lowered nighttime air and panel temperatures. Berglund et al. (1982) demonstrated the speed of response on a metal ceiling panel to be comparable to that of convection systems. However, in cooling operation, thermally massive embedded systems can offer cost- and energy-saving opportunities through nighttime precooling when electricity costs are typically lower, as are outdoor dry- and wet-bulb temperatures.

REFERENCES ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal.ashrae .org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org/bookstore. ASHAE. 1956. Thermal design of warm water ceiling panels. ASHAE Transactions 62:71. ASHAE. 1957. Thermal design of warm water concrete floor panels. ASHAE Transactions 63:239. ASHRAE. 1976. Energy calculations I—Procedures for determining heating and cooling loads for computerizing energy calculations. ASHRAE. 2013. Thermal environmental conditions for human occupancy. ANSI/ASHRAE Standard 55-2013. ASHRAE. 2013. Method of testing for rating ceiling panels for sensible heating and cooling. ANSI/ASHRAE Standard 138-2013. Berglund, L., R. Rascati, and M.L. Markel. 1982. Radiant heating and control for comfort during transient conditions. ASHRAE Transactions (88): 765-775. Buckley, N.A. 1989. Application of radiant heating saves energy. ASHRAE Journal 31(9):17-26. CEN. 2011. Water based surface embedded heating and cooling systems— Part 4: Installation. Standard EN 1264-4:2011. European Committee for Standardization, Brussels. CEN. 2004. Ventilation for buildings—Chilled ceilings—Testing and rating. Standard EN 14240:2004. European Committee for Standardization, Brussels. CEN. 2013. Free hanging heating and cooling surfaces for water with a temperature below 120°C. Standard EN 14037:2013. European Committee for Standardization, Brussels. DOE. 2002. Energy consumption characteristics of commercial building HVAC systems, vol. III: Energy savings potential. TIAX ref. no. 6837000, for Building Technologies Program, U.S. Department of Energy. Contract no. DE-AC01-96CE23798. Washington, D.C. Fanger, P.O. 1972. Thermal comfort analysis and application in environmental engineering. McGraw-Hill, New York. Feng, J., S. Schiavon, and F. Bauman. 2013a. Impact of solar heat gain on radiant floor cooling system design. Proceedings of the 11th REHVA World Congress-CLIMA 2013, Prague, Czech Republic. Feng, J., S. Schiavon, and F. Bauman. 2013b. Cooling load differences between radiant and air systems. Energy and Buildings 65:310-321. Feng, J., F. Bauman, and S. Schiavon. 2014. Experimental comparison of zone cooling load between radiant and air systems. Energy and Buildings 84:152-159. dx.doi.org/10.1016/j.enbuild.2014.07.080. Feng, J., F. Chuang, F. Borrelli, and F. Bauman. 2015. Model predictive control of radiant slab systems with evaporative cooling sources. Energy and Buildings 87:199-210. Hansen, E.G. 1985. Hydronic system design and operation. McGraw-Hill, New York.

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Radiant Heating and Cooling

6.21

ISO. 2012. Building environment design—Design, dimensioning, installation and control of embedded radiant heating and cooling systems. Standard 11855-2012. International Organization for Standardization, Geneva. ISO. 2014. Building environment design—Design, test methods and control of hydronic radiant heating and cooling panel systems. Standard 185662014. International Organization for Standardization, Geneva. Kalisperis, L.N. 1985. Design patterns for mean radiant temperature prediction. Department of Architectural Engineers, Pennsylvania State University, University Park. Kalisperis, L.N., and L.H. Summers. 1985. MRT33GRAPH—A CAD program for the design evaluation of thermal comfort conditions. Tenth National Passive Solar Conference, Raleigh, NC. Kilkis, B.I. 1998. Equipment oversizing issues with hydronic heating systems. ASHRAE Journal 40(1):25-31. Kollmar, A., and W. Liese. 1957. Die Strahlungsheizung, 4th ed. R. Oldenburg, Munich. Lindstrom, P.C., D. Fisher, and C. Pedersen. 1998. Impact of surface characteristics on radiant panel output. ASHRAE Research Project RP-876, Final Report. Liu, X., X. Xiang, K. Zhao, and Y. Jiang. 2015. Cooling performance comparison of radiant floor system and all-air system with solar radiation. 6th International Building Physics Conference, IBPC 2015. Energy Procedia 78:2322-2327. MacCluer, C.R., M. Miklavcic, and Y. Chait. 1989. The temperature stability of a radiant slab-on-grade. ASHRAE Transactions 95(1):1001-1009. Meierhans R., and B.W. Olesen. 2002. Art museum in Bregenz—Soft HVAC for a strong architecture. ASHRAE Transactions 108(2). Min, T.C., L.F. Schutrum, G.V. Parmelee, and J.D. Vouris. 1956. Natural convection and radiation in a panel heated room. ASHAE Transactions 62: 337. NFPA. 2016. Installation of sprinkler systems. Standard 13-2016. National Fire Protection Association, Quincy, MA. NFPA. 2014. National electrical code®. Standard 70-2014. National Fire Protection Association, Quincy, MA. NRB. 1981. Indoor climate. Technical Report 41. The Nordic Committee on Building Regulations, Stockholm. Oldewurtel, F., A. Parisio, C. N. Jones, D. Gyalistras, M. Gwerder, V. Stauch, B. Lehmann, and M. Morari. 2012. Use of model predictive control and weather forecasts for energy efficient building climate control. Energy and Buildings 45:15-27 Parmelee, G.V., and R.G. Huebscher. 1947. Forced convection, heat transfer from flat surfaces. ASHVE Transactions 53:245. Ritter, T.L., and B.I. Kilkis. 1998. An analytical model for the design of inslab electric heating panels. ASHRAE Transactions 104(1B):1112-1115. Sartain, E.L., and W.S. Harris. 1956. Performance of covered hot water floor panels, part I—Thermal characteristics. ASHAE Transactions 62:55. Scheatzle, D.G. 2003. Establishing a baseline data set for the evaluation of hybrid (radiant/convective) HVAC systems. ASHRAE Research Project RP-1140, Final Report. Schutrum, L.F., and C.M. Humphreys. 1954. Effects of non-uniformity and furnishings on panel heating performance. ASHVE Transactions 60:121. Schutrum, L.F., and J.D. Vouris. 1954. Effects of room size and non-uniformity of panel temperature on panel performance. ASHVE Transactions 60:455. Schutrum, L.F., G.V. Parmelee, and C.M. Humphreys. 1953a. Heat exchangers in a ceiling panel heated room. ASHVE Transactions 59:197. Schutrum, L.F., G.V. Parmelee, and C.M. Humphreys. 1953b. Heat exchangers in a floor panel heated room. ASHVE Transactions 59:495. Siegenthaler, J. 1995. Modern hydronic heating. Delmar Publishers, Boston. Simmonds, P. 2006. Radiant cooled floors—Operation and control dependent upon solar radiation. ASHRAE Transactions 112(1):358-367.

Steinman, M., L.N. Kalisperis, and L.H. Summers. 1989. The MRT-correction method—An improved method for radiant heat exchange. ASHRAE Transactions 95(1):1015-1027. Sugawara, F., M. Nobushisa, and H. Miyazawa. 1996. Comparison of miteallergen and fungal colonies in floor dust in Seoul (Korea) and Koriyama (Japan) dwellings. Journal of Architecture, Planning and Environmental Engineering, Architectural Institute of Japan 48:35-42. TSI. 1994. Fundamentals of design for floor heating systems (in Turkish). Turkish Standard 11261. Turkish Standards Institute, Ankara. Walton, G.N. 1980. A new algorithm for radiant interchange in room loads calculations. ASHRAE Transactions 86(2):190-208. Watson, R.D., K.S. Chapman, and J. DeGreef. 1998. Case study: Seven-system analysis of thermal comfort and energy use for a fast-acting radiant heating system. ASHRAE Transactions 104(1B):1106-1111. Wilkes, G.B., and C.M.F. Peterson. 1938. Radiation and convection from surfaces in various positions. ASHVE Transactions 44:513.

BIBLIOGRAPHY ALI. 2003. Fundamentals of panel heating and cooling, short course. ASHRAE Learning Institute. Babiak, J., B. Olesen, and D. Petras. 2007. REHVA guidebook no. 7: Low temperature heating and high temperature cooling. Federation of European Heating, Ventilation and Air Conditioning Associations, Brussels. BSR/ASHRAE. 2003. Method of test for determining the design and seasonal efficiencies of residential thermal distribution systems. Draft Standard 152P, 2nd Public Review. Buckley, N.A., and T.P. Seel. 1987. Engineering principles support an adjustment factor when sizing gas-fired low-intensity infrared equipment. ASHRAE Transactions 93(1):1179-1191. Chapman, K.S., and P. Zhang. 1995. Radiant heat exchange calculations in radiantly heated and cooled enclosures. ASHRAE Transactions 101(2): 1236-1247. Chapman, K.S., J. Ruler, and R.D. Watson. 2000. Impact of heating systems and wall surface temperatures on room operative temperature fields. ASHRAE Transactions 106(1). Hanibuchi, H., and S. Hokoi. 2000. Simplified method of estimating efficiency of radiant and convective heating systems. ASHRAE Transactions 106(1). Hogan, R.E., Jr., and B. Blackwell. 1986. Comparison of numerical model with ASHRAE designed procedure for warm-water concrete floor-heating panels. ASHRAE Transactions 92(1B):589-601. Jones, B.W., and K.S. Chapman. 1994. Simplified method to factor mean radiant temperature (MRT) into building and HVAC system design. ASHRAE Research Project RP-657, Final Report. Kilkis, B.I. 1993. Computer-aided design and analysis of radiant floor heating systems. Paper no. 80. Proceedings of Clima 2000, London (Nov. 1-3). Kilkis, B.I. 1993. Radiant ceiling cooling with solar energy: Fundamentals, modeling, and a case design. ASHRAE Transactions 99(2):521-533. Kilkis, B.I., S.S. Sager, and M. Uludag. 1994. A simplified model for radiant heating and cooling panels. Simulation Practice and Theory Journal 2:61-76. Kilkis, B.I., A.S.R. Suntur, and M. Sapci. 1995. Hybrid HVAC systems. ASHRAE Journal 37(12):23-28. Nall, D. 2013. Thermally active floors. ASHRAE Journal 36(1-3). Ramadan, H.B. 1994. Analysis of an underground electric heating system with short-term energy storage. ASHRAE Transactions 100(2):3-13. Sprecher, P., B. Gasser, O. Böck, and P. Kofoed. 1995. Control strategy for cooled ceiling panels. ASHRAE Transactions 101(2). Watson, R.D., and K.S. Chapman. 2002. Radiant heating and cooling handbook. McGraw-Hill, New York. For additional literature on high-temperature radiant heating, see the Bibliography in Chapter 16.

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COMBINED HEAT AND POWER SYSTEMS Terminology ............................................................................... 7.2 CHP System Concepts ............................................................... 7.3 Performance Parameters ........................................................... 7.5 Fuel-to-Power Components ....................................................... 7.9 Thermal-to-Power Components............................................... 7.24

Thermal-to-Thermal Components ............................................ Electrical Generators and Components................................... System Design .......................................................................... Codes and Installation ............................................................. Economic Evaluation ...............................................................

C

On-site CHP systems are small compared to typical central station power plants. DG systems are inherently modular, which makes distributed power highly flexible and able to provide power where and when it is needed. DG and CHP systems can offer significant benefits, depending on location, rate structures, and application. Typical advantages of an on-site CHP plant include improved power reliability and quality, reduced energy costs, increased predictability of energy costs, lowered financial risk, use of renewable energy sources, reduced emissions, and faster response to new power demands because capacity additions can be made more quickly. CHP system efficiency is not as simple as adding outputs and dividing by fuel inputs. Nevertheless, using what is normally waste exhaust heat yields overall efficiencies (O) of 50 to 70% or more (for a definition of overall efficiency, see the section on Performance Parameters). CHP can operate on a topping, bottoming, or combined cycle. Figure 1 shows an example of topping and bottoming configurations. In a topping cycle, energy from the fuel generates shaft or electric power first, and thermal energy from the exiting stream is recovered for other applications such as process heat for cooling or heating systems. In a bottoming cycle, shaft or electric power is generated last from thermal energy left over after higher-level thermal energy has been used to satisfy thermal loads. A typical topping cycle recovers heat from operation of a prime mover and uses this thermal energy for the process (cooling and/or heating). A bottoming cycle recovers heat from the process to generate power. A combined cycle uses thermal output from a prime mover to generate additional shaft power (e.g., combustion turbine exhaust generates steam for a steam turbine generator). Grid-isolated CHP systems, in which electrical power output is used on site to satisfy all site power and thermal requirements, are sometimes referred to as total energy systems. Grid-parallel CHP systems, which are actively tied to the utility grid, can, on a contractual or tariff basis, exchange power with or reduce load on (thus reducing capacity demand) the public utility. This may eliminate or lessen the need for redundant on-site back-up generating capacity and allows operation at maximum thermal efficiency when satisfying

OMBINED heat and power (CHP) is the simultaneous production of electrical or mechanical power and useful thermal energy from a single energy source. By capturing and using the recovered thermal energy from an effluent stream that would otherwise be rejected to the environment, CHP (or cogeneration) systems can operate at utilization efficiencies greater than those achieved when heat and power are produced in separate processes and with potentially separate fuel sources, thus contributing to sustainable building solutions. Recovered thermal energy from fuel used in reciprocating engines, steam or combustion turbines (including microturbines, which are typically less than 500 kW power generation capacity), Stirling engines, or fuel cells can be used in the following applications: • Direct heating: exhaust gases or coolant fluids are used directly for drying processes, to drive an exhaust-heat-driven absorption chiller, to regenerate desiccant materials in a dehumidifier, or to operate a bottoming cycle • Indirect heating: exhaust gases or coolant fluids are used to heat a secondary fluid (e.g., steam or hot water) for devices, to generate power, or to power various thermally activated technologies • Latent heat: extracting the latent heat of condensation from a recovered flow of steam when the load served allows condensation (e.g., a steam-to-water exchanger) instead of rejecting the latent heat to a cooling tower (e.g., a full condensing turbine with a cooling tower) There are many potential applications, including base-load power, peaking power where on-site power generation (distributed generation) is used to reduce the demand or high on-peak energy charges imposed by the electric energy supplier, back-up power, remote power, power quality, and CHP, providing both electricity and thermal needs to the site. Usually, customers own the smallscale, on-site power generators, but third parties may own and operate the equipment. Table 1 provides an overview of typical applications, technologies and uses of distributed generation (DG) and CHP systems. The preparation of this chapter is assigned to TC 1.10, Cogeneration Systems.

7.32 7.40 7.41 7.47 7.48

Table 1 Applications and Markets for DG/CHP Systems DG Technologies (in electrical power output capacity)

Standby Power

Reciprocating engines: 93°C) or low-pressure steam [5275

57 to 78

$3700

5275 to 14 000 8790 to >19 340

71 to 100 114 to 128 185 to 227

$4400 to $5000 $3700 to $4200 $4800 to $5500

350 to >10 550 350 to 14 000 11 430

128 to 185 — 114 to 569

$4400 + engine maintenance $3500+ depending on size $4800 to $5500

(COP 0.55 to 0.70)

11 430

128 to 284

$4800 to $5500

(COP 0.60 to 0.75)

11 430

128 to 227

$4800 to $5500

*Maintenance costs courtesy of Johnson Controls, Inc./York International. Typical annual activities include changing oil filter, oil filter analysis and motor checks. Costs do not include cleaning tubes, eddy current testing, or complete oil or refrigerant replacement. Approximate pricing for those additional items follows: • Cleaning evaporator or condenser tubes as required, use $1200.

in the system. Traditional systems use a 7 K t, resulting in a flow rate of 2.4 L/min per kilowatt (40 g/kJ) of refrigeration. Because of the cost of the distribution system piping and pumping energy, modern chilled-water systems operate at lower supply water temperatures and/or higher return water temperatures to allow a larger t to be achieved, thereby reducing chilled-water flow per kilowatt of capacity. For systems involving stratified chilled-water storage, a practical lower limit is 4°C because of water density considerations; however, chemical additives can suppress this temperature below – 2°C. For ice storage systems, supply water temperatures as low as 1°C have been used. Multiple air-conditioning loads connected to a central chilledwater system provide economic advantages and energy conservation opportunities over a decentralized system approach. In addition, central plants afford the opportunity to use refrigerants such as ammonia that may be impractical for use in individual buildings. The size of air-conditioning loads served, diversity among the loads, and their distance from the chilling plant are principal factors in determining the feasibility of large central plants. The distribution system pipe capacity is directly proportional to the operating temperature difference between the supply and return lines, and it benefits additionally from increased diversity in the connected loads. For extremely large district systems, several plants may be required to meet the loads economically and provide redundancy. In some areas, plants over 70 000 kW are common for systems exceeding a total connected load of 175 000 kW. Another reason for multiple plants is that single plants over 100 000 kW can require large piping headers (over 1200 mm in diameter) in the plant, as well as large distribution headers in streets already congested with other utilities, making piping layout problematic. Multiple plants use smaller distribution piping and thus increase system constructability, redundancy, and reliability. An economic evaluation of piping and pumping costs versus chiller power requirements can establish the most suitable supply and return water temperatures. When sizing piping and calculating

• Eddy current testing for under 1760 kW chiller, use $1700; for larger chillers, $2500. • Complete oil replacement: contact vendor’s service department. For all absorbers, a bromide test is conducted twice per year. Costs do not include chemicals.

pumping cost, the heat load on the chiller generated by the frictional heating of the flowing fluid should be considered because most of the pumping power adds to the system heat load. For high chiller efficiency, it is often more efficient to use isolated auxiliary equipment for special process requirements and to allow the central plant supply water temperature to float up at times of lower load. As with heating plants, optimum chilled-water control may require a combination of temperature modulation and flow modulation. However, the designer must investigate the effects of higher chilled-water supply temperatures on chilled-water secondary system distribution flows and air-side system performance (humidity control) before applying this to individual central water plants.

Thermal Storage Both hot- and chilled-water thermal storage can be implemented for district systems. In North America, the current economic situation primarily results in chilled-water storage applications. Depending on plant design and loading, thermal storage can reduce chiller equipment requirements and lower operating costs. By shifting part of the chilling load, chillers can be sized closer to the average load than the peak load. Shifting some or all of the refrigeration load to off peak reduces the on-peak electrical demand load while using the same (or slightly larger) chiller machine capacity. Because many utilities offer lower rates (and perhaps some rebates) during off-peak periods, operating costs for electrically driven chillers can be substantially reduced. Both ice and chilled-water storage have been applied to districtsized chiller plants. In general, the largest systems (>250 GJ capacity) use chilled-water (CHW) storage and small- to moderate-sized systems use ice storage. Storage capacities in the 120 to 400 GJ range are now common and systems have been installed up to 1600 GJ for district cooling systems. When CHW storage is feasible, be careful not to reduce the chilled-water temperature below 4°C, to allow proper temperature stratification in the thermal energy storage (TES) tank. A TES tank charging temperature lower than 3.9°C

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without low-temperature additives will result in mixing in the tank, loss of tank stratification, and possibly disturbance of the system supply temperature and the storage concept. For these reasons, most chilled-water plants are designed for a 4.4°C supply temperature and a 13.3°C or greater return temperature. Some plants may be designed for lower temperatures by using multiple cascaded heat exchangers in series, such as highrise towers; however, note that each 0.6 K reduction in supply temperature increases chiller-specific energy consumption (i.e., kW/ kW) by approximately 2%. Some plants are designed for a higher return temperature (15.5°C or higher) to increase the t, but this requires a great deal of additional coordination with the design of the building’s HVAC system to ensure the system operates per design intent. In Europe, several cooling systems use naturally occurring underground aquifers (caverns) for storage of chilled water. Selection of the storage configuration (chilled-water steel tank above grade, chilled-water concrete tank below grade, ice direct, ice indirect) is often influenced by space limitations. Depending on the system design temperatures, chilled-water storage requires four to six times the volume of ice storage for the same capacity. For chilled-water storage, the footprint of steel tanks (depending on height) can be less than concrete tanks for the same volume (Andrepont 1995); furthermore, the cost of above-grade tanks is usually less than below-grade tanks. Chapter 51 has additional information on thermal storage; for thermal storage specifically in district cooling and heating, also see Phetteplace et al. (2013a, 2013b), respectively.

Auxiliaries Numerous pieces of auxiliary support equipment related to the boiler and chiller operations are not unique to the production plant of a DHC system and are found in similar installations. Some components of a DHC system deserve special consideration because of their critical nature and potential effect on operations. Although instrumentation can be either electronic or pneumatic, electronic instrumentation systems offer the flexibility of combining control systems with data acquisition systems. This combination brings improved efficiency, better energy management, and reduced operating staff for the central heating and/or cooling plant. For systems involving multiple fuels and/or thermal storage, computer-based controls are indispensable for accurate decisions about boiler and chiller operation. Boiler feedwater treatment has a direct bearing on equipment life. Condensate receivers, filters, polishers, and chemical feed equipment must be accessible for proper management, maintenance, and operation. Depending on the temperature, pressure, and quality of the heating medium, water treatment may require softeners, alkalizers, and/or demineralizers for systems operating at high temperatures and pressures. Equipment and layout of a central heating and cooling plant should reflect what is required for proper plant operation and maintenance. The plant should have an adequate service area for equipment and a sufficient number of electrical power outlets and floor drains. Equipment should be placed on housekeeping pads to protect the bases from spills or leaks. Figure 4 presents a typical layout for a large hot-water/chilled-water plant. Notice that the layout provides space for future expansion as well as storage of spare parts. The control room is typically close to the operating equipment for ease of visual inspection. Other functions that should be considered include adequate space for maintenance, chemical treatment laboratory and chemical storage, and operator conveniences such as locker rooms and lunch rooms. Designers must follow the requirements of ASHRAE Standard 15 for ventilation as well as laying out the equipment room, and should coordinate with the architect regarding tight-sealing doors, minimizing penetrations, etc.

Expansion Tanks and Water Makeup. The expansion tank is usually located in the central plant building. To control pressure, either air or nitrogen is introduced to the air space in the expansion tank. To function properly, the expansion tank must be the single point of the system where no pressure change occurs. Multiple, airfilled tanks may cause erratic and possibly harmful movement of air though the piping and should be avoided. Although diaphragm expansion tanks eliminate air movement, the possibility of hydraulic surge should be considered. On large chilled-water systems, a makeup water pump generally is used to makeup water loss. The pump is typically controlled from level switches on the expansion tank or from a desired pump suction pressure. A conventional water meter on the makeup line can show water loss in a closed system. This meter also provides necessary data for water treatment. The fill valve should be controlled to open or close and not modulate to a very low flow, so that the water meter can detect all makeup. For systems with thermal energy storage (TES) tanks, consider providing a rapid-fill connection from either the domestic water system or an adjacent fire hydrant connection. Use caution if connecting to the fire line, because this water would be untreated and raw, and would require chemical treatment; hence, the chemical treatment system should also be sized for a rapid-fill event. Air Venting. Air in water systems can cause major issues for district systems, because it acts as insulator for heat transfer and increases corrosion in system components (e.g., chillers, pipes, heat exchangers). Trapped air causes air pockets in the CHW system that can accumulate at fittings and prohibit flow, causing the system to become air bound. Sources of air in a closed system include • • • •

Makeup water containing normal amounts of dissolved air Air trapped in the system after initial filling Diffusion Air ingress caused by negative pressure

Several devices are used to eliminate air from CHW systems. Air exists within the system in three forms: (1) stagnant bubbles, (2) entrained in the flow, and (3) dissolved. Large, stagnant bubbles can be removed by manual or automatic venting if enough pressure is available. If bubble size exceeds the air vent passage, a capillary problem occurs and prevents bubble removal. In this case, use either a surface tension breaker or a large-bore vent (13 mm ). Centrifugal separators are ideally suited for continuous gas venting in district plants.

Fig. 4

Layout for Hot-Water/Chilled-Water Plant

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District Heating and Cooling For proper separator function, velocity through the separator should not exceed 0.03 m/s and should allow for water turbulence. Some separators contain a zone to collect dirt and blow it down. Dissolved air (invisible) is difficult to remove through separators or air vents, but can be removed using a vacuum-pressure degasser, in which a fraction of the circulated water is put into a vacuum, allowing separation and removal of dissolved gas. Emission Control. In all cases, the cleanest fuel should be used to minimize emissions and environmental impacts. Emission treatment equipment, including electrostatic precipitators, baghouses, and scrubbers, is required to meet emission standards for coal-fired or solid-waste-fired operations. Proper control is critical to equipment operation, and it should be designed and located for easy access by maintenance personnel. A baghouse gas filter provides good service if gas flow and temperature are properly maintained. Because baghouses are designed for continuous online use, they are less suited for cyclic operation. Cyclical heating and cooling of the fabric significantly reduces the useful life of the bags through acidic condensation. Using an economizer to preheat boiler feedwater and help control flue gas temperature may enhance baghouse operation. Contaminants generated by plant operation and maintenance, such as washdown of floors and equipment, may need to be contained. Local codes and regulations may also require low-NOx burners on gas- or oil-fired boilers or engine generators. Chapter 28 of the 2013 ASHRAE Handbook—Fundamentals and Chapter 30 of this volume have information on air pollution and its control.

2.2

CHILLED-WATER DISTRIBUTION DESIGN CONSIDERATIONS

Water distribution systems are designed for either constant flow (variable return temperature) or variable flow (constant return temperature). The design decision between constant- or variable-volume flow affects the (1) selection and arrangement of the chiller(s), (2) design of the distribution system, and (3) design of the customer connection to the distribution system. Unless very unusual circumstances exist, most systems large enough to be considered in the district category are likely to benefit from variable-flow design. See Chapter 13 for additional information on pumping.

12.11 In constant-flow design, chillers arranged in parallel have decreased entering water temperatures at part load; thus, several machines might have needed to run simultaneously, each at a reduced load, to produce the required chilled-water flow. In this case, chillers in series were better because constant flow could be maintained though the chilled-water plant at all times, with only the chillers required for producing chilled water energized. Legacy constant-flow systems should be analyzed thoroughly when considering multiple chillers in a parallel arrangement, because the auxiliary electric loads of condenser water pumps, tower fans, and central plant circulating pumps are a significant part of the total energy input. Contemporary control systems mitigate the reason for constant flow through the chillers, and variable flow should be used for larger systems.

Variable Flow Variable-flow design can significantly reduce energy use and expand the capacity of the distribution system piping by using diversity. To maintain a high temperature differential at part load, the distribution system flow rate must track the load imposed on the central plant. Multiple parallel pumps or, more commonly, variablespeed pumps can reduce flow and pressure, and lower pumping energy at part load. Terminal device controls should be selected to ensure that variable flow objectives are met. Correctly sized terminal unit flow-throttling (two-way) valves, and especially pressureindependent control valves (PICV), provide the continuous high return temperature needed to correlate the system load change to a system flow change. Systems in each building are usually two-pipe, with individual in-building pumping. In some cases, the pressure of the distribution system may cause flow through the in-building system without inbuilding pumping. Distribution system pumps can provide total building system pumping if (1) the distribution system pressure drops are minimal, and (2) the distribution system is relatively short-coupled (1000 m or less). To implement this pumping method, the total flow must be pumped at a pressure sufficient to meet the requirements of the building with the largest pressure differential

Constant Flow In the past, constant chilled-water flow was applied only to smaller systems where simplicity of design and operation were important and where distribution pumping costs were low, before variable- and adjustable-speed drives were available and affordable. (However, for new systems, designers should refer to ASHRAE Standard 90.1, which requires variable-flow pumping.) Chillers were also arranged in series to handle higher system design temperature differentials. Flow rate through a full-load distribution system depended on the type of constant-flow system used. A common technique connected the building and its terminal units across the distribution system. The central plant pump circulated chilled water through air-side terminal units controlled by three-way valves (constant-volume direct primary pumping). Balancing problems could occur in this design when many separate flow circuits were interconnected (Figure 5). Constant-flow distribution was also applied to in-building (secondary or tertiary) circuits with separate pumps. This arrangement isolates the flow balance problem between buildings. In this case, flow through the distribution system could be significantly lower than the sum of the flows needed by the terminal if the in-building system supply temperature were higher than the distribution system supply temperature (Figure 5). The water temperature rise in the distribution system was determined by the connected in-building systems and their controls.

Fig. 5 Constant-Flow Primary Distribution with Secondary Pumping

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2020 ASHRAE Handbook—HVAC Systems and Equipment (SI)

requirement. Consequently, all buildings on the system should have their pressure differentials monitored and transmitted to the central plant, where pump speeds are adjusted to provide adequate pressure to the building with the lowest margin of pressure differential (i.e., hydraulically most remote or the critical node). If the designer has control over the design of each in-building system, this pumping method can be achieved in a reasonable manner. In retrofits where existing buildings under different ownership are connected to a new central plant, coordination is difficult and individual building pumps are more practical. When buildings have separate circulating pumps, use hydraulic isolating piping and pumping design to ensure that two-way control valves are subjected only to the differential pressure established by the in-building pump. Figure 6 shows a connection using in-building pumping with hydraulic isolation from the primary loop. When in-building pumps are used, all series interconnections between the distribution system pump and the in-building pumps should be removed. Without adequate instrumentation and controls, a series connection can cause the distribution system return to operate at a higher pressure than the distribution system supply and disrupt flow to adjacent buildings. Series operation often occurs during improper use of three-way mixing valves in the distribution-tobuilding connection. In very large systems, distributed pumping may be used. Under this approach, the distribution pumps in the central plant are eliminated and relocated to the buildings, so all the electrical energy is borne by the pumps in the user buildings. Where the distribution network piping constitutes a significant pressure loss (systems covering a large area), this design allows the distributed pumps in the buildings to be sized for just the pressure loss imposed at that particular location. Ottmer and Rishel (1993) found that this approach reduces total chilled-water pump power by 20 to 25% in very large systems. It is best applied in new construction where the central plant and distributed building systems can be planned fully and coordinated initially. Note that this system is not the best approach for a system that expects dynamic growth, such as a college and university campus, because adding a large load anywhere in the system adversely affects the pressure drop of each distribution pump. Furthermore, it is a common design approach to oversize distribution piping in distributed pumping systems to keep the distribution pressure drop low, to reduce the pressure drop and size of the building distribution pumps.

Fig. 6 Variable-Flow Primary/Secondary Systems

Usually, a positive pressure must be maintained at the highest point of the system at all times. This height determines the static pressure at which the expansion tank operates. Excessively tall buildings that do not isolate the in-building systems from the distribution system can impose unacceptable static pressure on the distribution system. To prevent excessive operating pressure in distribution systems, heat exchangers have been used to isolate the in-building system from the distribution system to act as pressure interceptors. To ensure reasonable temperature differentials between supply and return temperatures, flow must be controlled on the distribution system side of the heat exchanger. In high-rise buildings, all piping, valves, coils, and other equipment may be required to withstand high pressure. Where system static pressure exceeds safe or economical operating pressure, either the heat exchanger method or pressure-sustaining valves in the return line with check valves in the supply line may be used to minimize pressure. However, the pressure-sustaining/check valve arrangement may overpressurize the entire distribution system if a malfunction of either valve occurs.

Chilled-Water System Design Guidelines Guidelines for plant design and operation include the following: • Variable-speed pumping saves energy and should be considered for distribution system pumping. • Design chilled-water systems to optimize the temperature differential of greater than 7 to 9 K. A 9 to 13 K maximum temperature differential with 7 K minimum temperature differential can be achieved with this design. • Avoid using constant-flow chilled-water systems. • Investigate using chillers arranged in series for larger temperature differentials or increased chiller efficiency. • Larger central chilled-water plants with high distribution system pressure drops can benefit from variable flow in the distribution system. This can be achieved with variable primary, primary/ secondary, or primary/secondary/tertiary pumping systems. • Size the distribution system for a low overall total pressure loss. Short-coupled distribution systems (1000 m or less) can be used for a total pressure loss of 60 to 120 kPa. With this maximum differential between any points in the system, size the distribution pumps to provide the necessary pressure to circulate chilled water through the in-building systems, eliminating the need for inbuilding pumping systems. This decreases the complexity of operating central chilled-water systems. Newer controls on chillers enable all-variable-flow systems. Check with the manufacturer about minimum flows on chiller evaporators to achieve stable operation over all load ranges. • All two-way valves must have proper close-off ratings and a design pressure drop of at least 20% of the maximum design pressure drop for controllability. Commercial quality automatic temperature control valves generally have low shutoff ratings; but industrial valves can achieve higher ratings. Three-way control valves should be avoided except to accommodate minimum pump flow and turndown. See Chapter 43 for more information on control valves. • Although chillers can easily produce colder water, the lower practical limit for chilled-water supply temperatures is 4°C. Temperatures below that should be carefully analyzed for energy usage, although systems with thermal energy storage may operate advantageously at lower temperatures.

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District Heating and Cooling

3.

DISTRIBUTION SYSTEM

3.1

HYDRAULIC CONSIDERATIONS

12.13 fore, it is prudent to lay out distribution systems with as few offsets as possible. Some designers prefer two 45° elbows in lieu of a single 90° elbow, assuming that the pressure drop is lower. As pressure drop calculations show, this is not so.

Objectives of Hydraulic Design Although the distribution of a thermal utility such as hot water encompasses many of the aspects of domestic hot-water distribution, many dissimilarities also exist; thus the design should not be approached in the same manner. Thermal utilities must supply sufficient energy at the appropriate temperature and pressure to meet consumer needs. Within the constraints imposed by the consumer’s end use and equipment, the required thermal energy can be delivered with various combinations of temperature and pressure. Computeraided design methods are available for thermal piping networks (Bloomquist et al. 1999; COWIconsult 1985; Rasmussen and Lund 1987; Reisman 1985). Using these methods allows rapid evaluation of many alternative designs. General steam system design can be found in Chapter 11, as well as in IDHA (1983) and Phetteplace et al. (2013b). For water systems, see Chapter 13, IDEA (2008b), IDHA (1983), and Phetteplace et al. (2013a, 2013b). Licensed for single user. © 2020 ASHRAE, Inc.

Water Hammer The term water hammer is used to describe several phenomena that occur in fluid flow. Although these phenomena differ in nature, they all result in stresses in the piping that are higher than normally encountered. Water hammer can have a disastrous effect on a thermal utility by bursting pipes and fittings and threatening life and property. In steam systems (IDHA 1983), water hammer is caused primarily by condensate collecting in the bottom of the steam piping. Steam flowing at velocities 10 times greater than normal water flow picks up a slug of condensate and accelerates it to a high velocity. The slug of condensate subsequently collides with the pipe wall at a point where flow changes direction. To prevent this type of water hammer, condensate must be prevented from collecting in steam pipes by using proper steam pipe pitch and adequate condensate collection and return facilities. Water hammer also occurs in steam systems because of rapid condensation of steam during system warm-up. Rapid condensation decreases the specific volume and pressure of steam, which precipitates pressure shock waves. This form of water hammer is prevented by controlled warm-up of the piping. Valves should be opened slowly and in stages during warm-up. Large steam valves should be provided with smaller bypass valves to slow the warm-up. Water hammer in hot- and chilled-water distribution systems is caused by sudden changes in flow velocity, which causes pressure shock waves. The two primary causes are pump failure (i.e., power failure) and sudden valve closures. A simplified method to determine maximum resultant pressure may be found in Chapter 22 of the 2013 ASHRAE Handbook—Fundamentals. More elaborate methods of analysis can be found in Fox (1977), Stephenson (1981), and Streeter and Wylie (1979). Preventive measures include operational procedures and special piping fixtures such as surge columns.

Pressure Losses Frictional pressure losses occur at the interface between the inner wall of a pipe and a flowing fluid due to shear stresses. In steam systems, these pressure losses are compensated for with increased steam pressure at the point of steam generation. In water systems, pumps are used to increase pressure at either the plant or intermediate points in the distribution system. Calculation of pressure loss is discussed in Chapters 3 and 22 of the 2013 ASHRAE Handbook— Fundamentals. Hydraulic calculations reveal that a great deal of the system pressure drop is caused by pipe fittings and offsets; there-

Pipe Sizing Ideally, the appropriate pipe size should be determined from an economic study of the life-cycle cost for construction and operation. In practice, however, this study is seldom performed because of the effort involved. Instead, criteria that have evolved from practice are frequently used for design. These criteria normally take the form of constraints on the maximum flow velocity or pressure drop. Chapter 22 of the 2013 ASHRAE Handbook—Fundamentals provides velocity and pressure drop constraints. Noise generated by excessive flow velocities is usually not a concern for thermal utility distribution systems outside of buildings. For steam systems, maximum flow velocities of 60 to 75 m/s are recommended (IDHA 1983). For water systems, Europeans use the criterion that pressure losses should be limited to 100 Pa per metre of pipe (Bøhm 1988). Other studies indicate that higher levels of pressure loss may be acceptable (Stewart and Dona 1987) and warranted from an economic standpoint (Bøhm 1986; Koskelainen 1980; Phetteplace 1989). For chilled-water systems (excluding condenser water piping) in buildings, the requirements of ASHRAE Standard 90.1-2013 must be followed for water velocity and piping pressure drops. When establishing design flows for thermal distribution systems, consider the diversity of consumer demands (i.e., the various consumers’ maximum demands do not occur at the same time). Thus, the heat supply and main distribution piping may be sized for a maximum load that is somewhat less than the sum of the individual consumers’ maximum demands. For steam systems, Geiringer (1963) suggests diversity factors of 0.80 for space heating and 0.65 for domestic hot-water heating and process loads. Geiringer also suggests that these factors may be reduced by approximately 10% for high-temperature water systems. Werner (1984) conducted a study of the heat load on six operating low-temperature hot-water systems in Sweden and found diversity factors ranging from 0.57 to 0.79, with the average being 0.685. Use caution when applying these diversity factors, because the size and number of loads in a system affect the values.

Network Calculations Calculating flow rates and pressures in a piping network with branches, loops, pumps, and heat exchangers can be difficult without the aid of a computer. Methods have been developed primarily for domestic water distribution systems (Jeppson 1977; Stephenson 1981). These may apply to thermal distribution systems with appropriate modifications. Computer-aided design methods usually incorporate methods for hydraulic analysis as well as for calculating heat losses and delivered water temperature at each consumer. Calculations are usually carried out in an iterative fashion, starting with constant supply and return temperatures throughout the network. After initial estimates of the design flow rates and heat losses are determined, refined estimates of the actual supply temperature at each consumer are computed. Flow rates at each consumer are then adjusted to ensure that the load is met with the reduced supply temperature, and the calculations are repeated. Most calculations are performed for a static or steady-state condition and have several input assumptions (e.g., temperatures, pressures, flows, loads) for a specific moment in time. Some computer modeling software packages are dynamic and take real-time inputs of flow and pressure from field instrumentation to efficiently control and optimize the speed of the distribution pumps.

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Condensate Drainage and Return in Steam Systems Condensate forms in operating steam lines as a result of heat loss. When a steam system’s operating temperature is increased, condensate also forms as steam warms the piping. At system startup, these loads usually exceed any operating heat loss loads; thus, special provisions should be made. To drain the condensate, steam piping should slope toward a collection point called a drip station. Drip stations are located in access areas or buildings where they are accessible for maintenance. Steam piping should slope toward the drip station at least 2 mm/m. If possible, the steam pipe should slope in the same direction as steam flow; otherwise, increase the pipe size to at least one size greater than would normally be used. This reduces the flow velocity of the steam and provides better condensate drainage against the steam flow. Drip stations should be spaced no further than 150 m apart in the absence of other requirements. Drip stations consist of a short piece of pipe (called a drip leg) positioned vertically on the bottom of the steam pipe, as well as a steam trap and appurtenant piping. The drip leg should be the same diameter as the steam pipe. The length of the drip leg should provide a volume equal to 50% of the condensate load from system start-up for steam pipes of 100 mm diameter and larger and 25% of the startup condensate load for smaller steam pipes (IDHA 1983). Steam traps should be sized to meet the normal load from operational heat losses only. Start-up loads should be accommodated by manual operation of the bypass valve. Steam traps are used to separate the condensate and noncondensable gases from the steam. For drip stations on steam distribution piping, use inverted bucket or bimetallic thermostatic traps. Some steam traps have integral strainers; others require separate strainers. Ensure that drip leg capacity is adequate when thermostatic traps are used because they always accumulate some condensate. If it is to be returned, condensate leaving the steam trap flows into the condensate return system. If steam pressure is sufficiently high, it may be used to force the condensate through the condensate return system. With low-pressure steam or on systems where a large pressure exists between drip stations and the ultimate destination of the condensate, condensate receivers and pumps must be provided. High-pressure condensate from drip traps may be sparged into lowpressure condensate mains with a sparge tube (essentially, the highpressure condensate flashes in the sparge tube). Schedule 80 steel piping is recommended for condensate lines because of the extra allowance for corrosion it provides. Steam traps have the potential of failing in an open position, thus nonmetallic piping must be protected from live steam where its temperature/ pressure would exceed the limitations of the piping. Nonmetallic piping should not be located so close to steam pipes that heat losses from the steam pipes could overheat it. Additional information on condensate removal may be found in Chapter 11. Information on sizing condensate return piping may be found in Chapter 22 of the 2013 ASHRAE Handbook—Fundamentals.

3.2

THERMAL CONSIDERATIONS

Thermal Design Conditions Three thermal design conditions must be met to ensure satisfactory system performance: 1. The “normal” condition used for the life-cycle cost analysis determines appropriate insulation thickness. Average values for the temperatures, burial depth, and thermal properties of the materials are used for design. If the thermal properties of the insulating material are expected to degrade over the useful life of the system, make appropriate allowances in the cost analysis.

2. Maximum heat transfer rate determines the load on the central plant due to the distribution system. It also determines the temperature drop (or increase, in the case of chilled-water distribution), which determines the delivered temperature to the con sumer. For this calculation, each component’s thermal conductivity must be taken at its maximum value, and the temperatures must be assumed to take on their extreme values, which would result in the greatest temperature difference between the carrier medium and the soil or air. The burial depth is normally at its lowest value for this calculation. During operation, none of the thermal capabilities of the materials (or any other materials in the area influenced thermally by the system) must exceed design conditions. To satisfy this objective, each component and the surrounding environment must be examined to determine whether thermal damage is possible. A numerical heat transfer analysis may be necessary in some cases. 3. The conditions of these analyses must be chosen to represent the worst-case scenario from the perspective of the component being examined. For example, in assessing the suitability of a coating material for a metallic conduit, the thermal insulation is assumed to be saturated, the soil moisture is at its lowest probable level, and the burial depth is maximum. These conditions, combined with the highest anticipated pipe and soil temperatures, give the highest conduit surface temperature to which the coating could be exposed. Heat transfer in buried systems is influenced by the thermal conductivity of the soil and by the depth of burial, particularly when the insulation has low thermal resistance. Soil thermal conductivity changes significantly with moisture content; for example, Bottorf (1951) indicated that soil thermal conductivity ranges from 0.14 W/(m·K) during dry soil conditions to 2.16 W/(m·K) during wet soil conditions. For details on calculating thermal effects on district energy distribution piping, see Phetteplace et al. (2013a, 2013b).

Thermal Properties of Pipe Insulation and Soil Uncertainty in heat transfer calculations for thermal distribution systems results from uncertainty in the thermal properties of materials involved. Generally, the designer must rely on manufacturers’ data to obtain approximate values. The data in this chapter should only be used as guidance in preliminary calculations until specific products have been identified; then specific data should be obtained from the manufacturer of the product in question. Insulation. Insulation provides the primary thermal resistance against heat loss or gain in thermal distribution systems. Thermal properties and other characteristics of insulations normally used in thermal distribution systems are listed in Table 7. Material properties such as thermal conductivity, density, compressive strength, moisture absorption, dimensional stability, and combustibility are typically reported in ASTM standards for the respective material. Some properties have more than one associated standard For example, thermal conductivity for insulation material in block form may be reported using ASTM Standards C177, C518, or C1114. For piping containing hot media, thermal conductivity for insulation material fabricated or molded for use on piping is measured using ASTM Standard C335. Chyu et al. (1997a, 1997b, 1998a, 1998b) studied the effect of moisture on the thermal conductivity of insulating materials commonly used in underground district energy systems (ASHRAE research project RP-721). The results are summarized in Table 8. The insulated pipe was immersed in water maintained at 8 to 38°C to simulate possible conduit water temperatures during a failure. The fluid temperature in the insulated pipe ranged from 2 to 230°C. All insulation materials were tested unfaced and/or unjacketed to simulate installation in a conduit.

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District Heating and Cooling

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Table 7 Comparison of Commonly Used Insulations in Underground Piping Systems Calcium Silicate Type I/II ASTM C533

Urethane Foam

Mineral Fiber/ Preformed Glass Fiber Type 1 ASTM C547

Cellular Glass ASTM C552

conductivitya

Thermal (Values in parentheses are maximum permissible by ASTM standard listed), W/(m·K) Mean temp. = 40°C 0.048 0.022 0.057 (0.052) 90°C 0.054 (0.066/0.078) 0.024 0.067 (0.064) 150°C 0.059 (0.073/0.083) 0.080 (0.078) 200°C 0.066 (0.080/0.088) 0.092 (0.093) Density (max.), kg/m3

107 to 147

128 to 176 kg/m3

430

450

100 at 5% deformation

65

N/A

2%

N/A

2%

0

5

25

240/360

Maximum temperature, °C Compressive strength

(min.),b

650 kg/m2

Dimensional stability, linear shrinkage at maximum use temperature Flame spread Smoke index Water absorption

0.038 (0.036) 0.043 (0.045) 0.048 (0.057) (0.074)

120

0

0

50

As-shipped moisture content, 20% max. (by mass)

0.5

Water vapor sorption, 5% max. (by mass)

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aThermal

conductivity values in this table are from previous editions of this chapter and have been retained as they were used in examples. Thermal conductivity of insulation may vary with temperature, temperature gradient, moisture content, density, thickness, and shape. ASTM maximum values given are comparative for establishing quality control compliance, and are suggested for preliminary calculations where actual values are not available. They may not represent installed performance of insulation under actual conditions that may differ substantially from test conditions. The manufacturer should be able to supply appropriate design values. bCompressive strength for cellular glass shown is for flat material, capped as per ASTM Standard C240.

Table 8 Characteristics Heating Test

Effect of Moisture on Underground Piping System Insulations

Polyurethanea Pipe temp. 1.8 to 127°C Water bath 8 to 38°C

Length of submersion time to reach 70 days steady-state k-value

Cellular Glass

Mineral Woolb

Fibrous Glass

Pipe temp. 1.8 to 215°C Water bath 8 to 38°C

Pipe temp. 1.8 to 230°C Water bath 8 to 38°C

Pipe temp. 1.8 to 230°C Water bath 8 to 38°C

c

10 days

2h

Effective k-value increase from dry conditions after steady state achieved in submersion

14 to 19 times at steady state. Estimated water content of insulation 70% (by volume).

Up to 50 times at steady 52 to 185 times. Insulation Avg. 10 times, process unsteady.c Insulation showed state. Insulation completely completely saturated at steady evidence of moisture zone on saturated. state. inner diameter.

Primary heat transfer mechanism

Conduction

c

Length of time for specimen to return to dry steady-state k-value after submersion

Pipe at 127°C, after 16 days Pipe at 215°C, 8 h moisture content 10% (by volume) remaining

Pipe at 230°C, 9 days

Pipe at 193°C, 6 days

Cooling Test

Pipe temp. 2.8°C Water bath at 11°C

Pipe temp. 2.2°C Water bath 8 to 14°C

Pipe temp. 1.8 to 7°C Water bath 13°C

Insulation 1.8 to 230°C Water bath 8 to 38°C

Data recorded at 4 days constant at 12 days

6 days

1/2 h

14 times. Insulation completely saturated at steady state.

20 times. Insulation completely saturated at steady state.

Length of submersion time to reach 16 days steady-state conditions for k-value

Conduction and convection Conduction and convection

Effective k-value increase from dry conditions after steady state achieved

2 to 4 times. Water None. No water penetration. absorption minimal, ceased after 7 days.

Primary heat transfer mechanism

Conduction

Conduction

Conduction and convection Conduction and convection

Length of time for specimen to return to dry steady-state k-value after submersion

Pipe at 3.3°C, data curve extrapolated to 10+ days

Pipe at 0.6°C, no change

Pipe at 1.8°C, data curve extrapolated to 25 days

Pipe at 1.8°C, 15 days

Source: Chyu et al. (1997a, 1997b; 1998a, 1998b). aPolyurethane material tested had a density of 46 kg/m3. bMineral wool tested was a preformed molded basalt designed for pipe systems operating up to 650°C. It was specially formulated to withstand the Federal Agency Committee (FAC) 96 h boiling water test. See Phetteplace (2013b) for the FAC test protocol.

cCracks

Painting chilled-water piping before insulating is recommended in areas of high humidity. Insulations used today for chilled water include polyurethane, polyisocyanurate, phenolics, cellular plastics, and fiberglass. No insulation system is totally water- and vaportight; thus, water from the atmosphere can enter the insulation system in small amounts, which can cause corrosion of the pipe. Chloride ions

present in most atmospheric water exacerbate corrosion. The best way to minimize corrosion is to make the insulation system highly water resistant by using a closed-cell insulation material coupled with a high-performance vapor retarder, and painting the pipe exterior with a strong rust-preventative coating (two-part epoxy) before insulating. This is good engineering practice and most insula-

formed on heating for all samples of cellular glass insulation tested. Flooded heat loss mechanism involved dynamic two-phase flow of water through cracks; period of dynamic process was about 20 min. Cracks had negligible effect on thermal conductivity of dry cellular glass insulation before and after submersion. No cracks formed during cooling test.

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Table 9 Soil Thermal Conductivities Thermal Conductivity, W/(m·K)

Soil Moisture Content (by mass)

Sand

Silt

Clay

Low, 20%

0.29 1.87 2.16

0.14 1.30 2.16

0.14 1.00 2.16

tion manufacturers suggest this, but it may not be in their literature. In addition, a good vapor retarder is required on the exterior of the insulation. When assessing the potential for pipe corrosion under insulation, the pipe operating temperature must be considered. Pipes operating permanently or predominantly at temperatures below about 10°C or above 121°C are not prone to corrosion. Pipes operating between these temperatures, particularly if they are subject to temperature cycling in use or from frequent shutdowns, are most susceptible to corrosion. Soil. If a soil analysis is available or can be done, the thermal conductivity of the soil can be estimated from published data [e.g., Farouki (1981); Lunardini (1981)]. The thermal conductivity factors in Table 9 may be used as an estimate when detailed information on the soil is not known. Because dry soil is rare in most areas, only assume a low moisture content for system material design, or where it can be validated for calculation of heat losses in the normal operational condition. Values of 1.4 to 1.7 W/(m·K) are commonly used where soil moisture content is unknown. Because moisture migrates toward a chilled pipe, use a thermal conductivity value of 2.16 W/(m·K) for chilled-water systems in the absence of any sitespecific soil data. For steady-state analyses, only the thermal conductivity of the soil is required. If a transient analysis is required, the specific heat and density are also required.

3.3

METHODS OF HEAT TRANSFER ANALYSIS

Because heat transfer in piping is not related to the load factor, it can be a large part of the total load. The most important factors affecting heat transfer are the difference between ambient and fluid temperatures and the thermal insulation. For direct-buried piping, the ambient temperature is the earth or soil temperature. For example, the extremes might be a 150 mm, insulated, 200°C water line in 5°C earth with 100 to 200 W/m loss; and a 150 mm, uninsulated, 13°C chilled-water return in 16°C earth with 10 W/m gain. The former requires analysis to determine the required insulation and its effect on the total heating system; the latter suggests analysis and insulation needs might be minimal. Other factors that affect heat transfer are (1) depth of burial, related to the earth temperature and soil thermal resistance; (2) soil thermal conductivity, related to soil moisture content and density; and (3) distance between adjacent pipes. Computing transient heat gains or losses in underground piping systems requires using numerical methods that approximate any physical problem and include factors such as the effect of temperature on thermal properties. For most designs, numerical analyses may not be warranted; one exception is where the potential exists to thermally damage components of the distribution system or something adjacent to it. Also, complex geometries may require numerical analysis. Albert and Phetteplace (1983), Minkowycz et al. (1988), and Rao (1982) have further information on the theory of methods of numerical analysis, and many commercial software packages are available for conducting numerical analyses. Steady-state calculations are appropriate for determining the annual heat loss/gain from a buried system if average annual earth temperatures are used. Steady-state calculations may also be appropriate for worst-case analyses of thermal effects on materials. Steady-state calculations for a one-pipe system may be done without a computer,

but it becomes increasingly difficult for a two-, three-, or four-pipe system. The following steady-state methods of analysis use resistance formulations developed by Phetteplace and Meyer (1990) that simplify the calculations needed to determine temperatures within the system. Each type of resistance is given a unique subscript and is defined only when introduced. In each case, the resistances are on a unit length basis so that heat flows per unit length result directly when the temperature difference is divided by the resistance.

Calculation of Undisturbed Soil Temperatures Before any heat loss/gain calculations may be conducted, the undisturbed soil temperature at the site must be determined. The choice of soil temperature is guided primarily by the type of calculation being conducted; see the section on Thermal Design Conditions. For example, if the purpose of the calculation is to determine whether a material will exceed its temperature limit, the maximum expected undisturbed ground temperature is used. The appropriate choice of undisturbed soil temperature also depends on the location of the site, time of year, depth of burial, and thermal properties of the soil. Some methods for determining undisturbed soil temperatures and suggestions on appropriate circumstances to use them are as follows: 1. Use the average annual air or groundwater temperature to approximate the average annual soil temperature. This estimate is appropriate when the objective of the calculation is to yield average heat loss over the yearly weather cycle. Mean annual air temperatures may be obtained from various sources of climatic data (e.g., Chapter 14 of the 2017 ASHRAE Handbook—Fundamentals). 2. Use the maximum/minimum air temperature as an estimate of the maximum/minimum undisturbed soil temperature for pipes buried at a shallow depth. This approximation is an appropriate conservative assumption when checking the temperatures to determine if the temperature limits of any of materials proposed for use will be exceeded. Maximum and minimum expected air temperatures may be calculated from Equations (1) and (2) and setting the depth to zero. 3. For systems that are buried more deeply, maximum/minimum undisturbed soil temperatures may be estimated as a function of depth, soil thermal properties, and prevailing climate. This estimate is appropriate when checking temperatures in a system to determine whether the temperature limits of any of the materials proposed for use will be exceeded. The following equations may be used to estimate the minimum and maximum expected undisturbed soil temperatures: Maximum temperature = ts,z = tms + As e Minimum temperature = ts,z = tms – As e

– z   

– z   

(1) (2)

where ts,z z   tms As

= = = = = =

temperature, °C depth, m annual period, 365 days thermal diffusivity of the ground, m2/day mean annual surface temperature, °C surface temperature amplitude, °C

Values for the climatic constants tms and As may be found at tc62 .ashraetcs.org/pdf/ASHRAE_Climatic_Data.pdf for all worldwide weather stations included in the CD accompanying the 2009 ASHRAE Handbook—Fundamentals. Thermal diffusivity for soil may be calculated as follows: 86.4k s 24  3600k s  = ------------------------------------------------------------- = ----------------------------------------------------1000 s  c s + c w  w  100    s  c s + 4.18  w  100  

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District Heating and Cooling

12.17 transfer resistance at the ground surface. The effective thickness is calculated as follows:

where

s cs cw w ks

= = = = =

soil density, kg/m3 dry soil specific heat, kJ/(kg·K) specific heat of water = 4.18 kJ/(kg·K) moisture content of soil, % (dry basis) soil thermal conductivity, W/(m·K)

 = ks h

where

Because the specific heat of dry soil is nearly constant for all types of soil, cs may be taken as 0.73 kJ/(kg·K).

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4. For buried systems, the undisturbed soil temperatures may be estimated for any time of the year as a function of depth, soil thermal properties, and prevailing climate. This temperature may be used in lieu of the soil surface temperature normally called for by the steady-state heat transfer equations when estimates of heat loss/gain as a function of time of year are desired. Substituting the undisturbed soil temperatures at the pipe depth allows the steady-state equations to be used as a first approximation to the solution to the actual transient heat transfer problem with its annual temperature variations at the surface. The following equation may be used to estimate the undisturbed soil temperature at any depth at any point during the yearly weather cycle (ASCE 1996). (Note: The argument for the sine function is in radians.) ts,z = tms – As e

– z   

2   –  lag   sin ------------------------------ – z ----- 

(5)

 = effective thickness of fictitious soil layer, m ks = thermal conductivity of soil, W/(m·K) h = convective heat transfer coefficient at ground surface, W/(m2 ·K)

The effective thickness calculated with Equation (5) is simply added to the actual burial depth of the pipes in calculating the soil thermal resistance using Equations (6), (7), (20), (21), and (27).

Uninsulated Buried Pipe For this case (Figure 7), an estimate for soil thermal resistance may be used. This estimate is sufficiently accurate (within 1%) for the ratios of burial depth to pipe radius indicated next to Equations (6) and (7). Both the actual resistance and the approximate resistance are presented, along with the depth/radius criteria for each. 2

ln  2d  r o  Rs = -----------------------2k s

(4)

where

 = Julian date, days lag = phase lag of soil surface temperature, days

Use Equation (3) to calculate soil thermal diffusivity. Values for the climatic constants tms, As, and lag may be found at tc62.ashraetcs .org/pdf/ASHRAE_Climatic_Data.pdf for all worldwide weather stations included in the CD accompanying the 2009 ASHRAE Handbook—Fundamentals. Phetteplace et al. (2013b) contains equations that may be used to calculate the climatic constants for any weather data set, real or contrived. Equation (4) does not account for latent heat effects from freezing, thawing, or evaporation. However, for soil adjacent to a buried heat distribution system, the equation provides a good estimate, because heat losses from the system tend to prevent the adjacent ground from freezing. For buried chilled-water systems, freezing may be a consideration; therefore, systems that are not used or drained during the winter months should be buried below the seasonal frost depth. For simplicity, the ground surface temperature is assumed to equal the air temperature, which is an acceptable assumption for most design calculations. If a more accurate calculation is desired, use the following method to compensate for the convective thermal resistance to heat transfer at the ground surface.

12

ln {  d  r o  +   d  r o  – 1  } Rs = --------------------------------------------------------------------------2k s

for d/ro > 2

for d/ro > 4

(7)

where Rs ks d ro

= = = =

thermal resistance of soil, (m·K)/W thermal conductivity of soil, W/(m·K) burial depth to centerline of pipe, m outer radius of pipe or conduit, m

The pipe’s thermal resistance is included if it is significant compared to the soil resistance. The thermal resistance of a pipe or any concentric circular region is given by ln  r o  r i  Rp = ---------------------2k p

where Rp = thermal resistance of pipe wall, (m·K)/W kp = thermal conductivity of pipe, W/(m·K) ri = inner radius of pipe, m

Convective Heat Transfer at Ground Surface Heat transfer between the ground surface and the ambient air occurs by convection. In addition, heat transfer with the soil occurs due to precipitation and radiation. The heat balance at the ground surface is too complex to warrant detailed treatment in the design of buried district heating and cooling systems. However, in some circumstances, approximations that included impacts beyond an average convective heat transfer coefficient were found. For example, McCabe et al. (1995) observed significant temperature variations caused by the type of surfaces over district heating and cooling systems. Phetteplace et al. (2013a) contains methods to approximate the impacts of surface type. Normally, only convection is considered, and as a first approximation, an effective thickness of a fictitious soil layer may be added to the burial depth to account for the effect of the convective heat

(6)

Fig. 7

Uninsulated Buried Pipe

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2020 ASHRAE Handbook—HVAC Systems and Equipment (SI)

Example 2. Consider an uninsulated, 75 mm Schedule 40 PVC chilledwater supply line carrying 7°C water. Assume the pipe is buried 1 m deep in soil with a thermal conductivity of 1.7 W/(m·K), and no other pipes or thermal anomalies are in close proximity. Assume the average annual soil temperature is 15°C. ri ro d ks kp

= = = = =

ering that the uncertainties in the material properties are likely greater than 5%, it is usually appropriate to neglect minor resistances such as those of piping and jacket materials if insulation is present.

Buried Pipe in Conduit with Air Space

39.1 mm = 0.0391 m 44.5 mm = 0.0445 m 1m 1.7 W/(m·K) 0.17 W/(m·K)

Solution: Calculate the thermal resistance of the pipe using Equation (8):

Systems with air spaces (Figure 8) may be treated by adding an appropriate resistance for the air space. For simplicity, assume a heat transfer coefficient of 17 W/(m2 ·K) (based on the outer surface area of the insulation), which applies in most cases. The resistance caused by this heat transfer coefficient is then Ra = 1/(17  2  roi) = 0.0094/roi

Rp = 0.12 (m·K)/W Calculate the thermal resistance of the soil using Equation (7). [Note: d/ro = 22 is greater than 4; thus Equation (7) may be used in lieu of Equation (6).] Rs = 0.36 (m·K)/W Calculate the rate of heat transfer by dividing the overall temperature difference by the total thermal resistance: tf – ts  7 – 15  q = ------------- = -------------------------------- = –17 W/m  0.12 + 0.36  Rt

(9)

where roi = outer radius of insulation, m Ra = resistance of air space, (m·K)/W

A more precise value for the resistance of an air space can be developed with empirical relations available for convection in enclosures such as those given by Grober et al. (1961). Consider the effect of radiation in the annulus when high temperatures are expected in the air space. For the treatment of radiation, refer to Siegel and Howell (1981).

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where Rt = total thermal resistance, (i.e., Rs + Rp in this case of pure series heat flow), (m·K)/W tf = fluid temperature, °C ts = average annual soil temperature, °C q = heat loss or gain per unit length of system, W/m

The negative result indicates a heat gain rather than a loss. Note that the thermal resistance of the fluid/pipe interface has been neglected, which is a reasonable assumption because such resistances tend to be very small for flowing fluids. Also note that, in this case, the thermal resistance of the pipe comprises a significant portion of the total thermal resistance. This results from the relatively low thermal conductivity of PVC compared to other piping materials and the fact that no other major thermal resistances exist in the system to overshadow it. If any significant amount of insulation were included in the system, its thermal resistance would dominate, and it might be possible to neglect that of the piping material.

Insulated Buried Pipe Equation (8) can be used to calculate the thermal resistance of any concentric circular region of material, including an insulation layer. When making calculations using insulation thickness, use actual rather than nominal thickness to obtain the most accurate results.

Example 4. Consider a 150 mm nominal diameter (168 mm outer diameter or 84 mm radius) high-temperature water line operating at 190°C. Assume the pipe is insulated with 65 mm of mineral wool with a thermal conductivity ki = 0.045 W/(m·K) at 100°C and ki = 0.052 W/(m·K) at 150°C. The pipe is encased in a steel conduit with a concentric air gap of 25 mm. The steel conduit is 3 mm thick and has a corrosion resistant coating approximately 3 mm thick. The pipe is buried 1.2 m deep to pipe centerline in soil with an average annual temperature of 15°C. The soil thermal conductivity is assumed to be 1.7 W/(m·K). Thermal resistances of the pipe, conduit, and conduit coating will be neglected. Solution: Calculate the thermal resistance of the pipe insulation. To do so, assume a mean insulation temperature of 120°C to establish its thermal conductivity, which is equivalent to assuming the insulation outer surface temperature is 50°C. Interpolating the data listed previously, insulation thermal conductivity ki = 0.048 W/(m·K). Then calculate insulation thermal resistance using Equation (8): ln   84 + 65   84  Ri = ------------------------------------------- = 1.90 (m·K)/W 2  0.048 Calculate the thermal resistance of the air space using Equation (9): Ra = 1/[107(84 + 65)  10–3] = 0.06 (m·K)/W

Example 3. Consider the effect of adding 25 mm of urethane foam insulation and a 3 mm thick PVC jacket to the chilled-water line in Example 2. Calculate the thermal resistance of the insulation layer from Equation (8) as follows: ln   44.5 + 25   44.5  Ri = ---------------------------------------------------- = 3.38 (m·K)/W 2  0.021

For the PVC jacket material, use Equation (8) again:

ln   3 + 44.5 + 25    44.5 + 25   Rj = ------------------------------------------------------------------------------- = 0.04 (m·K)/W 2  0.17

The thermal resistance of the soil as calculated by Equation (7) decreases slightly to Rs = 0.31 (m·K)/W because of the increase in the outer radius of the piping system. The total thermal resistance is now Rt = Rp + Ri + Rj + Rs = 0.12 + 3.38 + 0.04 + 0.31 = 3.85 (m·K)/W

Heat gain by the chilled-water pipe is reduced to about 2 W/m. In this case, the thermal resistance of the piping material and the jacket material could be neglected with a resultant error of bo

(27)

where Rts ao bo d

If the thermal conductivity of the pipe insulation is a function of temperature, assume an air temperature for the air space before starting calculations. Iterate the calculations if the air temperature calculated with Equation (30) differs significantly from the initial assumption.

The thermal resistance of the soil surrounding the trench is given by Equation (27): ln   3.5  1.25    1  1.3   Rts = --------------------------------------------------------------------------------- = 0.131 (m·K)/W 1.7  1.3   2  1  + 5.7  0.75

= = = =

thermal resistance of soil surrounding trench/tunnel, (m·K)/W width of trench/tunnel outside, m height of trench/tunnel outside, m burial depth of trench to centerline, m

Equations (26) and (27) can be combined with the equations already presented to calculate heat flow and temperature for trenches/tunnels. As with the conduits described in earlier examples, the heat transfer processes in the air space inside the trench/ tunnel are too complex to warrant a complete treatment for design purposes. The thermal resistance of this air space may be approximated by several methods [e.g., by using Equation (9)]. Thermal resistances at the pipe insulation/air interface can also be calculated from heat transfer coefficients as done in the section on Pipes in Air. If the thermal resistance of the air/trench wall interface is also included, use Equation (28): Raw = 1[2ht (a + b)]

(28)

0.25

The trench wall thermal resistance is calculated with Equation (26): Rw = 0.150/[2  1.7(1 + 0.7)] = 0.026 (m·K)/W

If the thermal resistance of the air/trench wall is neglected, the total thermal resistance on the soil side of the air space is Rss = Rw + Rts = 0.026 + 0.131 = 0.157 (m·K)/W Find a first estimate of the total heat loss using Equation (29):

 160 – 15   2.62 +  100 – 15   2.53 q = --------------------------------------------------------------------------------------- = 79.3 (m·K)/W 1 +  0.157  2.62  +  0.157  2.53  The first estimate of the air temperature in the trench is given by Equation (30): tta = 15 + (79.3  0.157) = 27.5°C

Refined estimates of the pipe insulation surface temperatures are then calculated using Equations (14) and (15): ti1 = 27.5 + [(160 – 27.5)(0.08/2.62)] = 31.5°C

where Raw = thermal resistance of air/trench wall interface, (m·K)/W ht = total heat transfer coefficient at air/trench wall interface, W/(m2 ·K)

ti2 = 27.5 + [(100 – 27.5)(0.09/2.53)] = 30.1°C

The total heat loss from the trench/tunnel is calculated from the following relationship:  t p1 – t s   R 1 +  t p2 – t s   R 2 q = --------------------------------------------------------------------1 +  R ss  R 1  +  R ss  R 2 

(29)

From these estimates, calculate the revised mean insulation temperatures to find the resultant resistance values. Repeat the calculation procedure until satisfactory agreement between successive estimates of the trench air temperature is obtained. Calculate the individual heat flows from the pipes with Equations (31) and (32). If the thermal resistance of the trench walls is added to the soil thermal resistance, the thermal resistance on the soil side of the air space is ln   3.5  1.25    0.7  1  Rss = --------------------------------------------------------------------------------- = 0.160 (m·K)/W 1.7   1  1.4  + 5.7  0.75

where Rl , R2 = thermal resistances of two-pipe/insulation systems within trench/tunnel, (m·K)/W Rss = total thermal resistance on soil side of air within trench/tunnel, (m·K)/W

Once the total heat loss has been found, the air temperature in the trench/tunnel may be found as tta = ts + qRss

(30)

where tta is air temperature in the trench/tunnel. The individual heat flows for the two pipes in the trench/tunnel are then q1 = (tp1 – tta)R1

(31)

q2 = (tp2 – tta)R2

(32)

0.25

The result is less than 2% higher than the resistance previously calculated by treating the trench walls and soil separately. In the event that the soil and trench wall material have significantly different thermal conductivities, this simpler calculation will not yield as favorable results and should not be used.

Pipes in Shallow Trenches The cover of a shallow trench is exposed to the environment. Thermal calculations for such trenches require the following assumptions: (1) the interior air temperature is uniform as discussed in the section on Pipes in Buried Trenches or Tunnels, and (2) the soil and the trench wall material have the same thermal conductivity. This assumption yields reasonable results if the thermal conductivity of the trench material is used, because most of the heat flows

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directly through the cover. The thermal resistance of the trench walls and surrounding soil is usually a small portion of the total thermal resistance, and thus the heat losses are not usually highly dependent on this thermal resistance. Using these assumptions, Equations (27) and (29) to (32) may be used for shallow trench systems.

other modes and thus may be neglected. Thermal resistances for cylindrical systems can be found from heat transfer coefficients derived using the following equations. Generally, piping systems used for thermal distribution have sufficiently low surface temperatures to preclude any significant heat transfer by thermal radiation.

Example 10. Consider a shallow trench having the same physical parameters and operating conditions as the buried trench in Example 9, except that the top of the trench is at grade level. Calculate the thermal resistance of the shallow trench using Equation (27):

Example 11. Consider a 150 mm nominal (168 mm outside diameter), high-temperature hot-water pipe that operates in air at 190°C with 65 mm of mineral wool insulation. The surrounding air annually averages 15°C (288 K). The average annual wind speed is 6 km/h. The insulation is covered with an aluminum jacket with an emittance  = 0.26. The thickness and thermal resistance of the jacket material are negligible. Because the heat transfer coefficient at the outer surface of the insulation is a function of the temperature there, initial estimates of this temperature must be made. This temperature estimate is also required to estimate mean insulation temperature.

ln   3.5  0.5    0.7  1  Rts = Rss = ------------------------------------------------------------------------------ = 0.0759 (m·K)/W 1.7   1  1.4  + 5.7  0.75

0.25

Use thermal resistances for the pipe/insulation systems from Example 9, and use Equation (24) to calculate q:  160 – 15   2.62 +  100 – 15   2.53 q = --------------------------------------------------------------------------------------- = 84.0 W/m 1 +  0.0759  2.62  +  0.0759  2.53 

From this, calculate the first estimate of the air temperature, using Equation (30): tta = 15 + (84.0  0.0759) = 21.4°C

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Then refine estimates of the pipe insulation surface temperatures using Equations (14) and (15): ti1 = 21.4 + [(160 – 21.4)(0.08/2.62)] = 25.6°C ti2 = 21.4 + [(100 – 21.4)(0.09/2.53)] = 24.2°C From these, calculate the revised mean insulation temperatures to find resultant resistance values. Repeat the calculation procedure until satisfactory agreement between successive estimates of the trench air temperatures are obtained. If the individual heat flows from the pipes are desired, calculate them using Equations (31) and (32).

Another method for calculating heat losses in a shallow trench assumes an interior air temperature and treats the pipes as pipes in air (see the section on Pipes in Air). Interior air temperatures in the range of 20 to 50°C have been observed in a temperate climate (Phetteplace et al. 1991). It may be necessary to drain the system or provide heat tracing in areas with significant subfreezing air temperatures in winter, especially for shallow trenches that contain only chilled-water lines (and thus have no source of heat) if the system is not in operation, or is in operation at very low flow rates.

Buried Pipes with Other Geometries Other geometries not specifically addressed by the previous cases have been used for buried thermal utilities. In some instances, the equations presented previously may be used to approximate the system. For instance, the soil thermal resistance for a buried system with a half-round clay tile on a concrete base could be approximated as a circular system using Equation (6) or (7). In this case, the outer radius ro is taken as that of a cylinder with the same circumference as the outer perimeter of the clay tile system. The remainder of the resistances and subsequent calculations would be similar to those for a buried trench/tunnel. The accuracy of such calculations varies inversely with the proportion of the total thermal resistance that the thermal resistance in question comprises. In most instances, the thermal resistance of the pipe insulation overshadows other resistances, and the errors induced by approximations in the other resistances are acceptable for design calculations with appropriate conservative assumptions applied.

Pipes in Air Pipes surrounded by gases may transfer heat via conduction, convection, and/or thermal radiation. Heat transfer modes depend mainly on the surface temperatures and geometry of the system being considered. For air, conduction is usually dominated by the

Solution: Assuming that the insulation surface is at 40°C as a first estimate, the mean insulation temperature is calculated as 115°C (313 K). Using the properties of mineral wool given in Example 4, the insulation thermal conductivity is ki = 0.047 W/(m·K). Using Equation (8), the thermal resistance of the insulation is Ri = 1.94 (m·K)/W. The forced convective heat transfer coefficient at the surface of the insulation can be found using the following equation (ASTM Standard C680):  1 hcv = 11.58  ------  d

 1  = 11.58  ---------  298

0.2

0.2

 2   ----------------   t oi + t a 

 2   ------------------  40 + 15

0.181

(toi – ta)0.266 (1 + 0.7935 V )0.5

0.181

(40 – 15)0.266 (1 + 0.7935  6)0.5

= 10.4 W/(m2 ·K) where d = outer diameter of surface, mm ta = ambient air temperature, °C V = wind speed, km/h The radiative heat transfer coefficient must be added to this convective heat transfer coefficient. Determine the radiative heat transfer coefficient as follows (ASTM Standard C680):  Ta – T s  hrad =  --------------------Ta – Ts 4

4

 313 – 288  hrad = 0.26  5.67  10–8 --------------------------------- = 1.60 W/(m2 ·K) 313 – 288 4

4

Add the convective and radiative coefficients to obtain a total surface heat transfer coefficient ht of 12.0 W/(m2 ·K). The equivalent thermal resistance of this heat transfer coefficient is calculated from the following equation: 1 1 Rsurf = ------------------ = ---------------------------------------------- = 0.09 (m·K)/W 2  0.149  13.53 2r oi h t With this, the total thermal resistance of the system becomes Rt = 1.94 + 0.09 = 2.03 (m·K)/W, and the first estimate of the heat loss is q = (190  15)/2.03 = 86.2 W/m. An improved estimate of the insulation surface temperature is toi = 190  (86.2  1.94) = 23°C. From this, a new mean insulation temperature, insulation thermal resistance, and surface resistance can be calculated. The heat loss is then 84.8 W/m, and the insulation surface temperature is calculated as 20°C. These results are close enough to the previous results that further iterations are not warranted. Note that the contribution of thermal radiation to the heat transfer could have been omitted with negligible effect on the results. In fact, the entire surface resistance could have been neglected and the resulting heat loss would have increased by only about 4%.

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District Heating and Cooling

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In Example 11, the convective heat transfer was forced. In cases with no wind, where the convection is free rather than forced, the radiative heat transfer is more significant, as is the total thermal resistance of the surface. However, in instances where the piping is well insulated, the thermal resistance of the insulation dominates, and minor resistances can often be neglected with little resultant error. By neglecting resistances, a conservative result is obtained (i.e., the heat transfer is overpredicted).

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Economical Thickness for Pipe Insulation A life-cycle cost analysis may be run to determine the economical thickness of pipe insulation. Because the insulation thickness affects other parameters in some systems, each insulation thickness must be considered as a separate system. For example, a conduit system or one with a jacket around the insulation requires a larger conduit or jacket for greater insulation thicknesses. The cost of the extra conduit or jacket material may exceed that of the additional insulation and is therefore usually included in the analysis. It is usually not necessary to include excavation, installation, and backfill costs in the analysis. A system’s life-cycle cost is the sum of the initial capital cost and the present worth of the subsequent cost of heat lost or gained over the life of the system. The initial capital cost needs only to include those costs that are affected by insulation thickness. The following equation can be used to calculate the life-cycle cost: LCC = CC + (qtuChPWF) (33) where LCC = present worth of life-cycle costs associated with pipe insulation thickness, $/m CC = capital costs associated with pipe insulation thickness, $/m q = annual average rate of heat loss, W/m tu = utilization time for system each year, s Ch = cost of heat lost from system, $/J PWF = present worth factor for future annual heat loss costs, dimensionless

The present worth factor is the reciprocal of the capital recovery factor, which is found from the following equation: i1 + i CRF = --------------------------n 1 + i – 1 n

(34)

where CRF = cost recovery factor, dimensionless i = interest rate n = useful lifetime of system, years

If heat costs are expected to escalate, the present worth factor may be multiplied by an appropriate escalation factor and the result substituted in place of PWF in Equation (33). Example 12. Consider a steel conduit system with an air space. The insulation is mineral wool with thermal conductivity as given in Examples 4 and 5. The carrier pipe is 100 mm NPS and operates at 175°C for the entire year (8760 h). The conduit is buried 1.2 m to the centerline in soil with a thermal conductivity of 1.7 W/(m·K) and an annual average temperature of 15°C. Neglect the thermal resistance of the conduit and carrier pipe. The useful lifetime of the system is assumed to be 20 years and the interest rate is taken as 10%. Solution: Find CRF from Equation (34): 0.10  1 + 0.10  CRF = ---------------------------------------- = 0.11746 20  1 + 0.10  – 1 20

The value Ch of heat lost from the system is assumed to be $10 per gigajoule. The following table summarizes the heat loss and cost data for several available insulation thicknesses.

Insulation thickness, mm 40 Insulation outer radius, mm 97 Insulation k, W/(m·K) 0.047 Ri, Eq. (8), (m·K)/W 1.80 Ra , Eq. (9), (m·K)/W 0.097 Conduit outer radius, m 0.136 Rs , Eq. (7), (m·K)/W 0.27 Rt , (m·K)/W 2.17 q, heat loss rate, W/m 73.7 Conduit system cost, $/m 75.50 LCC, Eq. (33), $/m 273.38

50 107 0.047 2.13 0.088 0.136 0.27 2.49 64.3 80.50 253.15

60 117 0.047 2.44 0.080 0.161 0.26 2.78 57.6 92.50 247.16

70 127 0.047 2.71 0.074 0.161 0.26 3.04 52.6 98.50 239.73

80 137 0.047 2.97 0.069 0.178 0.25 3.29 48.6 108.25 238.74

100 157 0.047 3.43 0.060 0.203 0.24 3.73 42.9 131.25 246.44

The table indicates that 80 mm of insulation yields the lowest lifecycle cost for the example. Because the results depend highly on the economic parameters used, they must be accurately determined.

3.4

EXPANSION PROVISIONS

All piping expands or contracts because of temperature changes, whether it contains chilled water, steam, or hot water. The piping’s length increases or decreases with its temperature. Field conditions and the type of system govern the method used to absorb the movement. Turns where the pipe changes direction must be used to provide flexibility. When the distance between changes in direction becomes too large for the turns to compensate for movement, add expansion loops at appropriate locations (this may require additional right-of-way). If field conditions allow, use the piping’s flexibility to allow expansion. Where space constraints do not allow expansion loops and/or changes in direction, mechanical methods, such as expansion joints or ball joints, must be used. However, because ball joints change the direction of a pipe, a third joint may be required to reduce the length between changes in direction. Chapter 46 covers the design of the pipe bends, loops, and the use of expansion joints. However, the chapter uses conservative stress values. Computer-aided design programs that calculate stress from pressure, thermal expansion, and weight simultaneously allow the designer to meet the requirements of ASME Standard B31.1. When larger pipe diameters are required, a computer program should be used to determine whether the pipe will provide the required flexibility. For example, Table 11 in Chapter 46 indicates that a 300 mm standard pipe with 300 mm of movement requires a 4.7 m wide by 9.4 m high expansion loop. One program recommends a 4.0 m wide by 7.9 m high loop; if equal height and width is specified, the loop is 7.2 m in each direction. Although the inherent flexibility of the piping should be used to handle expansion as much as possible, expansion joints must be used where space is too small to allow a loop to be constructed to handle the required movement. For example, expansion joints are often used in walk-through tunnels and vaults because there is seldom space to construct pipe loops. Either pipe loops or expansion joints can be used for aboveground, concrete shallow trench, directburied, and poured-envelope systems. The manufacturer of the conduit or envelope material should design loops and offsets in conduit and poured-envelope systems because clearance and design features are critical to the performance of both the loop and the pipe. All expansion joints require maintenance, and should therefore always be accessible for service. Joints in direct-buried and pouredenvelope systems and trenches without removable covers should be located in access ports. Cold springing is normally used when thermal expansion compensation is used. In DHC systems with natural flexibility, cold springing minimizes the clearance required for pipe movement only. The pipes are sprung 50% of the total amount of movement, toward the anchor. However, ASME Standard B31.1 does not allow cold springing in calculating the stresses in the piping. When expansion joints are used, they are installed in an extended position to achieve maximum movement. Contact the manufacturer of the expansion joint for the proper amount of extension.

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Fig. 12 Slide and Guide Detail

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In extremely hot climates, anchors may also be required, to compensate for pipe contraction when pipes are installed in high ambient temperatures and then filled with cold water. This can affect buried tees in the piping, especially at branch service line runouts to buildings. Crushable insulation may be used in the trench as part of the backfill, to compensate for the contractions. Anchors should be sized using computer-aided design software.

Pipe Supports, Guides, and Anchors For premanufactured conduit and poured-envelope systems, the system manufacturer usually designs the pipe supports, guides, and anchors in consultation with the expansion joint manufacturer, if such devices are used. For example, the main anchor force of an inline axial expansion joint is the sum of the pressure thrust (system pressure times the cross-sectional area of the expansion joint and the joint friction or spring force) and the pipe friction forces. Consult the manufacturer of the expansion device when determining anchor forces. Anchor forces are normally less when expansion is absorbed through the system instead of with expansion joints. Pipe guides used with expansion joints should be spaced according to the manufacturer’s recommendations. They must allow longitudinal or axial motion and restrict motion perpendicular to the axis of the pipe. Guides with graphite or low-friction fluorocarbon slide surfaces are often desirable for long pipe runs (Figure 12). In addition, these surface finishes do not corrode or increase sliding resistance in aboveground installations. Select guides to handle twice the expected movement, so they may be installed in a neutral position without the need for cold-springing the pipe. A computerized stress analysis program can help the designer calculate stresses and moments in the piping system to adequately size any anchors and anchor blocks to ensure compliance with ASME Standard B31.1.

3.5

DISTRIBUTION SYSTEM CONSTRUCTION

The combination of aesthetics, first cost, safety, and life-cycle cost naturally divide distribution systems into two distinct categories: aboveground and underground. The materials needed to ensure long life and low heat loss further classify DHC systems into low-, medium-, and high-temperature systems. The temperature range for medium-temperature systems is usually too high for materials used in low-temperature systems; however, the same materials used in high-temperature systems are typically used for medium-temperature systems. Because low-temperature systems have a lower temperature differential between the working fluid temperatures and the environment, heat loss is inherently less. In addition, the options for efficient insulation materials and inexpensive pipe materials that resist corrosion are much greater for lowtemperature systems. The aboveground system has the lowest first cost and the lowest life-cycle cost because it can be maintained easily and constructed with readily available materials. Generally, aboveground systems are acceptable where they are hidden from view or can be hidden by

landscaping. Poor aesthetics and the risk of vehicle damage to the aboveground system remove it from contention for many projects. Although the aboveground system is sometimes partially factory prefabricated, more typically it is entirely field fabricated of components such as pipes, insulation, pipe supports, and insulation jackets or protective enclosures that are commercially available. Other common systems that are completely field fabricated include walkthrough tunnel (see Figure 14), concrete surface trenches (see Figure 15), deep-burial small tunnels (see Figure 16), and underground systems that use poured insulation (see Figure 17) or rigid closedcell insulation (e.g., cellular glass) (see Figure 18) to form an envelope around the carrier pipes. Field-assembled systems must be designed in detail, and all materials must be specified by the project design engineer. Evaluation of the project site conditions indicates which type of system should be considered for the site. For instance, the shallow trench system is best where utilities that are buried deeper than the trench bottom need to be avoided and where the covers can serve as sidewalks. Directburied conduit, with a thicker steel casing coated in either epoxy or HDPE or wrapped in fiberglass-reinforced polymer/plastic (FRP), may be the only system that can be used in flooded sites where the conduit is in direct contact with groundwater. The conduit system is used where aesthetics is important. It is often used for short distances between buildings and the main distribution system, and where the owner is willing to accept higher life-cycle costs. Direct-buried conduit, concrete surface trench, and other underground systems must be routed to avoid existing utilities, which requires a detailed site survey and considerable design effort. Exterior surface temperature of the conduit is an important issue that must be addressed when selecting the appropriate construction materials and components that compose the conduit; temperature may be calculated using the methods previously outlined. In the absence of a detailed soil temperature distribution study, directburied heating systems should be spaced more than 4.5 m from other utilities constructed of plastic pipes because the temperature of the soil during dry conditions may be high enough to reduce the strength of plastic pipe to an unacceptable level. The strength of PVC at 60°C, for example, is only 20% of its room-temperature value. Rigid, extruded polystyrene insulation board may be used to insulate adjacent utilities from the impact of a buried heat distribution pipe; however, the temperature limit of the extruded polystyrene insulation must not be exceeded. Numerical analysis of the thermal problem may be required to ensure that the desired effect is achieved. Waterproofing the concrete trench or tunnel depends on the groundwater level. There are many methods of waterproofing belowground concrete structures, including membranes and concrete additives. Carefully address these issues so that the concrete structure outlives the piping system. Tunnels that provide walk-through or crawl-through access can be buried in nearly any location without causing future problems because utilities are typically placed in the tunnel. Regardless of the type of construction, it is usually cost effective to route distribution piping through the basements of buildings, but only after addressing liability issues. In laying out the main supply and return piping, consider redundancy of supply and return. If a looped system is used to provide redundancy, flow rates under all possible failure modes must be addressed when sizing and laying out the piping. Access ports for underground systems should be at critical points, such as where there are • High or low points on the system profile that vent trapped air or where the system can be drained • Elevation changes in the distribution system that are needed to maintain the required constant slope • Major branches with isolation valves • Steam traps and condensate drainage points on steam lines

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District Heating and Cooling • Mechanical expansion devices To facilitate leak location and repair and to limit damage caused by leaks, access points generally should be spaced no farther than 150 m apart. Pay special attention to the safety of personnel who come in contact with distribution systems or who must enter spaces occupied by underground systems. The regulatory authority’s definition of a confined space and the possibility of exposure to hightemperature or high-pressure piping can significantly affect the access design, which must be addressed by the designer. Gravity venting of tunnels is good practice, and access ports and tunnels should have lighting and convenience outlets to aid in inspection and maintenance of anchors, expansion joints, and piping.

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Piping Materials and Standards Supply Pipes for Steam and Hot Water. Adequate temperature and pressure ratings for the intended service should be specified for any piping. All piping, fittings, and accessories should be in accordance with ASME Standard B31.1 or with local requirements if more stringent. For steam and hot water, all joints should be welded and pipe should conform to either ASTM Standard A53 seamless or ERW, Grade B; or ASTM Standard A106 seamless, Grade B. Do not use ASTM Standard A53, Type F, because of its lower allowable stress and because of how its seams are manufactured. Mechanical joints of any type are not recommended for steam or hot water. Pipe wall thickness is determined by the maximum operating temperature and pressure. In the United States, most piping for steam and hot water is Schedule 40 for 250 mm NPS and below and standard for 300 mm NPS and above. Many European low-temperature water systems have piping with a wall thickness similar to Schedule 10. The reduced piping material in these systems means they are not only less expensive, but they also develop reduced expansion forces and thus require simpler methods of expansion compensation, a point to consider when choosing between a high- and low-temperature system. Welding pipes with thinner walls requires extra care and may require additional inspection. Also, extra care is required to avoid internal and external corrosion because the thinner wall provides a much lower corrosion allowance. Condensate Return Pipes. Condensate pipes require special consideration because condensate is much more corrosive than steam. This corrosiveness is caused by the oxygen that the condensate accumulates. The usual method used to compensate for condensate corrosiveness is to select steel pipe that is thicker than the steam pipe. For highly corrosive condensate, stainless steel and/or other corrosion-resistant materials should be considered. Materials that are corrosion resistant in air may not be corrosion resistant when exposed to condensate; therefore, a material with good experience handling condensate should be selected. FRP or glass-reinforced plastic (GRP) pipe used for condensate return has not performed well and should be avoided. Failed steam traps, pipe resin solubility, deterioration at elevated temperatures, and thermal expansion are thought to be the causes of the premature failures seen. Chilled-Water Distribution. For chilled-water systems, a variety of pipe materials, such as steel, ductile iron, reinforced concrete, HDPE, PVC, and FRP/GRP, have been used successfully. If ductile iron or steel is used, the designer must resolve the internal and external corrosion issue, which may be significant unless cement-lined ductile iron is used. Soil temperature is usually highest when chilled-water loads are highest; therefore, it is usually life-cycle cost-effective to insulate the chilled-water pipe. Phetteplace et al. (2013a) has more information on thermal design of chilled-water distribution systems, including the effects of surface type (e.g., pavement, grass). Life-cycle cost analysis often favors a factoryprefabricated product insulated with sprayed or injected foam with a waterproof casing or a field-fabricated system insulated with rigid

12.27 closed-cell insulation. If a nonmetallic product is selected, be careful to maintain an adequate distance from any high-temperature underground system that may be near. Damage to the chilled-water system would most likely occur from elevated soil temperatures when the chilled-water circulation stopped. Proper pressure relief devices should be designed into all systems to compensate for the increase in system pressure caused by increases in fluid temperatures. Uninsulated chilled-water piping placement should take into account the proximity of any heating piping, especially pipes that are poorly insulated and old. Consider increasing the distance between these piping systems, especially for nonmetallic piping. The heating distribution system should be at least 4.5 m from a chilled-water system containing plastic unless a detailed study of the soil temperature distribution indicates otherwise. Rigid extruded polystyrene insulation board may be used to insulate adjacent chilled-water lines from the effects of a buried heat distribution pipe; however, care must be taken not to exceed the temperature limits of the extruded polystyrene insulation; numerical analysis of the thermal problem may be required. Finally, at valve locations or when transitioning from ductile or plastic piping to steel at the buildings, flanged connections are usually best, but should be located inside the building or a small access port and not directly buried. Experience shows that buried flanges, or flanged connections such as to valves, are prone to leakage, are a weak link in the piping system, and should be avoided. Proper gasket selection and bolt torque are also critical to a leakfree system. Table 12 summarizes some of the important aspects of the various piping materials, and Figure 13 shows approximate relative costs of the most popular materials. The major advantages and disadvantages of each of these materials, as well as applicable standards when used for the carrier pipe, are as follows: • Steel Advantages: High strength and good flexibility, can be joined by welding for a high-integrity joint that can be inspected for quality control, widely available in all sizes, familiar material to most workforces. Disadvantages: Relatively high cost, highly susceptible to corrosion and will require corrosion protection. Skilled labor force required for welding. Slower installation, especially in larger diameters. Standards: ASTM A53/A53M, ASTM A106A/106M. • Copper Advantages: Good flexibility, can be joined by soldering for a high-integrity joint, corrosion resistant but may still require protection, familiar material to most workforces. Disadvantages: Expensive, only available/practical in small diameters (approximately 150 mm and smaller). Standards: ASTM B88. • Ductile Iron Advantages: Reasonable strength and flexibility, available in sizes from 100 mm to about 1.6 m, familiar material to many workforces. Faster installation. Disadvantages: Heavy, susceptible to corrosion and requires corrosion protection, can only be joined by mechanical joints, some mechanical joint designs require thrust blocks at all changes in direction. Fittings are expensive. Allowable leakage per standards. Standards: AWWA C151. • Cementitious Pipe Advantages: Reasonable strength, available in all sizes, familiar material to many workforces. Disadvantages: Heavy, poor flexibility, can only be joined by mechanical joints, thrust blocks are required, difficult to add branch line piping to. Lower pressure and velocity limits. Allowable leakage per standards.

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2020 ASHRAE Handbook—HVAC Systems and Equipment (SI) Table 12

Piping System

Relative Merits of Piping Materials Commonly Used for District Cooling Distribution Systems Carrier Pipe Joint Integrity

Welded steel

Excellent

Soldered copper Ductile iron

Medium Low

Joint Inspection NDT (x-ray, etc.), pressure testing Pressure testing Pressure testing

Cement pipe Low Pressure testing FRP Low/medium Pressure testing PVC Low Pressure testing HDPE High Pressure testing Large D = Medium Excellent Small D = Low Large D = Medium/high

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*Insulated

Insulated Joints Possible* Yes Yes No No Yes No Yes

Corrosion Resistance Low, requires protection Good Low, requires protection Excellent Excellent Excellent Excellent

Installation Skill Level

Installation Time

Strength under Burial Conditions

Relative Installed Cost

High

High

Excellent

High

Medium Low/medium

Medium Low

Good Very good

Small D = High Low/medium

Low/medium Medium Low/medium Medium

Low Low/medium Low Small D = Low

Good Low Low

Low/medium Low/medium Low

joints are not recommended for piping systems that have allowable leakage rates for joints.

Advantages: Low mass, very flexible, can be fusion welded for high-integrity joints, available in sizes up to 1.6 m. Leak free and fully restrained (no anchor blocks). Disadvantages: Low strength compared to steel and FRP results in significant wall thickness and thus cost in larger diameters. Increased wall thickness also reduces inside diameter, which results in higher pressure losses and may require larger sizes for the same flow rates. Larger-diameter fusion welding machines may be of limited availability. Cost fluctuates with petroleum pricing. Standards: AWWA C901, AWWA C906.

Aboveground Systems

Fig. 13 Relative Costs for Piping Alone, Uninsulated Cost data are from RSMeans-CostWorks® (www.rsmeansonline.com) for third-quarter 2012 water utilities.

Standards: AWWA C300, AWWA C301, AWWA C302, AWWA C303. • FRP Advantages: Low mass and high strength, available in all sizes. Disadvantages: Poor flexibility, can only be joined by cement/ field layup or mechanical joints, difficult to add branch line piping to, cemented joints must be kept clean and dry and may be slow to cure at low ambient temperatures, unfamiliar material to many workforces. Point of leakage may not be obvious. Standards: AWWA C950, ASTM D2996 e1. • PVC Advantages: Low mass, low cost, available in sizes up to 1.2 m. Disadvantages: Low strength and poor flexibility, loses strength very quickly at elevated temperatures and becomes brittle at low temperatures, can only be joined by cement or mechanical joints, cemented joints must be kept clean and dry, difficult to add branch line piping to. Water hammer will fracture piping. Requires thrust blocks and has lower velocity limits. Standards: AWWA C900, AWWA C905, ASTM D1785, ASTM D2241. • PE and HDPE

An aboveground system consists of a distribution pipe, insulation that surrounds the pipe, and a protective jacket that surrounds the insulation. In outdoor applications, the protective jacket should be metal with an 0.08 mm thick polysurlyn moisture barrier heatlaminated to the inside surface. Note that this moisture barrier is used to reduce the likelihood of corrosion on the interior surface of the jacket, is not the same as a vapor retarder, and does not replace a vapor retarder on cold systems. When the distribution system carries chilled water or other cold media, a vapor retarder is required for all types of insulation except cellular glass. In an ASHRAE test by Chyu et al. (1998a), cellular glass absorbed essentially zero water in a chilled-water application. In heating applications, the vapor retarder is not needed nor recommended; however, a reasonably watertight jacket is required to keep stormwater out of the insulation. The jacket material can be aluminum, stainless steel, galvanized steel, other metals, plastic sheet, a multilayered fabric and organic cement composite, or a combination of these. In outdoor applications, only metal or ultraviolet-resistant fabric and organic cement composites should be used. Structural columns and supports are typically made of wood, steel, or concrete. A crossbar is often placed across the top of the column when more than one distribution pipe is supported from one column. Sidewalk and road crossings require an elaborate support structure to elevate the distribution piping above traffic. Pipe expansion and contraction is taken up in loops, elbows, and bends. Manufactured expansion joints may be used, but they are usually not recommended because of a shorter life or a higher frequency of required maintenance than the remainder of the system. Supports that attach the distribution pipes to the support columns are commercially available as described in MSS Standards SP-58 and SP-69. The distribution pipes should have welded joints. An aboveground system has the lowest first cost and is the easiest to inspect and maintain; therefore, it has the lowest life-cycle cost. It is the standard against which all other systems are compared. Its major drawbacks are its poor aesthetics, its safety hazard if struck

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Underground Systems An underground system solves the problems of aesthetics and exposure to vehicles of the aboveground system; however, burying a system causes other problems with materials, design, construction, and maintenance that have historically been difficult to solve. An underground heat distribution system is not a typical utility like gas, domestic water, and sanitary systems, and requires an order of magnitude more design effort and construction inspection accuracy when compared to other utility distribution projects. The thermal effects and difficulty of keeping the insulation dry make it much more difficult to design and construct when compared to systems operating near the ambient temperature. Underground systems cost almost 10 times as much to build, and require much more to operate and maintain. Heat distribution systems must be designed for zero leakage and must account for thermal expansion, degradation of material as a function of temperature, high- pressure and transient shock waves, heat loss restrictions, and accelerated corrosion. In the past, resolving one problem in underground systems often created a new, more serious problem that was not recognized until premature failure occurred. Segan and Chen (1984) describe the types of premature failures that may occur if this guidance is ignored. Common types of underground systems are the walk-through tunnel, concrete surface trench, deep-buried small tunnel, poured insulation envelope, rigid closed-cell insulation, and conduit system. Walk-Through Tunnel. This system (Figure 14) consists of a field-erected tunnel large enough for a person to walk through after the distribution pipes are in place. It is essentially an aboveground system enclosed with a tunnel. The tunnel is buried deep enough to cover the top with earth, and is large enough for routine maintenance and inspection to be done easily without excavation. The preferred construction material for the tunnel walls and top cover is reinforced concrete. Masonry units and metal preformed sections have been used to construct the tunnel and top with less success, because of groundwater leakage and metal corrosion. The distribu-

Fig. 14

Walk-Through Tunnel

12.29 tion pipes are supported from the tunnel wall or floor with pipe supports that are commonly used on aboveground systems or in buildings. Some groundwater will penetrate the top and walls of the tunnel; therefore, a water drainage system must be provided. Usually, electric lights and electric service outlets are provided for ease of inspection and maintenance. This system has the highest first cost of all underground systems; however, it can have the lowest lifecycle cost because of its ease of maintenance, the ability to correct construction errors easily, and an extremely long life. If steam or HTHW piping is located in the tunnel, ambient temperatures may become extreme; such tunnels are typically either forced ventilated or gravity ventilated. Additional insulation may be required for chilled-water piping that shares the tunnel with steam or HTHW piping, because of the higher ambient temperatures. Selected pipe insulation material should be coupled with a high-performance vapor retarder and possibly a protective jacket. Shallow Concrete Surface Trench. This system (Figure 15) is a partially buried system. The floor is usually about 1 m below the surface grade. It is only wide enough for the carrier pipes and the pipe insulation plus some additional width to allow for pipe movement and possibly enough room for a person to stand on the floor. The trench usually is about as wide as it is deep. The top is constructed of reinforced concrete covers that protrude slightly above the surface and may also serve as a sidewalk. The floor and walls are usually cast-in-place reinforced concrete and the top is either precast or cast-in-place concrete. Precast concrete floor and wall sections have not been successful because of the large number of oblique joints and nonstandard sections required to follow the surface topography and to slope the floor for drainage. This system is designed to handle stormwater and groundwater that enters the system, so the floor is always sloped toward a drainage point, which typically is a steam trap pit, valve vault, or access port where a sump pump or other positive-drainage method is provided. Cross beams that attach to the side walls are preferred to support the carrier pipes. This keeps the floor free of obstacles that would interfere with drainage and allows the distribution pipes to be assembled before lowering them on the pipe supports. Also, floormounted pipe supports tend to corrode.

Fig. 15

Concrete Surface Trench

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The carrier pipes, pipe supports, expansion loops and bends, and insulation jacket are similar to aboveground systems, with the exception of pipe insulation. Experience with these systems indicates that flooding will occur several times during their design life; therefore, the insulation must be able to survive flooding and boiling and then return to near its original thermal efficiency. The pipe insulation is covered with a metal or plastic jacket to protect the insulation from abuse and from stormwater that enters at the top cover butt joints. Small inspection ports of about 300 mm diameter may be cast into the top covers at key locations so the system can be inspected without removing the top covers. All replaceable elements, such as valves, condensate pumps, steam traps, strainers, sump pumps, and meters, are located in valve vaults. The first cost of this system is among the lowest for underground systems because it uses typical construction techniques and materials. The life-cycle cost is often the lowest because it is easy to maintain, correct construction deficiencies, and repair leaks. Deep-Bury Tunnel. The tunnel in this system (Figure 16) is only large enough to contain the distribution piping, pipe insulation, and pipe supports. One type of deep-bury tunnel is the shallow concrete surface trench covered with earth and sloped independent of the topography. Because the system is covered with earth, it is essentially not maintainable between valve vaults without major excavation. All details of this system must be designed and all materials must be specified by the project design engineer. Because this system is not maintainable between valve vaults, great care must be taken to select materials that will last for the intended life and to ensure that the groundwater drainage system will function reliably. This system is intended to be used on sites where the groundwater elevation is typically lower than the bottom of the tunnel. The system can tolerate some groundwater saturation, depending on the watertightness of the construction and the capacity and reliability of the internal drainage. But even in desert areas, storms occur that expose underground systems to flooding; therefore, as with other types of underground systems, it must be designed to handle groundwater or stormwater that enters the system. The distribution pipe insulation must be of the type that can withstand flooding and boiling and still retain its thermal efficiency. Construction of this system is typically started in an excavated trench by pouring a cast-in-place concrete base that is sloped so intruding groundwater can drain to the valve vaults. The slope selected must also be compatible with the pipe slope requirements of the distribution system. The concrete base may have provisions for the supports for the distribution pipes, the groundwater drainage system, and the mating surface for the side walls. The side walls

Fig. 16

Deep-Bury Small Tunnel (Boxed Conduit)

may have provisions for the pipe supports if the pipes are not bottom supported. If the upper portion is to have cast-in-place concrete walls, the bottom may have reinforcing steel for the walls protruding upward. The pipe supports, distribution pipes, and pipe insulation are all installed before the top cover is installed. The groundwater drainage system may be a trough formed into the concrete bottom, a sanitary drainage pipe cast into the concrete bottom, or a sanitary pipe that is located slightly below the concrete base. The cover for the system is typically either of cast-in-place concrete or preformed sections such as precast concrete sections or half-round clay tile sections. The top covers must mate to the bottom and each other as tightly as possible to limit entry of groundwater. After the covers are installed, the system is covered with earth to match the existing topography. Poured Insulation. This system (Figure 17) is buried with the distribution system pipes encased in an envelope of insulating material and the insulation envelope covered with a thick layer of earth as required to match existing topography. This system is used on sites where the groundwater is typically well below the piping system. Like other underground systems, experience indicates that it will be flooded because the soil will become saturated with water several times during the design life; therefore, the design must accommodate flooding. The insulation material serves several functions. It may support the distribution pipes as bedding and backfill, with additional support as recommended by the product manufacturer, and it must support earth loads. The insulation must prevent groundwater from entering the interior of the envelope, and it must have long-term resistance to physical breakdown caused by heat and water. The insulation envelope must allow the distribution pipes to expand and contract axially as the pipes change temperature. In elbows, expansion loops, and bends, the insulation must allow formed cavities for lateral movement of the pipes, or be able to migrate around the pipe without significant distortion of the insulation envelope while still retaining the required structural load-carrying capacity. Pay special attention to corrosion of metal parts and water penetration at anchors and structural supports that penetrate the insulation envelope. Hot distribution pipes tend to drive moisture out of the insulation as steam; however, pipes used to distribute a cooling medium tend to condense water in the insulation, which reduces the insulation’s thermal resistance. A groundwater drainage system may be required, depending on the insulation material selected and severity of the groundwater; however, if such drainage is needed, it is a strong indicator that this is not the proper system for the site conditions.

Fig. 17

Poured Insulation System

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District Heating and Cooling This system is constructed by excavating a trench with a bottom slope that matches the desired slope of the distribution piping. The width of the bottom of the trench is usually the same as the width of the insulation envelope because it serves as a form. The distribution piping is then assembled in the trench and supported at the anchors and by blocks that are removed as the insulation is poured in place. The form for the insulation can be the trench bottom and sides, wooden forms, or sheets of plastic, depending on the type of insulation used and the site conditions. The insulation envelope is covered with earth to complete the installation. The project design engineer is responsible for finding an insulation material that fulfills all of the previously mentioned requirements. At present, no standards have been developed for insulation used in this type of application. Hydrophobic powders, which are a special type of pulverized rock treated to be water repellent, have been used successfully. The hydrophobic characteristic of this powder prevents water from dampening the powder and has some ability as a barrier for preventing water from entering the insulation envelope. This insulating powder typically has a much higher thermal conductivity than mineral wool or fiberglass pipe insulation; therefore, the thickness of the poured envelope must be significantly greater. In addition, Phetteplace et al. (1998) found installed heat losses to be much higher than would be expected using manufacturer’s data for one poured insulation material, and that actual installed dimensions were less than manufacturer’s recommendations in the majority of installations. This may have been the result of the design or construction errors, contractors purposely “shorting” on dimensions, or compaction during or after backfilling. The user is cautioned to verify dimensions during construction and also use appropriate in-place densities for the poured insulation material; do not use bulk or loose density. Measure installed density by a meaningful test for noncohesive soils (e.g., ASTM Standard D4253). Furthermore, thermal conductivity is a strong function of density and thus must also be measured at realistic in-place densities. As with any system, the designer should analyze the cost of the alternatives, including preinsulated (i.e., conduit) or field-insulated piping. Field-Installed, Direct-Buried Rigid Closed-Cell Insulation. In this system (Figure 18), a rigid closed-cell insulation is covered with an asphaltic jacket. The insulation supports the pipe. Oversized loops with internal support elements provide for expansion. The project design engineer must make provisions for pipe movement in the expansion loops. As shown in Figure 18, a drain should be installed to drain groundwater away. A waterproof jacket is recommended for all buried applications.

Fig. 18 Field-Installed, Direct-Buried Rigid Closed-Cell Insulation System

12.31 When used for heating applications, the dry soil condition must be investigated to determine if the temperature of the jacket exceeds the material’s allowable temperature (see the section on Methods of Heat Transfer Analysis). This is one of the controlling conditions to determine how high the carrier pipe fluid temperature can be without exceeding the temperature limits of the jacket material. Insulation thickness depends on the thermal operating parameters of the carrier pipe. For maximum system integrity under extreme operating conditions, such as groundwater flooding, the jacketing may be applied to the insulation segments in the fabrication shop with hot asphalt. As with other underground systems, experience indicates that this system may be flooded several times during its projected life.

Conduits The term conduit denotes an entire assembly, which consists of a carrier pipe, the pipe insulation, the casing, and the exterior casing coating (Figure 19). The conduit is assembled in a factory and shipped as unit called a conduit section. The pipe that carries the working medium is called a carrier pipe and the outermost perimeter enclosure is called a casing. Each conduit section is shipped in lengths up to 12.2 m. Elbows, tees, loops, and bends are factory prefabricated to match the straight sections. The prefabricated components are assembled at the construction site; therefore, a construction contract is typically required for trenching, backfilling, connecting to buildings, connecting to distribution systems, constructing valve vaults, and performing some electrical work associated with sump pumps, power receptacles, and lights. Much of the design work is done by the factory that manufactures the prefabricated sections; however, the field work must be designed and specified by the project design engineer or architect. Prefabricated components create a serious problem with accountability. For comparison, when systems are entirely field assembled, the design responsibility clearly belongs to the project design engineer, and system assembly and installation are clearly the responsibility of the construction contractor. When a condition arises where a conduit system cannot be built without modifying prefabricated components, or if the construction contractor does not follow the instructions from the prefabricator, a serious conflict of responsibility arises. The design engineer, as the engineer of record (EOR), assumes the responsibility of review and approval of any design work by others. For these reasons, it is imperative that the project design engineer or architect clearly delineate the responsibilities of the factory prefabricator.

Fig. 19 Conduit System Components

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Crushing loads have been used (erroneously) to size the casing thickness, assuming that corrosion was not a factor. However, corrosion rate is usually the controlling factor because the casing temperature can range from less than 40°C to more than 150°C, a range that encompasses the maximum corrosion rate of steel (Figure 20). As shown in the figure, the steel casing of a district heating pipe experiences corrosion rates several times that of domestic water pipes. The casing temperature varies with burial depth, soil conditions, carrier pipe temperature, and pipe insulation thickness. The casing must be strong and thick enough to withstand expansion and contraction forces and corrosion degradation. All insulation must be kept dry for it to maintain its thermal insulating properties; the exception is cellular glass in cold applications. Because underground systems may be flooded several times during their design life, even on sites that are thought to be dry, a reliable water intrusion removal system is necessary in the valve vaults. Two designs are used to ensure that the insulation performs satisfactorily for the life of the system. In the air space system, an annular air space between the pipe insulation and the casing allows the insulation to be dried out if water enters. In the water spread limiting (WSL) system, which has no air space, the conduit is designed to keep water from entering the insulation. If water enters one section, a WSL system slows or prevents its spread to adjacent sections of piping. The air space conduit system (Figures 21 and 22) should have an insulation that can survive short-term flooding without damage. The

Fig. 20 Corrosion Rate in Aggressive Environment Similar to Mild Steel Casings in Soil

Fig. 21 Conduit System with Annular Air Space and Single Carrier Pipe

conduit manufacturer usually runs a boiling test with the insulation installed in the typical factory casing. No U.S. standard has been approved for this boiling test; however, the U.S. government uses a Federal Agency Committee 96 h boiling test for conduit insulation [see Phetteplace et al. (2013b) for protocol]. The insulation must have demonstrated that it can be dried with air flowing through an annular air space, and it must retain nearly new thermal insulating properties when dried. Insulation fails because the bonding agents, called binders, that hold the principal insulation material in the desired shape degrade. The annular air space around the insulation, typically more than 25 mm wide, allows air to flow outside the insulation to dry it. Unfortunately, the air space has a serious detrimental side effect: it allows unwanted water to flow freely to other parts of the system. The WSL system (Figure 23) encloses the insulation in an envelope that will not allow water to contact the insulation. The typical insulation is polyurethane foam, which will be ruined if excess water infiltration occurs. Polyurethane foam is limited to a temperature of about 120°C for a service life of 30 years or more. Europe has the most successful of the WSL systems, which are typically used in low-temperature water applications. These systems, available in the United States as well, meet CEN European Standard EN 253 with regard to all major construction details. Standards also have been established for fittings (EN 448), preinsulated valves (EN 488), field joint assemblies (EN 489), and system design (EN 13941).

Fig. 22

Conduit System with Two Carrier Pipes and Annular Air Space

Fig. 23 Conduit System with Single Carrier Pipe and No Air Space (WSL)

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Fig. 24

12.33

Conduit Casing Temperature Versus Soil Thermal Conductivity

With this system, the carrier pipe, insulation, and casing are bonded together to form a single unit. Forces caused by thermal expansion are passed as shear forces to the mating component and ultimately to the soil; thus, no additional expansion provisions are required. Movement at changes in direction (elbows) may be accommodated for by cushions. This type of WSL system is feasible because of the small temperature differential and because the thinner carrier pipe wall creates smaller expansion forces because of its low cross-sectional area. Some systems rely on a watertight field joint between joining casing sections to extend the envelope to a distant envelope termination point where the casing is sealed to the carrier pipe. Other systems form the waterproof insulation envelope in each individual prefabricated conduit section, using the casing and carrier pipe to form part of the envelope and a waterproof bulkhead to seal the casing to the carrier pipe. In another type of construction, a second pipe fits tightly over the carrier pipe and seals the insulation between the second pipe and casing to achieve a watertight insulation envelope. Conduit Design Conditions. The following three design conditions must be addressed to have reasonable assurance that the system selected will have a satisfactory service life: • Maximum heat loss occurs when the soil is wettest and the conduit is shallowly buried (minimum burial depth), usually with about 0.6 m of earth cover. This condition represents the highest gross heat transfer and is used to size the distribution piping and equipment in the central heating plant. For heating piping, because the casing is coldest during start-up, the relative movement of the carrier pipe with respect to the casing may be maximum during this condition. • A dry-soil condition may occur when the conduit is buried deep. The soil plays a more significant role in the heat transfer than the pipe insulation because of the soil’s thickness (and thus, insulating value). The highest temperature of the insulation, casing, and casing coating occur during this condition. Paradoxically, the minimum heat loss occurs during this condition because the soil acts as a good insulator. This condition is used to select temperature-sensitive materials and to design for casing expansion. The relative movement of the carrier pipe with respect to the casing may be at minimum during this condition; however, if the casing is not restrained, its movement with respect to the soil will be at maximum. If restrained, the casing axial stresses and axial forces will

be highest and the casing allowable stresses will be lowest because of the high casing temperature. Buckling of the casing is possible in extreme situations. Figure 24A shows the effect of burial depth on casing temperature as a function of soil thermal conductivity for a typical system. Figure 24B shows the effect of insulation thickness on casing temperature, again as a function of soil thermal conductivity. Analysis of Figure 24A and 24B suggests some design solutions that could lower the effects of the dry-soil condition: reducing carrier pipe temperature, using thicker carrier pipe insulation, providing a device to keep the soil wet, or minimizing burial depth. However, if these solutions are not feasible or cost effective, a different type of material or an alternative system should be considered. Although it is possible that the soil will never dry out, given the variability of climate in most areas, it is likely that a drought will occur during the life of the system. Only one very dry condition can cause permanent damage to the insulation or other system components if the soil thermal conductivity drops below the assumed design value. On-site measurement of the driest soil condition likely to cause insulation damage is not feasible. As a result, the designer is left with the conservative choice of using the lowest thermal conductivity from Table 9 to calculate the highest temperature to be used to select materials. • The nominal or average condition occurs when the soil is at its average water content and the conduit is buried at the average depth. This condition is used to compute the yearly energy consumption from heat loss to the soil (see Table 9 for soil thermal conductivity and the previous section on Thermal Properties).

Cathodic Protection of Direct-Buried Conduits Corrosion is an electrochemical process that occurs when a corrosion cell is formed. A corrosion cell consists of an anode, a cathode, a connecting path between them, and an electrolyte (soil or water). The structure of this cell is the same as a dry-cell battery, and, like a battery, it produces a direct electrical current. The anode and cathode in the cell may be dissimilar metals, and because of differences in their natural electrical potentials, a current flows from anode to cathode. When current leaves an anode, it destroys the anode material at that point. The anode and cathode may also be the same material. Differences in composition, environment, temperature, stress, or shape makes one section anodic and an adjacent

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2020 ASHRAE Handbook—HVAC Systems and Equipment (SI)

section cathodic. With a connection path and the presence of an electrolyte, this combination also generates a direct electrical current and causes corrosion at the anodic area. If steel or ductile iron piping is directly buried, it must be coated or insulated to reduce the potential for corrosion. For ductile iron piping, polyethylene sleeves are available and provide some degree of corrosion protection, although not to the same degree as a continuous-coated welded-steel piping system. In addition to any coating or protective sleeve systems, it may be necessary or advisable to provide a cathodic protection system for uninsulated steel or ductile iron piping. Cathodic protection is a standard method used by the underground pipeline industry to further protect coated steel against corrosion but is not required for nonmetallic systems or preinsulated systems with a nonmetallic jacket. Cathodic protection systems are routinely designed for a minimum life of 20 years. Cathodic protection may be achieved by the sacrificial anode method or the impressed current method. Sacrificial anode systems are normally used with well-coated structures. A direct current is applied to the outer surface of the steel structure with a potential driving force that prevents the current from leaving the steel structure. This potential is created by connecting the steel structure to another metal, such as magnesium, aluminum, or zinc, which becomes the anode and forces the steel structure to be the cathode. The moist soil acts as the electrolyte. These deliberately connected materials become the sacrificial anode and corrode. If they generate sufficient current, they adequately protect the coated structure, and their low current output is not apt to corrode other metallic structures in the vicinity. Impressed current systems use a rectifier to convert an alternating current power source to usable direct current. The current is distributed to the metallic structure to be protected through relatively inert anodes such as graphite or high-silicon cast iron. The rectifier allows the current to be adjusted over the life of the system. Impressed current systems, also called rectified systems, are used on long pipelines in existing systems with insufficient coatings, on marine facilities, and on any structure where current requirements are high. They are installed selectively in congested pipe areas to ensure that other buried metallic structures are not damaged. The design of effective cathodic protection requires information on the diameter of both the carrier pipe and conduit casing, length of run, number of conduits in a common trench, and number of system terminations in access areas, buildings, etc. Soil from the construction area should be analyzed to determine the soil resistivity, or the ease with which current flows through the soil. Areas of low soil resistivity require fewer anodes to generate the required cathodic protection current, but the life of the system depends on the mass of anode material used. The design life expectancy of the cathodic protection must also be defined. All anode material is theoretically used up at the end of the cathodic protection system life. At this point, the corrosion cell reverts to the unprotected system and corrosion occurs at points along the conduit system or buried metallic structure. Anodes may be replaced or added periodically to continue the cathodic protection and increase the conduit life. A cathodically protected system must be electrically isolated at all points where the pipe is connected to building or access port piping and where a new system is connected to an existing system. Conduits are generally tied to another building or access piping with flanged connections. Flange isolation kits, including dielectric gaskets, washers, and bolt sleeves, electrically isolate the cathodically protected structure. If an isolation flange is not used, any connecting piping or metallic structure will be in the protection system, but protection may not be adequate. The effectiveness of cathodic protection can only be determined by an installation survey after the system has been energized. Cathodically protected structures should be tested at regular inter-

vals to determine the continued effectiveness and life expectancy of the system. Sacrificial anode cathodic protection is monitored by measuring the potential (voltage) between the underground metallic structure and the soil versus a stable reference. This potential is measured with a high resistance voltmeter and a reference cell. The most commonly used reference cell material is copper/copper sulfate. One criterion for protection of buried steel structures is a negative voltage of at least 0.85 V as measured between the structure surface and a saturated copper/copper sulfate reference electrode in contact with the electrolyte. Impressed current systems require more frequent and detailed monitoring than sacrificial anode systems. The rectified current and potential output and operation must be verified and recorded at monthly intervals. NACE Standard RPO169 has further information on control of external corrosion on buried metallic structures.

Leak Detection The conduit may require excavation to repair construction errors after burial. Various techniques are available for detecting leaks in district heating and cooling piping. They range from performing periodic pressure tests on the piping system to installing a sensor cable or wire along the entire length of the piping to continuously detect and locate leaks. Most sensor cable (located in the air gap of the conduit) and wire (located within the insulation) leak detection systems can be connected to a building automation system and monitored 24 h/day and alarm when moisture is detected. Pressure testing should be performed on all piping to verify integrity during installation and the life of the piping. Chilled-water systems should be pressure tested during the winter, and hot-water and steam systems tested during the summer. A leak is difficult to locate without the aid of a cable type leak detector. Finding a leak typically involves excavating major sections between valve vaults. Infrared detectors and acoustic detectors can help narrow down the location of a leak, but they do not work equally well for all underground systems. Also, they are not as accurate with underground systems as with an aboveground system. Chilled- and Hot-Water Systems. Chilled-water piping systems are usually insulated with urethane foam with a vaporproof jacket (HDPE, urethane, PVC, CPVC, etc.). Copper wires can be installed during fabrication to aid in detecting and locating liquid leaks. The wires may be insulated or uninsulated, depending on the manufacturer. Some systems monitor the entire wire length, whereas others only monitor at the joints of the piping system. The detectors either look for a short in the circuit using Ohm’s law or monitor for impedance change using time domain reflectometry (TDR). Steam, High-Temperature Hot-Water, and Other Conduit Systems. Air gap designs, which have a gap between the inner wall of the outer casing and the insulation, can have probes installed at the low points of drains or at various points to detect leaks. Leaks can also be detected with a continuous cable that monitors liquid leakage. The cable is installed at the bottom of the conduit with a minimum air gap required, typically 25 mm. Pull points or access ports are installed every 120 to 150 m on straight runs, with changes in direction reducing the length between pull points. Systems monitor either by looking for a short on the cable using Ohm’s law or by sensing the impedance on a coaxial cable using resistance temperature devices (RTDs). During installation, care must be taken to keep the system clean and dry to keep any contamination from the leak detection system that might cause it to fail. The system must be sealed airtight to prevent condensation from accumulating in the piping at the low points.

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Geotechnical Considerations Underground district heating or cooling distribution systems have more stringent burial requirements than most other building utilities. These piping systems normally have the coatings or jackets needed for corrosion protection, or insulation protection that must not be damaged by rocks, debris, or construction equipment. Thus, proper burial conditions must be established for the district heating or cooling distribution system to achieve its design life. Requirements vary, and manufacturers of the piping system should be able to provide guidance specific to their system. It is the EOR’s responsibility to ensure that these requirements are included in the contract documents. Preferably, a licensed geotechnical engineer familiar with local conditions should conduct a site survey before construction, to recommend any soil testing required and to develop the specifications for excavation and backfill. The geotechnical and structural engineer should also be responsible for designing any thrust blocks or anchors that are needed, based on forces provided by the EOR. In general, trenches must be overexcavated by at least 100 mm in depth to remove any unyielding material; overexcavation may need to be greater at the locations of the field joints, depending on the type of system and construction method. The overexcavation is generally filled with a select backfill material; normally, this is a sandy, noncohesive material free of any stones greater than 19 mm. Any unstable materials encountered in the excavation should be removed and properly backfilled and compacted. The selected backfill in the trench bottom should be prepared to achieve the minimum slope for the carrier pipe of 10 mm in 2.4 m (0.4%) and compacted to 95% of laboratory maximum density per ASTM Standard D698. Some of the methods of carrier pipe joining, such as welding, require a working area around the entire circumference of the field joint; one way to achieve this is to overexcavate under the pipe, and potentially even at the sides of the trench at the locations of the field joints. If this is done, be sure to fully compact the backfill material under the field joint area. Another method to provide working clearance for making the field joint is to block the piping up off the bottom of the trench during that process. When this method is used, care must be taken to block the pipe sufficiently to achieve proper alignment for joining and to emulate the pipe as it will ultimately lie on the sloped trench bottom. Once the field joints have been completed, remove the blocking and carefully and uniformly place the piping in the prepared trench bottom. The blocking should not be left in place, because this creates point loading on the piping and may contribute to differential settlement as well. In some situations when welded-steel piping is used, for example, it may be possible to join two sections of piping together adjacent to the trench and then lift the assembly in as a unit and thus reduce the work required in the trench. The elevation of the trench bottom must not have slope reversals between valve vaults and building entry locations. After the piping is placed in the trench and all field joints and pressure tests have been completed, immediately before backfilling, the elevation of the top invert of the pipe/jacket should be taken at each pipe section midpoint and field joint. These elevations should be recorded and subsequently transferred to the as-built drawings. Backfill of the piping should then be accomplished in layers of no more than 150 mm with the same select backfill material used for the pipe bedding. The selected backfill should be extended to approximately 300 mm above the top of the pipe or jacketing. Include buried utility warning tape in the trench at this depth. Compaction of this backfill material should also be to 95% of laboratory maximum density per ASTM Standard D698. Ensure that the backfill adequately fills the void created under the pipe and between the supply and return pipes. Also, take care not to damage any pipe coating or insulation jacketing material; if any such damage does occur, repair it according to

12.35 the pipe system manufacturer’s field repair instructions. Final backfill of the remainder of the trench should be accomplished using the native soil (removing any stones greater than75 mm), compacted in layers of no more than 150 mm. This final backfill should be compacted to 95% of laboratory maximum density per ASTM Standard D698 for noncohesive soils, or 90% of laboratory maximum density per ASTM Standard D698 for cohesive soils. Note that it is not advisable to complete the final backfilling with anything other than native soil, because the permeability of other substances may be much different than that of the native material. For example, using a permeable backfill material in an impermeable native soil is essentially placing the district heating or cooling system in a drainage ditch for surface water. Also, note that horizontal boring, jacking, and microtunneling have become popular methods of installing buried pipelines where the normal cut-and-cover methods described previously may be difficult or impossible, or simply cost prohibitive. These alternative burial methods preclude the use of protective backfilling, but often may not need such steps. However, metallic pipelines must still be protected from corrosion, and methods appropriate for the installation method must be used.

Valve Vaults and Entry Pits Manholes or valve vaults may be required on underground distribution systems to provide access to underground systems at critical points, such as high or low points that vent trapped air or where the system can be drained; where elevation changes in the distribution system are needed to maintain the required constant or nonreversing slope (a slope of 1:240 is a reasonable and achievable construction value); and to access isolation valves at major distribution system branches. Phetteplace et al. (2013b) contains a detailed discussion on valve vaults. The term vault is used to eliminate confusion with sanitary manholes. Valve vaults are important when the underground distribution system cannot be maintained without excavation, as is the case for all direct-buried district cooling piping systems. They allow for step elevation changes in the distribution system piping, while maintaining an acceptable slope on the system; they also allow the designer to better match the topography and avoid unreasonable and expensive burial depths. However, they also have a significant potential to generate maintenance requirements of their own, and any problems, if left uncorrected, can cascade into the adjacent buried portions of the distribution system. Valve vaults allow a user to isolate one segment of a system rather than analyze the entire system or a large section thereof. Isolation may be for routine maintenance as well as for operational problems or failures. This feature is important if the underground distribution system cannot be maintained between valve vaults without excavation. The optimum number of valve vaults is that which affords the lowest life-cycle cost and still meets all design requirements, typically around 204 m apart but usually no more than 150 m apart in the absence of other requirements, such as isolation valves for a building service. Valve vaults provide a space in which to put valves, steam traps, carrier pipe drains, carrier pipe vents, casing vents and drains, condensate pumping units, condensate cooling devices, flash tanks, expansion joints, groundwater drains, electrical leak detection equipment and wiring, electrical isolation couplings, branch line isolation valves, carrier pipe isolation valves, and flowmeters. Valve vaults allow for elevation changes in the distribution system piping while maintaining an acceptable slope on the system; they also allow the designer to better match the topography and avoid unreasonable and expensive burial depths. Penetrations. One of the basic functions of a valve vault is to provide a dry, corrosion-free environment for the piping and appurtenances in the valve vault. This means that the vault walls must not let groundwater enter the vault. Typical types of penetrations of the

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2020 ASHRAE Handbook—HVAC Systems and Equipment (SI)

vault walls are the district cooling piping, electrical service conduits, sump pump discharge pipes, and sanitary drains. All penetrations must have a method to provide a positive water seal between the vault wall and the pipe or electrical conduit. Often, a leak plate is welded to a steel sleeve that is cast into the concrete vault wall. Existing vault walls or precast valve vaults that do not have the leak plate cast into the wall may still be sealed acceptably, as long as the surface of the wall penetration is smooth and an adequate seal can be achieved. The annular space between the outside of the piping/conduit and the valve vault wall penetration is typically sealed with a link seal or other type of adjustable compressed rubber seal. These rubber seals work poorly where construction quality control is poor, because the holes in the vault wall are often the wrong diameter or are irregular. Piping that penetrates the vault wall at an angle causes especially difficult sealing problems for precast construction when the hole in the vault wall is not perpendicular to the conduit. In addition, for some nonmetallic systems, the casing of the district heating or cooling pipe tends to deflect when the seal exerts the radial compressive load it needs to achieve the water seal. When plastic piping or casings are used, a special design is needed to prevent radial deflection, or some other type of sealing method is required. Ponding Water. The most significant problem with valve vaults is that water ponds in them. Ponding water may be from either carrier pipe leaks or intrusion of surface or groundwater into the valve vault. When the hot- and chilled-water distribution systems share the same valve vault, plastic chilled-water lines often fail because ponded water heats the plastic to failure. Water gathers in the valve vault irrespective of climate; therefore, design strives to eliminate the water for the entire life of the underground distribution system. If the system depth allows, the most successful water removal systems are those that drain to sanitary or storm drainage systems; this technique is successful because the system is affected very little by corrosion and has no moving parts to fail. Backwater valves are recommended in the event of the drainage system backing up. Duplex sump pumps with lead-lag controllers and a failure annunciation system are used when storm drains and sanitary drains are not accessible or the collection point is below adjacent sewers. Because pumps have a history of frequent failures, duplex pumps help eliminate short cycling and provide standby pumping capacity. Steam ejector pumps can be used only if the distribution system is never shut down because the carrier pipe insulation can be severely damaged during even short outages. A labeled, lockable, dedicated electrical service should be used for electric pumps. The circuit label should indicate what the circuit is used for; it should also warn of the damage that will occur if the circuit is deenergized. Electrical components have experienced accelerated corrosion in the high heat and humidity of closed, unventilated vaults. A pump that works well at 10°C often performs poorly at 95°C and 100% rh. To resolve this problem, one approach specifies components that have demonstrated high reliability at 95°C and 100% rh with a damp-proof electrical service. The pump should have a corrosionresistant shaft (when immersed in water) and impeller and have demonstrated 200 000 cycles of successful operation, including the electrical switching components, at the referenced temperature and humidity. The pump must also pass foreign matter; therefore, the ability to pass a 10 mm ball should be specified. Another method drains the valve vaults into a separate sump adjacent to the vault. Then the pumps are placed in this sump, which is cool and more nearly a sump pump environment. Redundant methods may be necessary if maximum reliability is needed or future maintenance is questionable. The pump can discharge to the sanitary or storm drain or to a splash block near the valve vault. Water pumped to a splash block has a tendency to enter the vault, but this is not a significant problem if the vault construction joints have been sealed properly. Extreme caution must be used if the bottom of the valve vault has French drains, which allow groundwater to enter

the vault and flood the insulation on the distribution system during high-groundwater conditions. Adequate ventilation of the valve vault is also important. Crowding of Components. The valve vault must be laid out in three dimensions, considering standing room for the worker, wrench swings, the size of valve operators, variation between manufacturers in the size of appurtenances, and all other variations that the specifications allow with respect to any item placed in the vault. To achieve desired results, the vault layout must be shown to scale on the contract drawings. High Humidity. High humidity develops in a valve vault when it has no positive ventilation. Gravity ventilation is often provided, in which cool air enters the valve vault and sinks to the bottom. At the bottom of the vault, the air warms, becomes lighter, and rises to the top of the vault, where it exits. In the past, some designers used a closed-top valve vault with an exterior ventilation pipe with an elbow that directs the exiting air down. However, the elbowed-down vent hood tends to trap the exiting air and prevent gravity ventilation from working. Open structural grate tops are the most successful covers for ventilation purposes. Open grates allow rain to enter the vault; however, the techniques mentioned in the section on Ponding Water are sufficient to handle the rainwater. Open grates with sump basins have worked well in extremely cold climates and in warm climates. Some vaults have a closed top and screened, elevated sides to allow free ventilation. In this design, the solid vault sides extend slightly above grade; then, a screened window is placed in the wall on at least two sides. The overall above-grade height may be only 450 mm. High Temperatures. The temperature in the valve vault rises when no systematic way is provided to remove heat losses from the distribution system. The gravity ventilation rate is usually not sufficient to transport heat from the closed vaults. Part of this heat transfers to the earth; however, an equilibrium temperature is reached that may be higher than desired. Ventilation techniques discussed in the section on High Humidity can resolve the problem of high temperature if the heat loss from the distribution system is near normal. Typical problems that greatly increase the amount of heat released include • • • •

Leaks from a carrier pipe, gaskets, packings, or appurtenances Insulation that has deteriorated because of flooding or abuse Standing water in a vault that touches the distribution pipe Steam vented to the vault from partial flooding between valve vaults • Vents from flash tanks • Insulation removed during routine maintenance and not replaced To prevent heat release in a new system, a workable ventilation system must be designed. On existing valve vaults, the valve vault must be ventilated properly, all leaks corrected, and all insulation that was damaged or left off replaced. Commercially available insulation jackets that can easily be removed and reinstalled from fittings and valves should be installed. If flooding occurs between valve vaults, portions of the distribution system may have to be excavated and repaired or replaced. Vents from vault appurtenances that exhaust steam into the vault may have to be routed aboveground if the ventilation technique is insufficient to handle the quantity of steam exhausted. Deep Burial. When a valve vault is buried too deeply, (1) the structure is exposed to groundwater pressures, (2) entry and exit often become a safety problem, (3) construction becomes more difficult, and (4) the cost of the vault is greatly increased. Ideally, valve vault spacing should be less than 150 m (NAS 1975). If greater spacing is desired, use an accurate life-cycle analysis to determine spacing. The most common way to limit burial depth is to place the valve vaults closer together. Sawtooth-shaped steps in the distribution system slope are accommodated in the valve vault (i.e., the carrier pipes

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District Heating and Cooling enter the valve vault at a low elevation and leave at a higher elevation). If the slope of the distribution system is changed to more nearly match the earth topography, the valve vaults will be shallower; however, the allowable range of slope of the carrier pipes restricts this method. In most systems, the slope of the distribution system can be reversed in a valve vault, but not out in the system between valve vaults. The minimum slope for the carrier pipes is 4 mm/m. Lower slopes are outside the range of normal construction tolerance. If the entire distribution system is buried too deeply, the designer must determine the maximum allowable burial depth of the system and survey the topography of the distribution system to determine where the maximum and minimum depth of burial will occur. All elevations must be adjusted to limit the minimum and maximum allowable burial depths. Freezing Conditions. Failure of distribution systems caused by water freezing in components is common. The designer must consider the coldest temperature that may occur at a site and not the 99% or 99.6% condition used in building design (as discussed in Chapter 27 of the 2013 ASHRAE Handbook—Fundamentals). Drain legs or vent legs that allow water to stagnate are usually the cause of failure. Insulation should be on all items that can freeze, and it must be kept in good condition. Electrical heat tape and pipe-type heat tracing can be used under insulation. If part of a chilled-water system is in a ventilated valve vault, the chilled water may have to be circulated or be drained if not used in winter. Safety and Access. Some working fluids used in underground distribution systems can cause severe injury and death if accidentally released in a confined space such as a valve vault. The shallow valve vault with large openings is desirable because it allows personnel to escape quickly in an emergency. The layout of the pipes and appurtenances must allow easy access for maintenance without requiring maintenance personnel to crawl underneath or between other pipes. The goal of the designer is to keep clear work spaces for maintenance personnel so that they can work efficiently and, if necessary, exit quickly. Engineering drawings must show pipe insulation thickness; otherwise they will give a false impression of the available space. The location and type of ladder is important for safety and ease of egress. It is best to lay out the ladder and access openings when laying out the valve vault pipes and appurtenances as a method of exercising control over safety and ease of access. Ladder steps, when cast in the concrete vault walls, may corrode if not constructed of the correct material. Corrosion is most common in steel rungs. Either cast-iron or prefabricated, OSHA-approved, galvanized steel ladders that sit on the valve vault floor and are anchored near the top to hold it into position are best. If the design uses lockable access doors, the locks must be operable from inside or have some keyedopen device that allows workers to keep the key while working in the valve vault. Extremely deep vaults (greater than 6 m) may require an intermediate landing and caged ladder with fall protection to satisfy OSHA requirements. Vault Construction. The most successful valve vaults are those constructed of cast-in-place reinforced concrete. These vaults conform to the earth excavation profile and show little movement when backfilled properly. Leakproof connections can be made with mating tunnels and conduit casings, even though they may enter or leave at oblique angles. In contrast, prefabricated valve vaults may settle and move after construction is complete. Penetrations for prefabricated vaults, as well as the angles of entry and exit, are difficult to locate exactly. As a result, much of the work associated with penetrations is not detailed and must be done by construction workers in the field, which greatly lowers the quality and greatly increases the chances of a groundwater leak. Construction Deficiencies. Construction deficiencies that go unnoticed in buildings can destroy a cooling distribution system; therefore, the designer must clearly convey to the contractor that a valve vault does not behave like a sanitary manhole. A design that is

12.37 sufficient for a sanitary manhole will cause the appurtenances in district heating or cooling distribution system manholes to fail prematurely because many of the requirements mentioned previously are not provided. Construction of Systems Without Valve Vaults. Uninsulated systems have been designed where features such as valves are directly buried and then remotely operated, much the same as is frequently done for potable water distribution systems. For this approach to be successful for an insulated system, a continuous, high-integrity jacket system as well as the necessary seals for operable shafts, etc., must be provided. If the waterproofing is not entirely successful, groundwater will enter the system, causing corrosion and deterioration of the insulation’s thermal properties. Adequately waterproofing features (e.g., direct-burial valves, vents/ drains) is very difficult to do in the field. The standardized European preinsulated piping systems developed originally for low-temperature, hot-water systems are factory prefabricated.

4. CONSUMER INTERCONNECTIONS The thermal energy produced at the plant is transported via the distribution network and is finally transferred to the consumer. This end user interface connection is also commonly called an energy transfer station (ETS). When thermal energy (hot water, steam, or chilled water) is supplied, it may be used directly by the building HVAC system or process loads, or indirectly via a heat exchanger that transfers energy from one media to another. When energy is used directly, it may need to be reduced to pressure that is commensurate to the buildings’ systems. The design engineer must perform an analysis to determine which connection type is best. For commercially operated systems, a contract boundary or point of delivery divides responsibilities between the energy provider and the customer. This point can be at a piece of equipment, as in a heat exchanger with an indirect connection, or flanges as in a direct connection. It is highly recommended that a chemical treatment analysis be performed (regardless of the type of connection) to determine the compatibility of each side of the system (district and consumer) before energizing. Whether the connection is direct or indirect or heating or cooling, a cathodic isolation flange is recommended at the ETS to preserve system pipe integrity, especially when connecting to older buildings (Sperko 2009; Tredinnick 2008a). See Phetteplace et al. (2013a, 2013b) for additional information. There are traditionally two types of interfaces with a customer: direct or indirect connection. Advantages and disadvantages of each are summarized in Table 13. Chilled- and hot-water ETSs may be constant or variable flow, depending on the building’s HVAC system. However, variable flow is recommended for both chilled- and hot-water systems because of energy savings and the need to comply with current requirements of ASHRAE Standard 90.1. Typically, constant-flow systems are found in older buildings and may be converted to simulate a variable-flow system by blocking off the bypass line around the air handler heat exchanger coil three-way control valve. At low operating pressures, this potentially may convert a three-way bypass-type valve to a two-way modulating control valve. The valve actuator must be carefully analyzed, because shutoff requirements and control characteristics are totally different for a two-way valve compared to a three-way valve.

4.1

DIRECT CONNECTIONS

Because a direct connection offers no barrier between the district water and the building’s own system (e.g., air-handling unit cooling and heating coils, fan-coils, radiators, unit heaters, process loads), the water circulated at the district plant has the same quality as the

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2020 ASHRAE Handbook—HVAC Systems and Equipment (SI)

Fig. 25 Direct Connection of Building System to District Chilled Water with Building Pumps (Phetteplace et al. 2013a)

Table 13 Relative Merits of Direct and Indirect Consumer Interconnection Licensed for single user. © 2020 ASHRAE, Inc.

Issue

Direct Connection

Water quality

Indirect Connection

District system water is exposed to a building system that may have lower levels of treatment and filtering. Components in existing building systems may have scale and corrosion. Water consumption Leakage and consumption of district system water in building may be difficult to control and correct. Contractual Demarcation of consumer’s building system may not be clear. Cost Generally lower in overall cost because heat exchanger and (possibly) building pumps and controls are not needed. Reliability Failures in the building may cause problems or potentially even outages for the district system. Pressure isolation Building systems may need to be protected from higher pressure in a district system; for tall buildings, a district system may be subjected to higher pressures by the building system. t Potential for greater t because of absence of heat exchanger. In-building space requirements Low space requirements.

customer’s water (and vice versa). Direct connections, therefore, are at a greater risk of incurring damage or contamination based on the poor water quality of either party. Typically, district systems have contracts with water treatment vendors and monitor water quality continuously. This may not be the case with all consumers. A direct connection is often more economical than an indirect connection because the consumer is not burdened by the installation of heat exchangers, additional circulation pumps, or water treatment systems; therefore, investment costs are reduced and return temperatures identical to design values are possible. In general, consider using a direct connection under the following conditions: the building owner is the district system owner or they are related entities; control of first cost is important; buildings are generally low rise; building systems are new or in good condition; in-building space for interconnection is limited; and the building owner, if different than the district utility, respects the need for high t and will maintain the building systems accordingly and retrofit the building equipment where necessary to achieve adequate t. Figures 25 and 26 show a simple chilled-water direct connection using building circulation pumps and using the district cooling provider’s pumps, respectively. Figure 26’s method is preferred, because it is the simplest, with no control valve, but must have high return water temperatures at varying flows. Consequently, this method requires the building design engineer and controls contrac-

Water quality of the district system is isolated from building system and can be controlled. Water leakage is controlled by district heating utility. Clear delineation between consumer and district cooling utility equipment. Higher cost because of a heat exchanger and additional controls. District system is largely isolated from any problems in the building beyond the interconnection. The heat exchanger isolates the building system pressure from the district system pressure, and each may operate at their preferred pressures without influence from the other. Approach temperature in heat exchanger is a detriment to t. Additional space required for heat exchanger and controls

tor to implement a design that operates per the design intent. See the Temperature Differential Control section for further discussion regarding achieving high system t. Similarly, Figure 27 shows the simplest form of hot-water direct connection, where the district heating plant pumps water through the consumer building. This figure includes a pressure differential regulator (which may be required to reduce system differential pressure to meet any lower building system parameters), a thermostatic control valve on each terminal unit, and a pressure relief valve. Most commercial systems have a flowmeter installed as well as temperature sensors and transmitters to calculate the energy used. Pressure transmitters may be installed as input for plant circulating pump speed control. The location of each device may vary from system to system, but all of the major components are indicated. The control valve is the capacity regulating device that restricts flow to maintain either a water supply or return temperature on the consumer’s side. Particular attention must be paid to connecting high-rise buildings because they induce a static head that affects the design pressure of the entire system. Pressure control devices should be investigated carefully. It is not unusual to have a water-based district heating or cooling system with a mixture of direct and indirect connections in which heat exchangers isolate the systems hydraulically based on the location of the ETS in the building and the number of floors served.

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Fig. 26 Direct Connection of Building System to District Chilled Water Without Building Pumps

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(Phetteplace et al. 2013a)

the customer’s building (tertiary) pump. Figure 25 shows a primary/ secondary connection using an in-building pumping scheme. When tertiary pumps are used, all series connections between the district system pumps must be removed. A series connection can cause the district system return to operate at a higher pressure than the distribution system supply and disrupt normal flow patterns. Series operation usually occurs during improper use of three-way mixing valves in the primary to secondary connection. For similar reasons, booster pumps in chilled-water return lines back to the plant are discouraged.

4.2

Fig. 27

Direct Connection of Building System to District Hot Water

In a direct connected system, the pressure in the main distribution system must meet local building codes to protect the customer’s installation and the reliability of the district system. To minimize noise, cavitation, and control problems, constant-pressure differential control valves should be installed in the buildings. Special attention should be given to potential noise problems at the control valves. These valves must correspond to the design pressure differential in a system that has constantly varying distribution pressures because of load shifts. Multiple valves may be required to serve the load under all flow and pressure ranges. Industrial-quality valves and actuators should be used for this application. If the temperature supplied from the main distribution system is lower than that required in the consumer cooling systems, a larger temperature differential between supply and return occurs, thus reducing the required pipe size. The consumer’s desired supply temperature can be attained by mixing return water with district cooling supply water. Depending on the size and design of the main system, elevation differences, and types of customers and building systems, additional safety equipment (e.g., automatic shutoff valves on both supply and return lines) may be required. When buildings have separate circulation pumps, primary/secondary piping, and pumping, isolating techniques are used (crossconnection bridge between return and supply piping, decouplers, and bypass lines). This ensures that terminal unit two-way control valves are subjected only to the differential pressure established by

INDIRECT CONNECTIONS

Many of the components are similar to those used in the direct connection applications, with the exception that a heat exchanger performs one or more of the following functions: heat transfer, pressure interception, and buffer between potentially different quality water treatments. Identical to the direct connection, the rate of energy extraction in the heat exchanger is governed by a control valve that reacts to the building load demand. Once again, the control valve usually modulates to maintain a temperature set point on either side of the heat exchanger, depending on the contractual agreement between the consumer and the producer. The three major advantages of using heat exchangers are (1) the static head influences of a high-rise building are eliminated, (2) the two water streams are separated, and (3) consumers must make up all of their own lost water. The primary disadvantages of using an indirect connection are the (1) additional cost of the heat exchanger, (2) temperature loss caused by the approach of the heat exchanger, and (3) increased pumping pressure because of the addition of another heat transfer surface. Figure 28 shows a typical district chilled-water indirect connection, and Figure 29 shows a cascading district heating indirect connection where the domestic hot-water heat exchanger (HX) is supplied from the comfort HX outlet piping for greater t across the connection.

4.3

STEAM CONNECTIONS

Although higher pressures and temperatures are sometimes used, most district heating systems supply saturated steam at pressures between 35 and 1000 kPa (gage) to customers’ facilities. The steam is pretreated to maintain a neutral pH, and the condensate is cooled and discharged to the building sewage system (not preferred) or returned back to the central plant for recycling (preferred). Many consumers run the condensate through a heat exchanger to heat the domestic hot-water supply of the building before returning it to the

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2020 ASHRAE Handbook—HVAC Systems and Equipment (SI)

Fig. 28 Indirect Connection of Building System to District Chilled Water (Phetteplace et al. 2013a)

Fig. 29 Basic Cascading Indirect HeatingSystem Schematic

central plant or, sometimes, to the building drains if the district system does not have a condensate return pipe. This energy-saving process extracts the maximum amount of energy out of the delivered steam. Again, it is best to return the condensate back to the plant whenever the system configuration dictates, but there still are some vintage single-pipe steam systems in operation that are 100% makeup water systems. The designer should coordinate with the district energy provider to confirm the requirements. Interconnection between the district and the building is simple when the building uses the steam directly in heating coils or radiators or for process loads (humidification, kitchen, laundry, laboratory, steam absorption chillers, or turbine-driven devices). Other buildings extract the energy from the district steam via a steam-to-water heat exchanger to generate hot water and circulate it to the air-side terminal units. Typical installations are shown in Figures 30 and 31. Chapter 11 has additional information on building distribution piping, valving, traps, and other system requirements. The type of steam chemical treatment should be considered in applications for the food industry and for humidification (Tredinnick 2008b). Other components of the steam connection may include condensate pumps, flowmeters (steam and/or condensate), and condensate conductivity probes, which may dump condensate if contaminated by unacceptable debris. Often, energy meters are installed on both the steam and condensate pipes to allow the district energy supplier to determine how much energy is used directly and how much energy (condensate) is not returned back to the plant. Using cus-

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District Heating and Cooling

Fig. 30 District/Building Interconnection with Heat Recovery Steam System

12.41

Fig. 31 District/Building Interconnection with Heat Exchange Steam System Table 14

tomer energy meters for both steam and condensate is desirable for the following reasons: • Offers redundant metering (if the condensate meter fails, the steam meter can detect flow or vice versa) • Bills customer accordingly for makeup water and chemical treatment on all condensate that is not returned or is contaminated • Meter is in place if customer requires direct use of steam in the future • Assists in identifying steam and condensate leaks • Improves customer relations (may ease customer’s fears of overbilling because of a faulty meter) • Provides a more accurate reading for peak demand measurements and charges Monitor each level of steam pressure reduction as well as the temperature of the condensate. Where conductivity probes are used to monitor the quality of water returned to the steam plant, adequate drainage and cold-water quenching equipment may be required to satisfy local plumbing code requirements (temperature of fluid discharging into a sewer). The probe status should also be monitored at the control panel, to communicate high-conductivity alarms to the plant and, when condensate is being dumped, to notify the plant that a conductivity problem exists at a customer.

Building Conversion to District Heating Table 14 (Sleiman et al. 1990) summarizes the suitability or success rate of converting various heating systems to be served by a district hot-water system. As shown, the probability is high for

Conversion Suitability of Heating System by Type

Type of System Steam equipment One-pipe cast iron radiation Two-pipe cast iron radiation Finned-tube radiation Air-handling unit coils Terminal unit coils

Low

Medium

High

X X X X X

Hot-water equipment Radiators and convectors Radiant panels Unitary heat pumps Air-handling unit coils Terminal unit coils

X X X X X

Gas/oil-fired equipment Warm-air furnaces Rooftop units Other systems

X X X

Electric equipment Warm-air furnaces Rooftop units Air-to-air heat pumps Other systems

X X X X

Source: Sleiman et al. (1990).

water-based systems, lower for steam, and lowest for fuel oil or electric systems. Low-suitability systems usually require expensive replacement of the entire heating terminal and generating units with suitable water-based equipment, including piping, pumps, controls, and heat transfer media.

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12.42

2020 ASHRAE Handbook—HVAC Systems and Equipment (SI) 4.4 COMPONENTS

Heat Exchangers

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Aside from their isolation function, heat exchangers act as the line of demarcation between ownership responsibilities of the different components of an indirect system. They transfer thermal energy and act as pressure interceptors for the water pressure in high-rise buildings. They also keep fluids from each side (which may have different chemical treatments) from mixing. Figure 29 shows a basic building schematic including heat exchangers and secondary systems. Reliability of the installation is increased if multiple heat exchangers are installed. The number selected depends on the types of loads present and how they are distributed throughout the year. When selecting all equipment for the building interconnection, but specifically heat exchangers, the designer should • Size the unit’s capacity to match the given load and estimated load turndown as close as possible (oversized units may not perform as desired at maximum turndown; therefore, using several smaller units optimizes the installation). • Assess the critical nature of the load/operation/process to address reliability and redundancy. For example, if a building has 24 h process loads (i.e., computer room cooling, water-cooled equipment, etc.), consider adding a separate heat exchanger for this load. Also, consider operation and maintenance of the units. If the customer is a hotel, hospital, casino, or data center, select a minimum of two units at 50% load each to allow one unit to be cleaned without interrupting building service. Separate heat exchangers should be capable of automatic isolation during low-load conditions to increase part-load performance. • Determine customer’s temperature and pressure design conditions. Some gasket materials for plate heat exchangers have limits for low pressure and temperature. • Evaluate customer’s water quality (i.e., use appropriate fouling factor). • Determine available space and structural factors of the mechanical room. • Quantify design temperatures. The heat exchanger may require rerating at a higher inlet temperature during off-peak hours. • Calculate allowable pressure drop on both sides of heat exchangers. The customer’s side is usually the most critical for pressure drop. The higher the pressure drop, the smaller and less expensive the heat exchanger. However, the pressure drop must be kept in reasonable limits [100 kPa (gage) or below] if existing pumps are to be reused. For chilled-water connections, investigate the existing chiller evaporator pressure drop to assist in this evaluation. Size all heat exchangers with future expansion in mind. During selection, be aware that closer approach temperatures or low pressure drop require more heat transfer area and hence cost more and take up more space. Install strainers in front of any heat exchanger and control valve to keep debris from fouling surfaces or orifices. Plate, shell-and-coil, and shell-and-tube heat exchangers are all used for indirect connection. Whatever heat transfer device is selected must meet the appropriate temperature and pressure duty, and be stamped/certified accordingly as pressure vessels. Whether using heat exchangers for heating or cooling, it is advantageous and recommended that if one side of the unit has variable flow, the other side should as well. This ensures that minimum pump and thermal energy is used to satisfy a load (Tredinnick 2007). See Chapter 48 for more information on heat exchangers. Plate Heat Exchangers (PHEs). These exchangers, which are used for steam, hot-water, and chilled-water applications, are available as gasketed units and in two gasket-free designs (brazed and all- or semiwelded construction). All gasketed PHEs, also known as plate-and-frame heat exchangers (HXs), consist of multiple gas-

keted embossed metal plates bolted together between two end frames and sealed along the edges. Alternate plates are inverted and the gaps between the plates form the liquid flow channels. Fluids never mix because hot fluid flows on one side of the plate and cool fluid flows countercurrent on the other side. Ports at each corner of the end plates act as headers for the fluid. One fluid travels in the odd-numbered gaps and the other in the even-numbered gaps. Gaskets between the plates contain the two media in the plates and act as a boundary. Gasket failure does not cause the two media to mix; instead, the media leaks to the atmosphere. Gaskets can be either glued or clipped on. Gasketed PHEs are suitable for steam-toliquid and liquid-to-liquid applications. Designers should select the appropriate gasket material for the design temperatures and pressures expected. Plates are typically stainless steel; however, plate material can be varied based on the chemical makeup of the heat transfer fluids. Because PHEs require turbulent flow for proper heat transfer, pressure drops may be higher than those for comparable shell-andtube models. High efficiency leads to a smaller package. The designer should consider specifying that the frame be sized to hold 20% additional plates. PHEs require very little maintenance because the high velocity of the fluid in the channels keep surfaces clean from fouling. However, larger particles may become lodged in fine cavities between the plates and choke flow; automatic backflushing valves may be used to address this issue. PHEs generally have a cost advantage and require one-third to one-half the surface required by shell-and-tube units for the same operating conditions. PHEs are also capable of closer approach temperatures. PHEs are typically used for district heating and cooling with water and for cooling tower water heat recovery (free cooling). Double-wall plates are also available for potable-water heating, chemical processes, and oil quenching. PHEs have three to five times greater heat transfer coefficients than shell-and-tube units and can achieve 0.5 K approach, but for economic reasons the approach is traditionally 1 K. Gasketed PHEs can be disassembled in the field to clean the plates and replace the gaskets. Typical applications go up to 185°C and 2.8 MPa (gage). Plates are typically made from stainless steel, but are available in titanium for more corrosive uses such as seawater cooling. Brazed PHEs are suitable for steam, vapor, or water solutions. They feature a close approach temperature (within 1 K), large temperature drop, compact size, and a high heat transfer coefficient. Construction materials are stainless steel plates and sometimes frames brazed together with copper or nickel. Tightening bolts are not required as with the gasketed design. These units cannot be disassembled and cleaned; therefore, adequate strainers must be installed ahead of an exchanger and it must be periodically flushed clean in a normal maintenance program. Brazed PHEs usually peak at a capacity of under 60 kW (about 200 plates and 0.04 m3/s) and are suitable for 3 MPa and 225°C. Typical applications are district heating using hot water and refrigeration process loads. Doublewall plates are also available for domestic hot-water use. Avoid applications where the PHE may be exposed to large, sudden, or frequent changes in temperature and load, because of risk of thermal fatigue. Welded PHEs can be used in any application for which shell-andtube units are used that are outside the accepted range of gasketed PHE units, in liquid-to-liquid, steam-to-liquid, gas-to-liquid, gasto-gas, and refrigerant applications. Construction is very similar to gasketed units except gaskets are replaced with laser welds. Materials are typically stainless steel, but titanium, monel, nickel, and various alloys are available. Models have design ratings that range from 260°C at 1 MPa to 540°C at 6.7 MPa; however, they are available only in small sizes. Normally, these units are used in ammonia refrigeration and aggressive process fluids. They are more suitable to pressure pulsation or thermal cycling because they are thermal

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District Heating and Cooling fatigue resistant. A semiwelded PHE is a hybrid of the gasketed and the all-welded units in which the plates are alternatively sealed with gaskets and welds. Shell-and-Coil Heat Exchangers. These European-designed heat exchangers are suitable for steam-to-water and water-to-water applications and feature an all-welded-and-brazed construction. This counter/cross-flow heat exchanger consists of a hermetically sealed (no gaskets), carbon-steel pressure vessel with hemispherical heads. Copper or stainless helical tubes within are installed in a vertical configuration. This type of heat exchanger offers a high temperature drop and close approach temperature. Its vertical arrangement requires less floor space than other designs and has better heat transfer characteristics than shell-and-tube units. Shell-and-Tube Heat Exchangers. These exchangers are usually a multiple-pass design. The shell is usually constructed from steel and the tubes are often of U-bend construction, usually 20 mm (nominal) OD copper, but other materials are available. These units are ASME U-1 stamped for pressure vessels. Heat Exchanger Load Characteristics. To provide high t under multiple load conditions, variable flow is required on both sides of the heat exchanger (Perdue and Ansbro 1999; Skagestad and Mildenstein 2002; Tredinnick 2007). Without variable flow on the customer side, more water flow is required on the district side during reduced load. This condition results in both increased pumping for the district energy provider as well as reduced t. In addition, the customer side also experiences increased pumping costs without the use of variable flow. The specific degradation in t and the increases in flow depend on the actual heat exchanger selection, and can easily be determined for a specific heat exchanger by selection and sizing software available from the heat exchanger manufacturer. An example provided by Skagestad and Mildenstein (2002) for a 1500 kW design load indicates that, at 50% load and constant flow on the consumer side, 75% of the design flow would be required on the district cooling system side, compared to 45% if the consumer side used variable flow. However, consumer-side constant flow reduces the t from the design value of 8.3 K to just 5.6 K at 50% load; when variable flow is used on the consumer’s side of the PHE, the t actually increased from the design value of 8.3 K to 9.3 K. Another example of the need for variable-flow pumping on the consumer’s side of PHE is provided by Tredinnick (2007) for a 1758 kW application. In Figure 32, the consumer side of the heat exchanger has constant flow with the consumer-side design supply temperature of 5.6°C. The PHE has been sized such that, at 100% of design load, the district cooling return temperature is 12.2°C; thus, at maximum load the t is 7.7 K, assuming a 1.1 K approach. However, with constant flow on the consumer side at 50% of design load, over 83% of the peak design flow on the district cooling side is required and the district cooling return temperature has decreased to 9.8°C, thus lowering the t on the district cooling side to 5.3 K. Figure 33, also from Tredinnick (2007), shows the situation under identical conditions but with variable flow on the consumer side of the PHE. As before, with the consumer-side design supply temperature of 5.6°C and at 100% of design load, the PHE has been sized to yield a district cooling return temperature of 12.2°C; thus, at design conditions, the t is 7.7 K, again assuming 1.1 K approach. However, at 50% load with variable flow on the consumer’s side, only 51% of the design flow is required on the district cooling side, and the return temperature actually increases to 13.3°C. Thus, under this load condition, the t for the DCS increases from 5.3 K with constant flow to 8.8 K for variable consumer-side flow. Variable flow also saves electrical pump energy and aids in controlling comfort. These examples, as well as others [e.g., Perdue and Ansbro (1999)] should make clear the need for variable flow on the consumer side of a PHE in an indirect connection of district cooling. Typical constant-flow systems are found in older buildings and may be converted to simulate a variable-flow system by blocking off

12.43 the bypass line around the air handler heat exchanger coil three-way control valve. At low operating pressures, this potentially may convert a three-way bypass-type valve to a two-way modulating shutoff valve. Carefully analyze the valve actuator, because the shutoff requirements and control characteristics are totally different for a two-way valve than for a three-way valve. For more information on building conversion, see Skagestad and Mildenstein (2002). In theory, a cooling coil should have higher return water temperature when partially loaded than at full load, because the coil is oversized for the duty and thus has closer approach temperatures. In many real systems, as the load increases, the return water temperature tends to rise, and under low loads, the supply water temperature rises. Consequently, process or critical humidity control systems may suffer when connected to a system where return water temperature control is used to achieve high temperature differentials. Other techniques, such as separately pumping each CHW coil, may be used where constant supply water temperatures are necessary year round.

Flow Control Devices In commercial systems, after the flowmeter, control valves are the most important element in the interface with the district energy system because proper valve adjustment and calibration save energy. High-quality, industrial-grade control valves provide more precise control, longer service life, and minimum maintenance.

Fig. 32 Plate Heat-Exchanger Performance with Constant Flow on Customer Side and Customer-Side Supply Temperature of 5.6°C (Tredinnick 2007)

Fig. 33 Plate Heat-Exchanger Performance with Variable Flow on Customer Side and Customer-Side Supply Temperature of 5.6°C (Tredinnick 2007)

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All control valve actuators should take longer than 60 s to close from full open to mitigate pressure transients or water hammer, which occurs when valves slam closed. Actuators should also be sized to close against the anticipated system pressure so the valve seats are not forced open, thus forcing water to bypass and degrading temperature differential. The wide range of flows and pressures expected makes selection of control valves difficult. Typically, only one control valve is required; however, for optimal response to load fluctuations and to prevent cavitation, two valves in parallel are often needed. The two valves operate in sequence and for a portion of the load (i.e., one valve is sized for two-thirds of peak flow and the other sized for onethird of peak flow). The designer should review the occurrence of these loads to size the proportions correctly. The possibility of overstating customer loads complicates the selection process, so accurate load information is important. It is also important that the valve selected operates under the extreme pressure and flow ranges foreseen. Because most commercial-grade valves will not perform well for this installation, industrial-quality valves are typically specified. Electronic control valves should remain in a fixed position when a power failure occurs and should be manually operable. Pneumatic control valves should close upon loss of air pressure. A manual override on the control valves allows the operator to control flow. All chilled-water control valves must fail in the closed position. Then, when any secondary in-building systems are deenergized, the valves close and will not bypass chilled water to the return system. All steam pressure-reducing valves should close as well. Oversizing results in reduced valve and actuator life span and causes hunting. Select control valves having a wide range of control; low leakage; and proportional-plus-integral control for close adjustment, balancing, temperature accuracy, and response time. Control valves should have actuators with enough force to open and close under the maximum pressure differential in the system. The control valve should have a pressure drop through the valve equal to at least 10 to 30% of the static pressure drop of the distribution system. This pressure drop gives the control valve the “authority” it requires to properly control flow. The relationship between valve travel and capacity output should be linear, with an equal percentage characteristic. In hot-water systems, control valves are normally installed in the return line because the lower temperature in the line reduces the risk of cavitation and increases valve life. In chilled-water systems, control valves can be installed in either location; typically, however, they are installed the return line to reduce the potential for condensation on exposed external surfaces and to minimize water turbulence upstream of the flowmeter.

Instrumentation In many systems, where energy to the consumer is measured for billing purposes, temperature sensors assist in calculating the energy consumed as well as in diagnosing performance. Sensors and their transmitters should have an accuracy range commensurate to the accuracy of the flowmeter. In addition, pressure sensors are required for variable-speed pump control (water systems) or valve control for pressure-reducing stations (steam and water). Detailed recommendations on pressure, temperature, flow, and power transducers may be found in IDEA (2008a). Temperature sensors need to be located by the exchangers being controlled rather than in the common pipe. Improperly located sensors cause one control valve to open and others to close, resulting in unequal loads in the exchangers. Temperature sensors used in energy metering applications should be matched-pair, four-wire, 1000  for increased accuracy. Table 15 lists common measuring points and derivative parameters for remote monitoring and control of an indirect consumer interconnection. Other measuring points and derivative parameters may be required or recommended if the

Table 15 Measuring Points and Derivative Parameters for Remote Monitoring and Control of Indirect Consumer Interconnection Measured Point/ Parameter Temperature Pressure Differential pressure

Location District Cooling Side Supply, return Supply, return At building entrance, heat exchanger(s), control valve(s), strainer Supply or return Supply and return Yes

Flow rate Energy transfer Position of control valve(s) Variable-speed drive Yes (not typical) percentage(s)

Consumer Side Supply, return Supply, return Heat exchanger(s)

Optional Yes

district heating utility assumes some responsibility for operation of the building system or if the building and district cooling system belong to a single owner. Aside from the customary mechanical pressure gages that should be provided at the end-user interface for on-site diagnostics, pressure transducers are normally provided for remote monitoring and (often) control. Pressure transducers should be specified to provide accuracy of ±1% of full scale and typically resolution of 0.7 kPa (IDEA 2008a).

Controller The controller performs several functions, including recording demand and the amount of energy used for billing purposes, monitoring the differential pressure for plant pump control, energy calculations, alarming for parameters outside normal, and monitoring and control of all components. Often, the controller has a small battery power supply to preserve settings and billing information. Typical control strategies include regulating district flow to maintain the customer’s supply temperature (which results in a fluctuating customer return temperature) or maintaining the customer’s return temperature (which results in a fluctuating customer supply temperature). When controlling return flow for cooling, the effect on the customer’s ability to dehumidify properly with an elevated entering coil temperature should be investigated carefully.

Pressure Control Devices If the steam or water pressure delivered to the customer is too high for direct use, it must be reduced. Similarly, pressure-reducing or pressure-sustaining valves may be required if building height creates a high static pressure and influences the district system’s return water pressure. Water pressure can also be reduced by control valves or regenerative turbine pumps. The risk of using pressure-regulating devices to lower pressure on the return line is that if they fail, the entire distribution system is exposed to their pressure, and overpressurization will occur. In high-rise buildings, all piping, valves, coils, and other equipment may be required to withstand higher design pressures. Where system static pressure exceeds safe or economical operating pressure, either the heat exchanger method or pressure-sustaining valves in the return line may be used to minimize the effect of the pressure. Provide vacuum vents at the top of the building’s water risers to introduce air into the piping in case the vertical water column collapses.

Flow and Energy Metering All thermal energy or power delivered to customers or end users for billing or revenue by a commercially operated district energy system must be metered. The type of meter selected depends on the

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District Heating and Cooling

12.45 Table 16 Flowmeter Characteristics

Meter Type

Accuracy

Range of Control

Pressure Loss

Orifice plate

±1% to 5% full scale

3:1 to 5:1

High (>35 kPa)

±0.15% to 1% rate

30:1 to 100:1

Low (10:1 to 100:1

Low ( 50.5 > ha2 (see Figure 8). From Equation (16), the boundary airstream dry-bulb temperature is tab = 27 – (55.6 – 49.93)/1.00 = 21.33°C The boundary surface conditions are tsb = t a1 = 15.7°C

and

hsb = h a1 = 43.98 kJ/kg

From Equation (17), the boundary coolant temperature is trb = 13.0 – 0.367  1.00(27 – 21.8) = 10.92°C The cooling load for the dry surface part of the coil is now calculated from Equation (22b): qtd = 2.5  0.998  4.18(13.0 – 10.92) = 19.8 kW From Equation (18), the overall thermal resistance for the dry surface section is Ro = 0.013 + 0.004 + 0.006 = 0.023 (m2 ·K)/W From Equation (20), the mean temperature difference between air dry bulb and coolant for the dry surface section is  27 – 13.0  –  21.33 – 10.92  tm = ------------------------------------------------------------------------------- = 12.12 K ln   27 – 13.0    21.33 – 10.92   The dry surface area required is calculated from Equation (21): Ad = 19.8  0.023  1000/12.12 = 37.6 m2 From Equation (30), the cooling load for the wet surface section of the coil is qtw = 67.6 – 19.8 = 47.8 kW Knowing C, ha2 , and tr1, the surface condition at the leaving-air side of the coil is calculated by trial and error using Equation (24):

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23.14

2020 ASHRAE Handbook—HVAC Systems and Equipment(SI) C = 0.75 = (ts2 – 6.5)/(38.0 – hs2)

The numerical values for ts2 and hs2 are then determined directly by using a surface temperature chart (as shown in Figure 8 or in Figure 9 of AHRI Standard 410) and saturated air enthalpies from Table 2 in Chapter 1 of the 2017 ASHRAE Handbook—Fundamentals: ts2 = 10.8°C

and

hs2 = 31.16 kJ/kg

From Equation (25), the mean effective difference in air enthalpy between airstream and surface is

and psw = (Aw /Ao)(psw /Nri )Nri = (98/135)  65  4 = 189 Pa pst = (Aoi /Ao)psd + psw = (142/135)  (44 + 189) = 245 Pa total A more realistic p estimate of a coil operating at >70% wetted surface and a velocity > 2 m/s would be to consider the entire surface as wetted. Therefore, 65  4 = 260 Pa would be the coil’s operating airside static pressure. In summary, Aa Nri Aoi Ao qt tr2 ta2 SHR  psw  psd  pst

 49.93 – 43.81  –  37.93 – 31.16  hm = ------------------------------------------------------------------------------------------ = 6.44 kJ/kg ln   49.93 – 43.81    37.93 – 31.16   From Equation (27a), the wet surface area required is Aw = 47.8  1000  0.012  1.00/6.44 = 89 m2 From Equation (29), the net total surface area requirement for the coil is then Ao = 37.6 + 89 = 126.6 m2 external From Equation (31), the net air-side heat transfer exponent is c = 126.6/(1000  1.20  1.14  2.8  1.00  0.013) = 2.54

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From Equation (33), the enthalpy of saturated air corresponding to the effective surface temperature is 55.6 – 38.0 - = 36.4 kJ/kg hs = 55.6 – -------------------------– 2.54 1–e The effective surface temperature that corresponds to hs is then obtained from Table 2 in Chapter 1 of the 2017 ASHRAE Handbook— Fundamentals as ts = 13.0°C. The leaving-air dry-bulb temperature is calculated from Equation (34): ta2 = 13.0 + e–2.54(27 – 13.0) = 14.1°C The air-side sensible heat ratio is then found from Equation (35): 1.00  27 – 14.1  SHR = -------------------------------------- = 0.729 55.6 – 37.9 From Equation (1b), the calculated coil row depth Nrc to match job requirements is Nrc = Ao /AaFs = 135/(1.14  31.2) = 3.80 rows deep In most coil selection problems of this type, the initial calculated row depth to satisfy job requirements is usually a noninteger value. In many cases, there is sufficient flexibility in fluid flow rates and operating temperature levels to recalculate the required row depth of a given coil size to match an available integer row depth more closely. For this example, if the calculated row depth is Nrc = 3.8, and coils of three or four rows deep are commercially available, the coil face area, operating conditions, and fluid flow rates and/or velocities could possibly be changed to recalculate a coil depth close to either three or four rows. Although core tube circuitry has limited possibilities on odd (e.g., three or five) row coils, alternative coil selections for the same job are often made desirable by trading off coil face size for row depth. Most coil manufacturers have computer programs to run the iterations needed to predict operating values for specific coil performance requirements. The next highest integral row depth than computed is then selected for a commercially available coil with an even number of circuits, same end connected. For this example, assume that the initial coil selection requiring 3.8 rows deep is sufficiently refined that no recalculation is necessary, and that a 4-row coil with 4-pass coil circuitry is available. Thus, the installed row depth Nri is

8.

= = = = = = = = = = =

1.14 m2 coil face area 4 rows installed coil depth 142 m2 installed heat transfer surface area 135 m2 required heat transfer surface area 67.6 kW total refrigeration load 13.0°C leaving coolant temperature 13.9°C leaving-air dry-bulb temperature 0.740 air sensible heat ratio 189 Pa wet-coil surface air friction 44 Pa dry-coil surface air friction 245 Pa total coil surface air friction

DETERMINING REFRIGERATION LOAD

The following calculation of refrigeration load distinguishes between the true sensible and latent heat loss of the air, which is accurate within the data’s limitations. This division will not correspond to load determination obtained from approximate factors or constants. The total refrigeration load qt of a cooling and dehumidifying coil (or air washer) per unit mass of dry air is indicated in Figure 13 and consists of the following components: • The sensible heat qs removed from the dry air and moisture in cooling from entering temperature t1 to leaving temperature t2 • The latent heat qe removed to condense the moisture at the dewpoint temperature t4 of the entering air • The heat qw removed to further cool the condensate from its dew point t4 to its leaving condensate temperature t3 The preceding components are related by the following equation: qt = qs + qe + qw (36) If only the total heat value is desired, it may be computed by qt = (h1 – h2) – (W1 – W2)hw3 where h1 and h2 = enthalpy of air at points 1 and 2, respectively

Nri = 4 rows deep The amount of heat transfer surface area installed is Aoi = AaFs Nri = 1.14  31.2  4 = 142 m2 external The completely dry and completely wetted air-side frictions are, respectively, psd = (Ad /Ao)(psd Nri )Nri = (37/135)  40  4 = 44 Pa

Fig. 13 Psychrometric Performance of Cooling and Dehumidifying Coil

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(37)

Air-Cooling and Dehumidifying Coils

23.15 9.

W1 and W2 = humidity ratio at points 1 and 2, respectively hw3 = enthalpy of saturated liquid at final temperature t3

If a breakdown into latent and sensible heat components is desired, the following relations may be used. Latent heat may be found from qe = (W1 – W2)hfg4

(38)

where hfg4 = enthalpy representing latent heat of water vapor at condensing temperature t4

Sensible heat may be shown to be qs + qw = (h1 – h2) – (W1 – W2)hg4 + (W1 – W2)(hw4 – hw3) (39) or qs + qw = (h1 – h2) – (W1 – W2)(hfg4 + hw3)

(39a)

where

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hg4 = hfg4 + hw 4 = enthalpy of saturated water vapor at condensing temperature t4 hw 4 = enthalpy of saturated liquid at condensing temperature t4

The last term in Equation (39a) is the heat of subcooling the condensate from t4 to its final temperature t3. Then, qw = (W1 – W2)(hw4 – hw3)

(40)

The final condensate temperature t3 leaving the system is subject to substantial variations, depending on the method of coil installation, as affected by coil face orientation, airflow direction, and air duct insulation. In practice, t3 is frequently the same as the leaving wet-bulb temperature. Within the normal air-conditioning range, precise values of t3 are not necessary because the heat qt of condensate removed from the air usually represents about 0.5 to 1.5% of the total refrigeration cooling load. Example 4. Air enters a coil at 32°C db, 24°C wb; it leaves at 16°C db, 14.5°C wb; leaving-water temperature is assumed to be 12°C, which is between the leaving-air dew-point and coil surface temperatures. Find the total, latent, and sensible cooling loads on the coil with air at standard barometric pressure. Solution: Using Figure 1 (or the indicated equations) from Chapter 1 of the 2017 ASHRAE Handbook—Fundamentals, h1 h2 W1 W2 t4 hw4 hw3 hg4 hfg4

= = = = = = = = =

72.04 kJ/kg (dry air) 40.69 kJ/kg (dry air) 0.01557 kg (water)/kg (dry air) 0.00972 kg (water)/kg (dry air) 20.85°C dew point of entering air 4.186  20.85 = 87.28 kJ/kg 4.186  12.00 = 50.23 kJ/kg 2501 + (1.805  20.85) = 2538.63 kJ/kg hg4 – hw4 = 2451.35 kJ/kg

(32) (32) (35) (35) (39) (34) (34) (31) (31)

From Equation (37), the total heat is qt = (72.04 – 40.69) – (0.01557 – 0.00972)50.23 = 31.06 kJ/kg From Equation (38), the latent heat is qe = (0.01557 – 0.00972)2451.35 = 14.34 kJ/kg The sensible heat is therefore qs + qw = qt – qe = 31.06 – 14.34 = 16.72 kJ/kg The sensible heat may be computed from Equation (39a) as qs + qw = (72.04 – 40.69) – (0.01557 – 0.00972)2538.63 + (0.01557 – 0.00972)(87.28 – 50.23) = 16.72 kJ/kg The same value is found using Equation (39b). The subcooling of the condensate as a part of the sensible heat is indicated by the last term of the equation, 0.19 kJ/kg.

MAINTENANCE

If the coil is to deliver its full cooling capacity, both its internal and external surfaces must be clean. The tubes generally stay clean in pressurized water or brine systems. Tube surfaces can be cleaned in a number of ways, but are often washed with low-pressure water spray and mild detergent. Water coils should be completely drained if freezing is possible. When coils use built-up system refrigerant evaporators, oil can accumulate. Check and drain oil occasionally, and check for leaks and refrigerant dryness. Air Side. The best maintenance for the outside finned area is consistent inspection and service of inlet air filters. Surface cleaning of the coil with pressurized hot water and a mild detergent should be done only when necessary (primarily when a blockage occurs under severe fin-surface-fouling service conditions, or bacterial growth is seen or suspected). Pressurized cleaning is more thorough if done first from the air exit side of the coil and then from the air entry side. Foaming chemical sprays and washes should be used instead of high pressure on fragile fins, or when fin density is too restrictive to allow proper in-depth cleaning with pressurized water spray. In all cases, limit spray water temperature to below 65°C on evaporator coils containing refrigerant. In cases of marked neglect or heavy-duty use (especially in restaurants where grease and dirt have accumulated) coils must sometimes be removed and the accumulation washed off with steam, compressed air and water, or hot water. The surfaces can also be brushed and vacuumed. Best practice is to inspect and service the filters frequently. Also, condensate drain pan(s) and their drain lines, including open drain areas, should be kept clean and clear at all times. Water Side. The best service for the inside tube coil surface is keeping the circulated fluid (water or glycol) free of sediment, corrosive products, and biological growth. Maintaining proper circulated water chemistry and velocity and filtering out solids should minimize the water-side fouling factor. If large amounts of scale form when untreated water is used as coolant, chemical or mechanical (rod) cleaning of internal surfaces at frequent intervals is necessary. A properly maintained chilled-water system using a glycol solution as the circulated fluid is not considered to ever have waterside fouling, as such, but glycol solutions must be analyzed seasonally to determine alkalinity, percent concentration, and corrosion inhibitor condition. Consult a glycol expert or the manufacturer’s agent for detailed recommendations on proper use and control of glycol solutions. Refrigerant Side. Moisture content of the refrigerant should be checked yearly, and acidity of the compressor oil as often as monthly. For built-up direct-expansion (DX) systems, normal is 50 mg/kg of moisture for systems with mineral oil, and 100 mg/kg for some refrigerant types in systems with polyol ester (POE) compressor oil. The actual values are set by the compressor manufacturer, and should be checked and verified during operation by moisture-indicating sight-glass viewing, and yearly by laboratory sample analysis reports. Depending on temperature and velocity, excessive moisture coming to the cooling coil through refrigerant might cause internal freeze-up around the coil’s expansion valve. Generally, moisture in a system contributes to formation of acids, sludge, copper plating, and corrosion. Oil breakdown (by excessive compressor overheat) can form organic, hydrofluoric, and hydrochloric acids in the lubricant oil, all of which can corrode copper. The refrigerant must be of high quality and meet AHRI Standard 700 purity requirements. Application design ratings of the coil or coil bank are based on purity and dryness. This is a primary requirement of refrigerant evaporator coil maintenance. DX coil replacement often coincides with refrigerant change-out to a chlorine-free refrigerant. Special care should be taken to ensure

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2020 ASHRAE Handbook—HVAC Systems and Equipment(SI)

that the system is retrofitted in accordance with the compressor and refrigerant manufacturers’ written procedures.

10.

coil face or frontal area, m2 dry external surface area, m2 total external surface area, m2 exposed external prime surface area, m2 external secondary surface area, m2 wet external surface area, m2 ratio of external to internal surface area, dimensionless coil characteristic as defined in Equations (12) and (23), (kg·K)/kJ c = heat transfer exponent, or NTUa, as defined in Equations (31) and (32), dimensionless co = heat transfer exponent, as defined in Equation (7d), dimensionless cp = specific heat of humid air = 1.019 kJ/(kg·K) for cooling coils cpw = specific heat of water = 4.18 kJ/(kg·K) cr = specific heat of nonvolatile coolant, kJ/(kg·K) Di = tube inside diameter, mm Do = tube outside diameter, mm Ea = air-side effectiveness defined in Equation (7b), dimensionless Fs = coil core surface area parameter = (external surface area)/(face area) (no. of rows deep) f = convection heat transfer coefficient, W/(m2 ·K) h = air enthalpy (actual in airstream or saturation value at surface temperature), kJ/kg hm = mean effective difference of air enthalpy, as defined in Equation (25), kJ/kg k = thermal conductivity of tube material, W/(m·K) M = ratio of nonvolatile coolant-to-air temperature changes for sensible heat cooling coils, as defined in Equation (7e), dimensionless m = rate of change of air enthalpy at saturation with air temperature, kJ/(kg·K) Nr = number of coil rows deep in airflow direction, dimensionless psd = isothermal dry surface air-side at standard conditions (20°C, 101.325 kPa), Pa psw = wet surface air-side at standard conditions (20°C, 101.325 kPa), Pa q = heat transfer capacity, W qe = latent heat removed from entering air to condense moisture, kJ/kg qs = sensible heat removed from entering air, kJ/kg qt = total refrigeration load of cooling and dehumidifying coil, kJ/kg qw = sensible heat removed from condensate to cool it to leaving temperature, kJ/kg R = thermal resistance, referred to external area Ao, (m2 ·K)/W SHR = ratio of air sensible heat to air total heat, dimensionless t = temperature, °C tm = mean effective temperature difference, air dry bulb to coolant temperature, K tms = mean effective temperature difference, surface-to-coolant, K t m = mean effective temperature difference, air dry bulb to effective surface temperature ts , K Uo = overall sensible heat transfer coefficient, W/(m2 ·K) Va = coil air face velocity at 20°C, m/s W = air humidity ratio, kilograms of water per kilogram of air w = mass flow rate, kg/s y = ratio of nonvolatile coolant temperature rise to airstream enthalpy drop, as defined in Equation (13), (kg·K)/kJ  = fin effectiveness, as defined in Equation (6), dimensionless a = air density = 1.20 kg/m3 at 20°C at sea level Superscripts  = wet bulb  = dew point Aa Ad Ao Ap As Aw B C

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SYMBOLS

= = = = = = = =

Subscripts 1 = condition entering coil 2 = condition leaving coil a = airstream ab = air, dry/wet boundary ad = dry air aw = wet air b = dry/wet surface boundary d = dry surface e = latent f = fin (with R); saturated liquid water (with h) g = saturated water vapor i = installed, selected (with Ao , Nr) m = metal (with R) and mean (with other symbols) md = dry metal mw = wet metal o = overall (except for A) r = coolant rb = coolant dry/wet boundary s = surface (with d, t, and w) and saturated (with h) s = effective surface sb = surface dry/wet boundary t = tube (with R) and total (with s and q) td = total heat capacity, dry surface tw = total heat capacity, wet surface w = water (with ), condensate (with h and subscript number), and wet surface (with other symbols)

REFERENCES ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal.ashrae .org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org/bookstore. AHRI. 2001. Forced-circulation air-cooling and air-heating coils. ANSI/ AHRI Standard 410-01. Air-Conditioning, Heating, and Refrigeration Institute, Arlington, VA. AHRI. 2015. Specification for fluorocarbon refrigerants. Standard 7002015. Air-Conditioning, Heating, and Refrigeration Institute, Arlington, VA. Anderson, S.W. 1970. Air-cooling and dehumidifying coil performance based on ARI Industrial Standard 410-64. In Heat and mass transfer to extended surfaces, ASHRAE Symposium CH-69-3, pp. 22-28. ASHRAE. 2000. Methods of testing forced circulation air-cooling and air heating coils. Standard 33-2000. ASME. 2015. Code on nuclear air and gas treatment. ANSI/ASME Standard AG-1-2015. American Society of Mechanical Engineers, New York. ASME. 2015. Boiler and pressure vessel code. American Society of Mechanical Engineers, New York. Brown, G. 1954. Theory of moist air heat exchangers. Royal Institute of Technology Transactions 77, Stockholm, Sweden, 12. Kusuda, T. 1970. Effectiveness method for predicting the performance of finned tube coils. In Heat and mass transfer to extended surfaces, ASHRAE Symposium CH-69-3, pp. 5-14. McElgin, J., and D.C. Wiley. 1940. Calculation of coil surface areas for air cooling and dehumidification. Heating, Piping and Air Conditioning (March):195. McQuiston, F.C. 1981. Finned tube heat exchangers: State of the art for air side. ASHRAE Transactions 87(1):1077-1085. Paper CH-81-16-2. Mirth, D.R., S. Ramadhyani, and D.C. Hittle. 1993. Thermal performance of chilled water cooling coils operating at low water velocities. ASHRAE Transactions 99(1):43-53. Paper 3626. Mueller, A.C. 1998. Heat exchangers. Ch. 17 in Handbook of heat transfer, 3rd ed. W. Rohsenow, J. Hartnett, and Y. Cho, eds. McGraw-Hill, New York. Pedersen, C.O., D.E. Fisher, J.D. Spitler, and R.J. Liesen. 1998. Cooling and heating load calculation principles. ASHRAE. Shah, M.M. 1976. A new correlation for heat transfer during boiling flow through pipes. ASHRAE Transactions 82(2):66-75. Paper SE-2407.

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Air-Cooling and Dehumidifying Coils BIBLIOGRAPHY

Shah, M.M. 1978. Heat transfer, pressure drop, visual observation, test data for ammonia evaporating inside pipes. ASHRAE Transactions 84(2): 38-59.

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Shah, M.M. 1982. CHART correlation for saturated boiling heat transfer: Equations and further study. ASHRAE Transactions 88(1):185-196. Paper HO-2673. Webb, R.L. 1980. Air-side heat transfer in finned tube heat exchangers. Heat Transfer Engineering 1(3):33.

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Related Commercial Resources CHAPTER 24

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DESICCANT DEHUMIDIFICATION AND PRESSURE-DRYING EQUIPMENT Methods of Dehumidification ................................................... 24.1 Desiccant Dehumidification .................................................... 24.2 Liquid Desiccant Equipment .................................................... 24.3 Solid-Sorption Equipment ........................................................ 24.4 Rotary Solid-Desiccant Dehumidifiers .................................... 24.5 Equipment Ratings ................................................................... 24.7

Equipment Operating Recommendations ................................. 24.8 Applications for AtmosphericPressure Dehumidification ................................................. 24.10 Desiccant Drying at Elevated Pressure ................................. 24.12 Equipment Types .................................................................... 24.12 Applications ............................................................................ 24.13

D

• Dehumidifying buildings using chilled beams • Frost-free cooling for low-temperature process areas such as brewery cellars; blast freezers; and refrigerated warehouses • Offering frost-free dehumidification for processes that require air at a subfreezing dew-point humidity • Maintaining low dew points for electronics and battery production • Maintaining low-dew-point space conditions for surgical suites • Providing appropriate conditions for pneumatic conveying of hygroscopic materials • Providing appropriate conditions for engine test cells • Improving indoor air quality (IAQ) • Eliminating fog formation and ceiling condensation in ice rinks. • Minimizing frost build-up on refrigerated display cases in supermarkets. This chapter covers (1) the types of dehumidification equipment for liquid and solid desiccants, including high-pressure equipment; (2) performance curves; (3) variables of operation; and (4) some typical applications. Using desiccants to dry refrigerants is addressed in Chapter 8 of the 2018 ASHRAE Handbook—Refrigeration.

EHUMIDIFICATION is the removal of water vapor from air, gases, or other fluids. There is no pressure limitation in this definition, and sorption dehumidification equipment has been designed and operated successfully for system pressures ranging from subatmospheric to as high as 40 MPa. In common practice, dehumidification refers to equipment operating at atmospheric pressures and built to standards similar to other types of air-handling equipment. For drying gases under pressure, or liquids, the term dryer or dehydrator is normally used. This chapter mainly covers equipment and systems that dehumidify air rather than those that dry other gases or liquids. Both liquid and solid desiccants are used; they either adsorb water on the desiccant’s surface (adsorption) or chemically combine with water (absorption). Nonregenerative equipment uses hygroscopic salts such as calcium chloride, urea, or sodium chloride. Regenerative systems usually use a form of silica or alumina gel; activated alumina; molecular sieves; or lithium chloride, calcium chloride, or glycol solution. In regenerative equipment, the water removal mechanism is reversible. The choice of desiccant depends on installation requirements, equipment design, and chemical compatibility with the gas to be treated or impurities in the gas. Chapter 32 of the 2017 ASHRAE Handbook— Fundamentals has more information on desiccant materials and how they operate. Some applications of desiccant dehumidification include • Ventilating buildings with cooled and dried outdoor air • Addressing buildings with high internal humidity loads, such as pools, fitness centers, and beverage/food processing • Keeping buildings and HVAC systems dry to prevent mold growth • Lowering relative humidity to facilitate manufacturing and handling of hygroscopic materials • Lowering the dew point to prevent condensation on products manufactured in low-temperature processes • Providing protective atmospheres for heat treatment of metals • Controlling humidity in warehouses and caves used for storage • Preserving ships, aircraft, and industrial equipment that would otherwise deteriorate • Maintaining a dry atmosphere in a closed space or container, such as the cargo hold of a ship or numerous static applications • Eliminating condensation and subsequent corrosion • Speeding drying of heat-sensitive products, such as candy, seeds, and photographic film • Drying natural gas • Drying gases that are be liquefied • Drying instrument and plant air • Drying process and industrial gases • Dehydration of liquids

The preparation of this chapter is assigned to TC 8.12, Desiccant Dehumidification Equipment and Components.

Copyright © 2020, ASHRAE

1. METHODS OF DEHUMIDIFICATION Air may be dehumidified by (1) cooling it or increasing its pressure, reducing its capacity to hold moisture, or (2) removing moisture by attracting the water vapor with a liquid or solid desiccant. Frequently, systems use a combination of these methods to maximize operating efficiency and minimize installed cost. Figure 1 illustrates three methods to dehumidify with desiccant materials or equipment. Air in the condition at Point A is dehumidified

Fig. 1 Methods of Dehumidification

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Fig. 2

2020 ASHRAE Handbook—HVAC Systems and Equipment (SI)

Flow Diagram for Liquid-Absorbent Dehumidifier (A) Without and (B) With Extended Surface Contact Medium

Refrigerating air below its dew point is the most common method of dehumidification. This is advantageous when the gas is comparatively warm, has a high moisture content, and the desired outlet dew point is above 5°C. Frequently, refrigeration is combined with desiccant dehumidification to obtain an extremely low dew point at minimum cost.

concentration and temperature. Figure 4 presents this relationship for lithium chloride/water solutions in equilibrium with air at 101.325 kPa. The graph has the same general shape as a psychrometric chart, with the relative humidity lines replaced by desiccant concentration lines. To dehumidify air, a first airstream (either outdoor, return, or mixed air) contacts a cooled solution in the conditioner; water condenses into the desiccant solution from the air, and the solution is diluted. The diluted solution is continuously reconcentrated in the regenerator, where it is heated to elevate its water vapor pressure and equilibrium humidity ratio. A second airstream, usually outdoor air, contacts the heated solution in the regenerator; water evaporates from the desiccant solution into the air, and the solution is reconcentrated. Desiccant solution is continuously circulated between the conditioner and regenerator to complete the cycle. Liquid desiccant conditioners typically have high contact efficiency, so air leaves the conditioner at a temperature and humidity ratio very close to the entering temperature and equilibrium humidity ratio of the desiccant. Liquid desiccants are typically very effective antifreeze. As a result, liquid-desiccant conditioners can continuously deliver air at subfreezing temperatures without frosting or freezing problems. Lithium chloride/water solution, for example, has a eutectic point below –68°C; liquid desiccant conditioners using this solution can cool air to temperatures as low as –45°C.

Liquid Desiccants

Solid Sorption

Liquid desiccant conditioners (absorbers) contact the air with a liquid desiccant, such as a solution of lithium chloride or glycol (Figures 2A, 2B, and 3), to dehumidify and often cool that air. When the water vapor pressure of such a solution is lower than the partial pressure of water in the surrounding air, the solution will collect moisture, thereby decreasing its concentration and dehumidifying the air. The solution’s water vapor pressure is a function of its temperature and concentration. Solutions of higher desiccant concentration and/ or lower temperature result in lower water vapor pressures and have a stronger dehumidifying effect on the contacted air. Conversely, liquid desiccant regenerators also contact the air with the same liquid desiccant. When the water vapor pressure of such a solution is higher than the partial pressure of water in the surrounding air, the solution will reject moisture, thereby increasing its concentration and humidifying the air. Solutions of lower desiccant concentration and/or higher temperature result in higher water vapor pressures and have a humidifying effect on the contacted air. A simple way to show this relationship is to graph the humidity ratio of air in equilibrium with a liquid desiccant as a function of its

Solid sorption passes air through a bed of granular desiccant or through a structured packing impregnated with desiccant. Humid air passes through the desiccant, which when active has a vapor pressure below that of the humid air. This vapor pressure differential drives water vapor from the air onto the desiccant. After becoming loaded with moisture, the desiccant is reactivated (dried out) by heating, which raises the vapor pressure of the material above that of the surrounding air. With the vapor pressure differential reversed, water vapor moves from the desiccant to a second airstream called the reactivation air, which carries moisture away from the equipment.

and cooled to Point B. In a liquid-desiccant unit, air is simultaneously cooled and dehumidified directly from Point A to Point B. In a solid-desiccant unit, this process can be completed by precooling and dehumidifying from Point A to Point C, then desiccating from Point C to Point E, and finally cooling to Point B. It could also be done with solid-desiccant equipment by dehumidifying from Point A to Point D and then cooling from Point D to Point B.

Compression Compressing air reduces its capacity to hold moisture. The resulting condensation reduces the air’s moisture content in absolute terms, but produces a saturated condition: 100% relative humidity at elevated pressure. In atmospheric-pressure applications, this method is too expensive, but is worthwhile in pressure systems such as instrument air. Other dehumidification equipment, such as coolers or desiccant dehumidifiers, often follows the compressor to avoid problems associated with high relative humidity in compressed-air lines.

Cooling

2. DESICCANT DEHUMIDIFICATION Both liquid and solid desiccants may be used in equipment designed for drying air or other gases at atmospheric or elevated pressures. Regardless of pressure levels, basic principles remain the same. Desiccant capacity and actual dew-point performance depend on the specific equipment used, characteristics of the various desiccants, initial temperature and moisture content of the gas to be dried,

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Desiccant Dehumidification and Pressure-Drying Equipment

24.3

Fig. 3 Flow Diagram for Liquid-Absorbent Unit with Onboard Refrigeration System

Fig. 4 Lithium Chloride Equilibrium reactivation methods, etc. Factory-assembled units are available up to a capacity of about 38 m3/s. Greater capacities can be obtained with field-erected units.

2.1

LIQUID DESICCANT EQUIPMENT

Liquid desiccant dehumidifiers are shown in Figures 2 and 3. In Figure 2A, liquid desiccant is distributed onto a cooling coil, which acts as both a contact surface and a means of removing heat released when the desiccant absorbs moisture from the air. In Figure 2B, liquid desiccant is distributed onto an extended heat and mass transfer

surface (a packing material similar to that used in cooling towers and chemical reactors). Additionally, Figure 3 demonstrates a liquid desiccant air-conditioning system with an onboard refrigeration system, offering a packaged solution. The packing provides a great deal of surface for air to contact the liquid desiccant, allowing the desiccant to dehumidify and cool the air. It also uses a heat exchanger outside the airstream to cool the desiccant, thereby removing the heat of absorption and any other heat transferred from the air into the desiccant. Air can be passed through the contact surface vertically or horizontally to suit the best arrangement of air system equipment. Depending on the air and desiccant solution inlet conditions, air can be either simultaneously cooled and dehumidified; heated and dehumidified; heated and humidified; or cooled and humidified. When the enthalpy of the air is to be increased in the conditioner unit, heat must be added either by preheating the air before it enters the conditioner or by heating the desiccant solution with a second heat exchanger. Conversely, when the enthalpy of the air is to be decreased in the conditioner unit, heat must be removed with a heat exchanger. When the air is to be humidified, makeup water is automatically added to the desiccant solution to keep it at the desired concentration. When air is to be dehumidified, a regenerator to reject the collected moisture to an exhaust airstream and a mechanism for transporting the gathered water from the collector to the regenerator are required.

Moisture Removal In the conditioner, a pump continuously circulates the desiccant solution through a heat exchanger and distributes it over the packed bed contactor surface. The heat exchanger cools the desiccant solution either through an onboard refrigeration system or by transferring heat to an appropriate heat sink such as chilled water, ground water, or other low-temperature heat sink (e.g., even chilled ammonia for extremely low-temperature and low-dew-point applications).

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2020 ASHRAE Handbook—HVAC Systems and Equipment (SI)

This maintains the water vapor pressure of the desiccant used for conditioning lower than that of the air to be treated. Moisture is absorbed from or desorbed into the air because of the difference in water vapor pressure between the air and the desiccant solution. By controlling the temperature and concentration of the desiccant solution, the conditioner unit can deliver air at a precisely controlled temperature and humidity regardless of inlet air conditions. Therefore, a unit can dehumidify the air during humid weather and humidify it during dry weather. Thus, liquid desiccant conditioners can accurately control humidity without face-andbypass dampers or after-humidifiers. System performance can easily be controlled as process drying requirements change by altering temperature, concentration, or both to meet the new requirements. Similarly, solution strength can be easily monitored and adjusted while the unit is in operation. The solution strength can be monitored while in operation and can easily be adjusted to compensate for aging. In most cases, the solution retains its effectiveness for the life of the equipment, assuming that proper filtration is maintained.

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Heat Removal When a liquid desiccant absorbs moisture, heat is generated. This heat of absorption consists of the latent heat of condensation of water vapor at the desiccant temperature and the heat of solution (heat of mixing) of the condensed water and the desiccant. The heat of mixing is a function of the equilibrium relative humidity of the desiccant: a lower equilibrium relative humidity produces a greater heat of mixing. The total heat that must be absorbed by the desiccant solution consists of the (1) heat of absorption, (2) sensible heat associated with reducing the dry-bulb temperature of the air, and (3) residual heat carried to the conditioner by the warm, concentrated desiccant returning from the regenerator unit. This total heat is removed by cooling the desiccant solution in the conditioner heat exchanger (Figure 2B). Any coolant can be used, including cooling tower water, groundwater, seawater, chilled water or brine, and directexpansion or flooded refrigerants. Regenerator residual heat, generally called regenerator heat dumpback, can be substantially reduced by using a liquid-to-liquid heat exchanger to precool the warm, concentrated desiccant transferred to the conditioner using the cool, dilute desiccant transferred from the conditioner to the regenerator. This also improves the thermal efficiency of the system, typically reducing coolant and heat input by 10 to 15%.

Regeneration When the conditioner is dehumidifying, water is automatically removed from the liquid desiccant to maintain the desiccant at the proper concentration. Removal takes place in a separate regenerator. A small sidestream of the desiccant solution is transferred to the regenerator unit. In the regenerator, a separate pump continuously circulates the desiccant solution through a heat exchanger and distributes it over the packed bed contactor surface. The heat exchanger heats the desiccant solution with waste energy from the onboard compressor system or low-pressure steam or hot water (e.g., from a boiler, solar water heater, or waste heat source) so that the water vapor pressure of the desiccant used for regeneration is substantially higher than that of the outdoor air. Outdoor air is passed through the packing, and water evaporates into it from the desiccant solution, concentrating the solution. The hot, moist air from the regenerator is discharged to the outdoors. A sidestream of concentrated solution is transferred to the conditioner to replace the sidestream of weak solution transferred from the conditioner and completes the cycle. The regenerator’s water removal capacity is controlled to match the moisture load handled by the conditioner. This is accomplished by regulating heat flow to the regenerator heat exchanger to maintain

Fig. 5 Liquid Desiccant System with Multiple Conditioners a constant desiccant solution concentration. This is most commonly done by maintaining a constant solution level in the system with a level controller, but specific-gravity or boiling-point controllers are used under some circumstances. Regenerator heat input is regulated to match the instantaneous water removal requirements, so no heat input is required if there is no moisture load on the conditioner. When the conditioner is used to humidify the air, the regenerator fan and desiccant solution pump are typically stopped to save energy. The conditioner and regenerator can be in a single housing or in separate housings. If in a single housing, the installation, electrical wiring, and equipment cost can be lower than a unit with separate housings. If separated, they can be in different locations and connected by piping; this approach can can substantially lower ductwork cost and required mechanical space. Commonly, a single regenerator serves several conditioner units (Figure 5). In the simplest control arrangement, concentrated desiccant solution is metered to each conditioner at a fixed rate. The return flow of weak solution from each conditioner is regulated to maintain a constant operating level in the conditioner. A level controller on the regenerator regulates heat flow to the regenerator solution heater to maintain a constant volume of desiccant solution, and hence a practically constant solution concentration. The regenerator can be sized to match the dehumidification load of the conditioner unit or units. Regenerator capacity is affected by regenerator heat source temperatures (higher source temperatures increase capacity) and by desiccant concentration (higher concentrations reduce capacity). The relative humidity of air leaving the conditioner is practically constant for a given desiccant concentration, so regenerator capacity can be shown as a function of delivered air relative humidity and regenerator heat source temperature. Figure 6 is a normalized graph showing this relationship. For a given moisture load, various regenerator heat sources may be used if the regenerator is sized for the heat source selected. In many cases, the greater capital cost of a larger regenerator is paid back very quickly by reduced operating cost when a lower-cost or waste-heat source (e.g., condenser heat, especially from the onboard refrigeration system; solar hot water; process or turbine tail stream; jacket heat from an engine-driven generator or compressor) is used.

2.2

SOLID-SORPTION EQUIPMENT

Solid desiccants, such as silica gel, zeolites (molecular sieves), activated alumina, or hygroscopic salts, are generally used to dehumidify large volumes of moist air, and are continuously reactivated. Solid desiccants can also be used in (1) nonreactivated, disposable packages and (2) periodically reactivated desiccant cartridges. Disposable packages of solid desiccant are often sealed into packaging for consumer electronics, pharmaceutical tablets, and military supplies. Disposable desiccant packages rely entirely on vapor diffusion to dehumidify, because air is not forced through the desiccant. This method is used only in applications where there is no

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Fig. 6 Liquid Desiccant Regenerator Capacity anticipated moisture load at all (such as hermetically sealed packages) because the moisture absorption capacity of any nonreactivated desiccant is rapidly exceeded if a continuous moisture load enters the dehumidified space. Disposable packages generally serve as a form of insurance against unexpected, short-term leaks in small, sealed packages. Periodically reactivated cartridges of solid desiccant are used where the expected moisture load is continuous, but very small. A common example is the breather, a tank of desiccant through which air can pass, compensating for changes in liquid volume in petroleum storage tanks or drums of hygroscopic chemicals. Air dries as it passes through the desiccant, so moisture will not contaminate the stored product. When the desiccant is saturated, the cartridge is removed and heated in an oven to restore its moisture sorption capacity. Desiccant cartridges are used where there is no requirement for a constant humidity control level and where the moisture load is likely to exceed the capacity of a small, disposable package of desiccant. Desiccant dehumidifiers for drying liquids and gases other than air often use a variation of this reactivation technique. Two or more pressurized containers of solid desiccant are arranged in parallel, and air is forced through one container for drying, while desiccant in the other container is reactivated. These units are often called dual-tower or twin-tower dehumidifiers. Continuous reactivation dehumidifiers are the most common type used in high-moisture-load applications such as humidity control systems for buildings and industrial processes. In these units, humid process air is dehumidified in one part of the desiccant bed while a different part of the bed is dried for reuse by a second airstream (reactivation air). The desiccant generally rotates slowly between these two airstreams, so that dry, high-capacity desiccant leaving the reactivation air is always available to remove moisture from the process air. This type of equipment is generally called a rotary desiccant dehumidifier. It is most commonly used in building air-handling systems, and the section on Rotary Solid-Desiccant Dehumidifiers describes its function in greater detail.

2.3

ROTARY SOLID-DESICCANT DEHUMIDIFIERS

Operation Figure 7 illustrates the principle of operation and arrangement of major components of a typical rotary solid-desiccant dehumidifier. The desiccant can be beads of granular material packed into a bed, or it can be finely divided and impregnated throughout a structured medium. The structured medium resembles corrugated cardboard rolled into a drum, so that air can pass freely through flutes aligned lengthwise through the drum.

Fig. 7 Typical Rotary Dehumidification Wheel

Fig. 8 Effect of Changes in Process Air Velocity on Dehumidifier Outlet Moisture In both granular and structured-medium units, the desiccant itself can be either a single material, such as silica gel, or a combination, such as silica gel blended with zeolites. The wide range of dehumidification applications requires flexibility in selecting desiccants to minimize operating and installed costs. In rotary desiccant dehumidifiers, more than 20 variables can affect performance. In general, equipment manufacturers fix most of these to provide predictable performance in common applications for desiccant systems. Primary variables left to the system designer to define include the following for both process and reactivation air: • Inlet air temperature • Moisture content • Velocity at face of the desiccant bed In any system, these variables change because of weather, variations in moisture load, and fluctuations in reactivation energy levels. It is useful for the system designer to understand the effect of these normal variations on dehumidifier performance. Figures 8 to 12 show changes in process air temperature and moisture leaving a generic rotary desiccant dehumidifier as modeled by a finite difference analysis program (Worek and Zheng 1991). Commercial unit performance differs from this model because such units are generally optimized for very deep drying.

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Fig. 9 Effect of Changes in Process Air Inlet Moisture on Dehumidifier Outlet Moisture

Fig. 11

Effect of Changes in Process Air Inlet Moisture on Dehumidifier Outlet Temperature

Fig. 12 Effect of Changes in Reactivation Air Inlet Temperature on Dehumidifier Outlet Temperature

Fig. 10 Effect of Changes in Reactivation Air Inlet Temperature on Dehumidifier Outlet Moisture However, for illustration purposes, the model accurately reflects the relationships between the key variables. The desiccant used for the model is silica gel; the bed is a structured, fluted medium; the bed depth is 400 mm in the direction of airflow; and the ratio of process air to reactivation air is approximately 3:1. Process air velocity through the desiccant bed strongly affects leaving moisture. As shown in Figure 8, if the entering moisture is 8.0 g/kg and all other variables are held constant, the outlet moisture varies from 3.1 g/kg at 1.3 m/s to 5.7 g/kg at 3.6 m/s. Thus, air that passes through the bed more slowly is dried more deeply. Therefore, if air must be dried very deeply, a large unit (slower air velocities) must be used. Process air inlet moisture content affects outlet moisture: if air is more humid entering the dehumidifier, it will be more humid leaving the unit. For example, Figure 9 indicates that, for an inlet humidity of 8.0 g/kg, the outlet humidity is 5.0 g/kg. If inlet moisture content rises to 11.4 g/kg, the outlet humidity rises to 7.1 g/kg. Therefore, if constant outlet humidity is necessary, the dehumidifier needs capacity control unless the process inlet airstream does not

Fig. 13 Typical Performance Data for Rotary Solid Desiccant Dehumidifier vary in temperature or moisture throughout the year (a rare circumstance). Reactivation air inlet temperature changes the outlet moisture content of the process air. From 38 to 120°C, as more heat is added to the reactivation air, the desiccant dries more completely, which means that it can attract more moisture from the process air (see Figure 10). If reactivation air is only heated to 38°C, process outlet moisture is 7.1 g/kg, or only 0.9 g/kg lower than the entering humid-

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Desiccant Dehumidification and Pressure-Drying Equipment ity. In contrast, if reactivation air is heated to 93°C, the outlet moisture is 5.0 g/kg, so that almost 40% of the original moisture is removed. This relationship has two important consequences. If the design needs dry air, it is generally more economical to use high reactivation air temperatures. Conversely, if leaving humidity from the dehumidifier need not be especially low, inexpensive, low-grade heat sources (e.g., solar heat, waste heat, cogeneration heat, or rejected heat from refrigeration condensers) can be used to reactivate the desiccant. Process air outlet temperature is higher than the inlet air temperature primarily because the heat of sorption of moisture removed from the air is converted to sensible heat. The heat of sorption includes the latent heat of condensation of the removed moisture, plus additional chemical heat, which varies depending on the desiccant type and process air outlet humidity. Also, some heat is carried over to the process air from the reactivation sector because the desiccant is warm as it enters the relatively cooler process air. Generally, 80 to 90% of the temperature rise of process air is from the heat of sorption, and the balance is from heat carried over from reactivation. Process outlet temperature versus inlet humidity is shown in Figure 11. Note that as more moisture is removed (higher inlet humidity), outlet temperature rises. Air entering at room comfort conditions of 21°C, 8.0 g/kg leaves the dehumidifier at 31.7°C. If the dehumidifier removes more moisture, such as when the inlet humidity is 11.4 g/kg, outlet temperature rises to 34.4°C. The increase in temperature rise is roughly proportional to the increase in moisture removal. Process outlet temperature versus reactivation air temperature is shown in Figure 12, which shows the effect of increasing reactivation temperature when the moisture content of the process inlet air stays constant. If the reactivation sector is heated to elevated temperatures, more moisture is removed on the process side, so the temperature rise from latent-to-sensible heat conversion is slightly greater. In this constant-moisture inlet situation, if the reactivation sector is very hot, more heat is carried from reactivation to process as the desiccant mass rotates from reactivation to process. Figure 12 shows that if reactivation air is heated to 65°C, the process air leaves the dehumidifier at 29.4°C. If reactivation air is heated to 120°C, the process air outlet temperature rises to 31.6°C. The 2.2 K increase in process air temperature is primarily caused by the increase in heat carried over from reactivation. One consequence of this relationship is that desiccant equipment manufacturers constantly seek to minimize the “waste mass” in a desiccant dehumidifier, to avoid heating and cooling extra, nonfunctional material such as heavy desiccant support structures or extra desiccant that air cannot reach. Theoretically, the most efficient desiccant dehumidifier has an infinitely large effective desiccant surface combined with an infinitely low mass.

Use of Cooling In process drying applications, desiccant dehumidifiers are sometimes used without additional cooling because the temperature increase from dehumidification helps the drying process. In semiprocess applications such as controlling frost formation in supermarkets, excess sensible cooling capacity may be present in the system as a whole, so warm air from a desiccant unit is not a major consideration. However, in most other applications for desiccant dehumidifiers, provision must be made to remove excess sensible heat from process air after dehumidification. In a liquid-desiccant system, heat is removed by cooling the liquid desiccant itself, so process air emerges from the desiccant medium at the appropriate temperature. In a solid-desiccant system, cooling is accomplished downstream of the desiccant bed with cooling coils. The source of this cooling can affect the system’s operating economics.

24.7

In some systems, postcooling is accomplished in two stages, with cooling tower water as the primary source followed by compression or absorption cooling. Alternatively, various combinations of indirect and direct evaporative cooling equipment are used to cool the dry air leaving the desiccant unit. In systems where the latent and sensible loads peak at different times, the sensible cooling capacity of the basic air-conditioning system is sufficient to handle the process air temperature rise without additional equipment. Systems in moderate climates with high ventilation requirements often combine high latent loads in the morning, evening, and night with high sensible loads at midday, so desiccant subsystems to handle latent loads are especially economical.

Using Units in Series Solid-desiccant dehumidifiers are often used to provide air at low dew points. Applications requiring large volumes of air at moisture contents of 0.7 g/kg (–18°C dew point) are quite common and can be easily achieved by rotary desiccant units in a single pass beginning with inlet moisture contents as high as 6.5 g/kg (7.5°C dew point). Some solid-desiccant units commonly deliver air at 0.3 g/kg (–28°C dew point) without special design considerations. Where extremely low dew points must be achieved, or where air leakage inside the unit may be a concern, two desiccant dehumidifiers can be placed in series, with dry air from one unit feeding both process and reactivation air to a second unit. The second unit delivers very dry air, because there is reduced risk of moisture being carried over from reactivation to process air when dry air is used to reactivate the second unit.

Industrial Rotary Desiccant Dehumidifier Performance Figures 8 to 12 are based on the generalized model of a desiccant dehumidifier described by Worek and Zheng (1991). The model, however, differs somewhat from commercial products. Figure 13 shows typical performance of an industrial desiccant dehumidifier.

2.4

EQUIPMENT RATINGS

ASHRAE Standard 139 describes the parameters used in the desiccant industry to calculate desiccant dehumidifier performance: • • • •

Moisture removal capacity (MRC) Regeneration specific heat input (RSHI) Process outlet temperature Pressure drop through the wheel

These performance parameters can be obtained from any manufacturer by means of performance curves or selection software. However, they are generally rated at sea-level conditions. ASHRAE research project RP-1339 (Fumo and Mago 2011) investigated how the performance parameters are affected by altitude. The main objectives were to test a desiccant dehumidifier both at sea level and at altitude, and then to develop a general methodology that can be applied to any desiccant wheel to estimate performance at altitude. The test results showed that, by keeping process and regeneration mass flow rates, inlet temperature, and inlet humidity ratio constant between sea level and altitude, performance of a solid-desiccant dehumidifier at altitudes up to 1524 m is constant regardless of altitude, except for pressure drop through the desiccant wheel. Therefore, the MRC, RSHI, and process-out temperature obtained from the manufacturer’s performance data at sea level do not require correction for altitude. Note that, when air at altitude is near saturation, its humidity ratio can be above saturation at sea level. In this situation, determine the relative humidity of the air at altitude and use the same mass flow rate, temperature, and relative humidity with the manufacturer’s sea-level performance data when determining MRC and pro-

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cess outlet temperature. The humidity ratio change (grain depression) used to calculate MRC will be the same at altitude and should be subtracted from the inlet humidity ratio at altitude to determine the exit condition. The pressure drop through the wheel obtained from the manufacturer’s performance data at sea level must be corrected by the ratio of atmospheric pressure at sea level and altitude using the density ratio methodology: 0 Pz = ----- P0 z 1 - P0 Pz = -----------------------------------------------------------------– 5 5.2559  1 – 2.25577  10 Z  where

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= = = = =

wheel pressure drop at altitude, kPa wheel pressure drop at sea level, kPa moist air density at altitude, kg/m3 moist air density at sea level, kg/m3 elevation, m

Example 1. Find the exit air humidity ratio and wheel pressure drop for a dehumidifier operating at Denver International Airport with process air precooled to 12.78°C and 98.0% rh. Solution: Design inlet conditions are based on the 1% design conditions from Chapter 14 of the 2017 ASHRAE Handbook—Fundamentals. Site altitude: 1655 m Process flow rate 1359 m3/h Process inlet humidity ratio Wpi 11.08 g/kg Regeneration flow rate 454 m3/h Regeneration inlet temperature 19.9°C (before heater) Regeneration inlet humidity ratio Wri 13.17 g/kg (73.5% rh) Regeneration inlet temperature 120°C (after heater) Convert humidity ratios at altitude to equivalent values at sea level. 12.78°C 98% 9.04 g/kg 19.9°C 73.8% 10.75 g/kg

Obtain sea-level performance data from dehumidifier manufacturer using above psychrometric inputs. Process airstream out: Dry-bulb temperature: Humidity ratio Wpo Pressure drop Regeneration airstream out: Dry-bulb temperature: Humidity ratio Wro Pressure drop

1 -  214  = 262 Pa P r , z = -------------------------------------------------------------------------5.2559 –5 1 – 2.25577  10  1655 

EQUIPMENT OPERATING RECOMMENDATIONS

Desiccant equipment tends to be very durable if maintained properly, often operating at high efficiency 30 years after it was originally installed. Required maintenance is specific to the type of desiccant equipment, the application, and the installation. Each system requires a somewhat different maintenance and operational routine. The information in this section does not substitute for or supersede any recommendations of equipment manufacturers, and it is not a substitute for owners’ experience with specific applications.

Process Air Filters

The following example shows the method for rating desiccant equipment at altitude.

Process airstream: Dry-bulb temperature: Relative humidity Humidity ratio Wp Regeneration airstream: Dry-bulb temperature: Relative humidity Humidity ratio Wr

1 -  157  = 192 Pa P p , z = -------------------------------------------------------------------------5.2559 –5 1 – 2.25577  10  1655 

2.5

or

Pz P0 z 0 Z

MRCz = MRC0 = 11.09 kg/h Restate wheel pressure drops for altitude.

36.5°C 2.24 g/kg 157 Pa 48.9°C 31.06 g/kg 214 Pa

Calculate humidity ratio change and MRC, where MRC = m3/h × 0.0012 × Wp = kg/h Wp = 9.04 – 2.24 = 6.8 g/kg Wr = 31.06 – 10.75 = 20.31 g/kg MRC = 1360 × 0.0012 × 6.8 = 11.09 kg/h Restate outlet conditions for altitude. Wpo, z = Wpi, z – Wp = 11.08 – 6.8 = 4.27 g/kg Wro, z = Wri, z + Wr = 13.17 + 20.31 = 33.48 g/kg

Clean filters are the most important item in a maintenance routine. If a solid desiccant is clogged with particulates, or if a liquid desiccant’s sorption characteristics are changed by entrained particulates, the material may have to be replaced prematurely. Filters are much less expensive and much easier to change than the desiccant. Although each application is different, the desiccant usually must be replaced, replenished, or reconditioned after 5 to 10 years of operation. Without attention to filters, desiccant life can be reduced to 1 or 2 years of operation or less. Filters should be checked at least four times per year, and more frequently when airstreams are heavily laden with particulates. Differential pressure sensors across the filter bank to detect loading and alarm a building management system (BMS) may be an alternative to manual inspection. The importance of filter maintenance requires that filter racks and doors on desiccant systems be freely accessible and that enough space be allowed to inspect, remove, and replace filters. Optimal design ensures that filter locations, as well as the current condition of each filter, are clearly visible to maintenance personnel.

Reactivation/Regeneration Filters Air is filtered before entering the heater of a desiccant unit. If filters are clogged and airflow is reduced, unit performance may be reduced because there is not enough air to carry all the moisture away from the desiccant. If electrical elements or gas burners are used to heat the air, reducing airflow may damage the heaters. Thus, the previous suggestions for maintaining process air filters also apply to reactivation/regeneration filters.

Liquid-Phase Strainers In liquid-desiccant systems, clean strainers are an important item in a maintenance routine. Depending on the atmospheric cleanliness and air-phase filtration used, strainers in both the process and regeneration desiccant systems require rinsing every 3 to 6 months.

Reactivation/Regeneration Ductwork Air leaving the reactivation/regeneration section is hot and moist. When units first start up in high-moisture-load applications, the reactivation air may be nearly saturated and even contain water droplets. Thus, ductwork that carries air away from the unit should be corrosion resistant, because condensation can occur inside the ducts, particularly if the ducts pass through unheated areas in cool weather. If heavy condensation seems probable, the ductwork should be designed with drains at low points or arranged to let condensation flow out of the duct where the air is vented to the weather. The high

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Leakage All desiccant units produce dry air in part of the system. If humid air leaks into either the dry air ductwork or the unit itself, system efficiency is reduced. Energy is also wasted if dry air leaks out of the distribution duct connections. Therefore, duct connections for desiccant systems should be sealed tightly. In applications requiring very low dew points (below –12°C), the ductwork and desiccant system are almost always tested for leaks at air pressures above those expected during normal operation. In applications at higher dew points, similar leak testing is considered good practice and is recommended by many equipment manufacturers. Because desiccant equipment tends to be durably constructed, workers often drill holes in the dehumidifier unit casing to provide support for piping, ductwork, or instruments. Such holes eventually leak air, desiccant, or both. Designers should provide other means of support for external components so contractors do not puncture the system unnecessarily. Contractors installing desiccant systems should be aware that any holes made in the system must be sealed tightly using both mechanical means and sealant compounds. Sealants must be selected for long life at the working temperatures of the application and of the casing walls that have been punctured. For example, reactivation/regeneration sections often operate in a range from a cold winter ambient of –40°C to a heated temperature as high as 150°C. Process sections may operate in a range of –40°C at the inlet to 65°C at the outlet.

Airflow Indication and Control As explained in the section on Rotary Solid-Desiccant Dehumidifiers, performance depends on how quickly air passes through the desiccant; changes in air velocity affect performance. Thus, it is important to quantify the airflow rate through both the process and reactivation/regeneration parts of the unit. Unless both airflows are known, it is impossible to determine whether the unit is operating properly. In addition, if velocity exceeds the maximum design value, the air may carry desiccant particles or droplets out of the unit and into the supply air ductwork. Thus, manufacturers often provide airflow gages on larger equipment so the owner can be certain the unit is operating within the intended design parameters. Smaller equipment is not always provided with airflow indicators because precise performance may be less critical in applications such as small storage rooms. However, in any system using large equipment, or if performance is critical in smaller systems, unit airflow should be quantified and clearly indicated, so operating personnel can compare current flow rates through the system with design values. Many desiccant units are equipped with manual or automatic flow control dampers to control the airflow rate. If these are not provided with the unit, they should be installed elsewhere in the system. Airflows for process and reactivation/regeneration must be correctly set after all ductwork and external components are attached, but before the system is put into use.

Commissioning Heat and moisture on the dry-air side of desiccant equipment is balanced equally by the heat and moisture on the regeneration/ reactivation side. To confirm that a solid-desiccant system is operating as designed, the commissioning technician must measure airflow, temperature, and moisture on each side to calculate a mass balance. In liquid systems, these six measurements are taken on the process-air side. On the regenerator side, the liquid temperature is

24.9

read in the sump and at the spray head to confirm the regenerator’s heat transfer rate at peak-load conditions. If the dehumidification unit does not provide the means, the system should be designed to facilitate taking the readings that are essential to commissioning and troubleshooting. Provisions must be made to measure flow rates, temperatures, and moisture levels of airstreams as they enter and leave the desiccant. For liquid systems, provisions must be made for measuring the solution temperature and concentration at different points in the system. Four precautions for taking these readings at different points in a desiccant system follow. Airflow. Airflow instruments measure the actual volumetric flow rate, which must be converted to standard flow rate to calculate mass flow. Because temperatures in a desiccant system are often well above or below standard temperature, these corrections are essential. Air Temperature. Most airstreams in a desiccant system have temperatures between –20 and 150°C, but temperature can be widely varied and stratified as air leaves the desiccant in soliddesiccant systems. Air temperature readings must be averaged across the duct for accurate calculations. Readings taken after a fan tend to be more uniform, but corrections must be made for heat added by the fan itself. Process Air Moisture Leaving Dry Desiccant. In soliddesiccant equipment, air leaving the desiccant bed or wheel is both warm and dry: usually below 20% rh, often below 10%, and occasionally below 2%. Most low-cost instruments have limited accuracy below 15% rh, and all but the most costly instruments have an error of 2% rh. Consequently, to measure relative humidities near 2%, technicians use very accurate instruments such as manual dew cups or automated optical dew-point hygrometers. ASHRAE Standard 41.6 describes these instruments and procedures for their proper use. When circumstances do not allow the use of dew-point instruments, other methods may be necessary. For example, an air sample may need to be cooled to produce a higher, more easily measured relative humidity. Low humidity readings can be difficult to take with wet-bulb thermometers because the wet wick dries out very quickly, sometimes before the true wet-bulb reading is reached. Also, when the wet-bulb temperature is below the freezing point of water, readings take much longer, which may allow the wick to dry out, particularly in solid-desiccant systems where there may be considerable heat in the air leaving the desiccant. Therefore, wicks must be monitored for wetness. Many technicians avoid wet-bulb readings in air leaving a solid-desiccant bed, partly for these reasons, and partly because of the difficulty and time required to obtain average readings across the whole bed. Like air temperature, air moisture level leaving a solid-desiccant bed varies considerably; if taken close to the bed, readings must be averaged to obtain a true value for the whole air mass. When very low dew points are expected, the commissioning technician should be especially aware of limitations of the air-sampling system and the sensor. Even the most accurate sensors require more time to come to equilibrium at low dew points than at moderate moisture levels. For example, at dew points below –30°C, the sensor and air sample tubing may take many hours rather than a few minutes to equilibrate with the air being measured. Time required to come to equilibrium also depends on how much moisture is on the sensor before it is placed into the dry airstream. For example, taking a reading in the reactivation/regeneration outlet essentially saturates the sensor, so it will take much longer than normal to equilibrate with the low relative humidity of the process leaving air. Reactivation/Regeneration Air Moisture Leaving Desiccant. Air leaving the reactivation/regeneration side of the desiccant is warm and close to saturation. If the humidity measurement sensor is at ambient temperature, moisture may condense on its surface, dis-

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2020 ASHRAE Handbook—HVAC Systems and Equipment (SI)

torting the reading. It is good practice to warm the sensor (e.g., by taking the moisture reading in the warm, dry air of the processleaving airstream) before reading moisture in reactivation air. If a wet-bulb instrument is used, water for the wet bulb must be at or above the dry-bulb temperature of the air, or the instrument will read lower than the true wet-bulb temperature of the air.

Owners’ and Operators’ Perspectives Designers and new owners are strongly advised to consult other equipment owners and the manufacturer’s service department early in design to gain the useful perspective of direct operating experience (Harriman 2003).

2.6

APPLICATIONS FOR ATMOSPHERICPRESSURE DEHUMIDIFICATION

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Preservation of Materials in Storage Special moisture-sensitive materials are sometimes kept in dehumidified warehouses for long-term storage. Tests by the Bureau of Supplies and Accounts of the U.S. Navy (DOD 1987, 2004) concluded that 40% rh is a safe level to control deterioration of materials. Others (Sterling et al. 1985) have indicated that 60% rh is low enough to control microbiological attack. With storage at 40% rh, no undesirable effects on metals or rubber compounds have been noted. Some organic materials such as sisal, hemp, and paper may lose flexibility and strength, but they recover these characteristics when moisture is regained. Commercial storage relies on similar equipment for applications that include beer fermentation rooms, meat storage, and penicillin processing, as well as storage of machine tools, candy, food products, furs, furniture, seeds, paper stock, and chemicals. For recommended conditions of temperature and humidity, refer to Chapters 21 and 28 to 42 of the 2017 ASHRAE Handbook—Refrigeration.

an effort to improve indoor air quality. Large amounts of humid ventilation air carry enough moisture to upset the operation of highefficiency cooling equipment, which is generally designed to remove more sensible heat than moisture (Kosar et al. 1998). Removing excess moisture from the ventilation air with a ventilation dehumidification system improves both humidity control and cooling system effectiveness. For example, field tests suggest that when the environment is kept dry, occupants prefer warmer temperatures, which in turn saves cooling operational costs (Fischer and Bayer 2003). Also, cooling equipment is often oversized to remove ventilation-generated moisture. Predrying with a desiccant system may reduce the building’s construction cost, if excess cooling capacity is removed from the design (Spears and Judge 1997). Figures 14, 15, and 16 show the relative importance of moisture load from ventilation, how a commercial building can use a desiccant system to remove that load, and how such a system is applied in the field (Harriman et al. 2001).

Process Dehumidification Requirements for dehumidification in industrial processes are many and varied. Some of these processes are as follows: • Metallurgical processes, with controlled-atmosphere annealing of metals • Conveying hygroscopic materials • Film drying • Manufacturing candy, chocolate, and chewing gum • Manufacturing drugs and chemicals • Manufacturing plastic materials • Manufacturing laminated glass • Packaging moisture-sensitive products • Assembling motors and transformers • Solid propellant mixing • Manufacturing electronic components, such as transistors and microwave components • Processing poultry and other meats • Processing food and other powders

Fig. 14

Typical Peak Moisture Loads for Medium-Sized Retail Store in Atlanta, Georgia (Harriman et al. 2001)

For information about the effect of low-dew-point air on drying, refer to Chapters 20, 22, 25, and 30 of the 2019 ASHRAE Handbook—HVAC Applications.

Ventilation Air Dehumidification Over a full year, ventilation air loads a cooling system with much more moisture than heat. Except in desert and high-altitude regions, ventilation moisture loads in the United States exceed sensible loads by at least 3:1, and often by as much as 5:1 (Harriman et al. 1997). Consequently, desiccant systems are used to dehumidify ventilation air before it enters the main air-conditioning system. Drying ventilation air has gained importance because building codes mandate larger amounts of ventilation air than in the past, in

Fig. 15 Predrying Ventilation Air to Dehumidify a Commercial Building (Harriman et al. 2003)

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Desiccant Dehumidification and Pressure-Drying Equipment Ventilation dehumidification is most cost effective for buildings with high ventilation airflow rather than high sensible loads from internal heat or from heat transmitted through the building envelope. As a result, this approach is most common in densely occupied buildings such as schools, theaters, elder care facilities, large-scale retail buildings, and restaurants (Harriman 2003).

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Condensation Prevention Many applications require moisture control to prevent condensation. Airborne moisture condenses on cold cargo in a ship’s hold when it reaches a moist climate. Moisture condenses on a ship when the moist air in its cargo hold is cooled by the hull and deck plates as the ship passes from a warm to a cold climate. A similar problem occurs when aircraft descend from high, cold altitudes into a high dew point at ground level. Desiccant dehumidifiers are used to prevent condensation inside the airframe and avionics that leads to structural corrosion and failure of electronic components. In pumping stations and sewage lift stations, moisture condenses on piping, especially in the spring when the weather warms and water in the pipes is still cold. Dehumidification is also used to prevent airborne moisture from dripping into oil and gasoline tanks and open fermentation tanks. Electronic equipment is often cooled by refrigeration, and dehumidifiers are required to prevent internal condensation of moisture. Electronic and instrument compartments in missiles are purged with low-dew-point air before launching to prevent malfunctioning caused by condensation. Waveguides and radomes are also usually dehumidified, as are telephone exchanges and relay stations. For proper operation of their components, missile and radar sites depend largely on prevention of condensation on interior surfaces.

Dry Air-Conditioning Systems Cooling-based air-conditioning systems remove moisture from air by condensing it onto cooling coils, producing saturated air at a

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lower absolute moisture content. In many circumstances, however, there is a benefit to using a desiccant dehumidifier to remove the latent load from the system, avoiding problems caused by condensation, frost, and high relative humidity in air distribution systems. For example, low-temperature product display cases in supermarkets operate less efficiently when humidity in the store is high because condensate freezes on the cooling coils, increasing operating cost. Desiccant dehumidifiers remove moisture from the air, using rejected heat from refrigeration condensers to reduce the cost of desiccant reactivation. Combining desiccants and conventional cooling can lower installation and operating costs (Calton 1985). For information on the effect of humidity on refrigerated display cases, see Chapter 15 of the 2018 ASHRAE Handbook—Refrigeration and Chapter 2 of the 2019 ASHRAE Handbook—HVAC Applications. Air conditioning in hospitals, nursing homes, and other medical facilities is particularly sensitive to biological contamination in condensate drain pans, filters, and porous insulation inside ductwork. These systems often benefit from drying ventilation air with a desiccant dehumidifier before final cooling. Condensate does not form on cooling coils or drain pans, and filters and duct lining stay dry so that mold and mildew cannot grow inside the system. Refer to ASHRAE Standard 62.1 for guidance concerning maximum relative humidity in air distribution systems. Chapter 8 of the 2019 ASHRAE Handbook—HVAC Applications has information on ventilation of health care facilities. Hotels and large condominium buildings historically suffer from severe mold and mildew problems caused by excessive moisture in the building structure. Desiccant dehumidifiers are sometimes used to dry ventilation air so it can act as a sponge to remove moisture from walls, ceilings, and furnishings (AHMA 1991). Dehumidified ventilation air that positively pressurizes the building may help counter moist air infiltration. See Chapter 6 of the 2019 ASHRAE Handbook—HVAC Applications for more information on ventilating hotels and similar structures.

Fig. 16 Typical Rooftop Arrangement for Drying Ventilation Air Centrally, Removing Moisture Load from Cooling Units (Harriman et al. 2001)

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2020 ASHRAE Handbook—HVAC Systems and Equipment (SI)

Like supermarkets, ice rinks have large exposed cold surfaces that condense and freeze moisture in the air, particularly during spring and summer. Desiccant dehumidifiers remove excess humidity from air above the rink surface, preventing fog and improving both the ice surface and operating economics of the refrigeration plant. For recommended temperature and humidity for ice rinks, see Chapter 44 of the 2018 ASHRAE Handbook—Refrigeration.

Indoor Air Quality Contaminant Control

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Desiccant sorption is not restricted to water vapor. Both liquid and solid desiccants collect both water and large organic molecules at the same time (Hines et al. 1993). As a result, desiccant systems can be used to remove volatile organic compound (VOC) emissions from building air systems. In addition to preventing growth of mold, mildew, and bacteria by keeping buildings dry, desiccant systems can supplement filters to remove bacteria from the air itself. This is particularly useful for hospitals, medical facilities, and related biomedical manufacturing facilities where airborne microorganisms can cause costly problems. The usefulness of certain liquid and solid desiccants in such systems stems from their ability to either kill microorganisms or avoid sustaining their growth (Battelle 1971; SUNY Buffalo School of Medicine 1988).

Testing Many test procedures require dehumidification with sorption equipment. Frequently, other means of dehumidification may be used with sorbent units, but the low moisture content required can be obtained only by liquid or solid sorbents. Some typical testing applications are as follows: • • • • • •

Wind tunnels Spectroscopy rooms Paper and textile testing Bacteriological and plant growth rooms Dry boxes Environmental rooms and chambers

3.

DESICCANT DRYING AT ELEVATED PRESSURE

The same sorption principles that pertain to atmospheric dehumidification apply to drying high-pressure air and process or other gases. The sorbents described previously can be used with equal effectiveness.

3.1

EQUIPMENT TYPES

Absorption Solid absorption systems use a calcium chloride desiccant, generally in a single-tower unit that requires periodic replacement of the desiccant that is dissolved by the absorbed moisture. Normally, inlet air or gas temperature does not exceed 32 to 38°C saturated. The rate of desiccant replacement is proportional to the moisture in the inlet process flow. A dew-point depression of 10 to 20 K at pressure can be obtained when the system is operated in the range of 15 to 38°C saturated entering temperature and 700 kPa (gage) operating pressure. At lower pressures, the ability to remove moisture decreases in proportion to absolute pressure. Such units do not require a power source for operation because the desiccant is not regenerated. However, additional desiccant must be added to the system periodically.

Adsorption Drying with an adsorptive desiccant such as silica gel, activated alumina, or a molecular sieve usually incorporates regeneration

Fig. 17

Typical Performance Data for Solid Desiccant Dryers at Elevated Pressures

equipment, so the desiccant can be reactivated and reused. These desiccants can be readily reactivated by heat, purging with dry gas, or both. Depending on the desiccant selected, dew-point performance expected is in the range of –40 to –75°C measured at the operating pressure with inlet conditions of 32 to 38°C saturated and 700 kPa (gage). Figure 17 shows typical performance using activated alumina or silica gel desiccant. Equipment design may vary considerably in detail, but most basic adsorption units use twin-tower construction for continuous operation, with an internal or external heat source, with air or process gas as the reactivation purge for liberating moisture adsorbed previously. A single adsorbent bed may be used for intermittent drying requirements. Adsorption units are generally constructed in the same manner as atmospheric-pressure units, except that the vessels are suitable for the operating pressure. Units have been operated successfully at pressures as high as 40 MPa. Prior compression or cooling (by water, brine, or refrigeration) to below the dew point of the gas to be dried reduces the total moisture load on the sorbent, permitting the use of smaller drying units. The cost of compression, cooling, or both must be balanced against the cost of a larger adsorption unit. The many different dryer designs can be grouped into the following basic types: Heat-reactivated, purge dryers. Normally operating on 4 h (or longer) adsorption periods, these dryers are generally designed with heaters embedded in the desiccant. They use a small portion of dried process gas as a purge to remove the moisture liberated during reactivation heating (see Figure 18). Heatless dryers. These dryers operate on a short adsorption period (usually 60 to 300 s). Depressurizing gas in the desiccant tower lowers the vapor pressure, so adsorbed moisture is liberated from the desiccant and removed by a high purge rate of the dried process gas. Using an ejector reduces the purge gas requirements. Convection dryers. These dryers usually operate on 4 h (or longer) adsorption periods and are designed with an external heater and cooler as the reactivation system. Some designs circulate reactivation process gas through the system by a blower; others divert some or all of the process gas flow through the reactivation system before adsorption. Both heating and cooling are by convection. Radiation dryers. Also operating on 4 h (or longer) adsorption periods, radiation dryers are designed with an external heater and blower to force heated atmospheric air through the desiccant tower for reactivation. Desiccant tower cooling is by radiation to atmosphere.

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Desiccant Dehumidification and Pressure-Drying Equipment

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Equipment Testing Dry, high-pressure air is used extensively for testing refrigeration condensing units to ensure tightness of components and to prevent moisture infiltration. Similarly, dry inert gas is used in testing copper tubing and coils to prevent corrosion or oxidation. Manufacture and assembly of solid-state circuits and other electronic components require exclusion of all moisture, and final testing in dry boxes must be carried out in moisture-free atmospheres. Simulation of dry high-altitude atmospheres for testing aircraft and missile components in wind tunnels requires extremely low dew-point conditions.

REFERENCES

Fig. 18 Typical Adsorption Dryer for Elevated Pressures

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3.2

APPLICATIONS

Material Preservation Generally, materials in storage are preserved at atmospheric pressure, but a few materials are stored at elevated pressures, especially when the dried medium is an inert gas. These materials deteriorate when subjected to high relative humidity or oxygen content in the surrounding medium. Drying high-pressure air, subsequently reduced to 24 to 70 kPa (gage), has been used most effectively in pressurizing coaxial cables to eliminate electrical shorts caused by moisture infiltration. This same principle, at somewhat lower pressures, is also used in waveguides and radomes to prevent moisture film on the envelope.

Process Drying of Air and Other Gases Drying instrument air to a dew point of –40°C, particularly where air lines are outdoors or exposed to temperatures below the dew point of air leaving the aftercooler, prevents condensation or freeze-up in instrument control lines. To prevent condensation and freezing, it is necessary to dry plant air used for pneumatically operated valves, tools, and other equipment where piping is exposed to low ambient temperatures. Additionally, dry air prevents rusting of the air lines, which produces abrasive impurities, causing excessive wear on tools. Industrial gases or fuels such as natural gas are dried. For example, fuels (including natural gas) are cleaned and dried before storage underground to ensure that valves and transmission lines do not freeze from condensed moisture during extraordinarily cold weather, when the gas is most needed. Propane must also be clean and dry to prevent ice accumulation. Other gases, such as bottled oxygen, nitrogen, hydrogen, and acetylene, must have a high degree of dryness. In liquid oxygen and ozone manufacturing, air supplied to the process must be clean and dry. Drying air or inert gas for conveying hygroscopic materials in a liquid or solid state ensures continuous, trouble-free plant operation. Normally, gases for this purpose are dried to a –40°C dew point. Purging and blanketing operations in the petrochemical industry depend on using dry inert gas for reducing problems such as explosive hazards and the reaction of chemicals with moisture or oxygen.

ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal.ashrae .org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org/bookstore. AHMA. 1991. Mold and mildew in hotels and motels. Executive Engineers Committee Report. American Hotel and Motel Association, Washington, D.C. ASHRAE. 2006. Standard method for measurement of moist air properties. ANSI/ASHRAE Standard 41.6-1994 (RA 2006). ASHRAE. 2010. Ventilation for acceptable indoor air quality. ANSI/ ASHRAE Standard 62.1-2010. ASHRAE. 2015. Method of testing for rating desiccant dehumidifiers utilizing heat for the regeneration process. ANSI/ASHRAE Standard 139-2015. Battelle Memorial Institute. 1971. Project N-0914-5200-1971. Battelle Memorial Institute, Columbus, OH. Calton, D.S. 1985. Application of a desiccant cooling system to supermarkets. ASHRAE Transactions 91(1B):441-446. DOD. 1987. Military handbook—Covered storage. MIL-HDBK-1032/2. U.S. Department of Defense, Washington, D.C. DOD. 2004. Unified facilities criteria (UFC), design: Covered storage. UFC-4-442-01N. U.S. Department of Defense, Washington, D.C. Fischer, J.C., and C.W. Bayer. 2003. Report card on humidity control. ASHRAE Journal 45(5):30-39. Fumo, N., and P. Mago. 2011. Selection of desiccant equipment at altitude. ASHRAE Research Project RP-1339, Final Report. Harriman, L.G., III. 2003. 20 years of commercial desiccant systems: Where they’ve been, where they are now and where they’re going. Heating/Piping/Air Conditioning Engineering (June & July):43-54. Harriman, L.G., III, D. Plager, and D. Kosar. 1997. Dehumidification and cooling loads from ventilation air. ASHRAE Journal 39(11):37-45. Harriman, L.G., III, G. Brundrett, and R. Kittler. 2001. Humidity control design guide for commercial and institutional buildings. ASHRAE. Hines, A.L., T.K. Ghosh, S.K. Loyalka, and R.C. Warder, Jr. 1993. Investigation of co-sorption of gases and vapors as a means to enhance indoor air quality. Gas Research Institute, Chicago. Available from the National Technical Information Service, Springfield, VA. Order PB95104675. Kosar, D.R., M.J. Witte, D.B. Shirey, and R.L. Hedrick. 1998. Dehumidification issues of Standard 62-1989. ASHRAE Journal 40(5):71-75. Spears, J.W., and J.J. Judge. 1997. Gas-fired desiccant system for retail superstore. ASHRAE Journal 39(10):65-69. Sterling, E.M., A. Arundel, and T.D. Sterling. 1985. Criteria for human exposure to humidity in occupied buildings. ASHRAE Transactions 91(1B):611-622. SUNY Buffalo School of Medicine. 1988. Effects of glycol solutions on microbiological growth. Niagara Blower Report 03188. Worek, W., and W. Zheng. 1991. UIC IMPLICIT rotary desiccant dehumidifier finite difference program. University of Illinois at Chicago, Department of Mechanical Engineering.

BIBLIOGRAPHY ASHRAE. 1992. Desiccant cooling and dehumidification, L. Harriman, ed. ASHRAE. 2009. Method of test for rating desiccant-based dehumidification equipment. ANSI/ASHRAE Standard 174-2009.

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2020 ASHRAE Handbook—HVAC Systems and Equipment (SI) Lowenstein, A.I., and R.S. Gabruk. 1992. The effect of regenerator performance on a liquid-desiccant air conditioner. ASHRAE Transactions 98(1):704-711. Meckler, M. 1994. Desiccant-assisted air conditioner improves IAQ and comfort. Heating/Piping/Air Conditioning Engineering 66(10):75-84. Pesaran, A., and T. Penney. 1991. Impact of desiccant degradation on cooling system performance. ASHRAE Transactions 97(1):595-601. Vineyard, E.A., J.R. Sand, and D.J. Durfee. 2000. Parametric analysis of variables that affect the performance of a desiccant dehumidification system. ASHRAE Transactions 106(1):87-94.

ADDITIONAL INFORMATION ASHRAE Technical Committee 8.12 posts updated and additional information regarding desiccant equipment and systems on the committee’s subsection of the ASHRAE website, located at tc0812.ashraetcs.org.

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Bradley, T.J. 1994. Operating an ice rink year-round by using a desiccant dehumidifier to remove humidity. ASHRAE Transactions 100(1): 116-131. Collier, R.K. 1989. Desiccant properties and their effect on cooling system performance. ASHRAE Transactions 95(1):823-827. Harriman, L.G., III. 1990. The dehumidification handbook. Munters Cargocaire, Amesbury, MA. Harriman, L.G., III. 1996. Applications engineering manual for desiccant systems. American Gas Cooling Center, Arlington, VA. Harriman, L.G., III, and J. Judge. 2002. Dehumidification equipment advances. ASHRAE Journal 44(8):22-29 Jones, B.W., B.T. Beck, and J.P. Steele. 1983. Latent loads in low humidity rooms due to moisture. ASHRAE Transactions 89(1A):35-55. Lowenstein, A.I., and R.S. Gabruk. 1992. The effect of absorber design on the performance of a liquid-desiccant air conditioner. ASHRAE Transactions 98(1):712-720.

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Related Commercial Resources CHAPTER 25

MECHANICAL DEHUMIDIFIERS AND RELATED COMPONENTS Mechanical Dehumidifiers ............................................................................................................ 25.1 Controls and Sensors .................................................................................................................... 25.9 Installation and Service Considerations ....................................................................................... 25.9 Wraparound Heat Exchangers.................................................................................................... 25.10

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T

HE correct moisture level in the air is important for health and comfort. Controlling humidity and condensation is important to prevent moisture damage and mold or mildew development, thus protecting buildings and occupants, and preserving building contents This chapter covers mechanical dehumidification using a cooling process only, including basic dehumidifier models (with moisture removal capacity of less than 1.4 kg/h) used for home basements and small storage areaso, as well as larger sizes required for commercial applications. Requirements for testing and rating mechanical dehumidifiers’ energy efficiency are published in AHRI Standards 911 and 921. These dehumidifiers are used for applications where dew points of 1.7 to 4.4°C and above are maintained. For applications requiring dew points below 1.7°C and for other methods of dehumidification, see Chapter 24. Commercial applications for mechanical dehumidifiers include the following: • • • • • • • • • • • • • • •

Indoor swimming pools Makeup air treatment Ice rinks Dry storage Schools Hospitals Office buildings Museums, libraries, and archives Restaurants Hotels and motels Assisted living facilities Supermarkets Manufacturing plants and processes Water making Industrial drying

In addition, an air-to-air heat exchanger (e.g., heat pipe, coil runaround loop, fixed-plate heat exchanger, rotary heat exchanger) may be used to enhance moisture removal by a mechanical dehumidifier or air conditioner. The section on Wraparound Heat Exchangers discusses how dehumidification processes can be improved by using such a device. Other uses of air-to-air heat exchangers are covered in Chapter 26.

1.

MECHANICAL DEHUMIDIFIERS

Mechanical dehumidifiers remove moisture by passing air over a surface that has been cooled below the air’s dew point. This cold surface may be the exterior of a chilled-water coil or a direct-expansion refrigerant coil. To prevent overcooling the space (and avoid the need to add heat energy from another source), a mechanical dehumidifier also usually has means to reheat the air, normally using recovered and The preparation of this chapter is assigned to TC 8.10, Mechanical Dehumidification Equipment and Heat Pipes.

Copyright © 2020, ASHRAE

recycled energy (e.g., recovering heat from hot refrigerant vapor in the refrigeration circuit). Using external energy input for reheat is wasteful and is prohibited or limited in many countries (see ASHRAE Standard 90.1). A mechanical dehumidifier differs from a typical off-the-shelf air conditioner in that the dehumidifier usually has a much lower sensible heat ratio (SHR). The dehumidifier starts the compressor on a call for dehumidification, whereas an air conditioner starts the compressor on a call for sensible cooling. Typically, a room dehumidifier has an SHR of 0.6 or less, compared to a standard air-conditioning system of 0.8 SHR. Dehumidifiers must also allow condensation from the cooling coil to drain easily from the coils. They may need air velocities over the cooling coil lower than those for a typical air conditioner, to improve moisture runoff and minimize carryover of condensed moisture. In addition, the need to introduce code-mandated ventilation air may require that outdoor air be treated to avoid introducing excessive moisture. Basic strategies include precooling outdoor air entering the air-conditioning evaporator coil, or providing a separate system to provide properly conditioned outdoor air. For some lowdew-point (below 7°C) applications, mechanical dehumidification may be used as the first stage, with desiccant dehumidification for the final stage to maximize efficiency and minimize installed cost. Although the main purpose of a mechanical dehumidifier is to remove moisture from the air, many features can be incorporated for various applications, such as • Dehumidifying and cooling (no reheat) • Dehumidifying with partial reheat (leaving dry-bulb temperature is cooler than with a dehumidifier with full reheat) • Dehumidifying with full reheat • Dehumidifying with heat recovery to various heat sinks • Dehumidification capacity modulation • Reheat capacity modulation • Ventilation air introduction • Auxiliary space or water heating Often, mechanical dehumidifiers can be incorporated in a system to use waste heat from mechanical cooling (e.g., heat rejection to a swimming pool, whirlpool, domestic hot water, heat pump loop, chilled-water loop, or remote air-cooled condenser). Outdoor dehumidifiers should be protected against internal moisture condensation when winter conditions are severe, because of the higher dew-point temperature of air circulating in the unit.

Psychrometrics of Dehumidification Figures 1 and 2 show the process of air moving over the dehumidifying coil and being reheated. Many manufacturers size the dehumidifying coil for only part of the unit total airflow. This allows them to introduce outdoor air after the dehumidifying coil and allows more unit airflow for the condenser reheat coil to maintain proper system refrigeration pressures when rejecting refrigerant heat to the airstream. Figures 3 and 4 show a bypass air process.

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25.2

2020 ASHRAE Handbook—HVAC Systems and Equipment (SI)

Fig. 1

Dehumidification Process Points

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Fig. 3 Psychrometric Diagram of Typical Dehumidification Process with Bypass Air

Fig. 2 Psychrometric Diagram of Typical Dehumidification Process In Figures 1 and 2, air enters the dehumidifying coil at point A. The dehumidifying coil removes sensible heat (SH) and latent heat (LH) from the airstream. The dehumidified, cooled air leaves the coil at its saturation temperature at point B. The total heat removed (TH) is the net cooling capacity of the system. In reheating, the refrigerant (hot gas) rejects heat it has obtained from multiple sources. Sensible heat absorbed in the air-cooling process is rejected to air leaving the cooling coil. This air is at point C, which is the same dry-bulb temperature as the entering air minus the moisture content. This heat is also rejected into the airstream, raising the air temperature to point D. Also, nearly all electric power required to drive the refrigeration cycle is converted to heat. This portion of heat rejection raises the air leaving temperature to point E. This process assumes that all heat is rejected by the refrigerant reheat coil. Depending on refrigerant system complexity, any part of the total heat rejection can be diverted to other heat exchangers (condensers/desuperheaters). Dehumidifier supply air temperatures can be controlled between 10 and 35°C. However, system design should not rely on a mechanical dehumidifier as a dependable heat source for space heating, because heat is only available when the unit is operating. In Figures 3 and 4, air enters the dehumidifying coil at point A. The dehumidifying coil removes sensible heat (SH) and latent heat (LH) from the airstream. The dehumidified, cooled air leaves the coil at its saturation temperature at point B. The total heat removed (TH) is the net cooling capacity of the system. In Figure 3, the air at B mixes with the bypass air to get to point C. In reheating, the refrigerant (hot gas) rejects heat it has obtained

Fig. 4

Dehumidification Process Points with Bypass Air

from three sources. First, sensible heat absorbed in the air-cooling process is rejected to air leaving the cooling coil. This air is at point D, which is the same dry-bulb temperature as the entering air minus the moisture content. Next, the latent heat removal that causes the moisture to condense also adds heat to the hot refrigerant gas. This heat is also rejected into the airstream, raising the air temperature to point E. Finally, nearly all electric power required to drive the refrigeration cycle is converted to heat. This portion of heat rejection raises the air’s leaving temperature to point F. This process assumes that all heat is rejected by the refrigerant reheat coil. Depending on refrigerant system complexity, any part of the total heat rejection can be diverted to other heat exchangers (condensers/desuperheaters). Dehumidifier supply air temperatures can be controlled between 18 and 40°C. However, system design should not rely on a mechanical dehumidifier as a dependable heat source for space heating, because heat is only available when the unit is operating.

Residential Dehumidifiers Portable Dehumidifiers. These are smaller (usually less than 3.5 kW), simpler versions of commercial dehumidifiers. They are self-contained and easily movable. They are designed to be used in localized areas, such as basements or other high-moisture areas. As shown in Figure 5, a single fan draws humid room air through the cold coil, removing moisture that either drains into the water

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Mechanical Dehumidifiers and Related Components receptacle or passes through the cabinet into some other means of disposal. The cooled air passes through the condenser, reheating the air. Portable dehumidifiers ordinarily maintain satisfactory humidity levels in an enclosed space when the airflow rate and unit placement move the entire air volume of the space through the dehumidifier once an hour. Design and Construction. Portable dehumidifiers use hermetic motor-compressors; the refrigerant condenser is usually conventional finned tube. Refrigerant flow is usually controlled by a capillary tube, although some high-capacity dehumidifiers use an expansion valve. A propeller fan moves air through the unit at typical airflows of 70 to 120 L/s. The refrigerated surface (evaporator) is usually a bare-tube coil, although finned-tube coils can be used if they are spaced to allow rapid runoff of water droplets. Vertically disposed bare-tube coils tend to collect smaller drops of water, promote quicker runoff, and result in less condensate reevaporation compared to finned-tube or horizontally arranged bare-tube coils. Continuous bare-tube coils, wound in a flat circular spiral (sometimes with two coil layers) and mounted with the flat dimension of the coil in the vertical plane, are a good design compromise because they have most of the advantages of the vertical bare-tube coil. Evaporators are protected against corrosion by finishes such as waxing, painting, or anodizing (on aluminum). Waxing reduces the wetting effect that promotes condensate formation; however, tests on waxed versus nonwaxed evaporator surfaces show negligible loss of capacity. Thin paint films do not have an appreciable effect on capacity. Removable water receptacles, provided with most dehumidifiers, hold 7.5 to 11 L and are usually made of plastic to withstand corrosion. Easy removal and handling without spillage are important. Most dehumidifiers also provide either a means of attaching a flexible hose to the water receptacle or a fitting provided specially for that purpose, allowing direct gravity drainage to another means of disposal external to the cabinet. An adjustable humidistat (30 to 80%) automatically cycles the unit to maintain a preselected relative humidity. The humidistat may also provide a detent setting for continuous operation. Some models also include a sensing and switching device that automatically turns the unit off when the water receptacle is full. Dehumidifiers are designed to provide optimum performance at standard rating conditions of 27°C db room temperature and 60% rh. When the room is less than 18°C db and 60% rh, the evaporator may freeze. This effect is especially noticeable on units with a capillary tube.

Fig. 5

Typical Portable Dehumidifier

25.3 Some dehumidifiers are equipped with defrost controls that cycle the compressor off under frosting conditions. This control is generally a bimetal thermostat attached to the evaporator tubing, allowing dehumidification to continue at a reduced rate when frosting conditions exist. The humidistat can sometimes be adjusted to a higher relative humidity setting, which reduces the number and duration of running cycles and allows satisfactory operation at low-load conditions. Often, especially in the late fall and early spring, supplemental heat must be provided from other sources to maintain above-frosting temperatures in the space. Capacity and Performance Rating. Portable dehumidifiers are available with moisture removal capacities of 5 to 30 L per 24 h, and are operable from ordinary household electrical outlets (115 or 230 V, single-phase, 60 Hz). Input varies from 200 to 800 W, depending on the output capacity rating. AHAM Standard DH-1 establishes a uniform procedure for determining the rated capacity of dehumidifiers under specified test conditions and establishes other recommended performance characteristics. An industry certification program sponsored by AHAM covers the great majority of portable dehumidifiers and certifies dehumidification capacity. The U.S. Environmental Protection Agency (EPA) qualifies dehumidifiers to carry its ENERGY STAR® label if they remove the same amount of moisture as similarly sized standard units, but use at least 10% less energy. The EPA’s ENERGY STAR website provides additional information on qualifying products (EPA 2020). Whole-House Dehumidifiers. Whole-house dehumidifiers have higher moisture removal capacity than portable dehumidifiers. Blowers in whole-house dehumidifiers are typically more powerful because they draw air through the unit at higher external static pressures as compared to portable dehumidifiers. The design allows for connection of ductwork to the unit’s inlet and outlet. Moist air is typically drawn from either the return plenum of the HVAC system ductwork (Figure 6) or from a centrally located register. The air is dehumidified and then discharged into the supply ductwork of the HVAC system for distribution. To prevent reevaporation of moisture from the HVAC system evaporator coil, the dry air enters the system downstream of the coil.

Fig. 6 Typical Whole-House Dehumidifier Installation

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Fig. 7 Typical General-Purpose Dehumidifier Whole-house dehumidifiers typically integrate controls to allow operation of the HVAC system blower while dehumidifying the air; this allows the dry air to be distributed throughout the home. Some dehumidifiers are equipped with controls to bring in outdoor air, which can then be run through the dehumidifier before being distributed to the rest of the home. Sensors are typically located inside the dehumidifier to measure the temperature and relative humidity of air passing through the unit as opposed to the surrounding air. Controls on some models may provide for installation of a remote humidistat or humidity sensor. Drain tubing must typically be installed on the dehumidifier and routed to the nearest floor drain. Most models include an internal condensate overflow switch or provide for the installation of a field-installed overflow switch that will turn the unit off if it overflows. Units are typically sheltered from the elements and should not be installed in locations that can experience freezing temperatures. Whole-house dehumidifiers are typically located near the HVAC system for convenient access to the supply and return plenums and to drains. Systems may be installed in basements, closets, crawlspaces, or attics, so dehumidifier cabinets are typically insulated. When units are installed over a finished area, drain pans are commonly installed under the unit. Codes. Domestic dehumidifiers are designed to meet the safety requirements of UL Standard 474, Canadian Electrical Code, and ASHRAE Standard 15. UL-listed and CSA-approved equipment have a label or data plate indicating approval. UL also publishes the Electrical Appliance and Utilization Equipment Directory, which covers this type of appliance.

tion when introducing outdoor air before the evaporator coil. Some applications (e.g., indoor swimming pools) have a constant moisture load year round. The dehumidifier’s performance is based on the conditions of the air moving across the evaporator coil. Mixing outdoor air with space return air before the evaporator coil changes the dehumidifier’s performance to latent removal, which may be less than the space load and result in the inability to maintain the original design conditions. The unit’s latent capacity can be significantly reduced when outdoor air is introduced before the evaporator coil. Computerized controls can sense return air temperature and relative humidity. Remote wall-mounted sensors are also available. More sophisticated controls are desirable to regulate dew-point temperature and maintain the desired relative humidity in the space. General Considerations. Before considering installation of any type of dehumidification equipment, all latent loads should be identified and quantified. In many cases, this might lead to decisions that reduce the latent load. For example, a storage facility that does not have an adequate vapor retarder in the building envelope should be retrofitted before attempting to calculate the amount of moisture migration through the structure. The same approach should be taken to reduce the amount of uncontrolled air infiltration. Consider covering large water surfaces, such as vats, and/or providing a local exhaust hood to evacuate concentrated water vapor from where it is generated. Although these corrections seem to add cost to a project, the resulting reduced size of the dehumidifier and its lower operating cost often result in an attractive financial payback. Other special considerations include the following: • High volumes of outdoor air. A project may start out as suitable for a general-purpose dehumidifier. However, once outdoor air requirements are quantified to compensate for exhaust and to pressurize the facility, a general-purpose dehumidifier may no longer be applicable. The maximum acceptable portion of outdoor air for general-purpose dehumidifiers is limited, and depends on climatic conditions and the desired indoor conditions to be maintained. As a general rule, when outdoor air requirements exceed 20% of the dehumidifier’s total airflow, the manufacturer should be consulted to determine whether the equipment is suitable for the application. In many cases, a direct-exchange (DX) dedicated outdoor air system unit should be considered instead. • Low-return-air-temperature applications. When the return air temperature is below 18°C, consult the dehumidifier manufacturer to determine whether the equipment is suitable for the application. Recognize that defrost control might be required.

DX Dedicated Outdoor Air System (DOAS) Units

General-Purpose Dehumidifiers Basic components of general-purpose dehumidifiers are shown in Figure 7. An air filter is required to protect the evaporator. Dehumidifying coils, because of their depth and thoroughly wetted surfaces, are excellent dust collectors and not as easily cleanable as much thinner air-conditioning evaporator coils. However, the large amount of condensate has a self-cleaning effect. A bypass damper at the evaporator coil allows airflow adjustments for the evaporator without decreasing airflow for the reheat coil. Dehumidifying and reheat coils may operate at different airflows. The compressor may be isolated from the airstream or located in it. Locating the compressor in the airstream may make service more difficult, but this arrangement allows heat lost through the compressor casing to be provided to the conditioned space while reducing the size of the enclosure. During the cooling season, this compressor location reduces the unit’s sensible cooling capacity. Code-required outdoor air may be introduced between the evaporator and reheat coil. The amount of outdoor air should be controlled to not adversely affect the refrigeration system’s operation. Preheating outdoor air may be required in colder climates. Use cau-

A DX dedicated outdoor air system (DX-DOAS) unit is used to separately condition outdoor air brought into the building for ventilation or to replace air that is being exhausted. (As such, select a DX-DOAS unit based on its latent dehumidification capacity, not necessarily on its total air-conditioning capacity.) This conditioned outdoor air is then delivered either directly to each occupied space, to small HVAC units located in or near the space, or to a central air handler serving the spaces. Meanwhile, the local or central HVAC equipment is used to maintain space temperature. Treating outdoor air separately from recirculated return air makes it more straightforward to verify sufficient ventilation airflow and enables humidity control in the occupied spaces. Decoupling the latent load from the sensible using a DX-DOAS can make both the dehumidifier and sensible cooling equipment more efficient. DX-DOAS units use a vapor compression refrigeration cycle as part of the system to cool the air below its dew point and condense the moisture on a dehumidifying coil. AHRI Standard 921 establishes requirements for the testing and rating of moisture removal capacity and moisture removal effi-

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Mechanical Dehumidifiers and Related Components ciency of DX-DOAS units. ASHRAE Standard 198 prescribes test methods for rating DX-DOAS units. DX-DOAS units may require simultaneous heat rejection to the reheat coil and/or another condenser (air- or water-cooled), because it may not always be possible or warranted in the application to reject the total heat of rejection from the dehumidifying coil to the conditioned airstream. A rainproof air intake and cooling capacity modulation (or staging) are important. With constantly changing weather conditions, even throughout the day, compressor capacity must be adjusted to prevent coil freeze-ups. Basic components are shown in Figure 8. Auxiliary heating may be required for year-round operation in some climates. Water- and steam-heating coils should have freeze protection features. When using indirect-fired gas heaters, the combustion chamber should be resistant to condensation. DX-DOAS units may be interfaced with a building automation system (BAS) to control the unit’s on/off status and operating mode, because most spaces do not require continuous ventilation or replacement air. Air exhaust systems must also be synchronized with the DX-DOAS unit to maintain proper building pressurization. The inclusion of exhaust air energy recovery within DX-DOAS units provides the opportunity to transfer energy between the two airstreams. A typical arrangement is shown in Figure 9. For more information on DOAS, see Chapter 51.

Indoor Swimming Pool Dehumidifiers Indoor pools (natatoriums) are an efficient application for mechanical dehumidifiers. Humidity control is required 24 h a day,

Fig. 8 DX-DOAS Unit

Fig. 9 DX-DOAS Unit with Exhaust Air Heat/Energy Recovery

25.5 year-round. Dehumidifiers are available as single- and doubleblower units (see Figures 10 to 13). AHRI Standard 911 establishes requirements for testing and rating moisture removal capacity and moisture removal efficiency of indoor pool dehumidifiers. ASHRAE Standard 190 prescribes test methods for rating these units. The latent heat (LH) from dehumidification (see Figure 2) comes nearly exclusively from pool water (excluding humidity from makeup air and latent heat from large spectator areas). Loss of evaporation heat cools the pool water. By returning evaporation heat losses to the pool water, the sensible heat between points C and D of Figure 2 is not rejected into the supply air, which can reduce supply air temperature by approximately 8 K. ASHRAE Standard 90.1 requires that heated pools be provided with a vapor-retardant pool cover unless 60% of their energy for heating is site recovered. Use of compressor waste heat for pool water heating may satisfy this requirement. Energy Considerations. The pool water temperature, space temperature, and relative humidity conditions maintained at a facility directly affect occupant comfort, dehumidifier size, and operating costs. It is important that the design engineer understand the effects of changes to these conditions to properly establish realistic operating conditions for a given project. Operating temperatures can change dramatically, depending on the type of pool being designed. See Table 2 in Chapter 6 of the 2019 ASHRAE Handbook—HVAC Applications for additional information. The peak dehumidification load and peak energy conditions may not be concurrent in a natatorium. The peak dehumidification load generally occurs on a summer design day where the latent load from ventilation air is at its peak. The peak energy condition often is in the winter, when cold, dry ventilation air is a significant heating load and dehumidification credit to the space. In addition to extra heating costs, introducing more outdoor air than required by code in winter increases operating costs by driving space relative humidity levels down. This increases pool water evaporation and, consequently, pool water heating, makeup water, and pool chemical costs. Low humidity levels also increase bather discomfort through the evaporation effect across the surface of bathers’ skin, causing a chilling effect on leaving the pool. Fan energy can be a significant portion of the total energy consumed by a natatorium dehumidifier because the supply air fan operates constantly and delivers relatively high amounts of air. Many dehumidifiers have exhaust fans to maintain the natatorium at a negative pressure. This exhaust air is energy rich, and heat recovery should be considered. Dehumidifiers are available with heat recovery such as heat pipes, glycol-runaround loops, compressorized heat pumps, and plate heat exchangers that transfer heat from the energy-rich exhaust air to cold incoming ventilation air (see Chapter 26). Condensation. Note that the space dew point in most pools is above 17°C. Even when space conditions are properly maintained, there is still a significantly higher chance, compared to traditional conditioned spaces, that condensation might occur in the space or within the HVAC system. Condensation is possible in winter or summer. Most codes have adopted ASHRAE Standard 62.1 to determine the amount of outdoor air required for acceptable indoor air quality. Introducing outdoor air that is cooler than the room dew-point temperature could lead to condensation. This is especially important in areas that experience cold winters. If condensation might occur, preheating the outdoor air is required. Most dehumidifiers have a cooling mode. The units are generally designed to ensure supply air to the space is warmer than the space dew point. When adding additional cooling capacity to a dehumidifier, summer condensation is a concern if supply air could be cooled below the dew point of either the outdoor air or the space.

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Designers are encouraged to verify that all necessary measures to avoid condensation have been taken, especially when modifying or customizing standard products from manufacturers. Discharge air must always be supplied at a temperature higher than the room dewpoint temperature. Similarly, blended airstreams must not be allowed to result in temperatures cooler than the space dew point. Single-blower pool dehumidifiers (Figure 10) are similar to general-purpose dehumidifiers (see Figure 4), with the following exceptions:

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• Recovered heat from the refrigeration circuit can be used to provide heating to one or more pools or for domestic water preheat. • All components in the airstream must be corrosion-resistant to chloramine-laden air. Electrical panels must be fully isolated from the chloramine-laden airstream. • The pool water heater must be resistant to chlorinated water. • Cross-contamination prevention features in the pool water heater are not required, but should be considered to prevent introducing refrigerant oil to the pool water if a breach occurred. • Polyol ester (POE) refrigerant oil will damage polyvinyl chloride (PVC) piping. Figure 11 shows a double-blower pool dehumidifier with economizer dampers and a full-sized return fan located upstream of the evaporator coil. This configuration can provide up to 100% makeup air to maintain humidity levels, which can be attractive when the outdoor dew point is below the required indoor dew point during mild weather, or in climates when enough hours of dry- and wetbulb conditions are below the level to be maintained. Preheating outdoor air may be required to prevent condensation inside the mix-

Fig. 10 Typical Single-Blower Pool Dehumidifier

Fig. 11

ing box. Also, this configuration does not recover energy from the warm, moist exhaust air. During dehumidifying coil operation, the amount of makeup and exhaust air is limited by outdoor conditions, especially during the heating season. Cold makeup air may lower the mixed-air temperature to the point where the dehumidifying coil cannot extract any moisture. Be careful during economizer operation to not introduce more outdoor air than required by code, so as not to lower the space humidity below design levels, as discussed in the section on Energy Considerations. In some regions, it is economically attractive to remove moisture from the exhaust air to recover its latent heat. In this case, the dehumidifying coil is installed in the return air section. Figure 12 shows a double-blower pool dehumidifier with economizer dampers and return fan located downstream of the evaporator coil. A damper system can also be incorporated to exhaust before the evaporator coil during colder conditions and after the evaporator during warmer conditions. This configuration recovers energy from the warm, moist exhaust air; however, exhausting air from downstream of the evaporator coil also reduces the unit’s sensible capacity by the amount of the exhaust air. The ratio of return air to exhaust air must be considered to determine the unit’s capacity to remove moisture from the conditioned space. Figure 13 shows a different unit configuration that addresses concerns related to blower energy use during the various operating modes. This unit can operate with the supply blower only, or with the addition of one or two exhaust blowers. Most manufacturers also offer some means of air-to-air heat recovery between the exhaust and makeup airstreams (see Chapter 26). During cold weather, this arrangement preheats entering makeup air with heat recovered from the exhaust airstream. Latent heat recovery may not be practical, however, because it transfers moisture to the entering air, thus possibly increasing dehumidification requirements. Control systems should be compatible with building automation systems; however, the BAS must not disable dehumidifier operation because indoor pools always need some dehumidification, regardless of occupancy, and require specialized control sequences. General Considerations. The primary function of an indoor swimming pool dehumidifier is to maintain the space dew-point temperature at the design level year round and to provide adequate air circulation to comply with minimum air change rates. For more information on indoor swimming pool (natatorium) applications, see Chapter 6 of the 2019 ASHRAE Handbook—HVAC Applications. Types of Equipment. Indoor swimming pool dehumidifiers are available in single- and double-blower configurations (see Figures 10 to 13). Heat from the refrigeration circuit can be (1) used to

Typical Double-Blower Pool Dehumidifier with DX Coil in Supply Air Section

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reheat supply air, (2) used to heat pool water, or (3) rejected to the outdoors by an optional air- or water-cooled condenser. Equipment configurations are available to use the heat for any combination of these three purposes. The equipment can be located indoors or outdoors, and may be manufactured as a single package or as a split system with a remote condenser. Avoiding remote condensers with long refrigerant lines reduces refrigerant charge. Indoor, air-cooled condensers are typically equipped with a blower-type fan suitable for duct connections. An optional economizer allows for introduction of up to 100% of outdoor air (turning off the compressors) when conditions are appropriate. When selecting a dehumidifier for an indoor swimming pool application, several questions need to be addressed to ensure that the dehumidifier can maintain the space conditions: • In what mode of operation is the dehumidifier rated? • Does the rating include ventilation air, and what effect, if any, does it have on dehumidifier performance? • Does the unit include exhaust air, and what effect, if any, does it have on dehumidifier performance? • Does the cost of running a second fan offset the energy saved by the economizer? • Will the dehumidifier maintain the desired space conditions during all modes of operation?

25.7 Ice Rink Dehumidifiers Design for ice rink dehumidifiers is similar to that of generalpurpose dehumidifiers (see Figure 4). However, because of the lower temperatures, airflow and dehumidifying and reheat coils are selected in accordance with the following conditions: • The dehumidifying coil may or may not have an air bypass, depending on the location of makeup air intake and/or coil selection. • The dehumidifying coil may have means to defrost or to prevent frost formation. • Makeup air treatment is limited. For large spectator areas, special makeup air dehumidifiers may be required. General Considerations. For community ice rinks with small spectator areas (or none), it is customary to install two small dehumidifiers over the dasher boards in a diagonal arrangement, 3.7 to 4.6 m above the ice surface (Figure 14). Take care that discharge air from dehumidifiers is not directed toward the ice surface. Forced airflow at any temperature may damage the ice surface. Ice rinks with large spectator areas have different requirements. The spectator area is typically maintained at 21°C. To limit moisture migration to the ice sheet, space conditions must be maintained at 50% rh or less. The resulting dew-point temperature is then 10°C or less. The air temperature over the ice sheet in the dasher boards, however, is approximately 3 K lower than the air in the spectator area. With an air temperature of 18°C and a dew-point temperature

Fig. 12 Typical Double-Blower Pool Dehumidifier with DX Coil in Return Air Section

Fig. 13

Supply Blower and Double Exhaust Blower Pool Dehumidifier

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of 10°C or less, fog over the ice sheet cannot develop. As a general rule, mechanical ice rink dehumidifiers are most effective for condensation and fog control when dry-bulb space temperature is at least 8 K above the dew point. For additional fog and condensation prevention methods, see Chapter 44 of the 2018 ASHRAE Handbook—Refrigeration.

Industrial Dehumidifiers

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Industrial products and surfaces can sweat, causing corrosion, dimensional distortions, and other deterioration from excess humidity in the air. In some cases, this humidity is caused by leakage or infiltration; in other cases, it may be from indoor storage or processes. Normal air conditioning is not always adequate to control excess humidity, leading to material or structural damages and cool but damp working conditions. Industrial dehumidifiers use directexpansion refrigeration coils to remove moisture and thus lower the supply air dew point to • Prevent undesirable condensing or sweating on products and surfaces • Improve product/process quality • Help reduce building repair and maintenance costs • Provide a comfortably dry working environment (max. 60% rh) • Minimize biological pollutants such as mold spores Some mechanical dehumidifiers may also contribute to space heating and/or cooling, recover heat energy for processes, and/or recover water for allowable purposes. Sources of Humidity. Indoor air quality is affected by several key factors, which vary in importance depending on the location of the building and on the activity for which a building is designed. Relative humidity is usually a critical air quality factor, with high indoor relative humidity resulting from the following sources of moisture: • Increased quantities of humid outdoor air brought in to improve air quality • Openings, infiltration and permeation • Moisture produced by occupants • Moisture released by products or processes Moisture migrates from areas of high vapor pressure to areas of low vapor pressure. In the summer, when the outdoor air is warm and humid, moisture can find a path to the interior of a cooler or drier structure. This could be from openings in the building such as doors, infiltration through cracks and poorly sealed joints, or permeation in the case of low-quality or nonexistent vapor retarders. In many instances, the primary source of humidity is from outdoor air purposely brought into the structure to meet air quality standards, or to replace air being exhausted because of contamination. Occupants contribute to the moisture load through respiration and perspiration; amounts depend on the number of people and their activities. A worker involved in heavy lifting can generate seven

times the moisture of a co-worker seated at rest. In agricultural or laboratory structures, animals also produce a moisture load. Indoor materials or processes may give off moisture. Storing or cooking fruits, vegetables, or other foodstuffs may release moisture indoors. The presence of open tanks of water or the storing or handling of wet materials, such as wood, may also release moisture indoors. Moisture dissolved in indoor air will condense onto any surface that is at a temperature lower than the room air’s dew-point temperature. This can lead to quality and productivity problems or even to damage to the building and plant equipment. Rust and corrosion can affect metal surfaces, electrical controls and contacts, etc., which can lead to increased costs and even to potentially hazardous conditions. For typical conditions (temperatures and relative humidities) of industrial applications, see Table 1 of Chapter 15 of the 2019 ASHRAE Handbook—HVAC Applications.

Dehumidifiers for Controlled Environment Agriculture Dehumidifiers for controlled environment agriculture (CEA; also called indoor grow rooms) share many of the characteristics of general-purpose dehumidifiers (see Figures 4 and 7), but may have a few additional design considerations because of the specialized nature of their application: • Dehumidifiers for growing rooms in CEA are typically much larger than general-purpose or portable dehumidifiers, with 240 to 2300 L/day (10 to 90 kg/h) capacities • Use an evaporator bypass damper, variable airflow rate, and/or defrost cycle to ensure operation during cooler lights-off or endof-grow-cycle conditions, which can be as low as 18°C at 40% rh (4°C dp) • Use minimum efficiency reporting value (MERV) 11 to 13 filters to provide adequate mold spore reducing filtration (see ASHRAE Standard 52.2 for details on MERV filters) General Considerations. Because most grow rooms are sensible cooling primary load facilities, care is needed to ensure that heat gain from the dehumidification process can be managed by the existing or designed sensible cooling devices. A CEA dehumidifier can be selected with remote heat rejection to provide some or all of the sensible cooling required. This approach may achieve better energy management by ensuring that the facility is not forced to oversize or overuse sensible cooling devices to counter heat gain from dehumidification equipment in the space. Grow rooms can require washdown cycles. Ensure that selected equipment has washdown capabilities to reduce pathogen and mold growth, and that filters are easily accessible for regular replacement.

Tunnel Dryer Dehumidifier A general-purpose dehumidifier system removes moisture from a product inside the drying tunnel. Air at the desired temperature and humidity constantly recirculates around the product, continuously removing moisture. Tunnel dryer dehumidifier design is similar to that of a general-purpose dehumidifier (see Figure 4). Because of the closely controlled space humidity, temperature, and airflow, select the dehumidifying and reheat coils in accordance with the following conditions: • The dehumidifying coil must have variable apparatus dew-point control; it might also have an air bypass, depending on coil selection. • The reheat coil needs variable reheat control. • A means of external heat rejection is required (e.g., heat rejection to a heat pump loop, chilled-water loop, fluid cooler, or remote air-cooled condenser).

Fig. 14 Typical Installation of Ice Rink Dehumidifiers

Process Considerations. A precision control system is required to prevent overdrying by automatically shutting off or by reducing

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Mechanical Dehumidifiers and Related Components capacity when the moisture content in the product has reached the desired level. Continuous control of temperature and relative humidity improves the drying process and maintains product quality, and potentially can reduce operating cost compared to ventilation only. Users should be able to specify drying temperature to best preserve initial product quality. Design Considerations. The usual operating range for tunnel dryer dehumidifiers is 15 to 38°C db and 10 to 38°C dew point. An average air velocity of 2.54 m/s across the tunnel net free area is recommended for effective drying. Tunnels should be no longer than 9 m, to prevent saturated air at the tunnel’s end. Construction must be air- and vaportight to prevent moisture infiltration from surrounding spaces. Some products may produce corrosive contaminants, so tunnel dryer dehumidifier components should be corrosion resistant. For a typical installation of a tunnel dryer dehumidifier, see Figure 15.

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2.

CONTROLS AND SENSORS

Care is needed when locating the humidity-sensing device that triggers compressor operation. A poorly located sensor results in failure of the system to properly control the humidity levels. The sensor must be located in an area that gives an accurate reading and feedback of the conditions needing control. In larger-volume areas, multiple sensors with an averaging algorithm might provide the best result. Consider the following when selecting sensor locations: • If the supply fan cycles on/off with compressor operation, this scenario is likely to produce the least accurate sensor reading. Unless the sensor can be located in an area where the humidity levels are representative of the space when there is no air circulation, the supply fan should be set on a timer to operate on a regular cycle. This ensures air turnover and better monitoring and control of the conditioned space. • Natatorium dehumidifiers usually have sensors mounted on the unit because of the space high air turnover rate and continuous fan operation. Ensure that the air distribution system does not introduce dry supply air into the return; otherwise, the control system might incorrectly sense that the space is satisfied.

Fig. 15

Typical Tunnel Dryer Dehumidifier

25.9 • DOAS units typically have all sensors mounted on the unit and are configured to deliver neutral air. Some units have space-mounted temperature sensors to trigger a full cooling mode (no reheat) to contribute to the space’s sensible cooling requirements. This operational mode requires a space-mounted thermostat. • DOAS units are often configured with 100% recirculation or a percentage of recirculation mode operation, and are tasked for space cooling, heating, and humidity control. Proper location of these space-mounted temperature and humidity sensors is vital for the equipment to provide proper space control. Heat recovery devices require additional sensors for proper operation.

3.

INSTALLATION AND SERVICE CONSIDERATIONS

Equipment and sensors must be installed properly so that they function in accordance with manufacturers’ specifications. Interconnecting diagrams for the low-voltage control system should be documented for proper future servicing. Planning is important for installing large, roof-mounted equipment because special rigging is frequently required. The refrigerant circuit must be clean, dry, and leak-free. An advantage of packaged equipment is that proper installation minimizes the risk of field contamination of the circuit. Take care to properly install split-system interconnecting tubing (e.g., proper cleanliness, brazing, evacuation to remove moisture). Charge split systems according to the manufacturer’s instructions. Equipment must be located to avoid noise and vibration problems. Mount single-package equipment of over 70 kW capacity on concrete pads if vibration control is a concern. Large-capacity equipment should be roof mounted only after the roof’s structural adequacy has been evaluated. Additional installation guidelines include the following: • In general, install products containing compressors on solid, level surfaces. • Avoid mounting products containing compressors (e.g., remote units) on or touching the foundation of a building. A separate pad that does not touch the foundation is recommended to reduce noise and vibration transmission through the slab. • Do not box in outdoor air-cooled units with fences, walls, overhangs, or bushes. Doing so reduces the unit’s air-moving ability, reducing efficiency. • For a split-system remote unit, choose an installation site that is close to the indoor part of the system to minimize refrigerant charge and pressure drop in the connecting refrigerant tubing. • Contact the equipment manufacturer or consult the installation instructions for further information on installation procedures. Equipment should be listed or certified by nationally recognized testing laboratories to ensure safe operation and compliance with government and utility regulations. Equipment should also be installed to comply with agency standards’ rating and application requirements to ensure that it performs according to industry criteria. Larger and more specialized equipment often does not carry agency labeling. However, power and control wiring practices should comply with the National Electrical Code® (NFPA Standard 70). Consult local codes before design, and consult local inspectors before installation. A clear, accurate wiring diagram and well-written service manual are essential to the installer and service personnel. Easy, safe service access must be provided for cleaning, lubrication, and periodic maintenance of filters and belts. In addition, access for replacement of major components must be provided and preserved. Service personnel must be qualified to repair or replace mechanical and electrical components and to recover and properly recycle

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Fig. 16 Schematic of Dehumidification Enhancement with Wraparound Heat Pipe or dispose of any refrigerant removed from a system. They must also understand the importance of controlling moisture and other contaminants in the refrigerant circuit; they should know how to clean a hermetic system if it has been opened for service (see Chapter 7 of the 2018 ASHRAE Handbook—Refrigeration). Proper service procedures help ensure that the equipment continues operating efficiently for its expected life.

4.

Fig. 17

Enhanced Dehumidification Process with Wraparound Heat Pipe

WRAPAROUND HEAT EXCHANGERS

An air-to-air heat exchanger (heat pipe, coil runaround loop, fixedplate heat exchanger, or rotary heat exchanger) in a series (or wraparound) configuration can be used to enhance moisture removal by a mechanical dehumidifier, improving efficiency, and possibly allowing reduced refrigeration capacity in new systems. Other uses of airto-air heat exchangers are covered in Chapter 26. Air-to-air heat exchangers are used with a mechanical dehumidification system to passively move heat from one place to another. The most common configuration used for dehumidification is the runaround (or wraparound) configuration (Figure 16), which removes sensible heat from the entering airstream and transfers it to the leaving airstream. (Points A to E correspond to points labeled in Figure 17.) This improves the cooling coil’s latent dehumidification capacity. This method can be applied if design calculations have accounted for the condition of air entering the evaporator coil. In the runaround or wraparound configuration (Figure 16), one section of the air-to-air heat exchanger is placed upstream of the cooling coil and the other section is placed downstream of the cooling coil. The air is precooled before entering the cooling coil. Heat absorbed by the upstream section of the air-to-air heat exchanger is then transferred to air leaving the cooling coil (or supply airstream) by the downstream section. Sensible precooling by the air-to-air heat exchanger reduces the sensible load on the cooling coil, allowing an increase in its latent capacity (Figure 17). The combination of these two effects lowers the system SHR, much like the process described in the Mechanical Dehumidifiers section. Adding the air-to-air heat exchanger brings the condition of air entering the evaporator coil closer to the saturation line on the psychrometric chart (A to B). In new installations, this requires careful evaporator coil design that accounts for the actual range of air conditions after the air-to-air heat exchanger, which may differ significantly from the return air conditions. In retrofits, the duct-to-duct (or slide-in) configuration (Figure 18) is sometimes used. One section of the air-to-air heat exchanger is placed in the return airstream, and the other section in the supply airstream. This configuration, however, does not provide as much

Fig. 18 Slide-in Heat Pipe for Rooftop Air Conditioner Refit (Kittler 1996)

benefit as the wraparound configuration because (1) the upstream side of the heat exchanger is located upstream of where outdoor air enters the system, (2) the higher velocity reduces the effectiveness and increases the air-side pressure drop of the heat exchanger, and (3) it requires an additional filter upstream of the air-to-air heat exchanger. In retrofits, the lower entering-air temperature at the evaporator coil lowers the temperature of air leaving the evaporator coil. Evaporator coil capacity is reduced because of the lower entering wetbulb temperature, changing the system’s operating point. This must be analyzed to ensure that the mechanical refrigeration system still operates correctly. If evaporator coil freeze-up is possible, the system must include some means of deactivating the air-to-air heat exchanger or increasing airflow to prevent evaporator freezing. Some way to modulate the air-to-air heat exchanger’s capacity may be incorporated to better meet the load requirement of the mechanical dehumidifier. Adding an air-to-air heat exchanger typically improves the moisture removal capacity of an existing mechanical dehumidification

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system by allowing a lower supply air dew point, while providing some reheat without additional energy use. Proper design practices are necessary to ensure that the unit’s mechanical refrigeration system still operates efficiently with the new entering air conditions and additional air-side pressure drop. Also, the added pressure drop of the air-to-air heat exchanger is likely to reduce the airflow delivered, unless fan speed in increased. If increasing fan speed is necessary, verify that the fan motor can handle the added load. Figure 17 shows the dehumidification process when an air-to-air heat exchanger is added to an existing evaporator coil. Point A to C shows the cooling and dehumidification process of an existing DX evaporator coil, without the air-to-air heat exchanger. Point A to B shows precooling by the upstream section of heat exchanger. The process line from B to D (versus B to C, without the heat exchanger) shows how the evaporator coil’s dehumidification performance improves (lowering leaving air dew point, from C to D) if an air-to-air heat exchanger is added to the existing system, because the enthalpy of the air entering the evaporator is lowered. Point D to E illustrates that the heat removed from air upstream of the evaporator (A to B) is added back into air leaving the evaporator. The total amount of heat energy (enthalpy) removed in section A-B is equal to the amount of heat added in section D-E.

REFERENCES ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal.ashrae .org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org/bookstore. AHAM. 2008. Dehumidifiers. Standard DH-1-2008. Association of Home Appliance Manufacturers, Chicago, IL. AHRI. 2011. Performance rating of indoor pool dehumidifiers. Standard 911-2011. Air Conditioning, Heating, and Refrigeration Institute, Arlington, VA. AHRI. 2012. Performance rating of DX-dedicated outdoor air system. Standard 921-2011. Air Conditioning, Heating, and Refrigeration Institute, Arlington, VA. ASHRAE. 2013. Safety standard for refrigeration systems. ANSI/ASHRAE Standard 15-2013. ASHRAE. 2017. Method of testing general air-cleaning devices for removal efficiency by particle size. ANSI/ASHRAE Standard 52.2-2017.

25.11 ASHRAE. 2013. Ventilation for acceptable indoor air quality. ANSI/ ASHRAE Standard 62.1-2013. ASHRAE. 2013. Energy standard for buildings except low-rise residential buildings. ANSI/ASHRAE/IES Standard 90.1-2013. ASHRAE. 2013. Method of testing for rating indoor pool dehumidifiers. ANSI/ASHRAE Standard 190-2013. CSA. 2015. Canadian electrical code, part I (23rd edition): Safety standard for electrical installations. Standard C22.1-15. Canadian Standards Association, Toronto. EPA. 2020. ENERGY STAR®. www.energystar.gov. Kittler, R. 1996. Mechanical dehumidification control strategies and psychrometrics. ASHRAE Transactions 102(2):613-617. Paper SA-96-10-2. NFPA. 2014. National electrical code®. Standard 70-2014. National Fire Protection Association, Quincy, MA. UL. 2009. Dehumidifiers. ANSI/UL Standard 474-09. Underwriters Laboratories, Northbrook, IL. UL. 2011. Electrical appliances and utilization equipment directory. Underwriters Laboratories, Northbrook, IL.

BIBLIOGRAPHY ACCA. 2017. HVAC design for swimming pools and spas. ANSI/ACCA 10 Manual SPS. Air Conditioning Contractors of America, Arlington, VA. AHAM. Semiannually. Directory of certified dehumidifiers. Association of Home Appliance Manufacturers, Chicago, IL. Harriman, L., G. Brundrett, and R. Kittler. 2001. Humidity control design guide for commercial and institutional buildings. ASHRAE. IEC. 2018. Household and similar electric appliances—Safety—Part 2-40: Particular requirements for electrical heat pumps, air-conditioners and dehumidifiers. IEC Standard 60335-2-40:2018. International Electrotechnical Commission, Geneva. Kittler, R. 1989. Indoor natatorium design and energy recycling. ASHRAE Transactions 95(1):521-526. Paper CH-89-02-3. Kittler, R. 1994. Separate makeup air makes IAQ affordable. Mechanical Buyer & Specifier (June). Morris, W. 2003. The ABCs of DOAS. ASHRAE Journal (May). Morner, S., A. Hicks, and M. McDevitt. 2017. ASHRAE design guide for dedicated outdoor air systems. ASHRAE Research Project RP-1712. Murphy, J. 2006. Smart dedicated outdoor-air systems. ASHRAE Journal (July). Nevins, R., R.R. Gonzalez, Y. Nishi, and A.P. Gagge. 1975. Effect of changes in ambient temperature and level of humidity on comfort and thermal sensations. ASHRAE Transactions 81(2). Paper BO-2370 (RP 144). UL. 2017. Leakage current for appliances. ANSI/UL Standard 101-17. Underwriters Laboratories, Northbrook, IL.

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AIR-TO-AIR ENERGY RECOVERY EQUIPMENT Applications ............................................................................. 26.1 Basic Heat or Heat and Water Vapor Transfer Relations....... 26.2 Types of Air-to-Air Heat Exchangers ...................................... 26.5 Performance Ratings ............................................................. 26.19 Additional Technical Considerations .................................... 26.20 Comparison of Air-to-Air Heat or Heat and Mass Exchanger Characteristics ................................................................... 26.26

Use of Air-to-Air Heat or Heat and Mass Exchangers in Systems ........................................................................... 26.27

A

Table 1 Typical Applications for Air-to-Air Energy Recovery

IR-TO-AIR energy recovery is the process of recovering heat and/or moisture between two airstreams at different temperatures and humidities. This process is important in maintaining acceptable indoor air quality (IAQ) while keeping energy costs low and reducing overall energy consumption and carbon dioxide emission. This chapter describes various technologies for air-to-air energy recovery. Thermal and economic performance, maintenance, and related operational issues are presented, with emphasis on energy recovery for ventilation. Air-to-air energy recovery should be considered for every building in which energy is used to condition the air. This is consistent with ASHRAE’s strategic plan to support a sustainable built environment, and aligns with the United Nations International Panel for Climate Change’s call for action to reduce the emissions of carbon dioxide related to energy use. Energy can be recovered either in its sensible (temperature only) or latent (moisture) form, or a combination of both from multiple sources. Sensible energy can be extracted, for example, from outgoing airstreams in dryers, ovens, furnaces, combustion chambers, and gas turbine exhaust gases to heat supply air. Units used for this purpose are called sensible heat exchange devices or heat recovery ventilators (HRVs). Devices that transfer both heat and moisture are known as energy or enthalpy devices or energy recovery ventilators (ERVs). HRVs and ERVs are available for commercial and industrial applications as well as for residential and small-scale commercial uses. Air conditioners use significant energy to dehumidify moist airstreams. Excessive moisture in the air of a building can result in mold, allergies, and bacterial growth. ERVs can enhance dehumidification with packaged unitary air conditioners. Introducing outdoor or ventilation air is the primary means of diluting air contaminants to achieve acceptable indoor air quality. ERVs can cost-effectively provide large amounts of outdoor air to meet a building’s minimum ventilation requirements as prescribed in ASHRAE Standards 62.1 and 62.2. Types of ERVs include fixed-plate heat exchangers, rotary wheels, heat pipes, runaround loops, thermosiphons, and twin-tower enthalpy recovery loops. Performance is typically characterized by effectiveness; pressure drop, pumping, or fan power of fluids; cross flow (i.e., amount of air leakage from one stream to the other); and frost control (used to prevent frosting on the heat exchanger). Recovery efficiency, the ratio of output of a device to its input, is also often considered. In energy recovery ventilators, effectiveness refers to the ratio of actual energy or moisture recovered to the maximum possible amount of energy and/or moisture that can be recovered. Fluid stream pressure drops because of the friction between the fluid and solid surface, and because of the geometrical complexity of the flow passages. Pumping or fan power is the product of the fluid The preparation of this chapter is assigned to TC 5.5, Air-to-Air Energy Recovery.

Copyright © 2020, ASHRAE

Economic Considerations ...................................................... 26.35 Energy and/or Mass Recovery Calculation Procedure ........................................................................... 26.36 Symbols .................................................................................. 26.41

Method

Application

Comfort-to-comfort

Residences Offices Classrooms Retail Bars and restaurants Swimming pools Locker rooms Operating rooms Nursing homes Animal ventilation Plant ventilation Smoking exhaust Dryers Ovens Flue stacks Burners Furnaces Incinerators Paint exhaust Welding exhaust

Process-to-process and Process-to-comfort

volume flow rate and pressure drop. Economic factors such as cost of energy recovered and capital and maintenance cost (including pumping power cost) play a vital role in determining the economic feasibility of recovery ventilators for a given application.

1.

APPLICATIONS

Air-to-air energy recovery systems may be categorized according to their application as (1) process-to-process, (2) process-to-comfort, or (3) comfort-to-comfort. Some typical air-to-air energy recovery applications are listed in Table 1. In comfort-to-comfort applications, the energy recovery device lowers the enthalpy of the building supply air during warm weather and raises it during cold weather by transferring energy between the ventilation air supply and exhaust airstreams. Air-to-air energy recovery devices for comfort-to-comfort applications may be sensible heat exchange devices (i.e., transferring sensible energy only) or energy exchange devices (i.e., transferring both sensible energy and moisture). These devices are discussed further in the section on Additional Technical Considerations. When outdoor air humidity is low and the building space has an appreciable latent load, an ERV can recover sensible energy while possibly slightly increasing the latent space load because of water vapor transfer within the ERV. It is therefore important to determine whether the given application calls for HRV or ERV. HRVs are suitable when outdoor air humidity is low and latent space loads are high for most of the year, and also for use with swimming pools, chemical exhaust, paint booths, and indirect evaporative coolers.

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2020 ASHRAE Handbook—HVAC Systems and Equipment(SI)

ERVs are suitable for applications in schools, offices, residences and other applications that require year-round economical preheating and/or precooling of outdoor supply air. In process-to-process applications, heat is captured from the process exhaust stream and transferred to the process supply airstream. Equipment is available to handle process exhaust temperatures as high as 870°C. Process-to-process recovery devices generally recover only sensible heat and do not transfer latent heat, because moisture transfer is usually detrimental to the process. In cases involving condensable gases, less recovery may be desired to prevent condensation and possible corrosion. In process-to-comfort applications, waste heat captured from process exhaust heats building makeup air during winter. Typical applications include foundries, strip-coating plants, can plants, plating operations, pulp and paper plants, and other processing areas with heated process exhaust and large makeup air volume requirements. Although full recovery is usually desired in process-to-process applications, recovery for process-to-comfort applications must be modulated during warm weather to prevent overheating the makeup air. During summer, no recovery is required. Because energy is saved only in the winter and recovery is modulated during moderate weather, process-to-comfort applications save less energy annually than do process-to-process applications. Process-to-comfort recovery devices generally recover sensible heat only and do not transfer moisture between airstreams.

2.

Effectiveness When mass flow rates, temperatures, and humidities of the inlets and outlets are known, the sensible, latent, or total effectiveness of the exchanger for those operating conditions can be determined. Conversely, if the respective effectiveness of the exchanger at specific mass flow rates is known or rated, and the temperatures and humidities of the inlets are specified, it is possible to estimate conditions at the outlets, as long as confounding variables such as occurrence of condensation or internal leakages are insignificant. The general definition of transfer effectiveness , on which ASHRAE Standard 84 bases its definitions of effectiveness, is Actual transfer of moisture and/or energy  = ----------------------------------------------------------------------------------------------------------------Maximum possible transfer between airstreams

Referring to Figure 1, the gross sensible effectiveness s or sensible of an energy recovery ventilator is given as m· 2  c p 1 T 1 – c p 2 T 2  sensible = ------------------------------------------------------· m min  c p 1 T 1 – c p 3 T 3 

Fig. 1 Airstream Numbering Convention

(2a)

Referring to Figure 1, the gross latent effectiveness L or latent of an energy recovery ventilator is given as m· 2  h fg 1 W 1 – h fg 2 W 2  latent = -------------------------------------------------------------m· min  h fg 1 W 1 – h fg 3 W 3 

(2b)

Referring to Figure 1, the gross total effectiveness T or total of an energy recovery ventilator is given as m· 2  h 1 – h 2  total = --------------------------------m· min  h 1 – h 3 

BASIC HEAT OR HEAT AND WATER VAPOR TRANSFER RELATIONS

The second law of thermodynamics states that heat energy always transfers from a region of high temperature to one of low temperature. This law can be extended to say that mass transfer always occurs from a region of high vapor pressure to one of low vapor pressure. The conceptual energy recovery exchanger in Figure 1 facilitates this transfer (without mixing supply and exhaust airstreams) across a separating wall (shown by a thick horizontal line in Figure 1) made of a material that conducts heat and is permeable to water vapor. Heat is transferred when there is a difference in temperature between the two airstreams. Moisture is transferred when there is a difference in vapor pressure between the two airstreams. On a typical summer day, supply air at temperature, humidity, or enthalpy of x1 and mass flow rate ms enters the ERV, while exhaust air from the conditioned space enters at conditions x3 and m3. Because conditions at x3 are lower than conditions at x1, heat and mass transfer from the supply airstream to the exhaust airstream because of differences in temperature and vapor pressures across the separating wall. Consequently, the supply air exit properties decrease, while those of the exhaust air increase.

(1)

(2c)

where m· n m· min cp,n hfg Tn Wn hn

= = = = = = =

mass flow rate at station n, kg/s minimum of m· 2 and m· 3 , kg/s specific heat of dry air at station n, kJ/(kg·K) heat of vaporization of water, kJ/(kg·K) dry-bulb temperature at station n, °C humidity at station n, kgw/kgda enthalpy at station n, kJ/kg

Note that in Equations (2a), (2b), and (2c), the denominators express the theoretical maximum sensible, latent, or total energy transfer. An additional figure of merit, which characterizes the reduction in heating or cooling load associated with the supply air, is the enthalpy recovery ratio, given by h1 – h2 Enthalpy recovery ratio = ----------------h1 – h3

(3)

where h is enthalpy at station n, in kJ/kg. Equation (3) bears a superficial resemblance to Equation (2c) which defines gross total effectiveness, but when the supply and exhaust mass flows through the exchanger are different, the two metrics may have different values. Using the enthalpy recovery ratio when designing HVAC systems for compliance with ASHRAE Standard 90.1 is detailed in the section on Performance Ratings. The leaving supply air conditions and heat transfer rates easily can be estimated under ideal conditions: steady-state operation; no heat or moisture transfer between the heat exchanger and its surroundings; no cross-leakage; and no energy gains or losses from motors, fans, or frost control devices, with negligible condensation or frosting. Under such ideal conditions, the leaving supply temperature is

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Air-to-Air Energy Recovery Equipment m· 2 c p 1 T 1 –  sensible m· min  c p 1 T 1 – c p 3 T 3  T2 = --------------------------------------------------------------------------------------------------------m· 2 c p 2

26.3 Similarly, humidity transfer qw in the ERV can be estimated from (4a)

and the exhaust air temperature is m· 2 c p 3 T 3 +  sensible m· min  c p 1 T 1 – c p 3 T 3  T4 = --------------------------------------------------------------------------------------------------------m· c

(4b)

q w = m· 2  W 1 – W 2  = Q 2  2  W 1 – W 2 

(5c)

where qw is humidity transfer, in kgw/s In some cases, the total energy transfer qT or qtotal can be estimated from

2 p 2

q total = m· 2  h 1 – h 2  = Q 2  2  h 1 – h 2 

The leaving supply air humidity is m· 2 h fg 1 W 1 –  latent m· min  h fg 1 W 1 – h fg 1 W 3  W2 = --------------------------------------------------------------------------------------------------------------m· 2 h fg 2

(4c)

And the exhaust air humidity is (4d)

2 fg 2

The leaving supply air enthalpy is

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where qh is total energy transfer, kilowatts However, in the somewhat unusual case when the supply airstream gains sensible energy but loses water content, qtotal is best estimated from qtotal = qs + qL

m· 2 h fg 3 W 3 +  latent m· min  h fg 1 W 1 – h fg 3 W 3  W4 = --------------------------------------------------------------------------------------------------------------m· h

m· min -  h1 – h3  h2 = h 1 –  total ----------m· s

(4e)

And the exhaust air enthalpy is m· min -  h1 – h3  h4 = h 3 –  total ----------m· s

(4f)

Use Equations (4e) and (4f) with caution. When the enthalpies of the entering airstreams are very close but the temperatures and humidities are not, it is possible for the calculated total effectiveness to be greater than or less than both of the sensible or latent effectivenesses, or even to be greater than 100% or less than 0%. For example, in some conditions, the supply airstream flowing through an ERV may gain heat energy (+qs) from the adjoining stream but lose latent energy (–qL) if it transfers the water vapor to the adjoining stream. Therefore, it is generally best to determine the expected enthalpy of leaving air by first calculating the expected temperature and humidity of the leaving air. The sensible heat energy transfer qs in the energy recovery exchanger can be estimated from q s = m· 2  c p 1 T 1 – c p 2 T 2  = Q 2  2  c p 1 T 1 – c p 2 T 2 

(5a)

where qs = sensible heat transfer, kW cp = specific heat of air, kJ/(kg·K) Q2 = volume flow rate of supply outlet air, m3/s 2 = density of dry outlet supply air, kg/m3 t1, t2, t3 = inlet and exit temperatures of supply inlet, supply outlet, and exhaust inlet airstreams, respectively, K m· 2 = mass flow rate of supply air inlet, kg/s

The latent heat transfer qL in the energy recovery exchanger can be estimated from q L = m· 2  h fg 1 W 1 – h fg 2 W 2  = Q 2  2  h fg 1 W 1 – h fg 2 W 2  (5b) where qL = latent heat transfer, kW hfg = enthalpy or heat of vaporization of water, kJ/kg W1, W2, W3 = inlet and exit humidity ratios of the supply inlet, supply outlet, and exhaust inlet airstreams, respectively, kgw/kgda

(5d)

(5e)

Heat or energy exchange effectiveness as defined in Equations (1), (2a), (2b), and (2c) is used to characterize each type of energy transfer in air-to-air exchangers. For a given set of inlet properties and flow rates, knowledge of each effectiveness allows the designer to calculate the sensible, latent, and total energy transfer rates using Equations (5a), (5b), and (5d), respectively. These effectiveness values can be determined either from measured test data or using correlations that have been verified in the peer-reviewed engineering literature. These correlations can also be used to predict energy transfer rates and outlet air properties for operating conditions different from those used for certification purposes. Predicting effectiveness for noncertified operating conditions using certified test data is the most common use of correlations for HVAC designs. Although correlations are not available for all types of air-to-air exchangers under all operating conditions, they are available for the most common types of air-to-air exchanger under operating conditions which have no condensation or frosting.

Rate of Energy Transfer The rate of energy transfer depends on the operating conditions and the intrinsic characteristics of the energy exchanger, such as the geometry of the exchanger (parallel flow/counterflow/ cross-flow, number of passes, fins), thermal conductivity of walls separating the streams, and permeability of walls to various gases. As in a conventional heat exchanger, energy transfer between the airstreams is driven by cross-stream dry-bulb temperature differences. Energy is also transported piggyback-style between the streams by cross-stream mass transfer, which may include air, gases, and water vapor. In another mode of energy transfer, water vapor condenses into liquid in one of the two airstreams of the exchanger. The condensation process liberates the latent heat of condensation, which is transferred to the other stream as sensible heat; this two-step process is also called latent heat transfer. Latent energy transfer between airstreams occurs only when moisture is transferred from one airstream to another without condensation, thereby maintaining the latent heat of condensation. Once moisture has crossed from one airstream to the other, it may either remain in the vapor state or condense in the second stream, depending on the temperature of that stream. Rotating-wheel and permeable-walled flat-plate energy recovery units are used because of their moisture exchange function. Some cross-stream mass transfer may also occur through leakage even when such transfer is unintended. This may alter exchanger performance from its design value, but for most HVAC applications with exhaust air from occupied spaces, small transfers to the supply air are not important. However, these transfers can and should be evaluated during design, and in many cases can be controlled. Heat transfer differs in principle from mass transfer. Heat transfer only occurs when there is a temperature difference. In the case of air-

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2020 ASHRAE Handbook—HVAC Systems and Equipment(SI)

to-air exchange between the supply and exhaust airflow, heat transfer by conduction and convection only occurs when there is a temperature difference between these airstreams. Remember the following facts about heat/mass exchanger performance: • Effectiveness for moisture transfer may not equal the effectiveness for heat transfer. • Total energy effectiveness may not equal either the sensible or latent effectiveness. Net total energy transfer and effectiveness need careful examination when the direction of sensible (temperature-driven) transfer is opposite to that of latent (moisture or water vapor) transfer. ERV performance is expressed by the magnitudes of pumping power and sensible, latent, or total energy recovered. The energy recovered is estimated from the exit temperatures or humidity ratios, which are directly related to the effectiveness. Effectiveness is a function of two parameters: the number of transfer units (NTU) and thermal flow capacity ratio Cr . NTU = UA/Cmin

(6)

Cr = Cmin /Cmax

(7)

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where U = overall heat transfer coefficient, related to flow rates and dimensions of fluid flow path in heat exchanger, kW/(m2 ·K) A = area of heat exchanger, m2 Cmin, Cmax = minimum and maximum capacity rates, expressed for sensible as mscp,s and mecp,e

See Figure 8 for the variation of effectiveness with NTU for a rotary heat wheel. Variation of effectiveness is similar for plate exchangers. Example 1. Inlet supply air enters an ERV with a flow rate of 4.41 m3/s at 35°C and 20% rh. Inlet exhaust air enters with a flow rate of 4.27 m3/s at 24°C and 50% rh. Assume that the energy exchanger was tested under ASHRAE Standard 84, which rated the sensible heat transfer effectiveness at 50% and the latent (water vapor) transfer effectiveness at 50%. With the simplifying assumptions that the specific heat of air is 1 kJ/(kg·K) and the latent heat of vaporization is 2560 kJ/kg, determine the sensible, latent, and net energy transferred in the exchanger. Solution: From the psychrometric chart, the properties of air at 35°C and 20% rh are V1 = 0.8825 m3/kg

h1 = 54.2 kJ/kg

w3 = 0.0093 kg/kg of dry air

h3 = 48 kJ/kg

The mass flow rate at state 1 is obtained from Q v1

The sensible heat transfer qs in the exchanger is found from Equation (5a): q s = 5.0  1  35 – 1  29.5  = 27.5 kW The latent heat transfer qL in the exchanger is found from Equation (5b): q L = 5.0   2560   0.0071  –  2560   0.0080   = – 11.52 kW The net heat energy transfer qs in the exchanger is found from Equation (5d): q = q s + q L = 27.5 – 11.52 = 15.98 kW If the incoming outdoor air conditions had been at 35°C and 14% rh, then the net energy gained by the exhaust airstream would have been zero.

Fan Power It is important to estimate the fan power required to move the supply and exhaust airstream through the exchanger. The fan power Ps required by the supply air is estimated from Q s p s Ps = ----------------- } f

3

4.41 m /s { m 1 = ------1 = --------------------------------= 5.0 kg/s } 3 0.8825 m /kg

Similarly, the mass flow rate at state 3 is obtained from

Q e p e Pe = ------------------ } f

50  1 35 – 0.50  50    1 35 –  1 24  t 2 = ------------------------------------------------------------------------------------------ = 29.5C  50 1 The exit humidity of the supply airstream is found from Equation (4c):

(9)

where Ps Pe Qs Qe ps pe f

= = = = = = =

fan power for supply air, W fan power for exhaust air, W supply volume flow rate, m3/s exhaust volume flow rate, m3/s pressure drop of supply air caused by fluid friction, Pa pressure drop of exhaust air caused by fluid friction, Pa overall efficiency of fan and motor or product of fan and motor efficiencies

If pressure drops through the exchanger are not known at the operating conditions, they may be estimated from a standard pressure drop rating at a reference condition. Ideally, the density, viscosity, and Reynolds number of the air are known, which allows use of the following approach to estimate operating pressure drops. The density and viscosity of air vary with temperature, and therefore pressure drop through the exchanger varies with temperature as well as flow rate. The variation of viscosity with temperature is given by the Sutherland law as  T  3 / 2  To + S  ------ =  ------------------ T o  T + S  o

Q3 4.27 m 3 /s = 5.0 kg/s m 3 = ------ = -----------------------------v3 0.854 m 3 /kg Exit temperatures of the supply airstream can be obtained from Equation (4a):

(8)

The fan power Pe required by the exhaust air is estimated from

w1 = 0.0071 kg/kg of dry air

and the properties of air at 24°C and 50% rh are V3 = 0.854 m3/kg

w2 = {5.0(2560)0.0071 – 0.50(5.0)[(2560)0.0071 – (2560)0.0093]}/[(5.0)(2560)] = 0.0080 kgw/kgda

where  o T To S

= = = = =

dynamic viscosity, kg/(m·s) dynamic viscosity at the reference temperature, kg/(m·s) absolute temperature, K reference temperature, K constant = 110.4 K

The Reynolds number can be estimated from

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(10)

Air-to-Air Energy Recovery Equipment V  av D h Re D = -----------------------T

26.5 (11)

where  V Dh 

= = = =

air density, kg/m3 average velocity in flow channels, m/s hydraulic diameter of flow channels, m dynamic viscosity at average temperature T, kg/(m·s)

When flow through the exchanger is transitional or turbulent (i.e., when the Reynolds number ReD for airflow through the exchanger is in the range 5 × 103  ReD  105), then, treating air as an ideal gas, the pressure drop p at any temperature T is related to the pressure drop po at reference temperature To and is expressed as 0.25 p T  1.375  T m  1.75  ----o+S --------- =  ------------------      p o T+S To mo

(12)

where

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p po m mo

= = = =

pressure drop, Pa pressure drop at reference temperature, Pa mass flow rate at operating condition, kg/s mass flow rate at rating condition, kg/s

For fully developed laminar flow through an energy exchanger, i.e. where ReD < 2000, the corresponding dimensionless pressure drop relation similar to Equation (12) is given as + C  T o p T  3 / 2  1---------------------m   ------------- =  ----- m o  T o  1 + C  T  p o

(13)

The equations above cannot be used in the case of (1) flow channels that are not reasonably smooth; (2) flow Reynolds numbers that are out of range; (3) significant exchanger fouling caused by condensation, frost, or dust; (4) excessive nonuniform property distributions inside the exchanger; or (5) significant pressure deformation of the flow channels (e.g., some plate cross-flow exchangers). The total pumping power P of the ERV can be given as P = Ps + Pe

3.

(14)

TYPES OF AIR-TO-AIR HEAT EXCHANGERS

Ideal Air-to-Air Energy Exchange An ideal air-to-air energy exchanger • Allows temperature-driven heat transfer between participating airstreams • Allows partial-pressure-driven moisture transfer between the two streams • Minimizes cross-stream transfer of air, other gases (e.g., pollutants), biological contaminants, and particulates • Optimizes energy recovery performance to minimize pressure drop while providing reasonable cost, dimensions, and mass Heat transfer is an important energy recovery vehicle from airstreams that carry waste heat. The role of moisture transfer as an energy recovery process is less well known and merits explanation. Consider an air-to-air energy exchanger operating in a hot, humid climate in a comfort air-conditioning application. If the energy exchanger exchanges heat but not moisture, it cools outdoor ventilation air as it passes through the exchanger to the indoor space. Heat flows from the incoming outdoor air to the outgoing (and cooler) exhaust air drawn from the indoor conditioned space. This does very little to mitigate the high humidity carried into the indoor

space by the outdoor ventilation air and may even increase the relative humidity in the conditioned space, resulting in increased refrigeration and/or reheat to dehumidify the air and achieve acceptable comfort conditions. On the other hand, if the energy exchanger transfers both heat and moisture, the humid outdoor supply air transfers moisture to the less-humid inside exhaust air as the streams pass through the energy exchanger. The lower humidity of the entering ventilation air requires less energy input for comfort conditioning.

Fixed-Plate Heat Exchangers Plate exchangers are available in many configurations, materials, sizes, and flow patterns. Many have modules that can be arranged to handle almost any airflow, effectiveness, and pressure drop requirement. Plates are formed with spacers or separators (e.g., ribs, dimples, ovals) constructed into the plates or with external separators (e.g., supports, braces, corrugations). Airstream separations are sealed by folding, multiple folding, gluing, cementing, welding, or any combination of these, depending on the application and manufacturer. Ease of access for examining and cleaning heat transfer surfaces depends on the configuration and installation. Heat transfer resistance through the plates is small compared to the airstream boundary layer resistance on each side of the plates. Heat transfer efficiency is not substantially affected by the heat transfer coefficient of the plates. Aluminum is the most popular plate construction material because of its nonflammability and durability. Polymer plate exchangers may improve heat transfer by causing some turbulence in the channel flow, and are popular because of their corrosion resistance and cost-effectiveness. Steel alloys are used for temperatures over 200°C and for specialized applications where cost is not a key factor. Plate exchangers normally conduct sensible heat only; however, water-vapor-permeable materials, such as treated paper and microporous polymeric membranes, may be used to transfer moisture, thus providing total (enthalpy) energy exchange. Most manufacturers offer modular plate exchangers. Modules range in capacity from 0.01 to 5 m3s and can be arranged into configurations exceeding 50 m3s. Multiple sizes and configurations allow selections to meet space and performance requirements. Plate spacing ranges from 2.5 to 12.5 mm, depending on the design and application. Heat is transferred directly from the warm airstreams through the separating plates into the cool airstreams. Usually design, construction, and cost restrictions result in the selection of cross-flow exchangers, but additional counterflow patterns can increase heat transfer effectiveness. Normally, both latent heat of condensation (from moisture condensed as the temperature of the warm exhaust airstream drops below its dew point) and sensible heat are conducted through the separating plates into the cool supply airstream. Thus, energy is transferred but moisture is not. Recovering 80% or more of the available waste exhaust heat is possible. Fixed-plate heat exchangers can achieve high sensible heat recovery and total energy effectiveness because they have only a primary heat transfer surface area separating the airstreams and are therefore not inhibited by the additional secondary resistance (e.g., pumping liquid, in runaround systems or transporting a heat transfer medium) inherent in some other exchanger types. In a crossflow arrangement (Figure 2), they usually do not have sensible effectiveness greater than 75% unless two devices are used in series as shown in Figure 3. One advantage of the plate exchanger is that it is a static device with little or no leakage between airstreams. As velocity increases, the pressure difference between the two airstreams increases. High differential pressures may deform the separating plates and, if excessive, can permanently damage the exchanger, significantly reducing the airflow rate on the low-pressure side as well as the effectiveness

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Fig. 4

Fig. 3

Variation of Pressure Drop and Effectiveness with Airflow Rates for Membrane Plate Exchanger

and possibly causing excessive air leakage. This is not normally a problem because differential pressures in most applications are less than 1 kPa. In applications requiring high air velocities, high static pressures, or both, plate exchangers built to withstand these conditions are available and should be specified. Most sensible plate exchangers have condensate drains, which remove condensate and also wastewater in water-wash systems. Heat recovered from a high-humidity exhaust is better returned to a building or process by a sensible heat exchanger rather than an enthalpy exchanger if humidity transfer is not desired. Frosting can be controlled by preheating incoming supply air, bypassing part of the incoming air, recirculating supply air through the exhaust side of the exchanger, or temporarily interrupting supply air while maintaining exhaust. However, frost on cross-flow heat exchangers is less likely to block the exhaust airflow completely than with other types of exchangers. Generally, frost should be avoided unless a defrost cycle is included. Fixed-plate heat exchangers can be made from permeable membranes designed to maximize moisture and energy transfer between airstreams while minimizing air transfer. Suitable permeable microporous membranes for this emerging technology include cellulose, polymers, and other synthetic materials such as hydrophilic electrolyte. Hydrophilic electrolytes are made from, for example, sulphonation chemistry techniques and contain charged ions that

Typical Temperature Stratification at Outlets of Cross-Flow Heat Exchanger

attract polar water molecules; adsorption and desorption of water occur in vapor state. Airstreams exiting a cross-flow plate exchanger display temperature stratification when there is a temperature difference between the two airstreams (Figure 4). Enthalpy plate exchangers will also display humidity stratification. This should be considered if the temperature of one of these airstreams is to be measured (e.g., to control operation of downstream conditioning equipment or of a defrost mechanism). Heat exchanger effectiveness depends heavily on the airflow direction and pattern of the supply and exhaust airstreams. Parallelflow exchangers (Figure 5A), in which both airstreams move along heat exchange surfaces in the same direction, have a theoretical maximum effectiveness of 50%. Counterflow exchangers (Figure 5B), in which airstreams move in opposite directions, can have a theoretical effectiveness approaching 100%, but typical units have a lower effectiveness. Theoretical effectiveness for cross-flow heat exchangers is somewhat lower than for counterflow, and typical units have effectiveness of 50 to 70% (Figure 5C) and 60 to 85% for multiple-pass exchangers (Figure 5D). In practice, construction limitations favor designs that use transverse flow (or cross-flow) over much of the heat exchange surface (Figures 5C and 5D).

Rotary Air-to-Air Energy Exchangers A rotary air-to-air energy exchanger, or rotary enthalpy wheel, has a revolving cylinder filled with an air-permeable medium having a large internal surface area. Adjacent supply and exhaust airstreams each flow through half the exchanger in a counterflow pattern (Figure 6). Heat transfer media may be selected to recover sensible heat only or total (sensible plus latent) heat. Sensible heat is transferred as the medium picks up and stores heat from the hot airstream and releases it to the cold one. Latent heat is transferred as the medium adsorbs water vapor from the higher-humidity airstream and desorbs moisture into the lowerhumidity airstream, driven in each case by the vapor pressure difference between the airstream and energy exchange medium. Thus, the moist air is dried while the drier air is humidified. In total heat transfer, both sensible and latent heat transfer occur simultaneously. Sensible-only wheels (not coated with desiccant) can also transfer water via a mechanism of condensation and reevaporation driven by dew point and vapor pressure; the effectiveness varies strongly with conditions. Because rotary exchangers have a counterflow configu-

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Air-to-Air Energy Recovery Equipment ration and normally use small-diameter flow passages, they are quite compact and can achieve high transfer effectiveness. Air contaminants, dew point, exhaust air temperature, and supply air properties influence the choice of materials for the casing, rotor structure, and medium of a rotary energy exchanger. Aluminum, steel, and polymers are the usual structural, casing, and rotor materials for normal comfort ventilating systems. Exchanger media are fabricated from metal, mineral, or synthetic materials and provide either random or directionally oriented flow through their structures. Random-flow media are made by knitting wire into an open woven cloth or corrugated mesh, which is layered to the desired configuration. Aluminum mesh is packed in pie-shaped wheel segments. Stainless steel and monel mesh are used for high-temperature and corrosive applications. These media should only be used with clean, filtered airstreams because they plug easily. Random-flow media also require a significantly larger face area than directionally oriented media for a given airflow and pressure drop. Directionally oriented media are available in various geometric configurations. The most common consist of small (1.5 to 2 mm) air passages parallel to the direction of airflow. Air passages are very similar in performance regardless of their shape (triangular, hexagonal, parallel plate, or other). Aluminum foil, paper, plastic, and synthetic materials are used for low and medium temperatures. Stainless steel and ceramics are used for high temperatures and corrosive atmospheres. Media surface areas exposed to airflow vary from 300 to over 4000 m2m3, depending on the type of medium and physical configuration. Media may also be classified by their ability to recover sensible heat only or total heat. Media for sensible heat recovery are made of aluminum, copper, stainless steel, and monel. Media for total heat recovery can be from any of a number of materials and

Fig. 6 Rotary Air-to-Air Energy Exchanger

26.7 treated with a desiccant (typically zeolites, molecular sieves, silica gels, activated alumina, titanium silicate, synthetic polymers, lithium chloride, or aluminum oxide) to have specific moisture recovery characteristics. Cross-Leakage. Cross-leakage (mixing between supply and exhaust airstreams) can occur in all rotary energy exchangers by two mechanisms: carryover and seal leakage. Cross leakage can be reduced by placing the blowers so that they promote leakage of outdoor air to the exhaust airstream. A purge section also can be installed on the heat exchanger to reduce cross leakage. In many applications, recirculating some air is not a concern. However, critical applications such as hospital operating rooms, laboratories, and cleanrooms require stringent control of carryover. Carryover can be reduced to less than 0.1% of the exhaust airflow with a purge section but cannot be completely eliminated. The theoretical carryover of a wheel without a purge section is directly proportional to the speed of the wheel and the void volume of the medium (75 to 95% void, depending on type and configuration). For example, a 3 m diameter, 200 mm deep wheel with a 90% void volume operating at 14 rpm has a carryover volumetric flow of (3/2)2(0.2)(0.9)(14/60) = 0.3 m3s If the wheel is handling a 9 m3/s balanced flow, the percentage carryover is 0.3 -------  100 = 3.3% 9 The exhaust fan, which is usually located at the exit of the exchanger, should be sized to include leakage and purge airflows. Control. Two control methods are commonly used to regulate wheel energy recovery. In supply or exhaust air bypass control, the amount of supply air allowed to pass through the wheel establishes the supply air temperature. An air bypass damper, controlled by a wheel supply air discharge temperature sensor, regulates the proportion of supply air permitted to bypass the exchanger. The second method regulates the energy recovery rate by varying wheel rotational speed. The most frequently used variable-speed drives are (1) a silicon controlled rectifier (SCR) with variable-speed dc motor, (2) a constant-speed AC motor with hysteresis coupling, and (3) an AC frequency inverter with an AC induction motor. The dehumidification capacity and reheating effectiveness of a energy wheel can be varied by changing the wheel speed (Figure 7A) or by bypassing air around the wheel (Figure 7B) (Moffitt 2010). Figure 7A is a typical capacity curve for varying the wheel’s rotation speed when the outdoor air is cooler and drier than the exhaust air; when outdoor air conditions are cooler and more humid, capacity may increase at low speeds (Simonson et al. 2000).

Fig. 5 Plate Heat or Heat and Mass Exchanger Airflow Configurations

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Fig. 7 Latent and Sensible Effectiveness Versus (A) Wheel Speed and (B) Bypassed Air (Moffitt 2010)

Figure 8 shows the effectiveness , sensible heat transfer only, with balanced airflow, convection/conduction ratio less than 4, and no leakage or cross-flow, of a regenerative counterflow heat wheel versus number of transfer units (NTU). This simple example of a regenerative wheel also shows that regenerative counterflow rotary effectiveness increases with wheel speed (Cr is proportional to wheel speed), but there is no advantage in going beyond Cr /Cmin = 5 because the carryover of contaminants increases with wheel speed. See Kays and Crawford (1993) or Shah (1981) for details. Rotary energy or enthalpy wheels are more complex than heat wheels, but recent research has characterized their behavior using laboratory and field data (Johnson et al. 1998). A dead band control, which stops or limits the exchanger, may be necessary when no recovery is desired (e.g., when outdoor air temperature is higher than the required supply air temperature but below the exhaust air temperature). When outdoor air temperature is above the exhaust air temperature, the equipment operates at full capacity to cool incoming air. During very cold weather, it may be necessary to heat the supply air, stop the wheel, or, in small systems, use a defrost cycle for frost control. Rotary enthalpy wheels require little maintenance and tend to be self-cleaning because the airflow direction is reversed for each rotation of the wheel. The following maintenance procedures ensure best performance: • Clean the medium when lint, dust, or other foreign materials build up, following the manufacturer’s instructions. • Maintain drive motor and train according to the manufacturer’s recommendations. Speed-control motors that have commutators and brushes require more frequent inspection and maintenance than induction motors. Brushes should be replaced, and the commutator should be periodically turned and undercut. • Inspect wheels regularly for proper belt or chain tension. • Refer to the manufacturer’s recommendations for spare and replacement parts.

Coil Energy Recovery (Runaround) Loops A typical coil energy recovery loop (Figure 9) places extendedsurface, finned-tube water coils in the supply and exhaust airstreams of a building or process. The coils are connected in a closed loop by

Fig. 8

Effectiveness of Counterflow Regenerator as a Function of NTU (Shah 1981)

counterflow piping through which an intermediate heat transfer fluid (typically water or antifreeze solution) is pumped. Moisture must not freeze in the exhaust coil air passage. A dualpurpose, three-way temperature control valve prevents the exhaust coil from freezing. The valve is controlled to maintain the temperature of solution entering the exhaust coil at 5C or above. This condition is maintained by bypassing some of the warmer solution around the supply air coil. The valve can also ensure that a prescribed air temperature from the supply air coil is not exceeded. Coil energy recovery loops are highly flexible and well suited to renovation and industrial applications. The loop accommodates remote supply and exhaust ducts and allows simultaneous transfer of energy between multiple sources and uses. An expansion tank must be included to allow fluid expansion and contraction. A closed expansion tank minimizes oxidation when ethylene glycol is used. Standard finned-tube water coils may be used; however, these need to be selected using an accurate simulation model if high effectiveness and low costs are needed (Johnson et al. 1995). Integrating runaround loops in buildings with variable loads to achieve maximum benefits may require combining the runaround simulation with building energy simulation (Dhital et al. 1995). Manufacturers’ design curves and performance data should be used when selecting coils, face velocities, and pressure drops, but only when the design

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Air-to-Air Energy Recovery Equipment data are for the same temperature and operating conditions as in the runaround loop. When testing chilled-water coils for certification, AHRI Standard 410 requires testing in the range of 1.7 to 12.8°C. Care should be taken when using manufacturers’ selection programs for sizing coils for runaround loops. It is common to use a cooling coil design for these systems, but average water temperatures can vary outside the range for which these coils are typically used. For example, in commercial comfort conditioning, average water temperatures in a runaround loop can range from –4°C (–29°C outdoors, 21°C return) to 27°C (35°C outdoors, 24°C return) in a single application, whereas cooling coils are normally applied using a steady chilled-water supply designed from a narrower range. The coil energy recovery loop cannot transfer moisture between airstreams; however, indirect evaporative cooling can reduce the exhaust air temperature, which significantly reduces cooling loads. For the most cost-effective operation, with equal airflow rates and no condensation, typical effectiveness values range from 45 to 65%. The highest effectiveness does not necessarily give the greatest net life-cycle cost saving. Typically, the sensible heat effectiveness of a coil energy recovery loop is independent of the outdoor air temperature. However, when the capacity is controlled, the sensible heat effectiveness decreases. Coil energy recovery loops use coils constructed to suit their environment and operating conditions. For typical comfort-to-comfort applications, standard coil construction usually suffices. In processto-process and process-to-comfort applications, the effect of high temperature, condensable gases, corrosives, and contaminants on the coil(s) must be considered. Above 200°C, special construction may be required to ensure a permanent fin-to-tube bond. The effects of condensable gases and other adverse factors may require special coil construction and/or coatings. Chapters 23 and 27 discuss the construction and selection of coils in more detail. Complete separation of the airstreams eliminates crosscontamination between the supply and exhaust air. Coil energy recovery loops require little maintenance. The only moving parts are the circulation pump and three-way control valve. However, to ensure optimum operation, the air should be filtered, the coil surface cleaned regularly, the pump and valve maintained, and the transfer fluid refilled or replaced periodically. Fluid manufacturers or their representatives should be contacted for specific recommendations. The thermal transfer fluid selected for a closed-loop exchanger depends on the application and on the temperatures of the two airstreams. An inhibited ethylene glycol solution in water is common when freeze protection is required. These solutions break down to an acidic sludge at temperatures above 135°C. If freeze protection is needed and exhaust air temperatures exceed 135°C, a nonaqueous synthetic heat transfer fluid should be used. Heat transfer fluid manufacturers and chemical suppliers should recommend appropriate fluids.

Heat Pipe Heat Exchangers Figure 10 shows a typical heat pipe assembly. Hot air flowing over the evaporator end of the heat pipe vaporizes the working fluid. A vapor pressure gradient drives the vapor to the condenser end of the heat pipe tube, where the vapor condenses, releasing the latent energy of vaporization (Figure 11). The condensed fluid is wicked or flows back to the evaporator, where it is revaporized, thus completing the cycle. Thus the heat pipe’s working fluid operates in a closed-loop evaporation/condensation cycle that continues as long as there is a temperature difference to drive the process. Using this mechanism, the heat transfer rate along a heat pipe is up to 1000 times greater than through copper (Ruch 1976).

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Fig. 9 Coil Energy Recovery Loop

Fig. 10 Heat Pipe Assembly

Fig. 11 Heat Pipe Operation Energy transfer in heat pipes is often considered isothermal. However, there is a small temperature drop through the tube wall, wick, and fluid medium. Heat pipes have a finite heat transfer capacity that is affected by factors such as wick design, tube diameter, working fluid, and tube (heat pipe) orientation relative to horizontal. In current designs, a wick is replaced by circumferential grooves that facilitate capillary-action flow of condensed refrigerant back to the evaporator section. HVAC systems use copper or aluminum heat pipe tubes with aluminum or copper fins. Fin designs include continuous corrugated plate, continuous plain, and spiral. Modifying fin design and tube spacing changes pressure drop at a given face velocity. For process-to-comfort applications with large temperature changes, tubes and fins are usually constructed of the same material to avoid problems with different thermal expansions of materials. Heat pipe heat exchangers for exhaust temperatures below 220°C are most often constructed with aluminum or copper tubes and fins. Protective coatings allow inexpensive aluminum to replace exotic metals in corrosive atmospheres; these coatings have a minimal effect on thermal performance.

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Fig. 12 Heat Pipe Exchanger Effectiveness Heat pipe heat exchangers for use above 220°C are generally constructed with steel tubes and fins. The fins are often aluminized to prevent rusting. Composite systems for special applications may be created by assembling units with different materials and/or different working fluids. Selecting the proper working fluid for a heat pipe is critical to longterm operation. The working fluid should have high latent heat of vaporization, a high surface tension, and a low liquid viscosity over the operating range; it must be thermally stable at operating temperatures. Decomposition of the thermal fluids can form noncondensable gases that deteriorate performance. For low-temperature applications, gases such as helium can be used as working fluid; for moderate temperatures, commercial refrigerants and liquids such as water can be used; and for high temperatures, liquid metals such as sodium or mercury can be used. Heat pipe heat exchangers typically have no cross-contamination for pressure differentials between airstreams of up to 12 kPa. A vented double-wall partition between the airstreams can provide additional protection against cross contamination. If an exhaust duct is attached to the partition space, any leakage is usually withdrawn and exhausted from the space between the two ducts. Heat pipe heat transfer capacity depends on design and orientation. Figure 12 shows a typical effectiveness curves for various face velocities and rows of tubes. As the number of rows increases, effectiveness increases at a decreasing rate. For example, doubling the number of rows of tubes in a 60% effective heat exchanger increases the effectiveness to 75%. The effectiveness of a counterflow heat pipe heat exchanger depends on the total number of rows such that two units in series yield the same effectiveness as a single unit of the same total number of rows. Series units are often used to facilitate handling, cleaning, and maintenance. Effectiveness also depends on outdoor air temperature and the ratio of mass flow rates of the airstreams. Typically, heat capacity in the cooling season increases with a rise in outdoor air temperature. It has an opposite effect during the heating season. Effectiveness typically increases with the ratio of mass flow rates of the fluids (flow rate of the fluid with warmer entering temperature over that of cooler entering fluid temperature). The heat transfer capacity of a heat pipe increases roughly with the square of the inside diameter of the pipe. For example, at a given tilt angle, a 25 mm inside diameter heat pipe will transfer roughly 2.5 times as much energy as a 16 mm inside diameter pipe. Consequently, heat pipes with large diameters are used for larger-airflow

applications and where level installation is required to accommodate both summer and winter operation. Heat transfer capacity is virtually independent of heat pipe length, except for very short heat pipes. For example, a 1 m long heat pipe has approximately the same capacity as an 2 m pipe. Because the 2 m heat pipe has twice the external heat transfer surface area of the 1 m pipe, it will reach its capacity limit sooner. Thus, in some applications, it is more difficult to meet capacity requirements as the heat pipes become longer. A system can be reconfigured to a taller face height and more numerous but shorter heat pipes to yield the same airflow face area while improving system performance. Also, under limiting conditions, adding rows increases performance more proportionally. In typical HVAC comfort applications, the heat pipe capacity is not reached, and longer heat pipes can be used to recover more energy. The limit for a level (twoseason) 12.7 mm-diameter heat pipe, using R-410A and six rows deep, is 5080 mm in most HVAC comfort applications. The selection of fin design and spacing should be based on the dirtiness of the two airstreams and the resulting cleaning and maintenance required. For HVAC applications, fin spacing of 1.8 to 2.3 mm is common. Wider spacing (2.5 to 3.2 mm) is usually used for industrial applications. Plate-fin heat pipe heat exchangers can easily be constructed with different fin spacing for the exhaust and supply airstreams, allowing wider fin spacing on the dirty exhaust side. This increases design flexibility where pressure drop constraints exist and also prevents deterioration of performance caused by dirt build-up on the exhaust side surface. Changing the tilt of a heat pipe controls the amount of heat it transfers. Operating the heat pipe on a slope with the hot end below (or above) the horizontal improves (or retards) condensate flow back to the evaporator end of the heat pipe. This feature can be used to regulate the effectiveness of the heat pipe heat exchanger (Guo et al. 1998). Introducing the action of gravity makes the device a hybrid heat pipe/thermosiphon. Tilt control is achieved by pivoting the exchanger about the center of its base and attaching a temperature-controlled actuator to one end of the exchanger (Figure 13). Pleated flexible connectors attached to the ductwork allow freedom for the small tilting movement of only a few degrees. Tilt control may be desired • To change from supply air heating to supply air cooling (i.e., to reverse the direction of heat flow) during seasonal changeover • To modulate effectiveness to maintain desired supply air temperature (often required for large buildings to avoid overheating air supplied to the interior zone) • To decrease effectiveness to prevent frost formation at low outdoor air temperatures (with reduced effectiveness, exhaust air leaves the

Fig. 13 Heat Pipe Heat Exchanger with Tilt Control

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Other devices, such as face-and-bypass dampers and preheaters, can also be used to control the rate of heat exchange. Recent design improvements allow bidirectional heat transfer in heat pipes. Some heat pipe manufacturers have eliminated the need for tilting for capacity control or seasonal changeover. Once installed, the unit is removed only for routine maintenance. Capacity and frost control can also be achieved through bypassing airflow over the heat pipes, as for air coils. Bypassing airflow is the preferred method of control in applications where the average (annual) recovery efficiency ratio (RER) is important, because the reduced airflow through the heat pipe results in reduced air pressure losses during periods of modulation. Example 2. Sensible Heat Energy Recovery in a Heat Pipe. Outdoor air at 10°C enters a six-row heat pipe with a flow rate of 5 kgs and a face velocity of 2.5 ms. Exhaust air enters the heat pipe with the same velocity and flow rate but at 24°C. The pressure drop across the heat pipe is 150 Pa. The supply air density is 1.35 kgm3. The efficiency of the electric motor and the connected fan are 90 and 75%, respectively. Assuming the performance characteristics of the heat pipe are as shown in Figure 18, determine the sensible effectiveness, exit temperature of supply air to the space, energy recovered, and power supplied to the fan motor.To simplify the calculations, assume the specific heat is a constant of 1.0 kJ/(kg·K). Solution: From Figure 12, at face velocity of 2.5 ms and with six rows, the effectiveness is about 58%. Because the mass flow rate of the airstreams is the same and assuming their specific heat of 1 kJ(kg·K) is the same, then the exit temperature of the supply air to the space can be obtained from Equation (4a):  5   1 10 – 0.58  5   1   10 – 24  T2 = ------------------------------------------------------------------------------- = 18.1K 51 The sensible energy recovered can be obtained from Equation (5a), incorporating the simplification of specific heat to a single value, as

Fig. 14

Sealed-Tube Thermosiphons

qs = (5)(1)(18.12 – 10) = 40.6 kW The supply air fan power can be obtained from Equation (8) as 5 Ps = ---------- (150)[(0.9)(0.75)] = 823 W or 0.823 kW 1.35 Because there are two airstreams, neglecting the difference in the air densities of the airstreams, the total pumping power of the heat pipe is twice the above value (i.e., 1.65 kW).

Thermosiphon Heat Exchangers Two-phase thermosiphon heat exchangers are sealed systems that consist of an evaporator, a condenser, interconnecting piping, and an intermediate working fluid that is present in both liquid and vapor phases. Two types of thermosiphon are used: a sealed tube (Figure 14) and a coil type (Figure 15). In the sealed-tube thermosiphon, the evaporator and the condenser are usually at opposite ends of a bundle of straight, individual thermosiphon tubes, and the exhaust and supply ducts are adjacent to each other (this arrangement is similar to that in a heat pipe system). In coil-type thermosiphons, evaporator and condenser coils are installed independently in the ducts and are interconnected by the working fluid piping (this configuration is similar to that of a coil energy recovery loop). A thermosiphon is a sealed system containing a two-phase working fluid. Because part of the system contains vapor and part contains liquid, the pressure in a thermosiphon is governed by the liquid temperature at the liquid/vapor interface. If the surroundings cause a temperature difference between the regions where liquid and vapor interfaces are present, the resulting vapor pressure difference causes

Fig. 15 Coil-Type Thermosiphon Loops

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Fig. 16 Heat Recovered Across Evaporator and Condenser Coils for (A) Level Thermosiphon and (B) Thermosiphon with Elevated Condenser vapor to flow from the warmer to the colder region. The flow is sustained by condensation in the cooler region and by evaporation in the warmer region. The condenser and evaporator must be oriented so that the condensate can return to the evaporator by gravity (Figures 14 and 15). For coil-type thermosiphon loops (Figure 15), static liquid level or refrigerant charge has a significant effect on system performance. The static charge should be optimized: a low static charge leads to dryout in the evaporator and reduced system performance, and a high charge suppresses condensation in the condenser. When the coils are on the same vertical level, as in Figure 15B, a certain amount of suppressed condensation in the condenser is inevitable because a certain static liquid head is required in the bottom of the condenser to drive liquid through the loop into the evaporator (Wei Qu 2010). Figure 16A shows a typical temperature change profile along the vertical section of a condenser, where lower T indicates loss of energy recovery potential, and the thermosiphon manufacturer should quote performance as an average across the whole airstream. Wei Qu (2010) found that the condenser need not be placed much higher than the evaporator to reach full potential; Figure 16B shows that raising the condenser 25 to 30% of the coil face height above the evaporator can result in negligible variation in the temperature profile. In thermosiphon systems, a temperature difference and gravity force are required for the working fluid to circulate between the evaporator and condenser. As a result, thermosiphons may be designed to transfer heat equally in either direction (bidirectional), in one direction only (unidirectional), or in both directions unequally. Although similar in form and operation to heat pipes, thermosiphon tubes are different in two ways: (1) they have no wicks and hence rely only on gravity to return condensate to the evaporator, whereas heat pipes use capillary forces; and (2) they depend, at least initially, on nucleate boiling, whereas heat pipes vaporize the fluid from a large, ever-present liquid/vapor interface. Thus, thermosiphon heat exchangers may require a significant temperature difference to initiate boiling (Mathur and McDonald 1987; McDonald and Shivprasad 1989). Thermosiphon tubes require no pump to circulate the working fluid. However, the geometric configuration must be such that liquid working fluid is always present in the evaporator section of the heat exchanger. Thermosiphon loops differ from other coil energy recovery loop systems in that they require no pumps and hence no external power

Fig. 17 Eight-Row, Unidirectional System with 2 mm Fin Spacing; Static Charge of 80% (Mathur and McDonald 1986)

supply, and the coils must be appropriate for evaporation and condensation. Two-phase thermosiphon loops are used for solar water heating (Mathur 1990a) and for performance enhancement of existing (i.e., retrofit applications) air-conditioning systems (Mathur 1997a, 1997b, 1997c, 1997d). Two-phase thermosiphon loops can be used to downsize new air-conditioning systems and thus reduce the overall project costs. Figure 17 shows thermosiphon loop performance (Mathur and McDonald 1986) for an eightrow systemwith 2 mm fin spacing in unidirectional mode. Figure 18 shows the performance of the same system as a function of percent static charge for an overall temperature difference of 35 K.These systems have been designed with various working fluids like n-pentene and R-134a (Samba et al. 2013) with lower saturation pressures at atmospheric pressure, water (Mathur 1994) at negative operating pressures at typical ambient condition, and napthalene for high-temperature applications (Orr et al. 2016). These energy recovery systems are available to enhance system performance (similar to heat pipe heat exchangers) of existing systems (Cieslinski and Fiuk 2013; Mathur 1994; Yau 2007) or to downsize new equipment. Using two-phase thermosiphon loop systems

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26.13 Liquid-Desiccant Cooling Systems

Fig. 18 Typical Performance of Two-Phase Thermosiphon Loop as Function of Percent Static Charge: Eight Rows, Unidirectional System with 2 mm Fin Spacing; Overall Temperature Difference of 35 K Licensed for single user. © 2020 ASHRAE, Inc.

(Mathur and McDonald 1986)

or other energy recovery systems can improve the overall EERs of the systems and can earn significant LEED® points. These systems are also used for exhaust heat recovery for automotive application (Orr et al. 2016) and for cooling telecommunications systems (Samba et al. 2013). Example 3. Sensible Heat Energy Recovery in Two-Phase Thermosiphon Loop. Outdoor air at –12°C enters an eight-row. two-phase thermosiphon loop with a face area of 0.2 m2 at a face velocity of 2.2 m/s. Exhaust air enters the two-phase thermosiphon loop with the same velocity with same face area at 24°C. Pressure drop across the heat pipe is 0.18 kPa. The supply air density is 1.35 kg/m3. Efficiencies of the electric motor and connected fan are 90 and 75%, respectively. Assuming the performance characteristics of the two-phase thermosiphon as shown in Figure 26, determine the sensible effectiveness sensible, exit temperature of supply air to the space T2, sensible energy recovered qs, and power supplied to the fan motor ps. Solution: From the data given, the overall temperature difference between supply and exhaust airstreams is 36 K [i.e., 24 – (–12) = 36]. From Figure 17, at face velocity of 2.2 m/s and with eight rows, the effectiveness is about 52.5%. Because the mass flow rate of the airstreams is the same, and, for simplification, assuming specific heat is constant at 1.0 kJ/ (kg·K), then the exit temperature of the supply air to the space can be obtained as follows: t2 = –12 + 0.525[24 – (–12)] = 6.9°C The sensible energy recovered can be obtained from a variation of Equation (5a) as qs = (2.2)(0.2)(1.35)(1.0)[6.9 – (–12)] = 11.2 kW The supply air fan power can be obtained using Equation (8) as Qs = (2.2)(0.2) = 0.44 m3/s Ps = [(0.44)(0.18)]/[(0.9)(0.75)] = 117 W Because there are two airstreams, the total pumping power of the two-phase thermosiphon loop (neglecting the differences in air density between the airstreams) is approximately twice the value of Ps (i.e., 234 W).

Liquid-desiccant cooling systems (LDCSs) have great potential in reducing energy costs associated with meeting latent loads in high-occupancy commercial building applications. An LDCS can produce very dry air as long as the desiccant solution is supplied to its dehumidifier at low temperatures. However, an LDCS requires additional components: two or three plate-to-plate heat exchangers, a separate cooling tower, pump, fan, and a packed-bed humidifier and regenerator. When exposed to moist air at the proper temperature and concentration, some chemical solutions (e.g., lithium chloride, lithium bromide, calcium chloride, triethylene glycol) can either absorb water vapor from, or release water vapor to, the airstream. These chemicals, known as liquid desiccants, are used to dehumidify air with significantly less energy consumption than by a mechanical refrigeration system. Liquid desiccants also have the advantage of removing bacteria and dust from the air compared to solid desiccant wheels or mechanical vapor compression systems. It requires a lower-temperature (less than 82°C) heat source for regeneration, which can be provided by solar energy or waste heat from local equipment. Moisture transfer between the desiccant solution and process air is caused by the difference in vapor pressures at the desiccant surface and that of the air passing over it. The desiccant collects moisture when its vapor pressure is lower than that of the water vapor in the air, and releases it when its vapor pressure is higher. Absorption of water vapor releases heat energy, including the heat of vaporization and, in some cases, additional heat. After absorbing water vapor from the processes air, the solution temperature increases slightly, and its concentration drops as absorbed water vapor becomes part of the solvent of the solution. For cyclic operation, the desiccant solution is regenerated to restore its original concentration; this involves heating the solution with a lowgrade heat source and exposing the heated solution to an airstream, as shown in Figure 19 (Williams 2007). Often, outdoor air is used as a scavenging airstream in regeneration. Heat added to the desiccant solution for regeneration is of low grade and can be obtained from a source at no more than 71°C. Such low-temperature energy sources are often available as waste heat in many applications. As seen in Figure 19, the hot, humid process air to be dehumidified enters the packed-bed dehumidifier or absorber at state 1 and leaves at state 2. The concentrated desiccant solution enters the absorber at state 5 and leaves at state 6 as dilute solution after absorbing water vapor from the process airstream. The dilute solution now passes through a liquid-to-liquid plate heat exchanger to recover some of the heat energy from the concentrated solution exiting the regenerator. The dilute solution is heated further to the required temperature in the desiccant solution heating coil and is sprayed over the incoming outdoor scavenging air passing through the regenerator. For air-to-air energy recovery purposes, the exhaust air from the space can be used for greater effectiveness in place of outdoor scavenging air. Because of the high inlet temperature of the desiccant solution, its vapor pressure is relatively higher than that of the incoming air, causing moisture transfer from the solution to the air. The inlet temperature of the desiccant solution to the regenerator for the given state of the air at state 8 is such that the desiccant concentration is restored to its value at state 6 as it exits the regenerator. The concentrated desiccant solution leaving the regenerator transfers heat energy to the incoming dilute solution exiting the packedbed dehumidifier, in the liquid-to-liquid plate heat exchanger. Further temperature reduction of the desiccant solution, if required before it enters the absorber, can be accomplished by using cooling water from a cooling tower in the desiccant cooling coil. During extreme high-latent-load conditions, chilled water, rather than water from a cooling tower, may be necessary to cool the desiccant solution before it enters the absorber. The temperature of the solution at

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the inlet of the absorber strongly influences the air-dehumidifying capacity. The higher the concentration and lower the temperature of the solution, the higher the moisture absorbed and lower the humidity of the exiting air from the absorber.

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Types of Desiccant Solutions. Desiccant solutions used by different investigators, as well as their key features, are shown in Table 2. Lithium chloride (LiCl) is a good candidate material because it has good desiccant characteristics and does not vaporize in air at ambient temperature, but it is corrosive. Calcium chloride (CaCl2) is an equally good candidate for desiccant choice and is less expensive than LiCl. A 50%/50% combination of LiCl and CaCl2 is also being investigated. Triethylene glycol (TEG) has a low vapor pressure that enables effective dehumidification in the absorber, but it also makes it vulnerable to evaporation. Special coatings may be required to avoid corrosion, and efficient absorber design may be needed to prevent liquid or desiccant carryover. Fluid Flow Configuration. Four different types of fluid flow configurations are used by different investigators in the packed beds for absorbers and regenerators. • Counter flow • Parallel flow Table 2 Types of Liquid-Desiccant Solutions Desiccant Typical Concentration Type Range, % LiCl TEG LiBr CaCl2 KCOOH

34 to 43 92 to 98 42.6 to 57 40 to 43 72.8 to 74

Source: Jain and Bansal (2007).

Fig. 19

Operating Temperatures, °C 25 to 27 20.2 to 45.5 20.1 to 34.1 27.0 21.9 to 24.8

• Cross flow • Falling film Most applications are counter flow, because it provides the maximum enthalpy and humidity effectiveness for the packed beds. Types of Packing. Packing arrangements are basically either structured or random types. In some cases, vertical surfaces or tubes are used, over which the desiccant solution falls as a thin film. The purpose of packing is to provide the maximum surface area for the desiccant solution to come in contact with the process air. The greater the area, the greater the dehumidification. The main parameters affecting bed arrangement are the enhancement of heat and mass transfer, air pressure drop, and pumping power to circulate the fluids through the bed. The main advantage of structured packing is the low pressure drop, whereas that of random packing is high heat and mass transfer but with high pressure drop. Packing compactness is typically expressed in terms of packing density at, in units of m2/m3. Typically, packing densities exceeding 200 m2/m3 have been achieved in the field. A higher packing density allows for more dehumidification but also produces greater pressure drop for the fluid streams. In random packing arrangements, packing is made up of Berl saddles, Raschig rings, and saddles of various trademarked designs, whereas in structured packing, the packing consists of cellulose rigid media pads, wood grids, expanded metal lash packing, or double spiral rings.

Twin-Tower Enthalpy Recovery Loops One method of ventilation air enthalpy recovery is liquid desiccant twin-tower systems. In this air-to-liquid, liquid-to-air enthalpy recovery system, sorbent liquid circulates continuously between supply and exhaust airstreams, alternately contacting both airstreams directly in contactor towers (Figure 20). This liquid transports water vapor and heat. The sorbent solution is usually a halogen salt solution such as lithium chloride, calcium chloride, and water.

Typical Dehumidification of Outdoor Air by Liquid-Desiccant System (Williams 2007)

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Air-to-Air Energy Recovery Equipment

Fig. 20 Twin-Tower System Performing Indirect Evaporative Air Cooling: (A) Schematic and (B) Installation in Building Pumps circulate the solution between supply and exhaust contactor towers. Both vertical and horizontal airflow contactor towers are available, and both types can be combined in a common system. Contactor towers of both configurations are commonly available with airflow capacities of up to 40 m3/s. In the vertical configuration, the supply or exhaust air passes vertically through the contact surface counterflow to the sorbent liquid to achieve high contact efficiency. In the horizontal configuration, air passes horizontally through the contact surface cross flow to the sorbent liquid, yielding a slightly lower contact efficiency. The contactor surface is usually made of nonmetallic materials. Air leaving the contactor tower passes through demister pads, which remove any entrained sorbent solution. The halogen salt solution does not offgas or vaporize. Design Considerations. Twin-tower enthalpy recovery systems operate primarily in the comfort conditioning range; they are not suitable for high-temperature applications. During summer, these systems operate efficiently with building supply outdoor air temperatures as high as 46°C. Winter supply air temperatures as low as –40°C can be tolerated without freezing or frosting problems because the sorbent solution is an effective antifreeze at all useful concentrations. During weather conditions that are not advantageous for energy recovery, the system automatically shuts down the pump(s), thereby saving operating power. When outdoor conditions warrant energy recovery, the system restarts automatically. Because contactor supply and exhaust air towers are independent units connected only by piping, supply and exhaust air fans can be located

26.15 wherever desired. Contactor towers can operate with air inlet static pressure from –1.5 to 1.5 kPa, but require less than 0.4 kPa air-side pressure drop. The exhaust contactor tower may operate at a higher internal static pressure than the supply contactor tower without danger of exhaust-to-supply cross contamination. Particulate cross contamination does not occur because wetted particulates remain in the sorbent solution until they are filtered. Limited gaseous cross contamination may occur, depending on the solubility of the gas in the sorbent solution. Sorbent solutions, especially halide brines, are bactericidal. Lithium or calcium chloride as used in twin-tower systems is viricidal against all viruses. If the building or process exhaust contains large amounts of lint, animal hair, or other solids, filters should be placed upstream of the exhaust air contactor tower; if it contains large amounts of gaseous contaminants, such as chemical fumes and hydrocarbons, investigate the possibility of cross contamination as well as the possible effects of contaminants on the sorbent solution. When using twintower systems in controlled-humidity applications in colder climates, saturation effects (which can cause condensation, frosting, and icing in other types of equipment) may overdilute the twin towers’ sorbent solution. Heating the sorbent liquid supplied to the supply air contactor tower can prevent dilution, as shown in Figure 20A. Heating raises the discharge temperature and humidity of air leaving the supply contactor tower, thus balancing the humidity of the system and preventing overdilution. A thermostat sensing the leaving air temperature from the supply air contactor tower is commonly used to control the solution heater to deliver constant-temperature air regardless of outdoor temperature. Automatic addition of makeup water to maintain a fixed concentration of the sorbent solution enables the twin-tower system to deliver supply air at fixed humidity during cold weather. Thus, the system provides a constant supply air temperature and humidity without preheat or reheat coils or humidifiers. Any number of supply air towers can be combined with any number of exhaust air towers, as shown in Figure 20B. If a sufficient elevation difference exists between supply and exhaust towers, gravity may be used to return the sorbent solution from the upper tower(s) in lieu of pumped flow. Twin-tower energy recovery systems can be retrofitted to existing buildings. The towers are connected to each other with piping. The solution piping can be easily installed inside or outside the building. Twin-tower enthalpy recovery systems operate with only routine maintenance. Complete instructions on procedures as well as spare-parts lists are included in operating manuals relevant to each installation. Periodically, the circulation pumps, spray nozzles, liquid transfer controls, and mist eliminators need checking and adjusting or maintenance. Halide brine solutions are normally used as energy transfer media in twin-tower systems. Manufacturers provide technical support, including solution sampling services to monitor and report changes in concentration and pH, thus ensuring continued maximum performance.

Fixed-Bed Regenerators Fixed-bed regenerators (FBRs) are available in many configurations, sizes, and airflow patterns. In general, FBRs contain one or two cores for the purpose of energy transfer, and some means of reversing airflow through its core(s) to alternately store and release energy. In North America, double-core FBRs are more common; singlecore FBRs are available but have not been widely adopted. Doublecore FBRs are packaged with two cores and an airflow control module to alternate supply and exhaust air through each core. These units are used in large residential, commercial, and industrial applications with airflow capacity in the range of 0.025 to 50 m3/s. Cost and physical size are the main barriers between double-core FBRs and the residential home market.

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Fig. 23

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Fig. 22 Typical Temperature Profile of Fixed-Bed Regenerator Operating at 60 s Recovery Periods Unlike other energy recovery exchangers, in FBRs the outlet air temperatures and humidities are never steady state; instead, they rise and fall cyclically with each reversal of airflow. This poses challenges to test facilities, which generally seek to take measurement only when the device under test is operating in steady state with very small variations in outlet conditions. The latest edition of ASHRAE Standard 84 (in progress) introduces techniques for data reduction and humidity averaging to deal with this challenge. As more manufacturers enter the market, certified performance data may become available. Arrangement. Cores are typically split into smaller cells to facilitate transportation, handling, and cleaning. Plates are formed with spacers (e.g., ribs or dimples) constructed into the plates or by alternating corrugation angle between adjacent plates (direction of the angle spaces plates apart). Corrugation patterns and plate thickness varies depending on manufacturer. Plate spacing ranges from 2.5 to 13 mm. Aluminum cores are the most popular construction, but some manufacturers also offer polypropylene cores for corrosive or lightweight applications. Airflow control modules consist of a single deflector plate or a multiblade damper system. Early designs included pneumatic actuators controlled by solenoids and flow valves. Electric actuators were introduced to improve synchronization and to reduce switchover time of multiblade dampers to 0.5 to 1.5 s. Rapid

Single-Core Fixed-Bed Regenerator

switchover is of key importance to minimize exhaust air transfer between recovery periods. In Europe, various types of single-core devices are used for residential, apartment, condo, or renovation projects with limited ceiling space. Often, the only option is a room-based decentralized ductless ventilation system, which consists of several single-core FBRs installed in the exterior walls of a space. The individual devices are synchronized via interconnecting wiring or wireless communication to alternately store and release energy from a single honeycomb-shaped, ceramic core. Operation. Typically, the devices are staggered so half of the ventilation system is exhausting while the other half is bringing tempered outside air into the space. After a certain period of time, called the recovery period, the system switches direction to recharge the devices that have been in the outside airstream. The other half of the system now transfers energy to the outside air stream until the next recovery period. Many single-core FBRs are equipped with a relatively small fan, capable of changing its direction at the end of each recovery period to supply air to, or exhaust air from, the space. Depending on the manufacturer, the device is designed with two opposite fans in adjacent channels, airflow deflectors, and/or backdraft flaps to alternate exhaust and supply air through the core. Performance curves of the fan must be considered for high-rise buildings and locations with high winds. Stack effect and wind direction can be detrimental to supply airflow rate or effectiveness of single-core FBRs. Recovery periods in the range of 50 to 70 s are common to both single-core and double-core FBRs. Manufacturers report sensible

Fig. 21 Double-Core Fixed-Bed Regenerators

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effectiveness typically in the range of 80 to 90%. Latent recovery is also possible. With sensible-only FBRs, latent recovery can occur in condensing conditions (i.e., heating mode). FBRs also can be manufactured with desiccant treatment providing latent recovery in all conditions. Manufacturers report latent effectiveness in desiccanttreated FBRs in the 60 to 80% range, and in non-desiccant-treated FBRs, as high as 70% (depending on conditions). Longer recovery periods up to 3 h are used for free cooling mode. Some manufacturers vary the recovery period between 20 to 120 s to control discharge air temperature and/or humidity. Shorter recovery periods improve effectiveness, but the drawback is a higher exhaust air transfer ratio. Exhaust air transfer for outdoor double-core FBRs is reported by manufacturers to be in the range of 3 to 5%. Indoor applications with long ducts connecting the unit to the ambient suffer from higher exhaust transfer rates due to the volume of exhaust air drawn back from these ducts. Generally, duct length is kept under 8 m to limit increase in pressure drop and exhaust air transfer. An advantage of FBRs is the repeated exhaust regeneration, which allows this technology to operate in extremely cold climates below –40°C without the need of auxiliary defrost strategies. Typically, discharge air from a double-core FBR is condition by a heater, or an interlocked air handling unit before entering the space.

4.

PERFORMANCE RATINGS

Performance ratings should be based on tests performed in qualified laboratories that are staffed and instrumented to meet requirements of ASHRAE Standard 84 (in progress) and AHRI Standard 1061.It may be very difficult to adhere to any standard when field tests are made.

Performance Ratings for Air-to-Air Heat or Heat and Mass Exchangers ASHRAE Standard 84 (1) establishes a uniform method of testing for obtaining performance data; (2) specifies the data required, calculations to be used, and reporting procedures for testing each of seven independent performance factors and their uncertainty limits; and (3) guides selection of test equipment. The independent performance factors are sensible, latent, and total effectivenesses es, el , and et ; the enthalpy recovery ratio; supply and exhaust air pressure drops Ps and Pe; exhaust air transfer ratio (EATR), which characterizes the fraction of exhaust air transferred to the supply air; and outdoor air correction factor (OACF), which is the ratio of supply inlet to outlet airflow. An additional useful parameter, which can be evaluated only when fan efficiency is known or assumed, is the recovery efficiency ratio. AHRI Standard 1061 is an industry-established standard for rating air-to-air heat/energy exchanger performance for use in energy recovery ventilation equipment. This standard, based on ASHRAE Standard 84, establishes definitions, requirements for marking and nameplate data, and conformance conditions intended for the industry, including manufacturers, engineers, installers, contractors, and users. Standard temperature and humidity conditions at which equipment tests are to be conducted are specified for summer and winter conditions. Published ratings must be reported for the performance factors specified in ASHRAE Standard 84. The AHRI certification program using Standard 1061 is used to verify ratings published by manufacturers. At this time the scope of the certification program includes plate, rotary, and heat pipe exchangers. Coil energy recovery systems, twin-tower enthalpy recovery loops, and fixedbed regenerators are not in the scope of the certification program. Before 2020, AHRI Standard 1060 required the publication of ratings based on performance of the rated exchanger at two balanced airflow rates selected by the manufacturer, at standard winter and summer conditions:

26.17 Winter:

Outdoor air at t1 = 1.7°C and tw1 = 0.6°C Indoor (room) air at t3 = 21.1°C and tw3 = 14.4°C

and p2 – p3 = 0 Summer: Outdoor air at t1 = 35°C and tw1 = 25.6°C Indoor (room) air at t3 = 23.9°C and tw3 = 17.2°C and p2 – p3 = 0 These ratings applied to operation with the static pressure at supply air outlet equal to the static pressure at the exhaust air inlet. Ratings for the leakage metrics EATR and OACF also were to be provided at these static pressure conditions, as well as at two additional static pressure conditions selected by the manufacturer. As of 2020, the AHRI certification program verifies ratings generated by a manufacturer’s prediction software across a broad psychrometric range, at any combination of airflows and static pressures supported by the software. The program also verifies that all of the ratings required for a complete characterization of exchanger performance are provided. Some metrics can be provided optionally. The new energy recovery ratio (Equation [3]) is used to characterize the performance of the exchanger in relation to its use in a system. This metric is important because, as of 2019, ASHRAE Standard 90.1 requires that energy recovery systems have an enthalpy recovery ratio of at least 50% (i.e., a change in enthalpy of outdoor air supply equal to 50% of the difference between outdoor air and entering exhaust air enthalpies at design conditions). Standard 90.1 previously used 50% effectiveness at design conditions, but this metric measures performance of the exchanger and is less relevant when defining the performance of an enthalpy recovery system. It is more difficult to measure performance of a heat or heat/mass exchanger when condensation occurs. ASHRAE Standard 84 provides theoretical approaches, but in practice it is challenging to measure the rate at which condensate is generated, which is required to confirm the test meets steady-state requirements. CSA Standard C439-2018 provides a method to measure some energy and ventilation parameters of a complete ventilating unit equipped with heat or heat and mass recovery and active frost prevention or recovery means. However, this is intended for use with packaged ventilating units, and is not intended to characterize the conditions under which frost occurs in the exchanger itself.

Performance Ratings for Residential Ventilators with Air-to-Air Heat or Heat and Mass Exchangers Residential ventilation products are certified by the Home Ventilating Institute (HVI 2015). Heat recovery ventilators and energy recovery ventilators (H/ERVs) are certified for their ventilation performance at high speed and for their energy recovery performance at manufacturer’s selected airflows. The HVI Certified Products Directory (www.hvi.org/proddirectory) is useful for comparing product performance on a level playing field. H/ERVs must be tested at prescribed conditions in accordance with CSA Standard C439 at an HVI-designated laboratory to be eligible for HVI certification. HVI issues certified ratings for H/ERVs for airflow (net supply, gross exhaust, gross supply) and for energy recovery performance parameters include the following: • • • •

Sensible recovery efficiency (SRE) Adjusted sensible recovery efficiency (ASRE) Total sensible recovery efficiency (TRE) in cooling modes Adjusted total sensible recovery efficiency (ATRE) in cooling modes • Latent recovery moisture transfer • Very-low-temperature ventilation reductions

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2020 ASHRAE Handbook—HVAC Systems and Equipment(SI) cavities in the heat transfer medium is carried into the fresh airstream by the wheel’s rotation. The test estimate of the air leakage is characterized by two dimensionless parameters: the exhaust air transfer ratio (EATR) and outdoor air correction factor (OACF). During these tests, m2 = m3, while an inert tracer gas is introduced at station 3 (Shang et al. 2001a): c –c c3 – c1

2 1 EATR = ---------------

Fig. 24 Air Leakage in Energy Recovery Units Note that ratings for ASRE are to be used in energy-modeling software where fan power is a separate modeling input, to avoid counting fan power twice. Residential H/ERVs are certified as a packaged unit, whereas larger commercial H/ERVs typically carry certified ratings for energy recovery performance of just the core. Products must be listed in the online directory to be considered HVI certified (i.e., not listed means not certified).

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5.

ADDITIONAL TECHNICAL CONSIDERATIONS

The rated effectiveness of energy recovery units is obtained under balanced flow conditions (i.e., the flow rates of supply and exhaust airstreams are the same). However, these ideal conditions do not always exist, because of design for positive building pressure, air leakage, fouling, condensation, or frosting, and the following other factors.

Air Leakage Air leakage refers to any air that enters or leaves the supply or exhaust airstreams. Zero air leakage in either airstream would require m1 to equal m2 and m3 to equal m4. External air leakage occurs when the ambient air surrounding the energy recovery system leaks into (or exits) either or both airstreams. Internal air leakage occurs when holes or passages are open to the other airstream. Internal air leakage occurs when heat or energy exchanger design allows (1) tangential air movement in the wheel’s rotational direction and (2) air movement through holes in the barrier between airstreams. Under some pressure differentials, air leaks in and out of each airstream in nearly equal amounts, giving the illusion that there is no air leakage. Heat and water vapor transfer could appear to be greater than it is in reality. Leakage varies with heat exchanger type and design, airstream static pressure differences, and the physical condition of the heat exchanger (see Table 3). Air leakage is seldom zero because external and internal air pressures are usually different, causing air to leak from areas of higher pressure to those of lower pressure. Cross-leakage, cross-contamination, or mixing between supply and exhaust airstreams may occur in air-to-air heat exchangers and may be a significant problem if exhaust gases are toxic or odorous. Air leakage between incoming fresh air and outgoing exhaust air can be classified into two mechanisms: cross-flow and carryover. Cross-flow leakage is caused primarily by difference in static pressures between states 2 and 3 and/or between states 1 and 4, as shown in Figure 24. This is a major cause of cross-flow leakage, and underscores the importance of specifying precise locations for fans that circulate the airstreams. Cross-flow can also be caused by factors such as provisions for surging, geometrical irregularities, and local velocity distribution of the airstreams. Carryover occurs in rotary recovery units because of wheel rotation out of one airstream into the other. Exhaust air trapped in

(15)

where c1, c2, and c3 are the concentrations of inert gas at states 1, 2, and 3, respectively. EATR is the ratio of the concentration increase in supply air relative to the maximum concentration difference between stations 3 and 1. It is representative of the air mass leakage from exhaust air into supply air, when the air-to-air heat exchanger is operating at mass flow rates and static pressures equal to those at which the device was tested. m1 OACF= -----m2

(16)

where m1 and m2 are the mass flow rates of incoming fresh airstream at states 1 and 2, respectively. OACF gives the ratio of outdoor airflow required at the inlet (station 1) relative to the supply flow at 2 to compensate for air that leaks into or out of the exchanger, to meet the required net supply airflow to the building space, when the air-to-air heat exchanger is operating at mass flow rates and static pressures equal to those at which the device was tested. Ideal operating conditions exist when there is no air leakage between the streams, EATR is close to zero, and OACF approaches 1. Deviations from ideal conditions indicate air leakage between the airstreams, which complicates the determination of accurate values for pressure drop and effectiveness. Methods to estimate actual flow rates at states 1, 2, 3, and 4 when air leakage exists and the values of the parameters EATR and OACF are known are discussed in Friedlander (2003) and Moffitt (2003). EATR and OACF must be determined when evaluating any HRV or ERV in which leakage between airstreams occurs, when evaluating the actual flow rate of outdoor air supplied for a given ventilation requirement, and when estimating the capacity of ventilator fans, as shown by the following.

Air Capacity of Ventilator Fans For a given ventilation requirement Qv, the density-corrected volume flow rate capacity Qfan of the supply fan is greater than Q2 if air leakage exists (see Figure 24). In the following calculations, be sure to obtain the value for EATR that applies to the device when it operates at the airflow volumes and static pressure that will exist in the specific application. The leakage flow rate from the exhaust to the supply air Q3-2 can be estimated by the equation. EATR Q 3 – Q 2 = Q 2 --------------100

(17)

If the ventilation requirement is Qv, then the actual volume flow rate of supply air to the space is calculated as EATR Q 2 = Q v + Q 3 – 2 = Q v + Q 2 --------------100 This can be simplified as

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(18)

Air-to-Air Energy Recovery Equipment Qv Q 2 = -----------------------------------------1 –  EATR  100 

26.19 Controls (19)

which gives the quantity of air entering the space. Assuming steadystate conditions, and balanced flow through the energy recovery ventilator, the density corrected quantity of air leaving the building space (Q3) should be same as air supplied to the space Q2. Qv Q 3 = Q 2 = -----------------------------------------1 –  EATR  100 

(20)

Combining Equations (16) and (20) gives the required supply air inlet volume q1 as

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Q v  OACF  Q 1 = Q 2  OACF  = -----------------------------------------1 –  EATR  100 

(21)

Caution should be used with Equations (20) and (21) because in many practical applications, Q2 is not equal to Q3. Equation (21) gives the volume flow rate capacity of the intake, which equals the exhaust outlet for balanced airflow. (Fan capacity depends on location. This statement only applies to the supply fan at station 1 [blow-through]; its capacity would change at station 2 [draw-through].)

Pressure Drop Pressure drop for each airstream through an energy recovery unit depends on many factors, including exchanger design, mass flow rate, temperature, moisture, and inlet and outlet air connections. The pressure drop must be overcome by fans or blowers. Because the power required for circulating airstreams through the recovery unit is directly proportional to the pressure drop, the pressure drop through the energy recovery unit at its operating conditions should be known. The pressure drop may be used with the fan efficiency to characterize the energy used by the exchanger and in turn the efficiency (not effectiveness) of an application.

Maintenance The method used to clean a heat exchanger depends on the transfer medium or mechanism used in the energy recovery unit and on the nature of the material to be removed. Grease build-up from kitchen exhaust, for example, is often removed with an automatic water-wash system. Other kinds of dirt may be removed by vacuuming, blowing compressed air through the passages, steam cleaning, manual spray cleaning, soaking the units in soapy water or solvents, or using soot blowers. The cleaning method should be determined during design so that a compatible heat exchanger can be selected. Cleaning frequency depends on the quality of the exhaust airstream. Residential and commercial HVAC systems generally require only infrequent cleaning; industrial systems, usually more. Consult equipment suppliers about the specific cleaning and maintenance requirements of systems being considered.

Filtration Filters are recommended and should be placed in both the supply and exhaust airstreams to reduce fouling and thus the frequency of cleaning. Exhaust filters are especially important if the contaminants are sticky or greasy or if particulates can plug airflow passages in the exchanger. Supply filters eliminate insects, leaves, and other foreign materials, thus protecting both the heat exchanger and airconditioning equipment. Snow or frost can block the air supply filter and cause severe problems. Specify steps to ensure a continuous flow of supply air.

Heat exchanger controls may control frost formation or regulate the amount of energy transferred between airstreams at specified operating conditions. For example, ventilation systems designed to maintain specific indoor conditions at extreme outdoor design conditions may require energy recovery modulation to provide an economizer function, to prevent overheating ventilation supply air during cool to moderate weather or to prevent overhumidification. Modulation methods include tilting heat pipes, changing rotational speeds of (or stopping) heat wheels, or bypassing part of one airstream around the heat exchanger using dampers (i.e., changing the supply-toexhaust mass airflow ratio).

Fouling Fouling, an accumulation of dust or condensates on heat exchanger surfaces, reduces heat exchanger performance by increasing resistance to airflow, interfering with mass transfer, and generally decreasing heat transfer coefficients. Increased resistance to airflow increases fan power requirements and may reduce airflow. Increased pressure drop across the heat exchanger core can indicate fouling and, with experience, may be used to establish cleaning schedules. Reduced mass transfer performance (latent effectiveness) indicates fouling of permeable membranes or desiccant sorption sites. Heat exchanger surfaces must be kept clean to maximize system performance.

Corrosion Process exhaust frequently contains corrosive substances. If it is not known which construction materials are most corrosion resistant for an application, the user and/or designer should examine on-site ductwork, review literature, and contact equipment suppliers before selecting materials. A corrosion study of heat exchanger construction materials in the proposed operating environment may be warranted if installation costs are high and the environment is corrosive. Experimental procedures for such studies are described in an ASHRAE symposium (ASHRAE 1982). Often contaminants not directly related to the process are present in the exhaust airstream (e.g., welding fumes or paint carryover from adjacent processes). Moderate corrosion generally occurs over time, roughening metal surfaces and increasing their heat transfer coefficients. Severe corrosion reduces overall heat transfer and can cause cross-leakage between airstreams because of perforation or mechanical failure.

Condensation and Freeze-Up Condensation, ice formation, and/or frosting may occur on heat exchange surfaces. Ignoring entrance and exit effects, four distinct air/moisture regimes may occur as the warm airstream cools from its inlet condition to its outlet condition. First, there is a dry region with no condensate. Once the warm airstream cools below its dew point, there is a condensing region, which wets the heat exchange surfaces. If the heat exchange surfaces fall below freezing, the condensation freezes. Finally, if the warm airstream temperature falls below its dew point and is also below freezing temperature, sublimation causes frost to form. The locations of these regions and rates of condensation and frosting depend on the duration of frosting conditions; airflow rates; inlet air temperature and humidity; heat exchanger core temperature; heat exchanger effectiveness; geometry, configuration, and orientation; and heat transfer coefficients. Sensible heat exchangers, which are ideally suited to applications in which heat transfer is desired but humidity transfer is not (e.g., swimming pools, kitchens, drying ovens), can benefit from the latent heat released by the exhaust gas when condensation

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2020 ASHRAE Handbook—HVAC Systems and Equipment(SI)

occurs. One kilogram of moisture condensed transfers about 2440 kJ to the incoming air at room temperature. Equipment design should provide for continuous draining of the condensate from the unit. Transfer of latent heat occurs at the point of condensation, and little available energy remains in the condensate. Condensation increases the heat transfer rate and thus the sensible effectiveness; it can also significantly increase pressure drops in heat exchangers with narrow airflow passage spacing. Frosting fouls the heat exchanger surfaces, which initially improves energy transfer but subsequently restricts the exhaust airflow, which in turn reduces the energy transfer rate. In extreme cases, the exhaust airflow (and supply, in the case of heat wheels) can become blocked. Defrosting a fully blocked heat exchanger requires that the unit be shut down for an extended period. As water cools and forms ice, it expands, which may seriously damage the heat exchanger core unless the water is entirely removed. For frosting or icing to occur, an airstream must be cooled to a dew point below 0°C. Enthalpy exchangers transfer moisture from the airstream with higher moisture content (usually the warmer airstream) to the less humid one. As a result, frosting or icing occurs at lower supply air temperatures in enthalpy exchangers than in sensible heat exchangers. Cross-flow heat exchangers are more frost tolerant than other plate exchangers because frost blockage only occurs over a fraction of the exchanger exhaust airflow passages. For these reasons, some form of freeze control must be incorporated into heat exchangers that are expected to operate under freezing conditions. Frosting and icing can be prevented by preheating supply air or reducing heat exchanger effectiveness (e.g., reducing heat wheel speed, tilting heat pipes, bypassing some supply air around the heat exchanger). Alternatively, the heat exchanger may be periodically defrosted. Manufacturers of commercially available packaged heat or energy ventilators routinely incorporate frosting control strategies or mechanisms appropriate to their equipment and its targeted application. ASHRAE research project RP-543 discusses performance of several freeze control strategies (Phillips et al. 1989a, 1989b). ASHRAE research project RP-544 (Barringer and McGugan 1989a, 1989b) covers performance of enthalpy heat exchangers. Many effective defrost strategies have been developed for residential air-to-air heat exchangers, and may also be applied to commercial installations. Phillips et al. (1992) describe frost control strategies and their impact on energy performance in various climates. If cyclic frost methods are used, the defrost period should be long enough to completely remove the frost, ice, and water during each defrost period. Total heat exchangers transfer moisture from the more humid airstream to the less humid one. In winter, this results in an exhaust stream that is partially dehumidified as it is cooled. As a result, frosting or icing generally occurs at lower temperatures than in sensible-only heat exchangers. On the other hand, sensible heat exchangers can benefit from the latent heat released in a highhumid-exhaust airstream when condensation occurs. See the Bibliography for sources of more information on frost growth and control.

Frost Control Strategies for Air-to-Air Energy Recovery Systems In all cases, the potential for frost is in the warm-return-air side of the heat exchanger, which cools during a winter condition. For energy recovery devices, the exact outdoor air temperatures at which frosting will start depends on the temperature and humidity of the exhaust air and the type and performance of the energy recovery device. This temperature is known as the frost threshold tem-

perature. The frost threshold varies by specific technology, and significantly by whether water vapor is transferred between airstreams: • Flat Plate Heat Exchangers. In flat plate heat exchangers (Figures 2 and 4), warm return air cools as it passes through the heat exchanger. Moisture from the warm air condenses and creates water droplets on the media; as cool outdoor air enters the heat exchanger, the water droplets start to freeze, creating ice on the cold corner of the core. Figure 1 shows a typical temperature distribution of warm and humid; and cold and dry air streams. The circled area shows the potential problem area for frost formation. In general, frost buildup becomes a problem when outdoor air temperatures reach about –5°C for sensible devices, or approximately –20°C for total energy devices. Figure 3 shows frost threshold for energy (Afshin 2019) and heat recovery ventilators. • Energy Recovery Wheels. In energy recovery wheels (Figure 6), warm return air cools as it passes through the rotating wheel. Moisture from the warm air condenses and creates water droplets on the surface of the wheel. The water molecules are then absorbed by the desiccant on the enthalpy wheel. As the wheel rotates through the cold supply air, the water can start to freeze and form ice before the water droplets can be cleared from the wheel surface. Ice typically forms on the internal surface of the wheel, near the outdoor air intake which can cause damage and reduce airflow and performance. • Heat Pipe, Thermosiphon Loop, and Runaround Glycol Heat Exchangers. In these technologies, the return air is cooled by a fluid passing through a coil. (Figures 10 and 15) For frost to form, the coil fin surface must be freezing and below the return air dew point. For frost to build up, the return air must be cooled below freezing. Using the sensible effectiveness, the user can calculate the outdoor air temperature where frost would be expected. For example, with a sensible effectiveness of 50%, an outdoor air temperature of –21°C would cool 21°C return air to 0°C. There are a variety of different approaches that can be used to determine the frost threshold temperature, including measuring temperature, humidity, or a variation in pressure drop. A frost control strategy is then determined to either eliminate the frost once it builds up on the exchanger (before it becomes too severe), or to prevent frost formation from ever starting at all. The choice between frost control and frost prevention will depend on the robustness of the airto-air energy recovery device and the application. An effective frost control strategy is one that meets the requirements of the application while being cost-effective and device appropriate. Figure 25 shows the possible control strategies (Afshin 2019) for controlling frost for plate heat exchangers used as ERVs. In all cases of frost prevention, the amount of energy recovery is reduced and there will need to be additional heat added into the system. Energy efficiencies for each solution vary somewhat, and the best choice depends on application and the amount of upfront cost that can be tolerated. Frost Prevention. • Preheat coil to heat up exhaust air. Preheat frost control is a preventative strategy that can be used with any air-to-air energy recovery device. The objective is to prevent frost from occurring within the heat exchanger, while maintaining 100% or continuous ventilation. Energy recovery is reduced, because the difference in temperature between the preheated outside air and the return air has reduced. Heating coils (electric, steam, or hot water) are ductmounted or integrated into the unit in the outdoor airstream so that the entering outdoor air temperature is preconditioned to a temperature above the frost threshold for the technology being used. While preheat typically has higher upfront costs than other strategies, it can result in significant operating savings in climates where

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Air-to-Air Energy Recovery Equipment frost control is required for a long period of time (more than a few hours). The preheater should be sized for the coldest outdoor air design temperature and to the appropriate outdoor air temperature where the exhaust air frost threshold is reached based on the design conditions and airflows for the type of heat exchanger selected. Figure 3 shows frost threshold characteristics for energy and heat recovery ventilators. • Face and bypass damper (Figure 25D). Face and bypass frost control is a preventative strategy that can be used for flat plates, heat pipes, and rotary wheel exchangers, with the objective of preventing frost formation while maintaining 100% ventilation. As the outdoor air becomes colder, face and bypass dampers upstream of the heat exchanger modulate to reduce the amount of outdoor air flowing through the heat exchanger. This reduces the amount of energy recovered and keeps the exhaust temperature above the frost threshold. The supply and exhaust fans, as well as the outdoor air and exhaust air dampers, continue to operate during the frost control cycle. With this strategy there is no interrupted ventilation which makes it ideal for harsher environments or source control applications like laboratories. Furthermore, as both fans run continuously there is no depressurization of the building, eliminating the potential for combustion appliance backdraft into the occupied space. The lower leaving supply air temperature would potentially require post-conditioning, or some type of terminal reheat within the space to ensure that occupants remain comfortable under extreme conditions. A further advantage would be that the faceand-bypass can also be used for free cooling (economizer) and/or supply air temperature control in different modes of operation if required. • Performance modulation. The heat exchanger can be controlled to ensure that frost does not form in the return air, by reducing the effectiveness and, as a result, suppressing the frost threshold. For example, wheel rotational speed can be reduced, runaround pumps can be slowed, heat pipes can be equipped with control valves or tilted. Since the heat exchanger is recovering less energy, more heating is required in the process downstream. Defrosting. 1. Exhaust defrosting (Figure 25B). Exhaust-only defrost is one of the most cost-effective and simple strategies to implement. It periodically defrosts ice forming on the heat exchanger by shutting down the supply fan to remove the source of cold air, while using the warm exhaust air to heat up the exchanger. When the unit goes into a defrost cycle, the exhaust fan continues to operate, the supply fan is deenergized and the outdoor air damper closes. This method is most commonly used with flat plate or heat pipe heat exchangers and is ideal for source control applications where continuous exhaust is required. One drawback of this method is that ventilation is interrupted when the supply fan shuts down during the defrost cycle, which may not be acceptable as the equipment may not meet indoor air quality (IAQ) requirements as defined by ASHRAE Standard 62.1. This also creates negative indoor pressure, resulting in infiltration. 2. Recirculation defrost (Figure 25B). Recirculation defrost is also cost-effective and simple, commonly used in light commercial stand-alone air-to-air ERVs that are not used as a primary ventilation system. Ice formation on the heat exchanger is periodically defrosted by shutting down the exhaust fan, by closing the outdoor and exhaust air dampers, and by opening a recirculation air damper to remove the source of cold air. When the unit goes into a defrost cycle, the supply fan remains on to recirculate building exhaust air back into the occupied space. The exhaust air goes through the heat exchanger and provides defrosting in the absence of cold outdoor air. A drawback of this method is that ventilation is interrupted when the exhaust fan shuts down during the defrost cycle. This may not be acceptable in all applications

26.21

Fig. 25

Common Frost Control Strategies for Fixed-Plate Exchangers

and may not meet indoor air quality (IAQ) requirements as set out in ASHRAE Standard 62.1. It may however be acceptable in more moderate climates where freezing conditions occur only during unoccupied hours for a few hours per year.

Indirect Evaporative Air Cooling In indirect evaporative air cooling, exhaust air passing through a water spray absorbs water vapor until it becomes nearly saturated. As the water evaporates, it absorbs sensible energy from the air, lowering its temperature. The evaporatively cooled exhaust air can then be used to cool supply air through an air-to-air heat exchanger, which may be used either for year-round energy recovery or exclusively for its evaporative cooling benefits. This process follows a constant wetbulb line on a psychrometric chart: the airstream enthalpy remains nearly constant, moisture content increases, and dry-bulb temperature decreases. Indirect evaporative cooling has been used with heat pipe heat exchangers, runaround coil loop exchangers, and flat-plate heat exchangers for summer cooling (Dhital et al. 1995; Johnson et al. 1995; Mathur 1990a, 1990b, 1992, 1993; Scofield and Taylor 1986). Exhaust air or a scavenging airstream is cooled by passing it through a water spray, wet filter, or other wetted media, resulting in a greater overall temperature difference between the supply and exhaust or scavenging airstreams and thus more heat transfer. Energy recovery is further enhanced by improved heat transfer coefficients because of wetted exhaust-side heat transfer surfaces. No moisture is added to the supply airstream, and there are no auxiliary energy inputs other than fan and water pumping power. The coefficient of performance (COP) tends to be high, typically from 9 to 20, depending on available dry-bulb temperature depression. The dry-bulb temperature decrease in the exhaust airstream caused by evaporative cooling tends to be 85 to 95% of the maximum available difference between the exhaust air inlet dry-bulb and wet-bulb temperatures. Therefore, exhaust air evaporative cooling is usually most cost effective in hot, dry climates where the evaporator can be used frequently to obtain large exhaust air dry-bulb temperature depressions. Without a bypass scheme for either the evaporator or air-to-air heat exchanger, the net annual energy costs include the extra annual fan power for these devices as well as the benefit of evaporative cooling. Because less mechanical cooling is required with evaporative cooling, energy consumption and peak demand load are both reduced, yielding lower energy bills. Overall mechanical refrigeration system requirements are reduced, allowing use of smaller mechanical refrigeration systems. In some cases, the mechanical system may be eliminated. Chapter 41 of this volume and Chapter 53 of the 2019 ASHRAE Handbook—HVAC Applications have further information on evaporative cooling.

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2020 ASHRAE Handbook—HVAC Systems and Equipment(SI)

Fig. 26 Examples of Frost Threshold Temperatures for Energy and Heat Recovery Exchangers (Afshin 2019) Example 4. Indirect Evaporative Cooling Recovery. Room air at 30°C and 17°C wb ( = 1.2 kgm3) with a flow rate of 15 m3/s is used to precool 15 m3/s of supply outdoor air at 39°C and 20°Cwb ( = 1.13 kg/ m3) using an aluminum fixed-plate heat exchanger and indirect evaporative cooling. The evaporative cooler increases the exhaust air to 90% rh before it enters the heat exchanger. The sensible effectiveness is given by the manufacturer as 78%. Assuming EATR = 0 and OACF  1, determine the leaving supply air conditions and energy recovered, and check the energy exchange balance. Solution: First, determine the exhaust air condition entering the exchanger (i.e., after it is adiabatically cooled). Air at 30°C db, 17°C wb cools to 18°C db, 17°C wb as shown by the process line from point A to point 3 in Figure 27. In this problem the volumetric flows are equal, but the mass flows are not. 1. Because data on pressure drop are missing, skip this step. 2. Calculate the theoretical maximum heat transfer. Based on a preliminary assessment, the supply air is not expected to cool below its wet-bulb temperature of 20°C. Thus, use the denominator of sensible heat effectiveness Equation (2a) and, for simplification, assume the specific heat of 1.0 kJ/(kg·K). qmax (sensible) = (1.13 kgm3)(15 m3/s)[1.0 kJ(kg·K)](39 – 18) = 356 kW 3. Establish the sensible effectiveness. From manufacturer’s exchanger selection data for indirect evaporative coolers, an effectiveness of 78% is found to be appropriate. 4. Calculate actual energy transfer at the design conditions. qactual = (0.78)(356 kW) = 278 kW recovered 5. Calculate leaving air conditions. a. Leaving supply air temperature is

Indirect Evaporative Cooling Recovery (Example 4)

278 kW t4 = 18°C + ---------------------------------------------------------------------------------------- = 33.4°C 3 3  1.2 kg/m   15 m /s   1 kJ/(kg·°C)  6. Check performance. qs = (1.13 kg/m3)(15 m3/s)[1 kJ/(kg·K)](39 – 22.6K) = 278 kW recovered qe = (1.2 kg/m3)(15 m3/s)[1 kJ/(kg·K)](33.4 – 18K) = 278 kW recovered 7. Plot conditions on psychrometric chart (Figure 27), and confirm that no latent exchange occurred. Because EATR = 0 and OACF is 1, Steps 8 to 10 are not presented here.

Use of Economizer

– 278 kW t2 = 39°C + ------------------------------------------------------------------------------------------ = 22.6°C 3 3  1.13 kg/m   15 m /s   1 kJ/(kg·K)  b. Leaving exhaust air temperature is

Fig. 27

ASHRAE Standard 90.1 specifies the use of economizer whenever the conditions of the outdoor air, especially enthalpy, are lower than that of the return air. Typically, the use of economizer depends on factors such as the characteristics of the building envelope, num-

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26.23

Fig. 28 Economizer with Dessicant Wheel

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(Pieper 2015)

ber of occupants, internal loads, and the location of the building. An economizer may be required for buildings in U.S. climate zones other than 1a,1b, 2a, and 2b, provided the cooling capacity is higher than 11 tons in the moist region and 5 tons for the dry regions. Exceptions might exist for other regions for just a few hours of the year, which may not justify the use of economizer for economic considerations. A schematic of the economizer along with energy recovery wheel as shown in Figure 28. If the outdoor temperature appears better than that of the return air (Pieper 2015), there would be a call for the economizer to engage, but if the enthalpy of the outdoor air could be worse than that of the return air, the economizer should not be in operation. Enthalpy comparison prevents excess moisture being brought into the space. It is also important to note that not all outdoor air passes through the energy recovery device, in order to prevent excessive fan power associated with recovery device and bypass dampers. Dampers shown in Figure 28 have a wide range of flow rates such that the system may provide 100% outdoor air or less flow rate in combination with the return air through return air damper. The controls may be programmed to supply the required supply air conditions such that the mass flow rate and the enthalpy of the supply air satisfy the following equation. ms  hr – hs  = qt

(22)

where ms hr hs qt

= = = =

mass flow rate of the supply air, kg/s enthalpy of the room air, kJ/kgDA enthalpy of the supply air, kJ/kgDA total building load at the time, kW

Building loads vary each hour, as does the sensible heat ratio (SHR), however, for the peak load conditions, the supply air conditions are determined based on the supply air temperature (Ts) as shown in Figure 29. The letters X and Y in Figure 29 indicate the states of supply and room air, respectively, and that the SHR line AO on the protractor is parallel to the line XY. In Figure 29, it is assumed that SHR = 0.7, however it changes by the hour. The challenge lies in the ability of the control system (shown in Figure 28) to modulate the flow rates of the outdoor air and the return air.

6.

COMPARISON OF AIR-TO-AIR HEAT OR HEAT AND MASS EXCHANGER CHARACTERISTICS

It is difficult to compare different types of air-to-air energy recovery systems based on overall performance. They can be compared based on certified ratings such as sensible, latent, and total

Fig. 29 Supply Conditions for a Given Load qt and Given SHR effectiveness or on air leakage parameters. Comparing them on payback period or maximum energy cost savings requires accurate values of their capital cost, life, and maintenance cost, which vary from product to product for the same type of recovery system. Without such data, and considering the data available in the open literature such as that presented by Besant and Simonson (2003), use Table 3’s comparative data for common types of air-to-air energy recovery devices.

7.

USE OF AIR-TO-AIR HEAT OR HEAT AND MASS EXCHANGERS IN SYSTEMS

Characterizing System Efficiency of Heat or Energy Recovery Ventilators A measure of energy recovery ventilator performance is the relative magnitude of actual energy recovered and power supplied to fans to circulate the airstreams. The cost of power supplied to the fans depends on the pressure drop of airstreams, volume flow rate, and combined efficiency of the fan motor systems. The quality of power supplied to the fans is high, and its cost per unit energy is much higher than the quality and cost of energy recovered in the ventilator. The magnitude and costs of these two forms of energy vary over the year. Besant and Simonson (2003) suggest that a parameter such as recovery efficiency ratio (RER) may be introduced to characterize the efficiency of the recovery ventilators:



 rate of energy recovered  dt RER = --------------------------------------------------------------------------------------------------- rate of power supplied to fan motors  dt



(23)

RER is similar to the energy efficiency ratio (EER) for chillers or unitary air-conditioning equipment. Besant and Simonson (2003) also suggest that the entire system performance, including the recovery ventilator, can be represented by the ratio of COP and RER. RER is typically expressed in units of kJ/Wh; energy recovered can be sensible, latent, or total, in kilowatts, and energy expended is that spent for circulating exhaust and supply air through the energy recovery unit, expressed in watts. In AHRI Guideline V, RERTotal, RERsensible, and RERlatent are defined as  net total m· min  h 1 – h 3  RERTotal = ---------------------------------------------------------P blower + P comp

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(24a)

26.24

2020 ASHRAE Handbook—HVAC Systems and Equipment(SI) Table 3 Fixed Plate

Membrane Plate

Comparison of Air-to-Air Energy Recovery Devices Energy Wheel

Liquid Desiccant

Fixed-Bed Regenerator

—

Counterflow

50 and up

Counterflow Parallel flow 50 and up

—

25 and up

40 to 60b

45 to 65b

40 to 60

40 to 60b

80 to 90c

0

0

0

0

50 to 75b,d

55 to 80

25 to 60

15 to 35

—

—

40 to 75d

60 to 80 with dessicant coatingc 50 to 80c

2.5 to 5 100 to 300

2 to 5 100 to 300

2 to 4 150 to 500

1.5 to 3 150 to 500

2 to 4 150 to 500

1.5 to 2.2 170 to 300

1 to 2.5 50 to 300

0.5 to 10 0.99 to 1.1 –55 to 800

0.5 to 10 1 to 1.2 –55 to 800

0 to 1 0.99 to 1.01 –40 to 93

0 1.0 –45 to 500

0 1.0 –40 to 40

0 1.0 –40 to 46

3 to 5c 0.90 to 1c –55 to 60

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Airflow Counterflow Counterflow Counterflow arrangements Cross flowa Cross flowa Parallel flow Equipment size 25 and up 25 and up 25 to 35 000 range, L/s and up Typical sensible 50 to 75 55 to 75 65 to 80 effectiveness (ms = me), %c Typical latent ef0 25 to 60 50 to 80 fectiveness,* %c Total effective20 to 50 35 to 70 ness,* %c Face velocity, m/s 1 to 5 1 to 3 Pressure drop, 100 to 1000 100 to 500 Pa EATR, % 0 to 2 0 to 5 OACF 0.97 to 1.06 0.97 to 1.06 Temperature –60 to 800 –40 to 60 range, °C Typical mode Exchanger Exchanger of purchase only only Exchanger in Exchanger in case case Exchanger Exchanger and blowers and blowers Complete Complete system system Advantages No moving No moving parts parts Low pressure Low pressure drop drop Easily cleaned Low air leakage Moisture/ mass transfer

Exchanger only Exchanger in case Exchanger and blowers Complete system Moisture/ mass transfer Compact large sizes Low pressure drop Available on all ventilati on system platforms Limitations Large size at Few suppliers Supply air higher flow Long-term may require mainterates some furnance and ther cooling perforor heating mance Some EATR unknown without purge Heat rate control Bypass damp- Bypass damp- Bypass (HRC) methods ers and duct- ers and duct- dampers ing ing and wheel speed control

aRated

effectiveness values are for balanced flow conditions for cross flow. Effectiveness values increase slightly if flow rates of either or both airstreams are higher than flow rates at which testing is done.

Heat Wheel

Heat Pipe

Counterflow

Counterflow Parallel flow 50 and up

25 to 35 000 and up 65 to 80

Exchanger only Exchanger in case Exchanger and blowers Complete system Compact large sizes Low pressure drop Easily cleaned

Some EATR with purge

Runaround Coil Loop Thermosiphon

Exchanger only Coil only Exchanger in Complete case system Exchanger and blowers Complete system No moving parts except tilt Fan location not critical Allowable pressure differential up to 15 kPa

bData

Exchanger only Complete Exchanger in system case

Exhaust airstream can be separated from supply air Fan location not critical

No moving parts Latent transExhaust airfer from stream can be remote airseparated from streams supply air Efficient Fan location not microbiologcritical ical cleaning of both supply and exhaust airstreams Effectiveness Predicting Effectiveness Few suppliers limited by performan may be limited Maintenance pressure drop ce requires by pressure and perforand cost accurate drop and cost mance Few suppliers simulation Few suppliers unknown model

Bypass damp- Tilt angle down ers and wheel to 10% of speed control maximum heat rate

cData

—

Bypass Control valve valve or over full range pump speed control

Control valve or pump speed control over full range

Exchanger and damper Complete system

Few moving parts Defrost strategy not required Low maintenance Easily cleaned Moisture/mass transfer if dessicantcoated Indoor units may require additional airflow selector/ damper to control EATR Some EATR Bypass damper and ducting Recovery period timing

not based on third-party certified data. EATR = exhaust air transfer ratio based on typical range of third-party certifiedOACF = outdoor air correction factor

data. dFace

velocity of 1.27 to 2.54 m/s.

 net sensible m· min c p  t 1 – t 3  RERsensible = -------------------------------------------------------------------P blower + P comp

(24b)

 net latent m· min h fg   1 –  3  RERlatent = ---------------------------------------------------------------------P blower + P comp

(24c)

m· min h1 t1 1 Pcomp Pblower

= = = = = =

minimum mass flow rate of supply and exhaust airstreams enthalpy at state 1 temperature at state 1 humidity ratio at state 1 direct power input of recovery component blower power required for energy recovery component

where

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26.25

Fig. 31 Psychrometric Processes of Exchangers in Parallel Mode (Moffitt 2011)

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Pblower =

Selection of Heat or Energy Recovery Ventilators

Q blower supply P supply Q blower exhaust P exhaust --------------------------------------------------------- + ------------------------------------------------------------- fan/motor supply  fan/motor exhaust

(25)

where P = pressure loss of component for supply or exhaust airstreams, kPa Qblower supply=supply fan airflow, L/s Qblower exhaust=exhaust fan airflow, L/s fan/motor=combined fan and motor efficiency

Note that other alternatives (e.g., comparing operating points on a fan curve) that accurately characterize the additional fan power required by the component are acceptable means of obtaining blower power. For a flow rate of 470 L/s and typical values of airstream pressure drop and motor and fan efficiencies, RERsensible can range from 26 to 32 kJ/Wh. The calculated value of RER suggest that the energy recovered is about 7.3 to 8.8 times that used to operate the energy recovery unit. The quality of thermal energy is about one-third that of electric power, so the actual energy recovered is about three times the cost of energy spent for energy recovery. The combined efficiency (CEF) is the ratio of the net heating/ cooling delivered to the total electric power consumed. CEF can also be related to the sensible, latent, or total recovered energy and is related to the RER; energy efficiency ratio (EER), which is typically provided by manufacturers of unitary equipment; and load ratio Y. It represents the percentage of the system load (heating, cooling, humidification and/or dehumidification) met by the energy recovery component. The higher load ratio allows use of smaller equipment, thereby affecting capital costs. CEFcooling can be estimated as 1 CEFcooling = -------------------------------------------------------------Y c  RER +  1 – Y c   EER

(26)

Net cooling capacity of energy recovery unit Yc = ----------------------------------------------------------------------------------------------------------System net cooling capacity

(27)

Net cooling capacity EER = ---------------------------------------------------------------Total electric consumption

(28)

where

For a typical value of EER = 10 for a unitary system, and a load ratio of about 0.3, the CEFcooling can range from 15 to 19 kJ/Wh. Use of an energy recovery unit and its effects on system sizing and annual net savings for various climatic regions are discussed in AHRI Guideline V.

Heat and energy recovery ventilators are available as heat exchangers only or as a complete system, including the heat exchangers and fan/motor systems, as indicated in Table 3. Energy recovery is also available integrated into unitary air-conditioning equipment or in both standard and custom air-handling systems. Selection of such units is primarily dictated by the quantity of ventilation air. Several manufacturers have developed software or tables to help select these units. The user may have to determine the required fan size (see Example 9), if only the heat exchanger is to be purchased. However, the true overall system performance is the life-cycle cost, which takes into account the capital and maintenance costs. Because of lack of sufficient data on these factors, they are not presented in Table 3.

Systems with Multiple Energy Recovery Exchangers Multiple exchangers are often used in air-handling systems that have a cooling coil, to enhance dehumidification capability. The first air-to-air exchanger in the outdoor airstream is used to recover exhaust energy and reduce the required coil capacity. The second exchanger’s purpose is to increase the coil’s latent removal. This second exchanger is either a sensible heat exchanger or a passive desiccant wheel. A sensible exchanger (e.g., coil loop, plate exchanger, heat pipe, wheel) reheats the air as it leaves the cooling coil. This simultaneously precools the return air. The cooled return air that is exhausted then precools the outdoor air by passing through the first exchanger. A passive dehumidification wheel works similarly: it removes water vapor from air leaving the coil and transfers the vapor, with some heat, to the return air. The second exchanger can be in parallel with the first exchanger, as shown in Figure 30 (Moffitt 2011). The corresponding processes are shown in Figure 31 (Moffitt 2011); use of a passive desiccant wheel is shown on the right side of the figure. The second exchanger can also be placed in series with the cooling coil, as shown in Figure 32 (Moffitt 2011), where it has a similar effect. The first exchanger reduces the cooling capacity by recovery exhaust energy. The second exchanger improves latent cooling and reduces the sensible cooling delivered by the system. The corresponding processes are shown in Figure 33 (Moffitt 2011); use of a passive desiccant wheel is shown on the right side of the figure.

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Fig. 33

Psychrometric Processes of Exchangers in Series Mode

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(Moffitt 2011)

Fig. 30 Multiple Energy Recovery Exchangers in Parallel Mode

Fig. 32 Multiple Energy Recovery Exchangers in Parallel Mode (Moffitt 2011)

Using Air-to-Air Heat Exchangers to Modify the Latent Capacity Ratio of Cooling Coils Air-to-air energy recovery can be used in a series application where heat is transferred from one location in an airstream (where the heat is detrimental) to another location in the same airstream (where the heat is beneficial). The most common example of this application is in series energy recovery around a cooling coil for summer dehumidification. During dehumidification the warm,

humid outside air is cooled by a cooling coil so the dry bulb temperature reaches the desired dew point. This air is then reheated to reduce its relative humidity (for indoor air quality, IAQ), and to make it more comfortable before entering the building space. The in-series energy recovery device is placed in front of and after the cooling coil, where it provides precooling (reducing load on the cooling coil) and reheat (reducing load on the reheating equipment). Figure 34 provides an example and shows the driving force (left) and the enthalpy savings (right). In some cases, the outdoor air reaches saturation during precool, and humidity is removed before the air enters the cooling coil. The total energy consumption (kWh) and peak demand (kW) load are both reduced, yielding lower energy bills and healthier IAQ (Mathur 1997a, 1997b, 1997c, 1997d). Since there must be energy balance between the precool side and reheat side, the changes in enthalpy will always be equal and opposite. Where there is no condensation during precool, the T across the precool side will be equal and opposite to the reheat side. When condensation occurs during precool, the T across the precool side will be smaller than the reheat side. Even on direct-expansion (DX) systems where hot gas reheat is a resource created by the process, in-series energy recovery can be used to reduce compressor capacity and downsize equipment to reduce first costs, particularly where condensation takes place on more of the DX coil, which increases the cycle’s coefficient of performance (COP). Because both precooling and reheating are beneficial to the system, the RER in dehumidification applications is double that of an identical exchanger used in energy recovery. Also, the driving force is the dry-bulb temperature difference between the outdoor air and the cooling coil leaving temperature. For example, on a 32°C day and with 13°C air leaving the cooling coil, the driving force across the energy recovery device is 7°C. In contrast, a traditional air-to-air energy recovery device with outdoor air at 32°C and return air at 24°C is exposed to a driving temperature difference of only 8 K, a third of the in-series energy recovery device. Furthermore, when the energy recovery device is in economizer mode (outdoor air temperatures of about 13 to 24°C) the in-series recovery device is still recovering energy. The in-series energy recovery device should be sized for the highest dry-bulb temperature design condition, since this is when the device is exposed to the highest driving force and creates most reheat. This ensures the system will never provide too much reheat during the year. This approach is the simplest and is used when the primary goal is to downsize cooling equipment. The system designer must recognize that when the outdoor temperature is

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Fig. 34 Psychrometric Chart: In-Series Energy Recovery

Fig. 36 Flat Heat Recovery Device in Wraparound (In-Series) Configuration lower, that the driving force is also lower, and the device provides less reheat. If the design reheat is required during the whole cooling season, then supplemental reheat is still required after the inseries energy recovery device, albeit less than in a system without energy recovery. An alternative design strategy is to optimize energy recovery. In this strategy, the system designer selects a device with a higher effectiveness than in the previous strategy, and provides a means of modulating its performance. By modulating the performance during the warmest outdoor air conditions, excessive reheat is avoided. During cooler conditions the device operates at full performance to recover more energy. The device is usually sized with a mind to pressure drop, recognizing that the airstream passes through the device twice, and initial purchase cost and is optimized against annual energy savings. Examples of devices are wraparound heat pipes and thermosiphons (Figures 35 and 38), flat heat exchangers like wheels and heat pipes (Figure 36), and plate heat exchangers (Figure 37). Footprint and condensation management should be considered when designing the layout. Where a controllable device is required, options include: wraparound heat pipes and thermosiphons that use valves to control refrigerant flow; bypass around

Fig. 35 Wraparound (In-Series) Heat Pipe the device (but not through the cooling coil); or, for wheels, modulating wheel speed. Example 5. Precooling Air Reheater Dehumidifier. In this application, 1.6m3/s of outdoor supply air at 35 and 27°Cwb ( = 1.11 kg/m3) is used to reheat 1.6 m3/s of the same air leaving a cooling coil (exhaust) at 11.2 and 11.0°C wb using a sensible heat exchanger as a precooling air reheater. The reheated air is to be between 24 and 26°C. In this application, the warm airstream is outdoor air and the cold airstream is the same air after it leaves the cooling coil. The unit’s manufacturer lists its effectiveness as 58.4%. For EATR = 0 and OACF 1, determine the leaving precooled and reheated air conditions and energy recovered, and check the energy exchange balance. Solution: Step 1. Because data on pressure drop are missing, skip this step. Step 2. Calculate the theoretical maximum energy transfer.

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Fig. 37 Flat Plate Heat Exchanger in Wraparound (In-Series) Configuration

Fig. 39 Precooling Air Reheater Dehumidifier (Example 5) a. Precooler leaving air conditions Entering enthalpy, determined from the psychrometric chart for 35°C db and 27°C wb, is 85.2 kJ/kg.

– 24.7 1.11  1.6 

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h2 = 85.2 + ----------------------- = 71.3 kJ/kg The wet-bulb temperature for saturated air with this enthalpy is 23.5°C. This is point 2 on the psychrometric chart (Figure 39), which is near saturation. Note that this precooled air is further dehumidified by a cooling coil. b. Reheater leaving air conditions

24.7 1.11  1.6   1 

t4 = 11.2 + ------------------------------- = 25.1C Entering enthalpy, determined from the psychrometric chart for 11.2°C db and 11.0°C wb, is 32.0 kJ/kg.

24.7 1.11  1.6 

h4 = 32 + ----------------------- = 45.9 kJ/kg The wet-bulb temperature for air with this temperature and enthalpy is 16.3°C. Step 6. Check performance. qs = 1.11  1.6   85.2 – 71.3  = 24.7 kW precooling

Fig. 38 Wraparound (In-Series) Thermosiphon with a Valve for Modulation The air being reheated will have less mass than the outdoor air entering the precooler because moisture will condense from it as it passes through the precooler and cooling coil. Reheat is sensible heat only, so the denominator of Equation (2a) is used to determine the theoretical maximum energy transfer. To simplify the calculation, it is assumed that specific heat of dry air is constant at 1 kj/kg·K. qs = 1.11  1.6   1   35 – 11.2  = 42.3 kW saved Step 3. Establish the sensible effectiveness. The manufacturer gives the effectiveness as 58.4% at the designated operating conditions. Step 4. Calculate actual energy transfer at design conditions. qactual = 0.584  42.3  = 24.7 kW Step 5. Calculate leaving air conditions. Because condensation occurs as the outdoor airstream passes through the precooling side of the heat exchanger, use Equation (4e) to determine its leaving enthalpy, which is the inlet condition for the cooling coil. Sensible heat transfer Equation (5a) is used to determine the temperature of air leaving the preheat side of the heat exchanger.

qe = 1.11  1.6   1   25.1 – 11.2  = 24.7 kW reheat Step 7. Plot conditions on psychrometric chart (Figure 39). Because EATR = 0 and OACF  1, Steps 8 to 10 are not presented here.

Dessicant and Heat Wheel Systems The Center for Climate and Energy Solutions’ 2015 international conference on climate change emphasized reducing greenhouse gas emissions and requiring more efficient equipment, including HVAC systems (which consume significant electrical power) (C2ES 2015). Sensible and desiccant wheels can be used in combination with a heat pump direct evaporative cooler (DEC) to achieve this goal, as shown in the following cases. The heat dissipated in the condenser of a heat pump can be used to regenerate the desiccant wheel (Sheng et al. 2015) and has a potential of significant energy cost savings. At design conditions for Atlanta, Georgia, for a 100% replacement air system as shown in Figure 40, this approach can reduce cooling capacity requirements by 52% and net electrical shaft power requirements by more than 24% compared to the conventional vapor-compression system (Dhamshala 2016). Figure 41 shows an advanced desiccant system for 100% outdoor supply air (OSA) mode with heat and desiccant wheels and direct

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Fig. 40 Heat Pump Augmented by Heat and Dessicant Wheels

Fig. 41 Heat Pump Augmented by Evaporative Cooler and Heat and Dessicant Wheels

Fig. 42 System for Surgery Room (based on concepts presented in Murphy 2006)

evaporative cooler (DEC) in combination with a heat pump. This system could reduce cooling capacity requirements by 59% and net electrical shaft power requirements by close to 30% at the same design conditions as for Figure 40 (Dhamshala 2016). A simplified system suitable for surgery rooms (able to provide optimum temperature and humidity control) consists of a preheater, desiccant wheel with type 3 desiccant, and a cooling coil, as shown in Figure 42. The type 3 desiccant in the desiccant wheel has a large affinity to adsorb water vapor when the relative humidity of the air is high (as it is in states 6 and 7), whereas it gives off moisture at states 4 and 5 (which have lower humidity). The simplicity of this system not only reduces maintenance costs but can also save significant energy costs. A desiccant system consisting of two evaporative coolers and heat and desiccant wheels for a 100% replacement air application, as shown in Figure 43, could provide all the space summer cooling needs without the use of an air conditioner. Such a system can reduce net electrical power requirements by more than 90% at design conditions for the city of Phoenix, Arizona (Dhamshala 2016). The only electrical power required is for the fan power and rotation of the wheels. The magnitude of estimated electrical power savings should be of similar range for the entire dry (B) region of the ASHRAE climatic zones, with small differences. Note that the minuscule heat energy required (6.15 kW) in the heater can easily be obtained from an air-to-air heat exchanger using

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Fig. 43 Dessicant System of Evaporative Coolers and Heat and Dessicant Wheels outdoor air (which, at 39.4°C, exceeds the minimum temperature) before it enters the desiccant wheel. The cost of water supplied to the evaporative coolers is estimated to be negligible. Accurate estimates of energy cost savings over a year can be obtained through detailed computer simulations of systems for a specific building with given building loads using the hourly weather data. Highly specialized desiccant systems using a DOAS can be designed to serve the buildings with specific loads or needs (Moffitt 2015). Sultan et al. 2015 provide an overview of various widely used solid-desiccant air systems. In the case that outdoor air requirements are less than 100%, use of direct evaporative coolers adds more humidity into the supply air. An alternative to replace the direct evaporative cooler is the indirect evaporative cooler (M-cycle), as shown in Figure 44 for the cooling season, which has a potential to save 32% of cooling capacity, while the system shown in Figure 45 (suitable for winter heating) has a potential to save 22% of heating costs. These systems are particularly suitable for sustainable systems, where the emphasis is placed on the use of renewable energy resources such as solar, wind, and bioenergy. A PV/T panel system in combination with water-to-water heat pump, absorption chiller, and thermal storage can provide the energy requirements for cooling and heating seasons. However, these systems are also suitable if renewable energy resources are not available at the site. The heat energy required by the heater placed between the heat and desiccant wheels to regenerate the air can be obtained from the condenser of the air-conditioner, similar to the system shown in Figure 14. Sustainability. Recently, due to climate change, greater emphasis has been on reducing the greenhouse gas emissions (UN Climate Change Report; UNFCCC 2018), and on the rapid adoption of renewable energy resources. In light of the urgent need, the systems shown in Figures 44 and 45 pave the way to accommodate the use of renewable energy resources to meet the future goals of the HVAC systems to be compatible with smart grid applications. The desiccant and heat wheels shown in Figures 44 and 45 and not only perform the energy recovery from the outgoing exhaust air but also meet the peak loads of the building through proper selection of flow rates, mixing of outdoor air with exhaust air, and regenerating temperature. Several other options of these variations may be possible to suit the level of regenerating temperatures. Ge et al. (2015), Tu et al. (2015), and Zeng et al. (2014) find that two-stage desiccant systems can be designed to use low-grade heat energy resources, such as waste energy at the site or solar collectors, which can yield greater energy cost savings. To best use existing energy resources available at a specific temperature or grade, perform a series of second-law analyses on a variety of desiccant systems (including a two-stage unit) to find the maximum energy cost savings and to identify the components of the system that have largest exergy destruction (Enteria et al. 2013; Liu et al. 2016; Tu et al. 2015a, 2015b).

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Fig. 44

Dessicant and Heat Wheels with Indirect Evaporative Cooler (M-Cyle) Dhamshala et al. (2019)

Fig. 45 Dessicant and Heat Wheels with Humidifier for Winter Heating Dhamshala et al. (2019)

An emerging technology combining liquid-desiccant solutions (e.g., lithium chloride) with an evaporative cooler has the potential to save significant energy costs (Gao et al. 2015). A simplified schematic suitable for use in various climates is shown in Figure 46. Note that the minuscule heat energy required (6.15 kW) in the heater can easily be obtained from an air-to-air heat exchanger using outdoor air (which, at 39.4°C, exceeds the minimum temperature) before it enters the desiccant wheel. The cost of water supplied to the evaporative coolers is estimated to be negligible. Heat energy is required to regenerate the liquid desiccant solution; it can use waste heat energy, if available at the site, or use solar collectors. A combination of heat and desiccant wheels with this liquid-desiccant system might provide an economical system that is environmentally friendly.

8.

ECONOMIC CONSIDERATIONS

Air-to-air energy recovery systems are used in both new and retrofit applications. These systems should be designed for the maximum cost benefit or least life-cycle cost (LCC) expressed either over the service life or annually and with an acceptable payback period. The annualized system owning, operating, and maintenance costs are discussed in Chapter 38 of the 2019 ASHRAE Handbook— HVAC Applications. Although the capital cost and interest term in this method imply a simple value, it is in fact a complex function of the future value of money as well as all the design variables in the energy/heat exchanger. These variables include the mass of each material used, cost of forming these materials into a highly effective

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26.31 where Cs, init ITC Ce Tinc

= = = =

initial system cost investment tax credit for energy-efficient improvements cost of energy to operate the system for one period net income tax rate where rates are based on last dollar earned (i.e., marginal rates) = (local + state + federal rate) – (federal rate) (local + state rate) CRF = capital recovery factor i = effective discount rate adjusted for energy inflation n = total number of periods under analysis

The inverse of this term is usually called the return on investment (ROI). Well-designed energy recovery systems normally have a PP of less than 5 years, and often less than 3 years. Paybacks of less than 1 year are not uncommon in comfort-to-comfort applications in hot, humid climates, primarily because of the reduced size of cooling equipment required. Other economic factors include the following. Fig. 46 Schematic of Liquid Dessicant System and Evaporative Cooler

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NREL

energy/heat exchanger, cost of auxiliary equipment and controls, and cost of installation. The operating energy cost for energy recovery systems involves functions integrated over time, including variables such as flow rate, pressure drop, fan efficiency, energy cost, and energy recovery rate. The calculations are complex because air-heating and/or cooling loads are, for a range of supply temperatures, time dependent in most buildings. Time-of-use schedules for buildings can impose different ventilation rates for each hour of the day. Electrical utility charges often vary with time of day, amount of energy used, and peak power load. For building ventilation air-heating applications, the peak heat recovery rate usually occurs at the outdoor supply temperature at which frosting control throttling must be imposed. In addition to designing for the winter design temperature, heat recovery systems should be optimized for peak heat recovery rate, taking frost control into account. Overall exchanger effectiveness  should be high (see Table 3 for typical values); however, a high  implies a high capital cost, even when the exchanger is designed to minimize the amount of materials used. Energy costs for fans and pumps are usually very important and accumulate operating cost even when the energy recovery system must be throttled back. For building ventilation, throttling may be required much of the time. Thus, the overall LCC minimization problem for optimal design may involve 10 or more independent design variables as well as several specified constraints and operating conditions (see, e.g., Besant and Johnson [1995]). In addition, comfort-to-comfort energy recovery systems often operate with much smaller temperature differences than most auxiliary air-heating and cooling heat exchangers. These small temperature differences need more accurate energy transfer models to reach the maximum cost benefit or lowest LCC. Most importantly, recovered energy at design may be used to reduce the required capacity of heating and cooling equipment, which can be significant in both system performance/efficiency and economics. The payback period (PP) is best computed after the annualized costs have been evaluated. It is usually defined as Capital cost and interest PP = ----------------------------------------------------------------------------------Annual operating enery cost saved C s ,init – ITC = ------------------------------ CRF  i n  C e  1 – T inc 

(29)

System Installed Cost. Initial installed HVAC system cost is often lower for air-to-air energy recovery devices because mechanical refrigeration and fuel-fired heating equipment can be reduced in size. Thus, a more efficient HVAC system may also have a lower installed total HVAC cost. The installed cost of heat recovery systems becomes lower per unit of flow as the amount of outdoor air used for ventilation increases. Life-Cycle Cost. Air-to-air energy recovery cost benefits are best evaluated considering all capital, installation, operating, and energy-saving costs over the equipment life under normal operating conditions in terms of the life-cycle cost. As a rule, neither the most efficient nor the least expensive energy recovery device will be most economical. Optimizing the life-cycle cost for maximum net savings may involve many design variables, requiring careful cost estimates and use of an accurate recovery system model with all its design variables (see, e.g., Besant and Simonson [2000]). Energy Costs. The absolute cost of energy and relative costs of various energy forms are major economic factors. High energy costs favor high levels of energy recovery. In regions where electrical costs are high relative to fuel prices, heat recovery devices with low pressure drops are preferable. Amount of Recoverable Energy. Economies of scale favor large installations. Equipment is commercially available for air-toair energy recovery applications using 25 L/s and above. Grade of Exhaust Energy. High-grade (i.e., high-temperature) exhaust energy is generally more economical to recover than lowgrade energy. Energy recovery is most economical for large temperature differences between the waste energy source and destination. Coincidence and Duration of Waste Heat Supply and Demand. Energy recovery is most economical when supply coincides with demand and both are relatively constant throughout the year. Thermal storage may be used to store energy if supply and demand are not coincident, but this adds cost and complexity to the system. Proximity of Supply to Demand. Applications with a large central energy source and a nearby waste energy use are more favorable than applications with several scattered waste energy sources and uses. Operating Environment. High operating temperatures or the presence of corrosives, condensable gases, and particulates in either airstream results in higher equipment and maintenance costs. Increased equipment costs result from the use of corrosion- or temperature-resistant materials, and maintenance costs are incurred by an increase in the frequency of equipment repair and wash down and additional air filtration requirements. Effect on Pollution Control Systems. Removing process heat may reduce the cost of pollution control systems by (1) allowing less expensive filter bags to be used, (2) improving the efficiency of electronic precipitators, or (3) condensing out contaminant vapors,

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thus reducing the load on downstream pollution control systems. In some applications, recovered condensable gases may be returned to the process for reuse. Effect on Heating and Cooling Equipment. Heat recovery equipment may reduce the size requirements for primary utility equipment such as boilers, chillers, and burners, as well as the size of piping and electrical services to them. Larger fans and fan motors (and hence fan energy) are generally required to overcome increased static pressure loss caused by the energy recovery devices. Auxiliary heaters may be required for frost control. Effect on Humidifying or Dehumidifying Equipment. Selecting total energy recovery equipment results in the transfer of moisture from the airstream with the greater humidity ratio to the airstream with the lesser humidity ratio. This is desirable in many situations because humidification costs are reduced in cold weather and dehumidification loads are reduced in warm weather.

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9.

ENERGY AND/OR MASS RECOVERY CALCULATION PROCEDURE

The rate of energy transfer to or from an airstream depends on the rate and direction of heat transfer and water vapor (moisture) transfer. Under customary design conditions, heat and water vapor transfer will be in the same direction, but the rate of heat transfer often will not be the same as the rate of energy transfer by the crossstream flow of water vapor. This is because the driving potentials for heat and mass transfer are different, as are the respective wall resistances for the two types of transport. Both transfer rates depend on exchanger construction characteristics. The following general procedure may be used to determine the performance and energy recovered in air-to-air energy recovery applications at each operating condition. Step 1. Determine supply and exhaust air pressure drops ps and pe across exchanger. Request air pressure drops ps and pe across the heat or energy exchanger from the manufacturer, who may have certified AHRI Standard 1061 test condition data obtained using ASHRAE Standard 84 as a test procedure and analysis guide. These data may be extrapolated to non-AHRI conditions by the manufacturer using correlations such as Equations (12) or (13), if their restrictions are satisfied. For other flow conditions, somewhat different correlations may be more accurate to determine the pressure drop. Step 2. Calculate theoretical maximum moisture and energy transfer rates mmax and qmax. The airstream with the lower mass flow mmin limits heat and moisture transfer. Some designers specify and prefer working with airflows stated at standard temperature and pressure conditions. To correctly calculate moisture or energy transfer rates, the designer must determine mass flow rates. For this reason, the designer must know whether airflow rates are quoted for the entry conditions specified or at standard temperature and pressure conditions. If necessary, convert flow rates to mass flow rates (e.g., Ls or m3s at standard temperature and pressure to kg/s) and then determine which airstream has the minimum mass. The theoretical maximum sensible heat, latent heat and total energy rates are given by the denominators of Equations (2a), (2b), and (2c), respectively. The split between latent and sensible energy can be determined by plotting airstream conditions on a psychrometric chart as shown in Figure 47. Maximum sensible heat transfer is represented by a horizontal line drawn between the two dry-bulb temperatures, and maximum latent energy transfer is represented by the vertical line. Step 3. Establish moisture, sensible, and total effectiveness s , L , and t .

Fig. 47 Maximum Sensible and Latent Heat from Process A-B Each of these ratios is obtained from manufacturers’ product data using input conditions and airflows for both airstreams. The effectiveness for airflows depends on (1) exchanger construction, including configuration, heat transfer material, moisture transfer properties, transfer surface area, airflow path, distance between heat transfer surfaces, and overall size; and (2) inlet conditions for both airstreams, including pressures, velocities, temperatures, and humidities. In applications with unequal airflow rates, the enthalpy change will be higher for the airstream with the lesser mass flow. Step 4. Calculate actual moisture (latent) and energy (sensible, latent or total) transfer rates. The actual moisture, sensible heat, latent heat, and total energy rates are given by Equations (5a), (5b), (5c), and (5d or 5e), respectively. Step 5. Calculate leaving air properties for each airstream using Equations (4a), (4b), (4c), (4d), (4e), and (4f). With an enthalpy or moisture-permeable heat exchanger, moisture (and its inherent latent energy) is transferred between airstreams. With a sensible-only heat exchanger, if the warmer airstream is cooled below its dew point, the resulting condensed moisture transfers additional energy. When condensation occurs, latent heat is released, maintaining that airstream at a higher temperature than if condensation had not occurred. This higher air temperature (potential flux) increases the heat transfer to the other airstream. The sensible and total effectiveness are widely used because the energy flow in the condensate is relatively small in most applications. (Freezing and frosting are unsteady conditions that should be avoided unless a defrost cycle is included.) Equations (4c) and (4d) must be used to calculate the leaving air humidity conditions, and Equations (4e) and (4f) to calculate the enthalpy values for airstreams in which inherent latent energy transfer occurs. Equations (4a) and (4b) may be used for airstreams if only sensible energy transfer is involved. Step 6. Check energy transfer balance between airstreams. Equation (5a) can be used to estimate the sensible energy transfer rate into the supply airstream, and can be adapted to estimate the energy transfer rate out of the exhaust airstream by substituting T3 for T1, and substituting T4 for T2. Equation (5d) or (5e) can estimate the total energy for the two airstreams, using similar substitutions. Total energy transferred from one airstream should equal total heat transferred to the other. Calculate and compare the energy transferred to or from each airstream. Differences between these energy flows are usually because of measurement errors.

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Air-to-Air Energy Recovery Equipment Step 7. Plot entering and leaving conditions on psychrometric chart. Examine the plotted information for each airstream to verify that performance is reasonable and accurate. (Steps 8 to 10 apply only when EATR  0 and OACF  1.) Step 8. Obtain data on exhaust air transfer ratio (EATR > 0 and typically 0.05 > EATR > 0 for regenerative wheels). Request the EATR data from the manufacturer, who may have certified AHRI Standard 1061 test condition data obtained using ASHRAE Standard 84 as a test procedure and analysis guide. These data may be extrapolated to non-AHRI test conditions using correlations relating EATR to air pressure differences between the supply and exhaust and, for rotary regenerative wheels, carryover caused by wheel rotation. Shang et al. (2001a) show that, for regenerative wheels, a correlation may be developed between EATR and carryover ratio, Rc , and OACF, but for other air-to-air exchangers EATR will be very small or negligible.

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Step 9. Obtain data on outdoor air correction factor (OACF  1 and typically 0.9 < OACF < 1.1 for regenerative wheels). Request the OACF data from the manufacturer, who may have certified AHRI Standard 1061 test condition data obtained using ASHRAE Standard 84 as a test procedure and analysis guide. These data may be extrapolated to non-AHRI test conditions using correlations relating OACF to pressure differences (Shang et al. 2001b), for regenerative wheels; for other exchangers, OACF will be very nearly 1.0. Step 10. Correct the supply air ventilation rate, moisture transfer rate, and energy transfer rates for EATR  0 and OACF  1.0. Values of EATR significantly larger than zero and OACF significantly different than 1.0 imply the air-to-air exchanger is transferring air between the exhaust and supply airstreams. This transfer may be important, especially for some devices such as regenerative wheels. Shang et al. (2001b) show a method to correct the energy rates when EATR  0 and OACF  1. The procedure to correct the supply air ventilation rate is shown in Example 9. Example 6. Sensible Heat Recovery in Winter. Exhaust air at 23°C and 10% rh with a flow rate of 6 kg/s preheats an equal mass flow rate of outdoor air at –18°C and 60% rh ( = 1.39 kgm3) using an air-to-air heat exchanger with a measured effectiveness of 60%. Airflows are specified as mass flow rates at standard temperature and pressure (i.e., 15°C and 101.325 kPa). Assuming EATR = 0 and OACF  1, determine the leaving supply air temperatures and energy recovered, and check the heat exchange balance. To simplify the calculation, assume that specific heat of dry air is constant at 1 kJ/kg·K

26.33 a. Leaving supply air temperature t2 is given as 148 kW t2 = –18°C + ----------------------------------------------------- = 6.7°C  6 kg/s   1 kJ/(kg·K)  b. Leaving exhaust air temperature t4 is given as 148 kW t4 = 23°C – ----------------------------------------------------- = –1.7°C  6 kg/s   1 kJ/(kg·K)  6. Using Equation (5a), check performance. qs = (6 kg/s)[1 kJ/(kg·K)][6.7 – (–18)] = 148 kW saved 7. Plot conditions on psychrometric chart to confirm that no moisture exchange occurred (Figure 48). Because EATR = 0 and OACF  1, Steps 8 to 10 of the calculation procedure are not presented here. Example 7. Sensible Heat Recovery in Winter with Water Vapor Condensation. Exhaust air at 23°C and 28% rh ( = 1.2 kg/m3) and flow rate of 5 m3/s is used to preheat 4.5 m3/s of outdoor air at –10°C and 50% rh ( = 1.34 kg/m3) using a heat exchanger with a sensible effectiveness of 70%. Assuming EATR = 0 and OACF  1, determine the leaving supply air conditions and energy recovered, and check the energy exchange balance. Solution: The supply airstream has a lower airflow rate than the exhaust airstream, so it may appear that the supply airstream limits heat transfer. However, determination of mass flow rates for the given entry conditions shows that the mass flow rate of the supply airstream (6.03 kg/s) is slightly greater than that of the exhaust airstream (6.0 kg/s), so exhaust is the limiting airstream. Nevertheless, because the mass difference is negligible, it is convenient to use supply air volume as the limiting airstream. Also, to simplify the calculation, assume that specific heat of dry air is constant at 1 kj/kg·K. 1. Because data on pressure drop are missing, skip this step. 2. Calculate the theoretical maximum sensible heat transfer. The limiting airstream, the supply airstream, will be preheated in the heat exchanger, so it is not subject to condensation. Therefore, the denominator of Equation (2a) is used: qmax = (4.5 m3s)(1.34 kgm3)[1 kJ(kg·K)][23 – (–10)]°C= 199 kW 3. Select sensible effectiveness. From manufacturer’s literature and performance test data, the sensible effectiveness is determined to be 70% at the design conditions. 4. Calculate actual heat transfer at design conditions using Equation (5a): qs = (0.7)(199 kW) = 139 kW 5. Calculate leaving air conditions. a. Leaving supply air temperature is calculated by

Solution: Note: the numbers correspond to the steps in the calculation procedure. 1. Because data on pressure drop are missing, skip this step. 2. Calculate the theoretical maximum heat transfer. The two inlet conditions plotted on a psychrometric chart (Figure 48) indicate that, because the exhaust air has low relative humidity, latent energy transfer does not occur. Using the denominator of Equation (2a), the theoretical maximum sensible heat transfer rate qs is qmax = (6 kg/s)[1 kJ/(kg·K)][23 – (–18)] = 246 kW 3. Establish the sensible effectiveness. From manufacturer’s literature and certified performance test data, effectiveness is determined to be 60% at the design conditions. 4. Calculate actual heat transfer at given conditions. Using Equation (5a), qs = (0.6)(246) kW = 148 kW 5. Calculate leaving air conditions. Because no moisture or latent energy transfer will occur,

Fig. 48 Sensible Heat Recovery in Winter (Example 6)

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26.34

2020 ASHRAE Handbook—HVAC Systems and Equipment(SI)   139 kW -  = 13.1°C t2 = – 10°C +  -----------------------------------------------------------------------------------------3 3   1.34 kg/m   4.5 m /s   1 kJ/(kg·K)  

b. Because the dew point of exhaust air at inlet is 3.54°C, condensation occurs on the exhaust side, so the leaving exhaust air temperature cannot be determined using Equation (4b). The entering exhaust air enthalpy and humidity ratio are determined for the drybulb temperature of 23°C and 28% rh using a psychrometric chart and found to be h3 = 36 kJ/kg and w3 = 0.0052 kgkg. However, the leaving exhaust air enthalpy can be determined by 139 kW - = 12.83 kJkg h4 = 36 kJ/kg – -------------------------------------------------3 3  1.2 kg/m   5 m /s  Because the air will be saturated at the outlet of exhaust air, the dry-bulb or wet-bulb temperature and humidity ratio corresponding to an enthalpy of 12.83 kJ/kg is found to be t4 = 1.4°C and w4 = 0.0044 kg/kg. The rate of moisture condensed mw is mw = me (w3 – w4) = (6 kg/s)(0.0049 – 0.0042) = 0.0042 kg/s

Supply inlet (35°C db, 27°C wb) h1 = 85.2 kJ/kg w1 = 0.0194 kg/kg Exhaust inlet (23°C db, 17°C wb) h3 = 47.8 kJ/kg w3 = 0.0093 kg/kg The theoretical maximum sensible and total heat transfer rates can be obtained as follows: qmax (sensible) = (1.15 kg/m3)(4 m3/s)[1 kJ/(kg·K)](35 – 23) = 55.2 kW qmax (total energy) = (1.15 kg/m3)(4 m3/s)(85.2 – 47.8) = 172 kW 3. Determine supply sensible and total effectiveness. The manufacturer’s selection data for the design conditions provide the following effectiveness ratios: s = 70%

t = 56.7%

4. Calculate energy transfer at design conditions.

6. Check performance. qs = (134 kgm3)(4.5 m3s)[0.24 kJ(kg·°C)][13.1 – (–10)] = 139.3 kW saved

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The supply airstream is a lesser or limiting airstream for energy and moisture transfer. Determine entering airstream enthalpies and humidity ratio from psychrometric chart.

Neglect the enthalpy of the condensed water by adding the energy lost through condensation of vapor to the sensible heat lost of the exhausting air. qe = (1.2 kgm3)(5 m3s)[1 kJ(kg·K)][23 –1.4)] + (0.0042 kg/s)(2560 kJ/kg) = 140.35 kW saved, which is very close to 139.3 kW 7. Plot conditions on psychrometric chart (Figure 49). Note that moisture condenses in the exhaust side of the heat exchanger. Because EATR = 0 and OACF  1, Steps 8 to 10 of the calculation procedure are not presented here. Example 8. Total Heat Recovery in Summer. Exhaust air at 23°C and 17°C wb ( = 1.2 kg/m3 ) with a flow rate of 5 m3/s is used to precool 4 m3/s of supply outdoor air at 35°C and 27°C wb ( = 1.15 kg/m3 ) using a hygroscopic total energy exchanger. The sensible and total effectiveness for this heat exchanger are 70 and 56.7%, respectively. Assuming EATR = 0 and OACF  1, determine the leaving supply air conditions and energy recovered, and check the energy exchange balance. To simplify the calculation, assume that specific heat of dry air is constant at 1 kj/kg·K. Solution: 1. Because data on pressure drop are missing, skip this step. 2. Calculate the theoretical maximum heat transfer.

qt = (0.567)(172 kW) = 97.52 kW total recovered qs = –(0.7)(55.2 kW) = –38.64 kW sensible recovered qtat = 58.8 kW latent recovered 5. Calculate leaving air conditions. a. Supply air conditions – 38.64 kW t2 = 35°C + --------------------------------------------------------------------------------------- = 26.6°C 3 3  1.15 kg/m   4 m /s   1 kJ/(kg·K)  – 97.52 kW h2 = 85.2 kJ/kg + ------------------------------------------------------ = 64 kJ/kg 3 3  1.15 kg/m   4 m /s  From the psychrometric chart, the supply air humidity ratio and wet-bulb temperature are found to be w2 = 0.0145 and tw2 = 21.8°C. b. Exhaust air conditions 38.64 kW t4 = 23°C + ------------------------------------------------------------------------------------ = 29.4°C 3 3  1.2 kg/m   5 m /s   1 kJ/(kg·K)  97.52 kW h4 = 47.8 kJ/kg + -------------------------------------------------- = 64.1 kJ/kg = 64.1 kJ/kg 3 3  1.2 kg/m   5 m /s  From the psychrometric chart, the exhaust humidity ratio and wetbulb temperature are found to be w4 = 0.0134 and tw4 = 21.8°C. 6. Check total performance (Equation 5d). qt = (1.15 kg/m3)(4 m3/s)(85.2 – 64) = 97.52 kW saved qt = (1.2 kg/m3)(5 m3/s){[1 kJ/(kg · K)](29.4 – 23) + (0.013 – 0.0093)(2560 kJ/kg)} = 95.23 kW, which is close to 94.1 kW 7. Plot conditions on psychrometric chart (Figure 50). Because EATR = 0 and OACF  1, Steps 8 to 10 are not presented here. Example 9. Total Energy Recovery with EATR  0 and OACF  1.0. An ERV manufacturer claims a product has performance characteristics as shown here at 400 L/s: Pressure drop p = 225 Pa sensible = 0.73 latent = 0.68 total = 0.715

Fig. 49 Sensible Heat Recovery in Winter with Condensate (Example 6)

A building has a ventilation requirement of 400 L/s and exhaust air at 23°C and 17°C wb ( = 1.2 kg/m3) is used to precool supply outdoor air at 35°C and 27°C wb ( = 1.15 kg/m3).

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Air-to-Air Energy Recovery Equipment (a) Assuming EATR = 0 and OACF  1, determine the leaving supply air conditions and energy recovered, and check the energy exchange balance. (b) Assuming EATR = 5% and OACF = 1.05, determine the actual airflow rates. In both cases, to simplify the calculation, assume that specific heat of dry air is constant at 1 kj/kg·K. Solution: 1. From the manufacturer’s claims, at a flow rate of 400 Ls, the pressure drop  p = 225 Pa. Assuming the effective efficiency of the fan motor combination is about 0.6, the power Ps required to circulate the supply air can be obtained from Equation (8) as Ps = (400 L/s)(1 m31000 L)(225 Pa)(0.6) = 150 W or 0.15 kW Assuming the balanced flow the power required to circulate the exhaust air would be same, therefore the total power P required to circulate the airstreams would be twice this amount. P = 300 W or 0.3 kW

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2. Calculate the theoretical maximum heat transfer. Determine entering airstream enthalpies and humidity ratio from the psychrometric chart. Supply inlet (35°C db, 27°C wb) h1 = 85.2 kJ/kg

w1 = 0.0194 kg/kg

Exhaust inlet (23°C db, 17°C wb) h3 = 47.8 kJ/kg

w3 = 0.0093 kg/kg

The theoretical maximum heat transfer rates can be obtained as follows: qmax(sensible) = (1.15 kgm3)(400 Ls)(1 m31000 L)  [1 kJ/(kg · K)](35 – 23) = 5.52 kW

26.35 qmax(latent) = (1.15 kgm3)(400 Ls)(1 m31000 L)  (2560 kJ/kg)(0.0194 – 0.0093) = 11.89 kW qmax(total) = (1.15 kgm3)(400 Ls)(1 m31000 L)  (85.2 – 47.8) = 17.2 kW Note that sum of sensible and latent energy should equal the total energy. 3. Determine supply sensible and total effectiveness. From the manufacturer’s claims, at a flow rate of 400 L/s, s = 0.73, L = 0.68, and t = 0.715. 4. Calculate energy transfer at design conditions. qs = (0.73)(5.52 kW) = 4.03 kW sensible recovered qL = (0.68)(11.89 kW) = 8.09 kW latent recovered qt = (0.715)(17.2 kW) = 12.3 kW total recovered 5. Calculate leaving air conditions. a. Supply air conditions – 4.03 kW - = 26.2°C t2 = 35°C + -----------------------------------------------------------------------------------------3 3  1.15 kg/m   0.4 m /s   1 kJ/(kg·K)  – 12.3 kW h2 = 85.2 kJkg + ---------------------------------------------------------- = 58.5 kJkg 3 3  1.15 kg/m   0.4 m /s  From the psychrometric chart, the supply air humidity ratio and wet-bulb temperature are w2 = 0.0129 and tw2 = 20.2°C. b. Exhaust air conditions 4.03 kW t4 = 23°C + ---------------------------------------------------------------------------------------- = 31.4°C 3 3  1.2 kg/m   0.4 m /s   1 kJ/(kg·K)  12.3 kW h4 = 47.8 kJ/kg + ------------------------------------------------------- = 73.43 kJ/kg 3 3  1.2 kg/m   0.4 m /s  From the psychrometric chart, the exhaust humidity ratio and wetbulb temperature are found to be w4 = 0.0164, tw4 = 24.2°C. 6. Check total performance. qt = (1.15 kgm3)(0.4 m3s)(85.2 – 58.5) = 12.28 kW saved qt = (1.2 kgm3) (0.4 m3s){[1 kJ(kg·K)](31.4 – 23) + (0.0164 – 0.0093)(2560 kJ/kg)} = 12.7 kW, which is close to 12.28 kW 7. Plot conditions on psychrometric chart (Figure 51). 8. Obtain data on EATR. (Given: EATR = 5% or 0.05.)

Fig. 50 Total Heat Recovery in Summer (Example 7)

9. Obtain data on OACF. (Given: OACF = 1.05.) 10. Correct the supply air ventilation rate, the moisture transfer rate, and energy transfer rates EATR  0 and OACF  1.0. The net ventilation rate is 400 L/s and the EATR = 0.05; therefore, the actual flow rate Q2 to the space can be obtained from Equation (19) as Qv 400 L/s - = ------------------------- = 421 L/s Q3 = Q2 = ----------------------------------------1 –  EATR  100  1 – 5  100 Because OACF = 1.05, the actual flow rate Q1 of fresh air from outdoor can be calculated from Equation (21) as Q v  OACF   400 L/s   1.05  - = --------------------------------------- = 442 L/s Q1 = Q2(OACF) ----------------------------------------1 –  EATR  100  1 – 5  100 To balance the flow rates into the ERV, the actual air flow rates at states 3 and 4 are as shown in Figure 52.

Fig. 51 Total Energy Recovery with EATR  0 and OACF  1 (Example 8)

Supply and exhaust fan capacity should match the flows required at their locations. Example 9’s results are for balanced flow. Assuming flow rates in the ERV are same as the outdoor air ventilation requirements, then the effectiveness would be same as that for no air leakage. If air leaks at the inlet and outlet of the

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26.36

2020 ASHRAE Handbook—HVAC Systems and Equipment(SI)

Fig. 52

Actual Airflow Rates at Various State Points (Example 8)

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energy recovery ventilator, then the exit conditions of air temperature and humidity at states 3 and 4 can be calculated as those of the airstream mixture. For instance, the properties at state 2 would be those of an airstream mixture at state 2 for no air leakage and air quantity (Q32) at state 3. The error for these calculations should be less than 5%. Shang et al. (2001a) show a method to accurately estimate the energy rates when EATR  0 and OACF  1.

10. A cp C Ce Cp Cr Cs,int CRF h hfg i ITC ms me m· min n NTU p pp P q Q S t tw3 T U V w

= = = = = = = = = = = = = = = = = = = = = = = = = = = = =

SYMBOLS

area of recovery exchanger, m2 specific heat of moist air, kJ/(kg·K) capital cost cost of energy specific heat of air kJ/(kg·K) ratio of Cmin Cmax initial system capital cost capital recovery factor enthalpy, kJ/kg enthalpy of vaporization, kJ/kg arbitrary state, or discount rate income tax credit mass flow rate of supply moist air from outdoors, kg/s mass flow rate of exhaust moist air, kg/s minimum mass flow rate of supply and exhaust airstreams, kg/s number of years in economic consideration, years number of transfer units = UACmin pressure, Pa payback period, years pumping power, W or kW heat transfer rate, W or kW volume flow rate, m3/s reference temperature, K moist air temperature at state i, °C wet-bulb temperature of moist air at state 3, °C absolute temperature, K, or tax overall heat transfer coefficient, W/(m2 ·K) mean velocity, m/s humidity ratio

Greek Letters s = sensible effectiveness of heat or energy wheel L = latent effectiveness of energy wheel t = total effectiveness of ERV  = efficiency  = density, kg/m3  = volume fraction  = relative humidity  = rotational speed of the wheel, rpm Subscripts a = air e = exhaust side of heat/energy exchanger, exit or energy f = fan or fan motor combination if = threshold temperature of the outdoor air for freezing to occur in = indoor conditions of building space inc = increment

h L max min n

= = = = =

o p s t

= = = =

hydraulic latent maximum value minimum value station number indicating supply and exhaust inlets and outlets (see Figure 1) reference state or outlet constant pressure supply side or suction side total

REFERENCES ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal.ashrae .org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org/bookstore. Afshin, M., M. Tardif, and T. Rice. 2019. Frost control strategies for air-to-air energy recovery, ASHRAE Seminar 22, 2019 Annual Conference, Kansas City. AHRI. Forced circulation air-cooling & air-heating coils. AHRI Certification Program 410. AHRI. 2018. Performance rating air-to-air heat exchangers for energy recovery ventilation heat equipment. ANSI/AHRI Standard 1061-2018. Air-Conditioning, Heating, and Refrigeration Institute, Arlington, VA AHRI. 2011. Calculating the efficiency of energy recovery ventilation and its effect on efficiency and sizing of building HVAC systems. AHRI Guideline V-2011. Air-Conditioning, Heating, and Refrigeration Institute, Arlington, VA. Alahmer, A., S. Alsaqoor, and G. Borowski. 2019. Effect of parameters on moisture removal capacity in the desiccant cooling systems. Case Studies in Thermal Engineering 13:100364. ASHRAE. 1982. Symposium on energy recovery from air pollution control. ASHRAE Transactions 88(1):1197-1225. ASHRAE. 2019. Ventilation for acceptable indoor air quality. ANSI/ ASHRAE Standard 62.1-2019. ASHRAE. 2019. Ventilation and acceptable indoor air quality in low-rise residential buildings. ANSI/ASHRAE Standard 62.2-2019. ASHRAE. 2013. Method of testing air-to-air heat/energy exchangers. ANSI/ASHRAE Standard 84-2013. ASHRAE. 2016. Energy standard for buildings except low-rise residential buildings. ANSI/ASHRAE/IES Standard 90.1-2016. Barringer, C.G., and C.A. McGugan. 1989a. Development of a dynamic model for simulating indoor air temperature and humidity. ASHRAE Transactions 95(2):449-460. Barringer, C.G., and C.A. McGugan. 1989b. Effect of residential air-to-air heat and moisture exchangers on indoor humidity. ASHRAE Transactions 95(2):461-474. Besant, R.W., and A.B. Johnson. 1995. Reducing energy costs using runaround systems. ASHRAE Journal 37(2):41-47. Besant, R.W., and C.J. Simonson. 2000. Air-to-air energy recovery. ASHRAE Journal 42(5):31-38. Besant, R.W., and C. Simonson. 2003. Air-to-air exchangers. ASHRAE Journal 45(4):42-50. Buyukalaca, O., and T. Yilmaz. 2002. Influence of rotational speed on effectiveness of rotary-type heat exchanger. Heat and Mass Transfer 38(45):441-447. C2ES. 2015. Center for Climate and Energy Solutions. www.c2es.org. CSA Group. 2018. Laboratory methods of test for rating performance of heat/energy-recovery ventilators. C439-2018. Dhamshala, P. 2016. Modern practices in design of air-conditioning and refrigerated systems. The University of Tennessee, Chattanooga, Graphic Services, Chattanooga. Dhamshala, P., A. Byrd, J. Parker, J. Raines, and K. Gregg. 2019. Design of PV/T panels and modern air-conditioning systems for zero-energy building. Technology Symposium, University of Tennessee, Chattanooga. Dhital, P., R. Besant, and G.J. Schoenau. 1995. Integrating run-around heat exchanger systems into the design of large office buildings. ASHRAE Transactions 101(2):979-999.

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Air-to-Air Energy Recovery Equipment Enteria, N., H. Yoshino, R. Takaki, H. Yonekura, A. Satake, and A. Mochida. 2013. First and second law analysis of the developed solar-desiccant airconditioning system (SDACS) operation during the summer day. Energy and Buildings 60:239-251. Friedlander, M. 2003. How certified ratings can improve system design. Seminar at ASHRAE Winter Annual Meeting, Chicago. Gao, W.Z., Y.P. Cheng, A.G. Jiang, T. Liu, and K. Anderson. 2015. Experimental investigation on integrated liquid desiccant—Indirect evaporative air cooling system utilizing the Maisotesenko cycle. Applied Thermal Engineering 88:288-296. Ge, T.S., Y.J. Dai, R.Z. Wang, and Y. Li. 2015. Performance of two-stage rotary desiccant cooling system with different regeneration temperatures. Energy 81:556-566. Guo, P., D.L. Ciepliski, and R.W. Besant. 1998. A testing and HVAC design methodology for air-to-air heat pipe heat exchangers. International Journal of HVAC&R Research (now Science and Technology for the Built Environment) 4(1):3-26. HVI. 2015. HVI product performance certification procedure. Publication 920. Home Ventilating Institute. Jain, S., and P.K. Bansal. 2007. Performance analysis of liquid desiccant dehumidification systems. International Journal of Refrigeration 30(5): 861-872. John, D.A., and D. Elsberry. 2016. Wrap-around heat pipes in humid climates. ASHRAE Journal 58(11):28. Johnson, A.B., R.W. Besant, and G.J. Schoenau. 1995. Design of multi-coil run-around heat exchanger systems for ventilation air heating and cooling. ASHRAE Transactions 101(2):967-978. Johnson, A.B., C.J. Simonson, and R.W. Besant. 1998. Uncertainty analysis in the testing of air-to-air heat/energy exchangers installed in buildings. ASHRAE Transactions 104(1B):1639-1650. Kays, W.M., and M.E. Crawford, 1993. Convective heat and mass transfer, 3rd ed. McGraw-Hill, New York. Liu, X., T. Zhang, Y. Zheng, and R. Tu. 2016. Performance investigation and exergy analysis of two-stage desiccant wheel systems. Renewable Energy 86:877-888. Mathur, G.D. 1990a. Long-term performance prediction of refrigerant charged flat plate solar collector of a natural circulation closed loop. ASME HTD 157:19-27. Mathur, G.D. 1990b. Indirect evaporative cooling using heat pipe heat exchangers. ASME Symposium, Thermal Hydraulics of Advanced Heat Exchangers, ASME Winter Annual Meeting, Dallas. Mathur, G.D. 1992. Indirect evaporative cooling. Heating/Piping/Air Conditioning 64(4):60-67. Mathur, G.D. 1993. Retrofitting heat recovery systems with evaporative coolers. Heating/Piping/Air Conditioning 65(9):47-51. Mathur, G.D. 1997a. Performance enhancement of existing air conditioning systems. Proceedings of Intersociety Energy Conversion Engineering Conference, Honolulu, American Institute of Chemical Engineers, Paper #97367, pp. 1618-1623. Mathur, G.D. 1997b. Using heat pipe heat exchangers for reducing high energy costs of treating ventilation air. Proceedings of Intersociety Energy Conversion Engineering Conference, American Institute of Chemical Engineers, vol. 2, pp. 1447-1452. Mathur, G.D. 1997c. Predicting yearly energy savings using BIN weather data with heat pipe heat exchangers. Proceedings of Intersociety Energy Conversion Engineering Conference, American Institute of Chemical Engineers, vol. 2, pp. 1391-1396. Mathur, G.D. 1997d. Performance enhancements of existing air conditioning systems. Proceedings of Intersociety Energy Conversion Engineering Conference, American Institute of Chemical Engineers, vol. 3, pp. 1618-1623. Mathur, G.D. 2000. Controlling space humidity with heat pipe heat exchangers. Proceedings of Intersociety Energy Conversion Engineering Conference, American Institute of Chemical Engineers, vol. 2, pp. 835-842. Mathur, G.D., and T.W. McDonald. 1986. Simulation program for a twophase thermosiphon-loop heat exchanger. ASHRAE Transactions 92(2A): 473-485. Mathur, G.D., and T.W. McDonald. 1987. Evaporator performance of finned air-to-air two-phase thermosiphon loop heat exchangers. ASHRAE Transactions 98(2):247-257. McDonald, T.W., and D. Shivprasad. 1989. Incipient nucleate boiling and quench study. Proceedings of CLIMA 2000 1:347-352. Sarajevo, Yugoslavia.

26.37 Moffitt, R. 2003. Personal communication and reference, Trane Application Engineering Manual SYS-APM003-EN. Moffitt, R. 2010. Using energy recovery to improve dehumidification performance and control. Seminar 34, ASHRAE Winter Meeting, Orlando. Moffitt, R. 2011. Energy wheel capacity control. Trane. Mostafa, A.A., and M.M. Mausa. 2007. Heat pipe heat exchanger for heat recovery in air conditioning. Applied Thermal Engineering 27(4):795801. Moffitt, R. 2015. Dedicated outside air system with dual energy recovery used with distributed sensible cooling equipment. ASHRAE Transactions, Paper AT-15-C047. Murphy, J. 2006. Temperature and humidity control in surgery rooms. ASHRAE Journal, June. Phillips, E.G., R.E. Chant, B.C. Bradley, and D.R. Fisher. 1989a. A model to compare freezing control strategies for residential air-to-air heat recovery ventilators. ASHRAE Transactions 95(2):475-483. Phillips, E.G., R.E. Chant, D.R. Fisher, and B.C. Bradley. 1989b. Comparison of freezing control strategies for residential air-to-air heat recovery ventilators. ASHRAE Transactions 95(2):484-490. Phillips, E.G., D.R. Fisher, R.E. Chant, and B.C. Bradley. 1992. Freeze-control strategy and air-to-air energy recovery performance. ASHRAE Journal 34(12):44-49. Pieper, P. 2015. Outside air, economizers, and exhaust air energy recovery. ASHRAE Papers: 2015 ASHRAE Annual Conference, Atlanta, GA. Ruch, M.A. 1976. Heat pipe exchangers as energy recovery devices. ASHRAE Transactions 82(1):1008-1014. Scofield, M., and J.R. Taylor. 1986. A heat pipe economy cycle. ASHRAE Journal 28(10):35-40. Shah, R.K. 1981. Thermal design theory for regenerators. In Heat exchangers: Thermal-hydraulic fundamentals and design. S. Kakec, A.E. Bergles, and F. Maysinger, eds. Hemisphere Publishing, New York. Shang, W., M. Wawryk, and R.W. Besant. 2001a. Air crossover in rotary wheels used for air-to-air heat and moisture recovery. ASHRAE Transactions 107(2). Shang, W., H. Chen, R.W. Evitts, and R.W. Besant. 2001b. Frost growth in regenerative heat exchangers: Part I—Problem formulation and method of solution; Part II—Simulation and discussion. Proceedings of ASME International Mechanical Engineering Congress and Expo, November, New York. Sheng, Y., Y. Zhang, and G. Zhang. 2015. Simulation and energy savings analysis of high temperature heat pump coupling to desiccant wheel air conditioning system. Energy 83:583-596. Simonson, C.J., W. Shang, and R.W. Besant. 2000. Part-load performance of energy wheels: Part I—Speed control. ASHRAE Transactions 106(1): 286-300. Sultan, M., I.I. El-Sharkawy, T. Miyazaki, B.B. Saha, and S. Koyama. 2015. An overview of solid desiccant dehumidification and air conditioning systems. Renewable and Sustainable Energy Reviews 46:16-29. Tu, R., X.-H. Liu, and Y. Jiang. 2015a. Lowering the regeneration temperature of a rotary wheel dehumidification system using exergy analysis. Energy Conversion and Management 89:162-174. Tu, R., Y. Liu, and Y. Jiang. 2015b. Irreversible processes and performance improvement of desiccant wheel dehumidification and cooling systems using exergy. Applied Energy 145:331-344. UNFCCC. 2018. UN climate change annual report. United Nations Framework Convention on Climate Change. Williams, S. 2007. Assessment of liquid desiccant cooling technology. Master’s thesis, The University of Tennessee, Chattanooga. Yau, Y.H. 2007. Application of a heat pipe heat exchanger to dehumidification enhancement in a HVAC system for tropical climates—A baseline performance characteristics study. International Journal of Thermal Sciences 46(2):164-171. Zendehboudi, A., G. Angrisani, and X. Li. 2018. Parametric studies of silica gel and molecular sieve desiccant wheels: Experimental and modeling approaches. International Communications in Heat and Mass Transfer 91:176-186. Zeng, D.Q., H. Li, Y.J. Dai, and A.X. Xie. 2014. Numerical analysis and optimization of solar hybrid one-rotor two-stage desiccant cooling and heating system. Applied Thermal Engineering 73:474-483.

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BIBLIOGRAPHY AHRI. 2014. Selecting, sizing, and specifying packaged air-to-air energy recovery ventilation equipment. AHRI Guideline W. Air-Conditioning, Heating, and Refrigeration Institute, Arlington, VA. Andersson, B., K. Andersson, J. Sundell, and P.A. Zingmark. 1992. Mass transfer of contaminants in rotary enthalpy exchangers. Indoor Air 93(3): 143-148. ASHRAE. 1974. Symposium on heat recovery. ASHRAE Transactions 80(1):302-332. Beccali, M., F. Butera, R. Guanella, and R.S. Adhikari. 2002. Simplified models for the performance evaluation of desiccant wheel dehumidification. International Journal of Energy Research 27(1):17-29. Beccali, M., R.S. Adhikari, F. Butera, and V. Franzitta. 2014. Update on desiccant wheel model. International Journal of Energy Research 28(12):1043-1049. CSA. 1988. Standard methods of test for rating the performance of heatrecovery ventilators. CAN/CSA-C439-88. Canadian Standards Association, Rexdale, ON. Ciepliski, D.L., C.J. Simonson, and R.W. Besant. 1998. Some recommendations for improvements to ASHRAE Standard 84-1991. ASHRAE Transactions 104(1B):1651-1665. De Antonellis, S., M. Intini, and C.M. Joppolo. 2015a. Desiccant wheels effectiveness parameters: Correlations based on experimental data. Energy Buildings 103(15):296-306. De Antonellis, S., M. Intini, C.M. Joppolo, L. Molinaroli, and F. Romano. 2015b. Desiccant wheels for air humidification: An experimental and numerical analysis. Energy Conversion and Management 106:355-364. De Antonellis, S., and C.M. Joppolo. 2017. Simplified models for the evaluation of desiccant wheels performance. Desiccant Heating, Ventilating, and Air Conditioning Systems, pp. 63-85. N. Enteria, H. Awbi, and H. Yoshino, eds. Springer, Berlin. Dhamshala, P., and P. Tangirala. 2015. Use of heat and desiccant wheels in combination with a heat pump to reduce energy cost savings. Master’s thesis, The University of Tennessee, Chattanooga. Dehli, F., T. Kuma, and N. Shirahama. 1993. A new development for total heat recovery wheels. Energy Impact of Ventilation and Air Infiltration, 14th AIVC Conference, Copenhagen, Denmark, pp. 261-268. Inoue, T., and M. Monde. 2009. Operating limit of heat transport in twophase thermosiphon with connecting pipe. International Journal of Heat & Mass Transfer 52:4519-4524. Jouhara, H., and R. Meskimmon. 2010. Experimental investigation of wraparound loop heat pipe heat exchanger used in energy efficient air handling units. Energy 35(12):4592-4599. Jeffus, L. 2003. Refrigeration and air conditioning: An introduction to HVAC/R. Pearson Prentice Hall, Upper Saddle River, NJ. Kodama, D., R.Z. Wang, and Z,Z, Xia, 2006, Desiccant cooling air-conditioning: A review. Renewable and Sustainable Energy Reviews 10 (2):5577. Mathur, G.D. 1990. Indirect evaporative cooling using two-phase thermosiphon loop heat exchangers. ASHRAE Transactions 96(1):1241-1249. Mathur, G.D., and T.W. McDonald. 1987. Evaporator performance of finned air-to-air two-phase thermosiphon loop heat exchangers. ASHRAE Transactions 98(2):247-257. Mathur, G.D. 1996. Enhancing performance of an air conditioning unit system with a two-phase heat recovery loop retrofit. Proceedings of Intersociety Energy Conversion Engineering Conference, vol. 3, pp. 20272032. McDonald, T.W., and D. Shivprasad. 1989. Incipient nucleate boiling and quench study. Proceedings of CLIMA 2000 1:347-352. Sarajevo, Yugoslavia. Mostafa, A.A., and M.M. Mausa. 2007. Heat pipe heat exchanger for heat recovery in air conditioning. Applied Thermal Engineering 27(4):795801.

Ninomura, P.T., and R. Bhargava. 1995. Heat recovery ventilators in multifamily residences in the Arctic. ASHRAE Transactions 101(2):961-966. Noie-Baghban, S.H., and G.R. Majideian. 2000. Waste heat recovery using heat pipe heat exchanger (HPHE) for surgery rooms in hospitals. Applied Thermal Engineering 20(14):1271-1282. Panaras,,G., E. Mathioulakis,, V. Belessiotis, and N. Kyriakis. 2010. Theoretical and experimental investigation of the performance of a desiccant air-conditioning system. Renewable Energy 35(7):1368-1375. Panaras, G., E. Mathioulakis, V. Belessiotis, and N. Kyriakis. 2010. Experimental Validation of a simplified approach for a desiccant wheel model. Energy and Buildings 42(10):1719-1725. Ruivo, C.R, J.J. Costa, A.R. Figueiredo, A. Kodama. 2012. Effectiveness parameters for the prediction of the global performance of desiccant wheels-An assessment based on experimental data, Renewable Energy 38(1):181-187. Ruivo, C.R, A. Carrillo-Andres, J.J. Costa, F. Dominguez-Munoz. A new approach to the effectiveness method for the simulation of desiccant wheels with variable inlet states and airflow rates. Applied Thermal Engineering 58(1-2):670-678. Shang, W., and R.W. Besant. 2001. Energy wheel effectiveness evaluation: Part I—Outlet airflow property distributions adjacent to an energy wheel; Part II—Testing and monitoring energy wheels in HVAC applications. ASHRAE Transactions 107(2). Sheng, Y., Y. Zhang, Y. Sun, L. Fang, J. Nie, and L. Ma. 2014. Experimental analysis and regression prediction of desiccant wheel behavior in high temperature heat pump and desiccant wheel air-conditioning system. Energy and Buildings 80:358-365. Simonson, C.J., and R.W. Besant. 1997. Heat and moisture transfer in desiccant coated rotary energy exchangers: Part I—Numerical model; Part II—Validation and sensitivity studies. International Journal of HVAC&R Research (now Science and Technology for the Built Environment) 3(4): 325-368. Simonson, C.J., and R.W. Besant. 1998. Heat and moisture transfer in energy wheels during sorption, condensation and frosting. ASME Journal of Heat Transfer 120(3):699-708. Simonson, C.J., and R.W. Besant. 1998. Energy wheel effectiveness: Part I—Development of dimensionless groups; Part II—Correlations. International Journal of Heat and Mass Transfer 42(12):2161-2186. Simonson, C.J., D.L. Cieplisky, and R.W. Besant. 1999. Determining performance of energy: Part I—Experimental and numerical methods; Part II—Experimental data and validation. ASHRAE Transactions 105(1): 177-205. SMACNA. 1978. Energy recovery equipment and systems. Report. Sparrow, E.M., J.P. Abraham, and J.C.K. Tong. 2001. An experimental investigation on a mass exchanger for transferring water vapor and inhibiting the transfer of other gases. International Journal of Heat and Mass Transfer 44 (November):4313-4321. Sparrow, E.M., J.P. Abraham, J.C. Tong, and G.L. Martin. 2001. Air-to-air energy exchanger test facility for mass and energy transfer performance. ASHRAE Transactions 107(2):450-456. Stauder, F.A., and T.W. McDonald. 1986. Experimental study of a two-phase thermosiphon-loop heat exchanger. ASHRAE Transactions 92(2A):486497. Stauder, F.A., G.D. Mathur, and T.W. McDonald. 1985. Experimental and computer simulation study of an air-to-air two-phase thermosiphon-loop heat exchanger. ASME 85-WA/HT-15. Wu, X.P., P. Johnson, and A. Akabarzadah. 1997. Application of heat pipe heat exchangers to humidity control in air conditioning systems. Applied Thermal Engineering 17(6):561-568. Yau, Y.H. 2008. The use of a double pipe heat exchanger system for reducing energy consumption of treating ventilation air in an operating theatre—A full year energy consumption model simulation. Energy and Buildings 40(5):917-925.

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AIR-HEATING COILS Coil Construction and Design ................................................ 27.1 Coil Selection .......................................................................... 27.3

Installation Guidelines ............................................................ 27.4 Coil Maintenance .................................................................... 27.5

A

minimum coating designation of G90-U. Some corrosive air conditions may require stainless steel casings or corrosive-resistant coating, such as a baked phenolic applied by the manufacturer to the entire coil surface. Steam coil casings should be designed to accommodate thermal expansion of the tube core during operation (a floating core arrangement). Common core tube diameters vary from 8 up to 25 mm outside diameter (OD) and fin spacings from 1.4 to 6.4 mm. Fluid heating coils have a tube spacing from 20 to 45 mm and tube diameters from 8 to 16 mm OD. Steam coils have tube spacing from 30 to 75 mm and tube diameters from 13 to 25 mm OD. The most common arrangements are one- or two-row steam coils and two- to four-row hot-water coils. Fins should be spaced according to the application requirements, with particular attention given to any severe duty conditions, such as inlet temperatures and contaminants in the airstream. Tube wall thickness and the required use of alloys other than (standard) copper are determined primarily by the coil’s specified maximum allowable working pressure (MAWP) requirements. A secondary consideration is expected coil service life. Fin type, header, and connection construction also play a large part in this determination. All applicable local job site codes and national safety standards should be followed in the design and application of heating coils. Flow direction can strongly affect heat transfer surface performance. In air-heating coils with only one row of tubes, the air flows at right angles to the heating medium. Such a cross-flow arrangement is common in steam heating coils. The steam temperature in the tubes remains uniform, and the mean temperature difference is the same regardless of the direction of flow relative to the air. The steam supply connection is located either in the center or at the top of the inlet header. The steam condensate outlet (return connection) is always at the lowest point in the return header. When coils have two or more tube rows in the direction of airflow, such as hot-water coils, the heating medium in the tubes may be circuited in various parallel-flow and counterflow arrangements. Counterflow is the arrangement most preferred to obtain the highest possible mean temperature difference, which determines the heat transfer of the coil. The greater this temperature difference, the greater the coil’s heat transfer capacity. In multirow coils circuited for counterflow, water enters the tube row on the leaving air side of the coil.

IR-HEATING coils are used to heat air under forced convection. The total coil surface may consist of a single coil section or several coil sections assembled into a bank. The coils described in this chapter apply primarily to comfort heating and air conditioning using steam, hot water, refrigerant vapor heat reclaim (including heat pumps), and electricity. The choice between the various methods of heating depends greatly on the cost of the various available energy sources. For instance, in areas where electric power is cheaply available and heating requirements are limited, heat pumps are a very viable option. With available power and higher heat requirements, electric heat is used. If electric power is considerably expensive, steam or hot water generated using gasfired sources is used in larger buildings and district cooling. In smaller buildings, heat is supplied using gas furnaces, which are covered in Chapters 33 and 34. Water and steam heating are also widely used where process waste heat is available.

1.

COIL CONSTRUCTION AND DESIGN

Extended-surface coils consist of a primary and a secondary heattransfer surface. The primary surface is the external surface of the tubes, generally consisting of rows of round tubes or pipes that may be staggered or parallel (in-line) with respect to the airflow. Flattened tubes or tubes with other nonround internal passageways are sometimes used. The inside of the tube is usually smooth and plain, but some coil designs feature various forms of internal fins or turbulence promoters (either fabricated and then inserted, or extruded) to enhance fluid coil performance. The secondary surface is the fins’ external surface, which consists of thin metal plates or a spiral ribbon uniformly spaced or wound along the length of the primary surface. The intimate contact with the primary surface provides good heat transfer. Air-heating fluid and steam coils are generally available with different circuit arrangements and combinations that offer varying numbers of parallel water flow passes in the tube core. Copper and aluminum are the materials most commonly used for extended-surface coils. Tubing made of steel or various copper alloys is used where corrosive forces might attack the coils from inside or outside. The most common combination for low-pressure applications is aluminum fins on copper tubes. Low-pressure steam coils are usually designed to operate up to 350 kPa (gage). Higherstrength tube materials such as red brass, admiralty brass, or cupronickel assembled by brazed construction are usable up to 186°C water or 1 MPa (gage) saturated steam. Higher operating conditions call for electric welded stainless steel construction, designed to meet Section II and Section VIII requirements of the ASME Boiler and Pressure Vessel Code. Customarily, the coil casing consists of a top and bottom channel (also known as baffles or side sheets), two end supports (also known as end plates or tube sheets), and, on longer coils, intermediate supports (also known as center supports or tube sheets). Designs vary, but most are mounted on ducts or built-up systems. Most often, casing material is spangled zinc-coated (galvanized) steel with a The preparation of this chapter is assigned to TC 8.4, Air-to-Refrigerant Heat Transfer Equipment.

Copyright © 2020, ASHRAE

Steam Coils Steam coils are generally classified, similarly to boilers, by operating pressure: low (100 kPa) or high (>100 kPa). However, various organizations use other pressure classification schemes with differing divisions [e.g., low (100 kPa), medium (100 to 690 kPa), or high (690 kPa). Steam coils can also be categorized by operating limits of the tube materials: Standard steam High-pressure steam

1030 kPa 185°C 1625 kPa 204°C

copper tube special material [e.g., cupronickel (CuNi)]

Although these operating conditions are allowed by code, long exposures to them will shorten coil tube life. Leaks are less likely

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27.2

2020 ASHRAE Handbook—HVAC Systems and Equipment (SI)

Table 1 Preferred Operating Limits for ContinuousDuty Steam Coil Materials in Commercial and Institutional Applications Pressure, kPa

Material

Tube Wall Thickness, mm

35 35 to 100 100 to 200 200 to 345

Copper

345 to 515 515 to 690

Red brass

0.64 0.89

90/10 CuNi

0.89 1.25

690 to 1035 1035 to 1380

0.51 0.64 0.89 1.25

Licensed for single user. © 2020 ASHRAE, Inc.

Note: Red brass and CuNi may be interchanged, depending on coil manufacturer’s specifications.

when the coil tube core has thicker walls of higher-strength materials. Operational experience suggests preferred limits for continuousduty steam coils (tube OD ranging from 16 to 25 mm) in commercial and institutional applications, as shown in Table 1. Steam coils also can be categorized by type as basic steam, steam-distributing, or face-and-bypass. Basic steam coils generally have smooth tubes with fins on the air side. The steam supply connection is at one end and the tubes are pitched toward the condensate return, which is usually at the opposite end. For horizontal airflow, the tubes can be either vertical or horizontal. Horizontal tubes should be pitched within the casing toward the condensate return to facilitate condensate removal. Uniform steam distribution to all tubes is accomplished by careful selection of header size, its connection locations, and positioning of inlet connection distributor plates. Orifices also may be used at the core tube entrances in the supply header. Steam-distributing coils most often incorporate perforated inner tubes that distribute steam evenly along the entire coil. The perforations perform like small steam ejector jets that, when angled in the inner tube, help remove condensate from the outer tube. An alternative design for short coils is an inner tube with no distribution holes, but with an open end. On all coils, supply and return connections can be at the same end or at opposite ends of the coil. For long, low-pressure coils, supply is usually at both ends and the condensate return on one end only. Face-and-bypass steam coils have short sections of steam coils separated by air bypass openings. Airflow through the coil or bypass section is controlled by coil face-and-bypass dampers linked together. As a freeze protection measure, large installations use faceand-bypass steam coils with vertical tubes. For proper performance of all types of steam heating coils, air or other noncondensables in the steam supply must be eliminated. Equally important, condensate from the steam must easily drain from inside the coil. Air vents are located at a high point of the piping and at the coil’s inlet steam header. Whether airflow is horizontal or vertical, the coil’s finned section is pitched toward the condensate return connection end of the coil. Installers must give particular care in the selection and installation of piping, controls, and insulation necessary to protect the coil from freeze-up caused by incomplete condensate drainage. When entering air is at or below 0°C, the steam supply to the coil should not be modulated, but controlled as full on or full off. Coils located in series in the airstream, with each coil sized and controlled to be full on or completely off (in a specific sequence, depending on the entering air temperature), are not as likely to freeze. Temperature control with face-and-bypass dampers is also common. During part-load conditions, air is bypassed around the steam coil with full steam flow to the coil. In a face-and-bypass arrangement, highvelocity streams of freezing air must not impinge on the coil when

the face dampers are partially closed. The section on Overall Requirements in this chapter and the section on Heating Coils in Chapter 47 of the 2019 ASHRAE Handbook—HVAC Applications have more details.

Water/Aqueous Glycol Heating Coils Normal-temperature hot-water heating coils can be categorized as booster coils or standard heating coils. Booster (duct-mounted or reheat) coils are commonly found in variable-air-volume systems. They are one or two rows deep, have minimal water flow, and provide a small air temperature rise. Casings can be either flanged or slip-and-drive construction. Standard heating coils are used in runaround systems, makeup air units, and heating and ventilating systems. All use standard construction materials of copper tube and aluminum fins. High-temperature water coils may operate with up to 200°C water, with pressures comparable or somewhat higher than the saturated vapor temperature of the water supply. The temperature drop across the coil may be as high as 85 K. To safely accommodate these fluid temperatures and thermal stresses, the coil requires industrialgrade construction that conforms to applicable boiler and safety codes. These requirements should be listed in detail by the specifying engineer, along with the inspection and certification requirements and a compliance check before coil installation and operation. Proper water coil performance depends on eliminating air and on good water distribution in the coil and its interconnecting piping. Unless properly vented, air may accumulate in the coil circuits, which reduces heat transfer and possibly causes noise and vibration in the pipes. For this reason, water coils should be constructed with self-venting, drainable circuits. The self-venting design is maintained by field-connecting the water supply connection to the bottom and the water return connection to the top of the coil. Ideally, water is supplied at the bottom, flows upward through the coil, and forces any air out the return connection. Complete fluid draining at the supply connection indicates that coils are self draining and without air or water traps. Such a design ensures that the coil is always filled with water, and it should completely drain when it is required to be empty. Most manufacturers provide vent and drain fittings on the supply and return headers of each water coil. When water does get trapped in the coil core, it is usually caused by a sag in the coil core or by a nondraining circuit design. During freezing periods, even a small amount of water in the coil core can rupture a tube. Also, such a static accumulation of either water or glycol can corrode the tube over an extended period. Large multirow, multicircuited coils may not drain rapidly, even with self-draining circuitry; if they are not installed level, complete self draining will not take place. This problem can be prevented by including intermediate drain headers and installing the coil so that it is pitched toward the connections. To produce desired ratings without excessive water pressure drop, manufacturers use various circuit arrangements. A single-feed serpentine circuit is commonly used on booster coils with low water flows. With this arrangement, a single feed carrying the entire water flow makes a number of passes across the airstream. The more common circuit arrangement is called a full row feed or standard circuit. With this design, all the core tubes of a row are fed with an equal amount of water from the supply header. Others, such as quarter, half, and double-row feed circuit arrangements, may be available, depending on the total number of tubes and rows of the coil. Uniform flow in each water circuit is obtained by designing each circuit's length as equal to the other as possible. Generally, higher velocity provides greater capacity and more even discharge air temperature across the coil face, but with diminishing returns. To prevent erosion, do not exceed 1.8 m/s for copper coils. At higher velocities, only modest gains in capacity can be

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Air-Heating Coils

27.3

achieved at increasingly higher pumping power penalties. Above 2.4 m/s, any gain is negligible. Velocities with fluid flow Reynolds numbers (Re) between 2000 and 10 000 fall into a transition range where heat transfer capacity predictions are less likely to be accurately computed. Below Re = 2000, flow is laminar, where heat transfer prediction is again reliable, but coil capacity is greatly diminished and tube fouling can become a problem. For further insight on the transition flow effect on capacity, refer to Figure 16 of Air-Conditioning, Heating, and Refrigeration Institute (AHRI) Standard 410. Methods of controlling water coils to produce a uniform exit air temperature are discussed in Chapter 48 of the 2019 ASHRAE Handbook—HVAC Applications. In some cases, the hot water circulated may contain a considerable amount of sand and other foreign matter such as minerals. This matter should be filtered from the water circuit. Additionally, some coil manufacturers offer removable water header boxes (some are plates), or a removable plug at the return bends of each tube, allowing the tubes to be rodded clean. In an area where build-up of scale or other deposits is expected, include a fouling factor when computing heating coil performance. Hot-water coil ratings (AHRI Standard 410) include a 44 (mm2 ·K)/W fouling factor. Cupronickel, red brass, admiralty, stainless steel, and other tube alloys can protect against corrosion and erosion, which can be common in hot-water/glycol systems.

Volatile Refrigerant Heat Reclaim Coils A heat reclaim coil with a volatile refrigerant can function as a condenser either in series or parallel with the primary condenser of a refrigeration system. Heat from condensing or desuperheating vapor warms the airstream. It can be used as a primary source of heat or to assist some other form of heating, such as reheat for humidity control. In the broad sense, a heat reclaim coil functions at half the heat-dissipating capacity of its close-coupled refrigerant system’s condenser. Thus, a heat reclaim coil should be (1) piped to be upstream from the condenser and (2) designed with the assumption that some condensate must be removed from the coil. For these reasons, locate the coil outlet at the lowest point of the coil and trap it if this location is lower than the condenser inlet. Heat reclaim coils are normally circuited for counter-flow of air and refrigerant. However, most supermarket heat reclaim coils are two rows deep and use a cross-flow design. The section on AirCooled Condensers in Chapter 39 has additional information on this topic. Because refrigerant heat reclaim involves specialized heating coil design, a refrigerant equipment manufacturer is the best source for information on the topic.

Electric Heating Coils An electric heating coil consists of a length of resistance wire (commonly nickel/chromium) to which a voltage is applied. The resistance wire may be bare or sheathed in an electrically insulating layer, such as magnesium oxide, and compacted inside a finned steel tube. Sheathed coils are more expensive, have a higher air-side pressure drop, and require more space. A useful comparison for sizing is a heat transfer capacity of 130 kW/m2 of face area compared to 320 kW/m2 for bare resistance wire coils. However, the outer surface temperature of sheathed coils is lower, the coils are mechanically stronger, and contact with personnel or housing is not as dangerous. Coils with sheathed heating elements having an extended finned surface are generally preferred (1) for dust-laden atmospheres, (2) where there is a high probability of maintenance personnel contact, or (3) downstream from a dehumidifying coil that might have moisture carryover. Manufacturers can provide further information, including selection recommendations, applications, and maintenance instructions.

2.

COIL SELECTION

The following factors should be considered in coil selection: • Required duty or capacity considering other components • Temperature of air entering the coil and air temperature rise • Available heating media’s operating and maximum pressure(s) and temperature(s) • Space and dimensional limitations • Air volume, speed, distribution, and limitations • Heating media volume, flow speed, distribution, and limitations • Permissible flow resistances for both the air and heating media • Characteristics of individual designs and circuit possibilities • Individual installation requirements, such as type of control and material compatibility • Specified and applicable codes and standards regulating the design and installation. Load requirements are discussed in Chapters 17 and 18 of the 2017 ASHRAE Handbook—Fundamentals. Much is based on the choice of heating medium, as well as operating temperatures and core tube diameter. Also, proper selection depends on whether the installation is new, being modified, or a replacement. Dimensional fit is usually the primary concern of modified and replacement coils; heating capacity is often unknown. Air quantity is regulated by factors such as design parameters, codes, space, and size of the components. Resistance through the air circuit influences fan power and speed. This resistance may be limited to allow use of a given size fan motor or to keep operating expenses low, or because of sound level requirements. All of these factors affect coil selection. The air friction loss across the heating coil (summed with other series air pressure drops for system component such as air filters, cooling coils, grilles, and ductwork) determines the static pressure requirements of the complete air system. See Chapter 21 for selecting the fan component. Permissible resistance through the water or glycol coil circuitry may be dictated by the available pressure from a given size pump and motor. This is usually controlled within limits by careful selection of coil header size and the number of tube circuits. Additionally, the adverse effect of high fluid velocity in contributing to erosion/corrosion of the tube wall is a major factor in selecting tube diameter and the circuit. Heating coil performance depends on the correct choice of original equipment and proper application and installation. For steam coils, proper performance relies first on selecting the correct type of steam coil, and then the proper size and type of steam trap. Properly sized connecting refrigerant lines, risers, and traps are critical to heat reclaim coils. Heating coil thermal performance is relatively simple to derive. It only involves a dry-bulb temperature and sensible heat, without the complications of latent load and wet-bulb temperature for dehumidifying cooling performance. Even simpler, consult coil manufacturers’ catalogs for ratings and selection. Most manufacturers provide computerized coil selection and rating programs on request; some are certified accurate within 5% of an application parameter, representative of a normal application range. Many manufacturers participate in the AHRI Coil Certification Program, which approves application ratings that conform to all AHRI Standard 410 requirements, based on qualifying testing to ASHRAE Standard 33.

Coil Ratings Coil ratings are based on uniform face velocity. Nonuniform airflow may be caused by the system, such as air entering at odd angles, or by inadvertent blocking of part of the coil face. To obtain rated performance, the airflow quantity in the field must correspond to the design requirements and the velocity vary no greater than 20% at any point across the coil face.

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The industry-accepted method of coil rating is outlined in AHRI Standard 410. Test requirements for determining standard coil ratings are specified in ASHRAE Standard 33. AHRI application ratings are derived by extending the ASHRAE standard rating test results for other operating conditions, coil sizes, row depths, and fin count for a particular coil design and arrangement. Steam, water, and glycol heating coils are rated within the following limits (Table 1 of AHRI Standard 410), which may be exceeded for special applications. Air face velocity. 1 to 8 m/s, based on air density of 1.2 kg/m3 Entering air temperature. Steam coils: –29 to 38°C Water coils: –18 to 38°C Steam pressure. 14 to 1723 kPa (gage) at coil steam supply connection (pressure drop through steam control valve must be considered) Fluid temperatures. Water: 49 to 121°C Ethylene glycol and propylene glycol: –18 to 93°C Fluid velocities. Water: 0.1 to 2.4 m/s Ethylene glycol and propylene glycol: 0.1 to 1.8 m/s

Table 2 Tube Outside Diameter 16 mm 25 mm

Typical Maximum Condensate Loads Maximum Allowable Condensate Load, g/s Basic Coil 8.6 21.2

Steam Distributing Coil 5.0 12.0

applies here. For steam coils, the heat transfer coefficient of condensing steam must be calculated (see Chapter 4 of the 2017 ASHRAE Handbook—Fundamentals). For estimating the pressure drop of condensing steam, see Chapter 5 of the 2017 ASHRAE Handbook—Fundamentals. Parametric Effects. The heat transfer performance of a given coil can be changed by varying the airflow rate and/or the temperature of the heating medium, both of which are relatively linear. Understanding the interaction of these parameters is necessary for designing satisfactory coil capacity and control. A review of manufacturers’ catalogs and selection programs, many of which are listed in AHRI’s Directory of Certified Product Performance and Forced-Circulation Air-Cooling and Air-Heating Coils, shows the effects of varying these parameters.

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Overall Requirements Individual installations vary widely, but the following values can be used as a guide. The air face velocity is usually between 2.5 and 5 m/s. Delivered air temperature varies from about 22°C for ventilation only to about 65°C for complete heating. Steam pressure typically varies from 15 to 100 kPa (gage), with 35 kPa (gage) being the most common. A minimum steam pressure of 35 kPa (gage) is recommended for systems with entering air temperatures below freezing. Hot-water (or glycol) temperature for comfort heating is commonly between 80 and 95°C, with water velocities between 1.2 and 1.8 m/s. For high-temperature water, water temperatures can be over 200°C with operating pressures of 100 to 170 kPa over saturated water temperature. Water quantity is usually based on about 11 K temperature drop through the coil. Air resistance is usually limited to 100 to 150 Pa for commercial buildings and to about 250 Pa for industrial buildings. High-temperature water systems commonly have a water temperature between 150 and 200°C, with up to 80 K drop through the coil. Steam coils are selected with dry steam velocities not exceeding 30 m/s and with acceptable condensate loading per coil core tube depending on the type of steam coil. Table 2 shows some typical maximum condensate loads. Steam coil performance is maximized when the supply is dry, saturated steam, and condensate is adequately removed from the coil and continually returned to the boiler. Although steam quality may not significantly affect the coil’s heat transfer, the back-up effect of too rapid a condensate rate, augmented by a wet supply stream, can cause a slug of condensate to travel through the coil and condensate return. This situation can result in noise and possible damage. Complete mixing of return and outdoor air is essential to proper coil operation. The design of the air mixing damper or ductwork connection section is critical to the proper operation of a system and its air temperature delivery. Systems in which the air passes through a fan before flowing through a coil do not ensure proper air mixing. Dampers at the inlet air face of a steam coil should be the opposedblade type, which are better than in-line blades for controlling air volume and reducing individual blade-directed cold airstreams when modulating in low-heat mode. Heat Transfer and Pressure Drops. For air-side heat transfer and pressure drop, the information in Chapter 23 for sensible cooling coils is applicable. For water (or glycol) coils, Chapter 23’s information on water-side heat transfer and pressure drop also

3.

INSTALLATION GUIDELINES

Steam systems designed to operate at outdoor air temperatures below 0°C should be different from those designed to operate above it. Below 0°C, the steam air-heating system should be designed as a preheat and reheat pair of coils. The preheat coil functions as a nonmodulating basic steam coil, which requires full steam pressure whenever the outdoor temperature is below freezing. The reheat coil, typically a modulated steam-distributing coil, provides the heating required to reach the design air temperature. Above 0°C, the heating coil can be either a basic or steam-distributing type as needed for the duty. When the leaving air temperature is controlled by modulating steam supply to the coil, steam-distributing tube coils provide the most uniform exit air temperature (see the section on Steam Coils). Correctly designed steam-distributing tube coils can limit the exit air temperature stratification to a maximum of 3 K over the entire length of the coil, even when steam supply is modulated to a fraction of full-load capacity. Low-pressure steam systems and coils controlled by modulating steam supply should have a vacuum breaker or be drained through a vacuum-return system to ensure proper condensate drainage. It is good practice to install a closed vacuum breaker (where required) connected to the condensate return line through a check valve. This unit breaks the vacuum by equalizing the pressure, yet minimizes the possibility of air bleeding into the system. Steam traps should be located at least 300 mm below the condensate outlet to allow the coil to drain properly. Also, coils supplied with low-pressure steam or controlled by modulating steam supply should not be trapped directly to an overhead return line. Condensate can be lifted to overhead returns only when enough pressure is available to overcome the condensate head and any return line pressure. If overhead returns are necessary, the condensate must be pumped to the higher elevation (see Chapter 11). Water coils for air heating generally have horizontal tubes to avoid air pockets. Where water or glycol coils may be exposed to a freezing condition, drainability must be considered. If a coil is to be drained and then exposed to below-freezing temperatures, it should first be flushed with a nonfreeze solution. To minimize the danger of freezing in both steam and waterheating coils, the outdoor air inlet dampers usually close automatically when the fan is stopped (system shutdown). In steam systems with very cold outdoor air conditions (e.g., –30°C or below), it is desirable to fully open the steam valve when the system is shut

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Air-Heating Coils

27.5

down. If outdoor air is used for proportioning building makeup air, the outdoor air damper should be an opposed-blade design. Heating coils are designed to allow for expansion and contraction resulting from the temperature ranges in which they operate. Be careful not to impose strains from the piping to the coil connections, particularly on high-temperature hot-water applications. Expansion loops, expansion or three-elbow swing joints, or flexible connections usually provide the needed protection (see Chapter 11). Good practice supports banked coils individually in an angleiron frame or a similar supporting structure. With this arrangement, the lowest coil is not required to support the mass of the coils stacked above. This design also facilitates removing individual coils in a multiple-coil bank for repair or replacement. Heat reclaim coils depend on a closely located, readily available source of high-side refrigerant vapor. For example, supermarket rack compressors and the air handler’s coil section should be installed close to the store’s motor room. Most commonly, the heat reclaim coil section is piped in series with the rack’s condenser and sized for 50% heat extraction. Heat reclaim in supermarkets is discussed in Chapter 2 of the 2019 ASHRAE Handbook—HVAC Applications; also see Chapter 15 of the 2018 ASHRAE Handbook—Refrigeration. In heat pump applications, the heating coil is usually the indoor coil that is used for cooling in the summer. The heat pump system is normally optimized for cooling load and efficiency. Heating performance is a by-product of cooling performance, and any unsatisfied heating requirement is met using electric heating coils. The refrigerant circuiting is also specialized and involves driving the refrigerant in the reverse direction from the cooling mode. Heat pump coils also have an additional check valve bypass to the expansion device for operating in heat pump mode. For additional details about heat pump coil design and operation, see Chapters 9 and 49.

4.

COIL MAINTENANCE

Both internal and external surfaces must be clean for coils to deliver their full rated capacity. The tubes generally stay clean in glycol systems and in adequately maintained water systems. Should scale be detected in the piping where untreated water is used, chemical or mechanical cleaning of the internal tube surfaces is required. The need for periodic descaling can be minimized by proper boiler water treatment and deaeration. Internal coil maintenance consists primarily of preventing scale and corrosion in the coil core tubes and piping of potable-waterheating (including steam) systems. In its simplest form, this involves removing dissolved oxygen, maintaining deionized water, and controlling boiler water pH. Boiler water can be deaerated mechanically. Vacuum deaerating simultaneously removes oxygen and carbon dioxide. The last traces of oxygen can be removed

chemically by adding sodium sulfide. For steam coils, 100% dry steam contains 0% air. Good boiler water results in the absence of oxygen and a pH maintained at 10.5. If this is not practicable, a pH of 7 to 9 with a corrosion inhibitor is recommended. Because calcium carbonate is less soluble in water at higher pH, the inhibitor most often used for this purpose is sodium nitrate. Usually, the requirements for chemical treatment increase as the temperature of the coil’s return flow drops. With few exceptions, boiler water chemical treatment programs use only proportional feeding, which is the recommended way of maintaining a constant concentration at all times. Periodic batch or slug feeding of water treatment chemicals is not an accurate way to treat boiler water, particularly if the system has a high water makeup rate. Only use chemicals known to be compatible with the coil’s tube and connection metal(s). Chemical treatment of boiler water is complicated and has environmental effects. For this reason, it is important to consult a water treatment specialist to establish a proper boiler water treatment program. For further information on boilers, see Chapter 32. The finned surface of heating coils can sometimes be brushed and cleaned with a vacuum cleaner. Coils are commonly surfacecleaned annually using pressurized hot water containing a mild detergent. Reheat coils that contain their refrigerant charge should never be cleaned with a spray above 65°C. In extreme cases of neglect, especially in restaurants where grease and dirt have accumulated, coil(s) may need to be removed to completely clean off the accumulation with steam, compressed air and water, or hot water containing a suitable detergent. Pressurized cleaning is more thorough if first done from the coil’s air leaving side, before cleaning from the coil’s air entry side. Often, outdoor makeup air coils have no upstream air filters, so they should be visually checked on a frequent schedule. Overall, coils should be inspected and serviced regularly. Visual observation should not be relied on to judge cleaning requirements for coils greater than three rows deep because airborne dirt tends to pack midway through the depth of the coil.

REFERENCES AHRI. 2001. Forced-circulation air-cooling and air-heating coils. Standard 410-2001. Air-Conditioning, Heating, and Refrigeration Institute, Arlington, VA. AHRI. Semiannually. Forced-circulation air-cooling and air-heating coils. Air-Conditioning, Heating, and Refrigeration Institute, Arlington, VA. www.ahridirectory.org. AHRI. Semiannually. Directory of certified air-conditioning product performance. Air-Conditioning, Heating, and Refrigeration Institute, Arlington, VA. www.ahridirectory.org. ASHRAE. 2000. Methods of testing forced circulation air cooling and air heating coils. Standard 33-2000. ASME. 2015. Boiler and pressure vessel code. American Society of Mechanical Engineers, New York.

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Related Commercial Resources CHAPTER 28

UNIT VENTILATORS, UNIT HEATERS, AND MAKEUP AIR UNITS Unit Ventilators.................................................................................................................................. 1 Unit Heaters ....................................................................................................................................... 4 Makeup Air Units ............................................................................................................................... 9

1.

UNIT VENTILATORS

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A

HEATING unit ventilator is an assembly whose principal functions are to heat, ventilate, and cool a space by introducing outdoor air in quantities up to 100% of its rated capacity. The heating medium may be steam, hot water, gas, or electricity. The essential components of a heating unit ventilator are the fan, motor, heating element, damper, filter, automatic controls, and outlet grille, all of which are encased in a housing. An air-conditioning unit ventilator is similar to a heating unit ventilator; however, in addition to the normal winter function of heating, ventilating, and cooling with outdoor air, it is also equipped to cool and dehumidify during the summer. It is usually arranged and controlled to introduce a fixed quantity of outdoor air for ventilation during cooling in mild weather. The air-conditioning unit ventilator may be provided with a various of combinations of heating and airconditioning elements. Some of the more common arrangements include • • • • •

Combination hot- and chilled-water coil (two-pipe) Separate hot- and chilled-water coils (four-pipe) Hot-water or steam coil and direct-expansion coil Electric heating coil and chilled-water or direct-expansion coil Gas-fired furnace with direct-expansion coil

The typical unit ventilator has controls that allow heating, ventilating, and cooling to be varied while the fans operate continuously. In normal operation, the discharge air temperature from a unit is varied in accordance with the room requirements. The heating unit ventilator can provide ventilation cooling by bringing in outdoor air whenever the room temperature is above the room set point. Air-conditioning unit ventilators can provide refrigerated cooling when the outdoor air temperature is too high to be used effectively for ventilation cooling. Unit ventilators are available for floor mounting, ceiling mounting, and recessed applications. They are available with various airflow and capacity ratings, and the fan can be arranged so that air is either blown through or drawn through the unit. With direct-expansion refrigerant cooling, the condensing unit can either be furnished as an integral part of the unit ventilator assembly or be remotely located. Figure 1A shows a typical heating unit ventilator. The heating coil can be hot water, steam, or electric. Hot-water coils can be provided with face-and-bypass dampers for capacity control, if desired. Valve control of capacity is also available. Figure 1B shows a typical air-conditioning unit ventilator with a combination hot- and chilled-water coil for use in a two-pipe system. The preparation of the sections on Unit Ventilators and Unit Heaters is assigned to TC 6.1, Hydronic and Steam Equipment and Systems. The preparation of the section on Makeup Air Units is assigned to TC 5.8, Industrial Ventilation.

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This type of unit is usually provided with face-and-bypass dampers for capacity control. Figure 1C illustrates a typical air-conditioning unit ventilator with two separate coils, one for heating and the other for cooling with a four-pipe system. The heating coil may be hot water, steam, or electric. The cooling coil can be either a chilled-water coil or a direct-expansion refrigerant coil. Heating and cooling coils are sometimes combined in a single coil by providing separate tube circuits for each function. In such cases, the effect is the same as having two separate coils. Figure 1D illustrates a typical air-conditioning unit ventilator with a fan section, a gas-fired heating furnace section, and a directexpansion refrigerant coil section.

Application Unit ventilators are used primarily in schools, meeting rooms, offices, and other areas where the density of occupancy requires controlled ventilation to meet local codes. Floor-model unit ventilators are normally installed on an outer wall near the centerline of the room. Ceiling models are mounted against either the outer wall or one of the inside walls. Ceiling models discharge air horizontally. Best results are obtained if the unit can be placed so that the airflow is not interrupted by ceiling beams or surface-mounted lighting fixtures. Downdraft can be a problem in classrooms with large window areas in cold climates. Air in contact with the cold glass is cooled and flows down into the occupied space. Floor-standing units often include one of the following provisions to prevent downdraft along the windows (Figure 2): • Window sill heating uses finned radiators of moderate capacity installed along the wall under the window area. Heated air rises upward by convection and counteracts the downdraft by tempering it and diverting it upward. • Window sill recirculation is achieved by installing the return air intake along the window sill. Room or return air to the unit includes the window’s cold downdrafts, thereby removing them from the occupied area. • Window sill discharge directs a portion of the unit ventilator discharge air into a delivery duct along the sill of the window. The discharge air, delivered vertically at the window sill, is distributed throughout the room, and the upwardly directed air combats downdraft.

Selection Items to be considered in the application of unit ventilators are • • • • •

Unit air capacity Percent minimum outdoor air Heating and cooling capacity Cycle of control Location of unit

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Fig. 1 Typical Unit Ventilators

Fig. 2 Methods of Preventing Downdraft along Windows

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Unit Ventilators, Unit Heaters, and Makeup Air Units Mild-weather cooling capacity and number of occupants in the space are the primary considerations in selecting the unit’s air capacity. Other factors include state and local requirements, volume of the room, density of occupancy, and use of the room. The number Table 1 Typical Unit Ventilator Capacities Airflow, L/s 240 360 480 600 720

Heating Unit Ventilator Total Heating Capacity, kW

A/C Unit Ventilator Total Cooling Capacity, kW

10.7 15.6 20.4 25.3 30.2

5.6 8.4 11.2 14.0 16.8

28.3 qv  cp Q ti to

= = = = = =

heat required to heat ventilating air, W density of air at standard conditions = 1.2 kg/m3 air specific heat = 1.0 kJ/(kg·K) ventilating airflow, L/s required room air temperature, °C outdoor air temperature, °C qv = 1.2  1.0  600  (20/100)[21 – (–18)] = 5600 W

Total heating requirement: qt = q v + qs where qt = total heat requirement, W qs = heat required to make up heat losses, W qt = 5.6 + 7.0 = 12.6 W

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Ventilation cooling capacity:

of air changes required for a specific application also depends on window area, orientation, and maximum outdoor temperature at which the unit is expected to prevent overheating. Rooms oriented to the north (in the northern hemisphere) with small window areas require about 6 air changes per hour (ach). About 9 ach are required in rooms oriented to the south that have large window areas. As many as 12 ach may be required for very large window areas and southern exposures. These airflows are based on preventing overheating at outdoor temperatures up to about 13°C. For satisfactory cooling at outdoor air temperatures up to 15°C, airflow should be increased accordingly. These airflows apply principally to classrooms. Factories and kitchens may require 30 to 60 ach (or more). Office areas may need 10 to 15 ach. The minimum amount of outdoor air for ventilation is determined after the total air capacity has been established. It may be governed by local building codes, or it may be calculated to meet the ventilating needs of the particular application. For example, ASHRAE Standard 62.1 requires 3.8 to 5 L/s of outdoor air per occupant [0.3 to 0.9 L/(s·m2)] in lecture halls or classrooms, laboratories, and cafeterias, and 2.5 L/s per occupant in conference rooms. The heating and cooling capacity of a unit to meet the heating requirement can be determined from the manufacturer’s data. Heating capacity should always be determined after selecting the unit air capacity for mild-weather cooling. Capacity. Manufacturers publish the heating and cooling capacities of unit ventilators. Table 1 lists typical nominal capacities. Heating Capacity Requirements. Because a unit ventilator has a dual function of introducing outdoor air for ventilation and maintaining a specified room condition, the required heating capacity is the sum of the heat required to bring outdoor ventilation air to room temperature and the heat required to offset room losses. The ventilation cooling capacity of a unit ventilator is determined by the air volume delivered by the unit and the temperature difference between the unit discharge and the room temperature. Example. A room has a heat loss of 7.0 kW at a winter outdoor design condition of –18°C and an indoor design of 21°C, with 20% outdoor air. Minimum air discharge temperature from the unit is 15°C. To obtain the specified number of air changes, a 600 L/s unit ventilator is required. Determine the ventilation heat requirement, the total heating requirement, and the ventilation cooling capacity of this unit with outdoor air temperature below 15°C. Solution: Ventilation heat requirement: qv = cpQ(ti – to) where

qc = cpQ(ti – tf ) where qc = ventilation cooling capacity of unit, W tf = unit discharge air temperature, °C qc = 1.2  1.0  600(21 – 15) = 4320 W

Control Many cycles of control are available. The main difference in the various cycles is the amount of outdoor air delivered to the room. Usually, a room thermostat simultaneously controls both a valve and either damper or step controller, to regulate the heat supply, and an outdoor and return air damper. A thermostat in the airstream prevents discharge of air below the desired minimum temperature. Unit ventilator controls provide the proper sequence for the following stages: • Warm-up. All control cycles allow rapid warm-up by having the units generate full heat with the outdoor damper closed. Thus 100% of the room air is recirculated and heated until the room temperature approaches the desired level. • Heating and ventilating. As the room temperature rises into the thermostat’s operating range, the outdoor air damper partially or completely opens to provide ventilation, depending on the cycle used. Auxiliary heating equipment is shut off. As the room temperature continues to rise, the unit ventilator heat supply is throttled. • Cooling and ventilating. When the room temperature rises above the desired level, the room thermostat throttles the heat supply so that cool air flows into the room. The thermostat gradually shuts off the heat and then opens the outdoor air damper. The airstream thermostat frequently takes control during this stage to keep the discharge temperature from falling below a set level. There are three cycles of control commonly used for unit ventilators: • Cycle I. 100% outdoor air is admitted at all times, except during warm-up. • Cycle II. A minimum amount of outdoor air (normally 20 to 50%) is admitted during the heating and ventilating stage. This percentage is gradually increased to 100%, if needed, during the ventilation cooling stage. • Cycle III. Except during warm-up, a variable amount of outdoor air is admitted, as needed, to maintain a fixed temperature of air entering the heating element. The amount of air admitted is controlled by the airstream thermostat set low enough (often at 13°C) to provide cooling when needed. Air-conditioning unit ventilators can include any of the three cycles in addition to the mechanical cooling stage, where a fixed

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amount of outdoor air is introduced. The cooling capacity is controlled by the room thermostat. For maximum heating economy, the building temperature is reduced at night and during weekends and vacations. Several arrangements are used to accomplish this. One arrangement takes advantage of the natural convective capacity of the unit when the fans are off. This capacity is supplemented by cycling the fan with the outdoor damper closed as required to maintain the desired room temperature.

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2.

UNIT HEATERS

A unit heater is an assembly of elements with the main function of heating a space. The essential elements are a fan and motor, a heating element, and an enclosure. Filters, dampers, directional outlets, duct collars, combustion chambers, and flues may also be included. Some types of unit heaters are shown in Figure 3. Unit heaters can usually be classified in one or more of the following categories: • Heating medium. Media include (1) steam, (2) hot water, (3) gas indirect-fired, (4) oil indirect-fired, and (5) electric heating. • Type of fan. Three types of fans can be considered: (1) propeller, (2) centrifugal, and (3) remote air mover. Propeller fan units may be arranged to blow air horizontally (horizontal blow) or vertically (downblow). Units with centrifugal fans may be small cabinet units or large industrial units. Units with remote air movers are known as duct unit heaters. • Arrangement of elements. Two types of units can be considered: (1) draw-through, in which the fan draws air through the unit; and (2) blow-through, in which the fan blows air through the heating element. Indirect-fired unit heaters are always blow-through units.

Application Unit heaters have the following principal characteristics: • Relatively large heating capacities in compact casings • Ability to project heated air in a controlled manner over a considerable distance • Relatively low installed cost per unit of heat output • Application where an elevated sound level is permissible They are, therefore, usually placed in applications where the heating capacity requirements, physical volume of the heated space, or both, are too large to be handled adequately or economically by other means. By eliminating extensive duct installations, the space is freed for other use. Unit heaters are mostly used for heating commercial and industrial structures such as garages, factories, warehouses, showrooms, stores, and laboratories, as well as corridors, lobbies, vestibules, and similar auxiliary spaces in all types of buildings. Unit heaters may often be used to advantage in specialized applications requiring spot or intermittent heating, such as at outer doors in industrial plants or in corridors and vestibules. Cabinet unit heaters may be used where heated air must be filtered. Unit heaters may be applied to a number of industrial processes (e.g., drying and curing) that use heated air in rapid circulation with uniform distribution. They may be used for moisture absorption applications, such as removing fog in dye houses, or to prevent condensation on ceilings or other cold surfaces of buildings where process moisture is released. When such conditions are severe, unit ventilators or makeup air units may be required.

Selection The following factors should be considered when selecting a unit heater: • Heating medium to be used • Type of unit

• • • •

Location of unit for proper heat distribution Permissible sound level Rating of the unit Need for filtration

Heating Medium. The proper heating medium is usually determined by economics and requires examining initial cost, operating cost, and conditions of use. Steam or hot-water unit heaters are relatively inexpensive but require a boiler and piping system. The unit cost of such a system generally decreases as the number of units increases. Therefore, steam or hot-water heating is most frequently used (1) in new installations involving a relatively large number of units, and (2) in existing systems that have sufficient capacity to handle the additional load. High-pressure steam or high-temperature hot-water units are normally used only in very large installations or when a high-temperature medium is required for process work. Lowpressure steam and conventional hot-water units are usually selected for smaller installations and for those concerned primarily with comfort heating. Gas and oil indirect-fired unit heaters are frequently preferred in small installations where the number of units does not justify the expense and space requirements of a new boiler system or where individual metering of the fuel supply is required, as in a shopping center. Gas indirect-fired units usually have either horizontal propeller fans or industrial centrifugal fans. Oil indirect-fired units largely have industrial centrifugal fans. Some codes limit the use of indirect-fired unit heaters in some applications. Indirect-fired oil and gas units are of blow-through design to mitigate the possibility of combustion products entering the occupied space. Electric unit heaters are used when the cost of available electric power is lower than that of alternative fuel sources and for isolated locations, intermittent use, supplementary heating, or temporary service. Typical applications are ticket booths, security offices, factory offices, locker rooms, and other isolated rooms scattered over large areas. Electric units are particularly useful in isolated and untended pumping stations or pits, where they may be thermostatically controlled to prevent freezing. Type of Unit. Propeller fan units are generally used in nonducted applications where the heating capacity and distribution requirements can best be met by units of moderate output and where heated air does not need to be filtered. Horizontal-blow units are usually installed in buildings with low to moderate ceiling heights. Downblow units are used in spaces with high ceilings and where floor and wall space limitations dictate that heating equipment be kept out of the way. Downblow units may have an adjustable diffuser to vary the discharge pattern from a high-velocity vertical jet (to achieve the maximum distance of downward throw) to a horizontal discharge of lower velocity (to prevent excessive air motion in the zone of occupancy). Revolving diffusers are also available. Cabinet unit heaters are used when a more attractive appearance is desired. They are suitable for free-air delivery or low static pressure duct applications. They may be equipped with filters, and they can be arranged to discharge either horizontally or vertically up or down. Industrial centrifugal fan units are applied where heating capacities and space volumes are large or where filtration of the heated air or operation against static resistance is required. Downblow or horizontal-blow units may be used, depending on the requirements. Duct unit heaters are used where the air handler is remote from the heater. These heaters sometimes provide an economical means of adding heating to existing cooling or ventilating systems with ductwork. They require flow and temperature limit controls. Location for Proper Heat Distribution. Units must be selected, located, and arranged to provide complete heat coverage while main-

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Unit Ventilators, Unit Heaters, and Makeup Air Units

Fig. 3 Typical Unit Heaters

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taining acceptable air motion and temperature at an acceptable sound level in the working or occupied zone. Proper application depends on size, number, and type of units; direction of airflow and type of directional outlet used; mounting height; outlet velocity and temperature; and air volumetric flow. Many of these factors are interrelated. The mounting height may be governed by space limitations or by the presence of equipment such as display cases or machinery. The higher a downblow heater is mounted, the lower the temperature of air leaving the heater must be to force the heated air into the occupied zone. Also, the distance that air leaving the heater travels depends largely on the air temperature and initial velocity. A high discharge temperature reduces the area of effective heat coverage because of its buoyancy. High-temperature air is less dense than the cooler air of the space being heated. Unit heaters for high-pressure steam or high-temperature hot water should be designed to produce approximately the same leaving air temperature as would be obtained from a lower temperature heating medium. To obtain the desired air distribution and heat diffusion, unit heaters are commonly equipped with directional outlets, adjustable louvers, or fixed or revolving diffusers. For a given unit with a given discharge temperature and outlet velocity, the mounting height and heat coverage can vary widely with the type of directional outlet, adjustable louver, or diffuser. High-volume, low-speed fans in highceilinged areas are often used to augment air distribution and reduce temperature stratification. Other factors that may substantially reduce heat coverage include obstructions (such as columns, beams, or partitions) or machinery in either the discharge airstream or approach area to the unit. Strong drafts or other air currents also reduce coverage. Exposures such as large glass areas or outer doors, especially on the windward side of the building, require special attention; arrange units so that they blanket the exposures with a curtain of heated air and intercept the cold drafts. For area heating, place horizontal-blow unit heaters in exterior zones such that they blow either along the exposure or toward it at a slight angle. When possible, arrange multiple units so that the discharge airstreams support each other and create a general circulatory motion in the space. Interior zones under exposed roofs or skylights should be completely blanketed. Arrange downblow units so that the heated areas from adjacent units overlap slightly to provide complete coverage. For spot heating of individual spaces in larger unheated areas, single unit heaters may be used, but allowance must be made for the inflow of unheated air from adjacent spaces and the consequent reduction in heat coverage. Such spaces should be isolated by partitions or enclosures, if possible. Horizontal unit heaters should have discharge outlets located well above head level. Both horizontal and vertical units should be placed so that the heated airstream is delivered to the occupied zone at acceptable temperature and velocity. Outlet air temperature of free-air delivery unit heaters used for comfort heating should be 27 to 33 K higher than the design room temperature. When possible, locate units so that they discharge into open spaces, such as aisles, and not directly on the occupants. For further information on air distribution, see Chapter 20 of the 2017 ASHRAE Handbook—Fundamentals. Manufacturers’ catalogs usually include suggestions for the best arrangements of various unit heaters, recommended mounting heights, heat coverage for various outlet velocities, final temperatures, directional outlets, and sound level ratings. Sound Level in Occupied Spaces. Sound pressure levels in workplaces should be limited to values listed in Table 1 in Chapter 49 of the 2019 ASHRAE Handbook—HVAC Applications. Although the noise level is generated by all equipment within hearing dis-

tance, unit heaters may contribute a significant portion of noise level. Both noise and air velocity in the occupied zone generally increase with increased outlet velocities. An analysis of both the diverse sound sources and the locations of personnel stations establishes the limit to which the unit heaters must be held. Ratings of Unit Heaters. Steam or Hot Water. Heating capacity must be determined at a standard condition. Variations in entering steam or water temperature, entering air temperature, and steam or water flow affect capacity. Typical standard conditions for rating steam unit heaters are dry saturated steam at 14 kPa (gage) pressure at the heater coil, air at 15.6°C (101.3 kPa barometric pressure) entering the heater, and the heater operating free of external resistance to airflow. Standard conditions for rating hot-water unit heaters are entering water at 93.3°C, water temperature drop of 11.1 K, entering air at 15.6°C and 101.3 kPa barometric pressure, and the heater operating free of external resistance to airflow. Gas-Fired. Gas-fired unit heaters are rated in terms of both input and output, in accordance with the approval requirements of the American Gas Association. Oil-Fired. Ratings of oil-fired unit heaters are based on heat delivered at the heater outlet. Electric. Electric unit heaters are rated based on the energy input to the heating element. Effect of Airflow Resistance on Capacity. Unit heaters are customarily rated at free-air delivery. Airflow and heating capacity decrease if outdoor air intakes, air filters, or ducts on the inlet or discharge are used. The capacity reduction caused by this added resistance depends on the characteristics of the heater and on the type, design, and speed of the fans. As a result, no specific capacity reduction can be assigned for all heaters at a given added resistance. The manufacturer should have information on the heat output to be expected at other than free-air delivery. Effect of Inlet Temperature. Changes in entering air temperature influence the total heating capacity in most unit heaters and the final temperature in all units. Because many unit heaters are located some distance from the occupied zone, possible differences between the temperature of the air actually entering the unit and that of air being maintained in the heated area should be considered, particularly with downblow unit heaters. Higher-velocity units and units with lower vertical discharge air temperature maintain lower temperature gradients than units with higher discharge temperatures. Valve- or bypass-controlled units with continuous fan operation maintain lower temperature gradients than units with intermittent fan operation. Directional control of the discharged air from a unit heater can also be important in distributing heat satisfactorily and in reducing floor-toceiling temperature gradients. Filters. Air from propeller unit heaters cannot be filtered because the heaters are designed to operate with heater friction loss only. If dust in the building must be filtered, centrifugal fan units or cabinet units should be used. Chapter 29 has further information on air cleaners for particulate contaminants.

Control The controls for a steam or hot water unit heater can provide either (1) on/off operation of the unit fan, or (2) continuous fan operation with modulation of heat output. For on/off operation, a room thermostat is used to start and stop the fan motor or group of fan motors. A limit thermostat, often strapped to the supply or return pipe, prevents fan operation in the event that heat is not being supplied to the unit. An auxiliary switch that energizes the fan only when power is applied to open the motorized supply valve may also be used to prevent undesirable cool air from being discharged by the unit.

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Unit Ventilators, Unit Heaters, and Makeup Air Units Continuous fan operation eliminates both the intermittent blasts of hot air resulting from on/off operation and the stratification of temperature from floor to ceiling that often occurs during off periods. In this arrangement, a proportional room thermostat controls a valve modulating the heat supply to the coil or a bypass around the heating element. A limit thermostat or auxiliary switch stops the fan when heat is no longer available. One type of control used with downblow unit heaters is designed to automatically return the warm air, which would normally stratify at the higher level, down to the zone of occupancy. Two thermostats and an auxiliary switch are required. The lower thermostat is placed in the zone of occupancy and is used to control a two-position supply valve to the heater. An auxiliary switch is used to stop the fan when the supply valve is closed. The higher thermostat is placed near the unit heater at the ceiling or roof level where the warm air tends to stratify. The lower thermostat automatically closes the steam valve when its setting is satisfied, but the higher thermostat overrides the auxiliary switch so that the fan continues to run until the temperature at the higher level falls below a point sufficiently high to produce a heating effect. Indirect-fired and electric units are usually controlled by intermittent operation of the heat source under control of the room thermostat, with a separate fan switch to run the fan when heat is being supplied. For more information on automatic control, refer to Chapter 48 of the 2019 ASHRAE Handbook—HVAC Applications.

28.7 Unit heaters can be used to circulate air in summer. In such cases, the heat is shut off and the thermostat has a bypass switch, which allows the fan to run independently of the controls.

Piping Connections Piping connections for steam unit heaters are similar to those for other types of fan blast heaters. Unit heater piping must conform strictly to the system requirements, while allowing the heaters to function as intended. Basic piping principles for steam systems are discussed in Chapter 11. Steam unit heaters condense steam rapidly, especially during warm-up periods. The return piping must be planned to keep the heating coil free of condensate during periods of maximum heat output, and the steam piping must be able to carry a full supply of steam to the unit to take the place of condensed steam. Adequate pipe size is especially important when a unit heater fan is operated under on/off control because the condensate rate fluctuates rapidly. Recommended piping connections for unit heaters are shown in Figure 4. In steam systems, the branch from the supply main to the heater must pitch toward the main and be connected to its top to prevent condensate in the main from draining through the heater, where it might reduce capacity and cause noise. The return piping from steam unit heaters should provide a minimum drop of 250 mm below the heater, so that the water pressure required to overcome resistances of check valves, traps, and strainers does not cause condensate to remain in the heater.

Fig. 4 Hot Water and Steam Connections for Unit Heaters

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Dirt pockets at the outlet of unit heaters and strainers with 1.6 mm perforations to prevent rapid plugging are essential to trap dirt and scale that might affect the operation of check valves and traps. Always install strainers in the steam supply line if the heater has steam-distributing coils or is valve controlled. An adequate air vent is required for low-pressure closed gravity systems. The vertical pipe connection to the air vent should be at least 20 mm DN to allow water to separate from the air passing to the vent. If thermostatic instead of float-and-thermostatic traps are used in vacuum systems, a cooling leg must be installed ahead of the trap. In high-pressure systems, it is customary to continuously vent the air through a petcock (as shown in Figure 4C), unless the steam trap has a provision for venting air. Most high-pressure return mains terminate in flash tanks that are vented to the atmosphere. When possible, use pressure-reducing valves to allow heater operation at low pressure. Traps must be suitable for the operating pressure encountered. When piping is connected to hot-water unit heaters, it must be pitched to allow air to vent to the atmosphere at the high point in the piping. An air vent at the heater is used to facilitate air removal or to vent the top of the heater. The system must be designed for complete drainage, including placing nipple and cap drains on drain cocks when units are located below mains.

Maintenance Regular inspection, based on a schedule determined by the amount of dirt in the atmosphere, assures maximum operating economy and heating capacity. During design, ensure sufficient space for maintenance access to each component. Clean heating elements when necessary by brushing or blowing with high-pressure air or by using a steam spray. A portable sheet metal enclosure may be used to partially enclose smaller heaters for cleaning in place with air or steam jets. In some installations, however, it may be necessary to remove the heating element and wash it with a mild alkaline solution, followed by a thorough rinsing with water. Propeller units do not have filters and are, therefore, more susceptible to dust build-up on the coils. Dirt on fan blades reduces capacity and may unbalance the blades, which causes noise and bearing damage. Fan blades should be inspected and cleaned when necessary. Vibration and noise may also be caused by improper fan position or loose set screws. Place a fan guard on downblow unit heaters that have no diffuser or other device to catch the fan blade if it comes loose and falls from the unit. The amount of attention required by the various motors used with unit heaters varies greatly. Instructions for lubrication, in particular, must be followed carefully for trouble-free operation: excess lubrication, for example, may damage the motor, and an improper lubricant may cause the bearings to fail. Get instructions for care of the motor on any unit heater from the manufacturer and keep them at the unit. Fan bearings and drives must be lubricated and maintained according to the instructions specified by the manufacturer. If the unit is direct connected, inspect the couplings periodically for wear and alignment. V-belt drives should have all belts replaced with a matched set if one belt shows wear. Periodic inspections of traps, inspections of check and air valves, and the replacement of worn fans are other important maintenance functions. Strainers should be cleaned regularly. Filters, if included, must be cleaned or replaced when dirty.

3.

MAKEUP AIR UNITS

Description and Applications Makeup air units are designed to condition ventilation air introduced into a space or to replace air exhausted from a building, in compliance with ASHRAE Standard 62.1 as applicable. The air

exhausted may be from a process or general area exhaust, through either powered exhaust fans or gravity ventilators. The units may be used to prevent negative pressure within buildings or to reduce airborne contaminants in a space. If temperature and/or humidity in the structure are controlled, the makeup air system must have the capacity to condition the replacement air. In most cases, makeup air units must be used to supply this conditioned makeup air. The units may heat, cool, humidify, dehumidify, and/or filter incoming air. They may be used to replace air in the conditioned space or to supplement or accomplish all or part of the airflow needed to satisfy the heating, ventilating, or cooling airflow requirements. Makeup air can enter at a fixed flow rate or as a variable volume of outdoor air. It can be used to accomplish building pressurization or contamination reduction, and may be controlled in a manner that responds directly to exhaust flow. When compatible with the characteristics and class of the air being exhausted, makeup air units should also be connected to process exhaust with air-to-air heat recovery units, thermal wheels, heat pipes, or runaround coils (recovery loop exchangers) tested (when applicable) under ANSI/ ASHRAE Standard 84. Buildings under negative pressure because of inadequate makeup air may have the following symptoms: • Gravity stacks from unit heaters and processes back-vent. • Exhaust systems do not perform at rated volume. • The perimeter of the building is cold in winter because of high infiltration. • Severe drafts occur at exterior doors. • Exterior doors are hard to open or close completely. • Heating systems cannot maintain comfortable conditions throughout the building because the central core area becomes overheated. • Excessive condensation occurs. Other Applications. Spot Cooling. High-velocity air jets in the unit may be directed to working positions. During cold weather, supply air must be tempered or reduced in velocity to avoid overcooling workers. Door Heating. Localized air supply at swinging doors or overhead doors, such as for loading docks, can be provided by makeup air units. Heaters may blanket door openings with tempered air. The temperature may be reset from the outdoor temperature or with dual-temperature air (low when the door is closed and high when the door is open during cold weather). Heating may be arranged to serve a single door or multiple doors by an air distribution system. Door heating systems may also be arranged to minimize entry of insects during warm weather. Air curtain units for door openings should be tested and comply with ANSI/AMCA Standard 220-05, section C408.2.3.

Selection Makeup air systems used for ventilation may be (1) sized to balance air exhaust volumes or (2) sized in excess of the exhaust volume to dilute contaminants. In applications where contaminant levels vary, variable-flow units should be considered so that the supply air varies for contaminant control and the exhaust volume varies to track supply volume. In critical spaces, the exhaust volume may be based on requirements to control pressure in the space. Location. Makeup air units are defined by their location or the use of a key component. Examples are rooftop makeup air units, truss- or floor-mounted units, and sidewall units. Some manufacturers differentiate their units by heating mode, such as steam or direct gas-fired makeup air units. Rooftop units are commonly used for large single-story industrial buildings to simplify air distribution. Access (via roof walks) is more convenient than access to equipment mounted in the truss;

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truss units are only accessible by installing a catwalk adjacent to the air units. Disadvantages of rooftop units are (1) they increase foot traffic on the roof, thus reducing its life and increasing the likelihood of leaks; (2) inclement weather reduces equipment accessibility; and (3) units are exposed to weather. Makeup air units can also be placed around the perimeter of a building with air ducted through the sidewall. This approach limits future building expansion, and the effectiveness of ventilating internal spaces decreases as the building gets larger. However, access to the units is good, and minimum support is required because the units are mounted on the ground. Use caution in selecting the location of the makeup air unit and/or its associated combustion air source, to avoid introducing combustible vapors into the unit. Consult state and local fire codes for specific guidance. Heating and Cooling Media. Heating. Makeup air units are often identified by the heating or cooling medium they use. Heaters in makeup air systems may be direct gas-fired burners, electric resistance heating coils, indirect gas-fired heaters, steam coils, or hot-water heating coils. (Chapter 27 covers the design and application of heating coils.) Air distribution systems are often required to direct heat to spaces requiring it. Natural gas can be used to supply an indirect-fired burner, as for a large furnace. (Chapter 33 has more information, including detailed heater descriptions.) In a non-recirculating direct-fired heater, levels of combustion products generated by the heater are very low (CO < 5.0 mg/kg, NO2 < 0.5 mg/kg, and CO2 < 4000 mg/ kg) and are released directly into the airstream being heated. All air to a non-recirculating makeup air heater must be ducted directly from outdoor source. Non-recirculating direct gas-fired industrial air heaters are typically certified to comply with ANSI Standard Z83.4b/CSA3.7b-2006. In a recirculating makeup air heater, ventilation air to the heater must be ducted directly from an outdoor source to limit the concentration of combustion products in the conditioned space to a level below 25 mg/kg for CO, 3 mg/kg for NO2, and 5000 mg/kg for CO2. Recirculating direct gas-fired industrial air heaters are typically certified to comply with ANSI Standard Z83.18-2000. Installing carbon monoxide detectors to protect building occupants in the event of a heater malfunction is good engineering practice, and may be required by local codes. Hydronic heating sections in spaces requiring a fully isolated source (100%) of outdoor air must be protected from freezing in cold climates. Low-temperature protection includes two-position control of steam coils; careful selection of the water coil heating surface and control valves; careful control of water supply temperature; and use of an antifreeze additive. Cooling. Mechanical refrigeration with direct-expansion or chilled-water cooling coils, direct or indirect evaporative cooling sections, or well water coils may be used. Air distribution systems are often required to direct cooling to specific spaces that experience or create heat gain. Because industrial facilities often have high sensible heat loads, evaporative cooling can be particularly effective. An evaporative cooler helps clean the air, as well. A portion of the spray water must be bled off to keep the water acceptably clean and to maintain a low solids concentration. Chapter 41 of this volume and Chapter 53 of the 2019 ASHRAE Handbook—HVAC Applications cover evaporative cooling in more detail. Chapter 23 provides information on air-cooling coils. If directexpansion coils are used in conjunction with direct-fired gas coils, the cooling coils’ headers must be isolated from the airstream and directly vented outdoors. Filters. High-efficiency filters (approximately MERV 16 for near-HEPA performance) are not normally used in a makeup air unit

28.9 because of their relatively high cost. HVAC prefilters are generally in the MERV 6 to 13 range, depending on particulate removal needs. Designers should ensure that all filters are easy to change or clean. Appropriate washing equipment should be located near all washable filters. Throwaway filters should be sized for easy removal and disposal. Chapters 29 and 30 have more information on air filters and cleaners. Fans. Follow AMCA Standard 205-12 for fan selection. This standard addresses fans with a minimum impeller diameter of 125 mm, operating with a shaft power of at least 750 W, and with a total efficiency calculated according to one of the following fan test standards: • • • •

ANSI/AMCA Standard 210 (ANSI/ASHRAE Standard 51) ANSI/AMCA Standard 230 AMCA Standard 260 ISO Standard 5801

ASHRAE Standard 90.1-2013 requires a fan efficiency grade (FEG) 67, and that the fan should be sized and selected within 15% of its peak total efficiency (FEG 71 under AMCA Standard 205 and ICC [2015]). Fans should have variable-speed drives for possible energy savings or for use in variable-airflow systems.

Control Controls for a makeup air unit fall into the following categories: (1) local temperature controls, (2) airflow controls, (3) plant-wide controls for proper equipment operation and efficient performance, (4) safety controls for burner gas, and (5) building smoke control systems. For control system information, refer to Chapters 42, 42, and 48 of the 2019 ASHRAE Handbook—HVAC Applications. Safety controls for gas-fired units include components to properly light the burner and to provide a safeguard against flame failure. The heater and all attached inlet ducting must be purged with at least four air changes before initiating an ignition sequence and before reignition after a malfunction. A flame monitor and control system must be used to automatically shut off gas to the burner upon burner ignition or flame failure. Critical malfunctions include flame failure, supply fan failure, combustion air depletion, power failure, control signal failure, excessive or inadequate inlet gas supply pressure, excess air temperature, and gas leaks in motorized valves or inlet gas supply piping. Makeup air units should be interlocked with exhaust units to avoid overpressurization, and should include shutoff dampers with limit switches for when not in use. Damper leakage rates should be within limits set in ASHRAE Standard 90.1. These units should also be interlocked to the building’s fire alarm system to shut down in the case of a fire, where required by applicable codes. Consider using automatic safety shutoff valves on interconnecting piping systems where there are risks of overtemperature, overpressure, or gas leaks.

Applicable Codes and Standards A gas-fired makeup air unit must be designed and built in accordance with NFPA Standard 54 and the requirements of the owner’s insurance underwriter. Local codes must also be observed when using direct-fired gas makeup air units because some jurisdictions prohibit or restrict their use and may also require exhaust fans to be used while the unit is in operation. The following standard and codes and the sources in the References may also apply, depending on the application. AHRI. 2014. Performance rating of air-to-air exchangers for energy recovery ventilation equipment. ANSI/AHRI Standard 1060. Air-Conditioning, Heating, and Refrigeration Institute, Arlington, VA.

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ICC. 2015. International mechanical code® (IMC®). International Code Council, Falls Church, VA. ICC. 2015. International fuel gas code® (IFGC®). International Code Council, Falls Church, VA.

Commissioning Commissioning of makeup air systems is similar to that of other air-handling systems, requiring attention to • • • •

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• • • • • • • • • •

Equipment identification Piping system identification Belt drive adjustment Control system checkout in accordance with ASHRAE Guideline 11-2009 Documentation of system installation Lubrication Electrical system checkout for overload heater size and function Cleaning and degreasing of hydronic piping systems Pretreatment of hydronic fluids Setup of chemical treatment program for hydronic systems and evaporative apparatus Start-up of major equipment items by factory-trained technician Testing and balancing Planning of preventive maintenance program Instruction of owner’s operating and maintenance personnel in accordance with proposed ASHRAE Guideline GPC 1.3P, Building Operation and Maintenance Training for the HVAC&R Commissioning Process

Maintenance Basic operating and maintenance data required for makeup air systems may be obtained from the ASHRAE Handbook chapters covering the components. Specific operating instructions are required for makeup air heaters that require changeover from winter to summer conditions, including manual fan speed changes, air distribution pattern adjustment, or heating cycle lockout. Operating and maintenance documentation should comply with ASHRAE Guideline 4-2008 (RA13). Operations handling 100% outdoor air may require more frequent maintenance, such as changing filters, lubricating bearings, and checking the water supply to evaporative coolers/humidifiers. Filters on systems in locations with dirty air require more frequent changing, so a review may determine whether upgrading filter media would be cost-effective. More frequent cleaning of fans’ blades and heat transfer surfaces may be required in such locations to maintain airflow and heat transfer performance.

REFERENCES ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal.ashrae .org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org/bookstore. ACCA. 2005. Direct-fired MUA equipment. Technical Bulletin 109 (four parts). Air-Conditioning Contractors of America, Washington, D.C.

ACGIH. 2013. Industrial ventilation: A manual of recommended practice for design, 28th ed. American Conference of Governmental Industrial Hygienists, Cincinnati, OH. AHRI. 2009. Central-station air-handling units. ANSI/AHRI Standard 4302009. Air-Conditioning, Heating, and Refrigeration Institute, Arlington, VA. AMCA. 2012. Energy efficiency classification for fans. ANSI/AMCA Standard 205-12. Air Movement and Control Association, Arlington Heights, IL. AMCA. 2012. Laboratory methods of testing air circulating fans for rating and certification. ANSI/AMCA Standard 230-12. Air Movement and Control Association, Arlington Heights, IL. AMCA. 2013. Laboratory methods of testing induced flow fans for rating. Standard 260-13. Air Movement and Control Association, Arlington Heights, IL. ANSI. 2012. Laboratory methods of testing air curtains for aerodynamic performance ratings. Standard 220-05, Section C408.2.3. American National Standards Institute, New York. ANSI. 2012. Recirculating direct gas-fired industrial heaters. Standard Z83.18-2012. American National Standards Institute, New York. ANSI/CSA. 2013. Non-recirculating direct gas-fired industrial air heaters. ANSI Standard Z83.4/CSA3.7-2013. American National Standards Institute, New York. ASHRAE. 2013. Ventilation for acceptable indoor air quality. ANSI/ ASHRAE Standard 62.1-2013. ASHRAE. 2013. Method of testing air-to-air heat/energy exchangers. ANSI/ ASHRAE Standard 84-2013. ASHRAE. 2013. Energy standard for buildings except low-rise residential buildings. ANSI/ASHRAE/IESNA Standard 90.1-2013. ASHRAE. 2006. Energy conservation in existing buildings. ASHRAE/ IESNA Standard 100-2006. ASHRAE. 2009. Field testing of HVAC controls components. Guideline 112009. ICC. 2015. International green construction code (IgCC). International Code Council, Washington, D.C. ISO. 2007. Industrial fans—Performance testing using standardized airways. Standard 5801:2007. International Organization for Standardization, Geneva.

BIBLIOGRAPHY Brown, W.K. 1990. Makeup air systems—Energy saving opportunities. ASHRAE Transactions 96(2):609-615. Paper SL-90-05-1. Bridgers, F.H. 1980. Efficiency study—Preheating outdoor air for industrial and institutional applications. ASHRAE Journal 22(2):29-31. Burroughs, H.E. 2005. Filtration and building security. ASHRAE Journal April 2005:24-29. Gadsby, K.J., and T.T. Harrje. 1985. Fan pressurization of buildings—Standards, calibration and field experience. ASHRAE Transactions 91(2B): 95-104. Paper HI-85-03-1. Holness, G.V.R. 1989. Building pressurization control: Facts and fallacies. Heating/Piping/Air Conditioning (February). NFPA. 2018. National fuel gas code. Standard 54-2018. National Fire Protection Association, Quincy, MA. Persily, A. 1982. Repeatability and accuracy of pressurization testing. In Thermal Performance of the Exterior Envelopes of Buildings II, Proceedings of ASHRAE/DOE Conference. ASHRAE SP 38:380-390. PG&E. 2002. Improving commercial kitchen ventilation system performance: Optimizing makeup air. Design Guide 2.

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Related Commercial Resources CHAPTER 29

AIR CLEANERS FOR PARTICULATE CONTAMINANTS Terminology ............................................................................. 29.1 Atmospheric Aerosols .............................................................. 29.2 Aerosol Characteristics ........................................................... 29.2 Air-Cleaning Applications ....................................................... 29.2 Mechanisms of Particle Collection .......................................... 29.3 Evaluating Air Cleaners .......................................................... 29.3

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T

HIS chapter discusses removal of contaminants from both ventilation and recirculated air used for conditioning building interiors. Complete air cleaning may require removing of airborne particles, microorganisms, and gaseous contaminants, but this chapter only covers removal of airborne particles and briefly discusses bioaerosols. Chapter 47 of the 2019 ASHRAE Handbook— HVAC Applications covers the removal of gaseous contaminants. The total suspended particulate concentration in applications discussed in this chapter seldom exceeds 2 mg/m3 and is usually less than 0.2 mg/m3 of air (Chapter 11 of the 2018 ASHRAE Handbook— Refrigeration). This is in contrast to flue gas or exhaust gas from processes, where dust concentrations typically range from 200 to 40 000 mg/m3. Chapter 26 discusses exhaust gas control. Most air cleaners discussed in this chapter are not used in exhaust gas streams because of the extreme dust concentration, large particle size, high temperature, high humidity, and high airflow rate requirements that may be encountered in process exhaust. However, the air cleaners discussed here are used extensively in supplying makeup air with low particulate concentration to industrial processes.

1.

TERMINOLOGY

Definitions Aerodynamic Diameter. The diameter of a spherical particle with a density of 1000 kg/m3 and the same settling velocity as the irregular particle of interest. Arrestance. A measure of the ability of an air-cleaning device with efficiencies less than 20% in the size range of 3.0 to 10.0 µm to remove loading dust from the air passing through the device. Dust Holding Capacity. The total weight of synthetic loading dust captured by an air-cleaning device over all of the incremental dust loading steps. Particle Size Removal Efficiency. The fraction or percentage of particles retained by an air cleaner for a given particle-size range. Particulate Matter (PM). Solid and/or liquid particles of various sizes suspended in ambient air. Penetration. The fraction or percentage of particles that pass through an air cleaner for a given particle-size range. Resistance to Airflow. Difference in absolute (static) pressure between two points in a system. This parameter is often called pressure drop.

Acronyms AFI. Air Filter Institute. AHRI. Air-Conditioning, Heating, and Refrigeration Institute. ASME. American Society of Mechanical Engineers. BI/BO. Bag-in/bag-out. CEN. Comité Européen de Normalisation. The preparation of this chapter is assigned to TC 2.4, Particulate Air Contaminants and Particulate Contaminant Removal Equipment.

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Air Cleaner Test Methods ........................................................ 29.4 Types of Air Cleaners ............................................................... 29.6 Filter Types and Performance ................................................. 29.6 Selection and Maintenance ...................................................... 29.9 Air Cleaner Installation ......................................................... 29.11 Safety Considerations ............................................................. 29.12

DEHS. Di-ethyl-hexyl-sebacate. DHC. Dust-holding capacity. DOP. Dioctyl phthalate. EN. European norm (European standard). EPA. Environmental Protection Agency. EB. Existing buildings. ETS. Environmental tobacco smoke. HEPA. High-efficiency particulate air. IEST. Institute of Environmental Sciences and Technology. IPA. Isopropyl alcohol. ISO. International Organization for Standardization. LCC. Life-cycle cost. LEED. Leadership in Energy and Environmental Design. MERV. Minimum efficiency reporting value. MIL-STD. U.S. Military standard. MPPS. Most penetrating particle size. MSHA. Mine Safety and Health Administration. NAFA. National Air Filtration Association. NC. New construction. NIOSH. National Institute for Occupational Safety and Health. NIST. National Institute of Standards and Technology. PAO. Polyalphaolefin. PM. Particulate matter. PSE. Particle size removal efficiency. PSL. Polystyrene latex. UL. Underwriters Laboratory. ULPA. Ultra-low particulate air. VAV. Variable air volume.

2.

ATMOSPHERIC AEROSOLS

Atmospheric dust is a complex mixture of smokes, mists, fumes, dry granular particles, bioaerosols, and natural and synthetic fibers. When suspended in a gas such as air, this mixture is called an aerosol. A sample of atmospheric aerosol usually contains soot and smoke, silica, clay, decayed animal and vegetable matter, organic materials in the form of lint and plant fibers, and metallic fragments. It may also contain living organisms, such as mold spores, bacteria, and plant pollens, which may cause diseases or allergic responses. (Chapter 11 of the 2017 ASHRAE Handbook—Fundamentals contains further information on atmospheric contaminants.) A sample of atmospheric aerosol gathered at any point generally contains materials common to that locality, together with other components that originated at a distance but were transported by air currents or diffusion. These components and their concentrations vary with the geography of the locality (urban or rural), season of the year, weather, direction and strength of the wind, and proximity of dust sources. Aerosol sizes range from 0.01 µm and smaller for freshly formed combustion particles and radon progeny; to 0.1 µm for aged cooking and cigarette smokes; and 0.1 to 10 µm for airborne dust, microor-

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ganisms, and allergens; and up to 100 µm and larger for airborne soil, pollens, and allergens. Concentrations of atmospheric aerosols generally peak at submicrometre sizes and decrease rapidly as the particulate size increases above 1 µm. For a given size, the concentration can vary by several orders of magnitude over time and by location, particularly near an aerosol source, such as human activities, equipment, furnishings, and pets (McCrone et al. 1967). This wide range of particulate size and concentration makes it impossible to design one cleaner for all applications.

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3.

AEROSOL CHARACTERISTICS

The characteristics of aerosols that most affect air filter performance include particle size and shape, mass, concentration, and electrical properties. The most important of these is particle size. Figure 3 in Chapter 11 of the 2017 ASHRAE Handbook—Fundamentals gives data on the sizes and characteristics of a wide range of airborne particles that may be encountered. Particle size in this discussion refers to aerodynamic particle size. Particles less than 0.1 µm in diameter are generally referred to as ultrafine-mode or nanoparticles, those between 0.1 and 2.5 µm are termed fine mode, and those larger than 2.5 µm as coarse mode. Whereas ultrafine- and fine-mode particles may be formed together, fine- and coarse-mode particles typically originate by separate mechanisms, are transformed separately, have different chemical compositions, and require different control strategies. Vehicle exhaust is a major source of ultrafine particles. Ultrafines are minimally affected by gravitational settling and can remain suspended for days at a time. Fine-mode particles generally originate from condensation or are directly emitted as combustion products. Many microorganisms (bacteria and fungi) either are in this size range or produce components this size. These particles are less likely to be removed by gravitational settling and are just as likely to deposit on vertical surfaces as on horizontal surfaces. Coarse-mode particles are typically produced by mechanical actions such as erosion and friction. Coarse particles are more easily removed by gravitational settling, and thus have a shorter airborne lifetime. For industrial hygiene purposes, particles  5 µm in diameter are considered respirable particles (RSPs) because a large percentage of this size range has been shown to reach the alveolar region of the lungs. A cutoff of 5.0 µm includes 80 to 90% of the particles that can reach the functional pulmonary region of the lungs (James et al. 1991; Phalen et al. 1991). Willeke and Baron (1993) described a detailed aerosol sampling technique for RSPs, including the use of impactors. See also the discussion in the section on Sizes of Airborne Particles in Chapter 11 of the 2017 ASHRAE Handbook— Fundamentals. In the United States, particulate matter (PM) levels in outdoor air are regulated by the Environmental Protection Agency (EPA). Outdoor PM is a complex mixture of small particles and liquid droplets; their effects on health are related to size. Smaller particles have more impact because they penetrate deeper into the respiratory system. Currently, there are two regulated size ranges: • PM10 (i.e., all particles that have aerodynamic diameters less than or equal to 10 μm according to the EPA); typical sources include road dust and industrial emissions • PM2.5 (i.e., all PM with aerodynamic diameter less than or equal to 2.5 μm); the main sources are industrial emissions and combustion exhaust from automobiles, power plants, and heating systems PM1, which is not regulated, consists of fine and ultrafine particles of less than 1 μm aerodynamic diameter. The U.S. Clean Air Act requires the EPA to impose both primary standards designed to protect public health, and secondary stan-

Table 1 U.S. EPA Standards for Particulate Matter in Outdoor Air Type of Standard Primary Secondary

Time Period Applicable

PM10, μg/m3

PM2.5, μg/m3

24 h 1 yr 24 h 1 yr

150 — 150 —

35 12 35 15

Source: EPA (2015).

dards intended to protect against adverse environmental effects. The limits presently in place (Federal Register 2013) are shown in Table 1. There is no filtration requirement for areas of noncompliance. Bioaerosols are a diverse class of particles of biological origin. They include bacteria, fungal spores, fungal fragments, pollen grains, subpollen particles, viruses, pet- and pest-associated allergens, and plant debris (Fröhlich-Nowoisky et al. 2016). They are of particular concern in indoor air because of their association with allergies and asthma and their ability to cause disease. Chapters 10 and 11 of the 2017 ASHRAE Handbook—Fundamentals contain more detailed descriptions of these contaminants. Bioaerosols range in size from 0.01 to 100 µm. Single bacterial cells are approximately 1 µm or less in aerodynamic diameter; however, they are often transported as larger bacterial cell agglomerates (~2 to 5 µm) or attached to other biological and abiotic particles. Unicellular fungal spores are generally 2 to 5 µm in aerodynamic diameter and multicellular fungal spores are larger than 10 µm (Després et al. 2012, Qian et al. 2012). Fungal fragments are typically less than 1 µm in size (Mensah-Attipoe et al. 2016). Pollen grains range in size from 10 to 100 µm (Després et al. 2012), whereas subpollen particles span ~0.01 µm to several micrometers in size (Taylor et al. 2004). Individual viruses can be as small as 0.02 to 0.03 µm, but are typically carried on larger particles (Verreault et al. 2008). Pet- and pest-associated allergens are approximately 1 to over 20 µm (O’Meara and Tovey 2000).

4.

AIR-CLEANING APPLICATIONS

Different air-cleaning applications require different degrees of air cleaning effectiveness. In industrial ventilation, only removing the larger dust particles from the airstream may be necessary for cleanliness of the structure, protection of mechanical equipment, and employee health. In other applications, surface discoloration must be prevented. Unfortunately, the smaller components of atmospheric aerosol are the worst offenders in smudging and discoloring building interiors. Electronic air cleaners or medium- to highefficiency filters are required to remove smaller particles, especially the respirable fraction, which often must be controlled for comfort and health reasons. In cleanrooms or when radioactive or other dangerous particles are present, high- or ultrahigh-efficiency filters should be selected. For more information on cleanrooms, see Chapter 19 of the 2019 ASHRAE Handbook—HVAC Applications. Major factors influencing filter design and selection include (1) degree of air cleanliness required, (2) specific particle size range or aerosols that require filtration, (3) aerosol concentration, (4) resistance to airflow through the filter, (5) design face velocity to achieve published performance, (6) change-out cycle requirements, (7) energy consumption requirements, (8) special disposal mandates, and (9) resistance to certain conditions (physical, chemical, or biological).

5.

MECHANISMS OF PARTICLE COLLECTION

In particle collection, air cleaners made of fibrous media rely on the following five main principles or mechanisms (Figure 1):

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Fig. 1 Collection Mechanisms for Filter Fiber Straining. The coarsest kind of filtration strains particles through an opening smaller than the particle being removed. It is most often observed as the collection of large particles and lint on the filter surface. The mechanism is not adequate to achieve the filtration of submicrometre aerosols through fibrous matrices, which occurs through other physical mechanisms, as follows. Inertial Impingement. When particles are large or dense enough that they cannot follow the airstream around a fiber, they cross over streamlines, hit the fiber, and remain there if the attraction is strong enough. With flat-panel and other minimal-media-area filters having high air velocities (where the effect of inertia is most pronounced), the particle may not adhere to the fiber because drag and bounce forces are so high. In this case, a viscous coating (preferably odorless and nonmigrating) is applied to the fiber to enhance retention of the particles. This adhesive coating is critical to metal mesh impingement filter performance. Interception. Particles follow the airstream close enough to a fiber that the particle contacts the fiber and remains there mainly because of van der Waals forces (i.e., weak intermolecular attractions between temporary dipoles). The adhesion process depends on air velocity through the media being low enough not to dislodge the particles, and is therefore the predominant capture mechanism in extended-media filters such as bag and deep-pleated rigid cartridge types. Diffusion. The path of very small particles is not smooth but erratic and random within the airstream. This is caused by gas molecules in the air bombarding them (Brownian motion), producing an erratic path that brings the particles close enough to a media fiber to be captured by interception. As more particles are captured, a concentration gradient forms in the region of the fiber, further enhancing filtration by diffusion and interception. The effects of diffusion increase with decreasing particle size and media velocity. Electrostatic Effects. Particle or media electrostatic charge can produce changes in dust collection affected by the electrical properties of the airstream. Some particles may carry a natural charge. Passive electrostatic (without a power source) filter fibers may be electrostatically charged during their manufacture or (in some materials) by mainly dry air blowing through the media. Charges on the particle and media fibers can produce a strong attracting force if opposite. Efficiency is generally considered to be highest when the media is new and clean.

6.

EVALUATING AIR CLEANERS

In addition to criteria affecting the degree of air cleanliness, factors such as cost (initial investment, maintenance, and energy effectiveness), space requirements, and airflow resistance have led to the development of a wide variety of air cleaners. Comparisons of different air cleaners can be made from data obtained by standardized test methods. The distinguishing operating characteristics are particle size removal efficiency, resistance to airflow, service life, and life-cycle

29.3 cost. Efficiency measures the ability of the air cleaner to remove particles from an airstream. Minimum efficiency during the life of the filter is the most meaningful characteristic for most filters and applications. Resistance to airflow (or simply resistance) is the static pressure drop differential across the filter at a given face velocity. The term static pressure differential is interchangeable with pressure drop and resistance if the difference of height in the filtering system is negligible. Life-cycle cost is the evaluation of device performance in the application in terms of overall cost, along with filter service life, including element cost, energy consumption, maintenance, disposal, etc. Air filter testing is complex, and no individual test adequately describes all filters. Ideally, performance testing of equipment should simulate operation under actual conditions and evaluate the characteristics important to the equipment user. Wide variations in the amount and type of particles in the air being cleaned make evaluation difficult. Another complication is the difficulty of closely relating measurable performance to the specific requirements of users. Recirculated air tends to have a larger proportion of lint than does outdoor air. However, performance tests should strive to simulate actual use as closely as possible. Arrestance. A standardized ASHRAE synthetic dust consisting of various particle sizes and types is fed into the test airstream to the air cleaner and the mass fraction of the dust removed is determined. In the ASHRAE Standard 52.2 test, summarized in the segment on Air Cleaner Test Methods in this chapter, this measurement is called synthetic dust mass arrestance to distinguish it from other efficiency values. The indicated mass arrestance of air filters, as determined in the arrestance test, depends greatly on the particle size distribution of the test dust, which, in turn, is affected by its state of agglomeration. Therefore, this filter test requires highly standardized test dust, dust dispersion apparatus, and other elements of test equipment and procedures. This test is particularly suited to distinguish between the many types of low-efficiency air filters in the minimum efficiency reporting value (MERV) 1 to 4 categories. These are generally roughing filters such as automatic rolls, metal washables, or screen mesh filters used for gross concentration removal of debris and very large particles. It does not adequately distinguish between higherefficiency filters. ASHRAE Atmospheric Dust-Spot Efficiency. This method evaluated discoloration (staining) of targets in upstream versus downstream sampling. The dust-spot efficiency method is no longer an ASHRAE standard of test; it was replaced with particle-sizespecific testing under Standard 52.2. Fractional Efficiency or Penetration. Defined-size particles are fed into the air cleaner and the percentage removed by the cleaner is determined, typically by a photometer, optical particle counter, or condensation nuclei counter. In fractional efficiency tests, the use of defined-particle-size aerosols results in an accurate measure of the particle size versus efficiency characteristic of filters over a wide atmospheric size spectrum. This method led to the ASHRAE Standard 52.2 test, in which a polydispersed challenge aerosol (e.g., potassium chloride) is metered into the test duct as a challenge to the air cleaner. Air samples taken upstream and downstream are drawn through an optical particle counter or similar measurement device to obtain removal efficiency versus particle size at a specific airflow rate in 12 designated particle size ranges (0.3 to 10 m). High-efficiency particulate air (HEPA) testing, specifically the dioctyl phthalate (DOP) or polyalphaolefin (PAO) test for HEPA filters, is widely used for production testing in very small particle size ranges. For more information on the DOP test, see the DOP Penetration Test section. Fractional efficiency is the measure of particles removed from the upstream concentration by a known device. The testing can be done in the field with atmospheric conditions, or in a laboratory with known particle challenges.

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Penetration efficiency typically is the reference characteristic in HEPA filter testing, using DOP or PAO and determining the amount of specific particle penetration through a tested element. Dust-Holding Capacity. Dust-holding capacity of air cleaners is the reported amount of synthetic dust retained in an air cleaner at the end of the test period. Atmospheric dust-holding capacity is a function of environmental conditions as well as variability of atmospheric dust (size, shape, and concentration), and is therefore impossible to duplicate in a laboratory test. Artificial dusts are not the same as atmospheric dusts, so dust-holding capacity as measured by these accelerated tests is different from that achieved in life-cycle cost evaluations and should not be used to compare filter life expectancies. Differences in laboratory dust-holding capacity can result from variability of test aerosols, variabilities of the tested filter elements, measurement device tolerances, and atmospheric condition.

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7.

AIR CLEANER TEST METHODS

Air cleaner test methods have been developed by the heating and air-conditioning industry, the automotive industry, the atomic energy industry, and government and military agencies. Several tests have become standard in general ventilation applications in the United States. In 1968, the test techniques developed by the U.S. National Bureau of Standards (now the National Institute of Standards and Technology [NIST]) and the Air Filter Institute (AFI) were unified (with minor changes) into a single test procedure, ASHRAE Standard 52-1968. Dill (1938), Nutting and Logsdon (1953), and Whitby et al. (1956) give details of the original codes. After multiple revisions, ASHRAE Standard 52-1968 was discontinued in 2009 and replaced by ASHRAE Standard 52.2. ASHRAE Standard 52.2-2012 contains minimum efficiency reporting values (MERVs) for air cleaner particle size efficiency. In 2008, the standard incorporated arrestance testing from the discontinued ASHRAE Standard 52.1. Table 3 provides an approximate cross-reference for air cleaners tested under ASHRAE Standards 52.1 and 52.2. Currently, there is no ASHRAE standard for testing electronic air cleaners.

Arrestance Test ASHRAE Standard 52.2 defines synthetic test dust as a compounded test dust consisting of (by mass) 72% ISO Standard 12 103-A2 fine test dust, 23% powdered carbon, and 5% no. 7 cotton linters. A known amount of the prepared test dust is fed into the test unit at a known and controlled concentration. The amount of dust in the air leaving the test filter is determined by passing the entire airflow through a high-efficiency final filter (98%) downstream of the test filter and measuring the gain in filter mass. The synthetic mass arrestance is calculated using the masses of the dust captured on the final high-efficiency filter and the total dust fed. Atmospheric aerosol particles range from a small fraction of a micrometre to tens of micrometres in diameter. The artificially generated dust cloud used in the ASHRAE arrestance method is considerably coarser than typical atmospheric dust. It tests the ability of a filter to remove the largest atmospheric aerosol particles and gives little indication of filter performance in removing the smallest particles. However, where the mass of dust in the air is the primary concern, this is a valid test because most of the mass is contained in the larger particles. Where extremely small particles (such as respirable sizes) are involved, arrestance rating does not differentiate between filters.

Dust-Holding Capacity (DHC) Test Synthetic test dust (as described in the preceding section) is fed to the filter in accordance with ASHRAE Standard 52.2 procedures. The pressure drop across the filter (its resistance to airflow) rises as

dust is fed. The test normally ends when resistance reaches the maximum operating resistance set by the manufacturer. However, not all filters of the same type retain collected dust equally well. The test, therefore, requires that arrestance be measured at least four times during dust loading and that the test be terminated when two consecutive arrestance values of less than 85%, or one value equal to or less than 75% of the maximum arrestance, have been measured. The ASHRAE dust-holding capacity is, then, the integrated amount of dust held by the filter up to the time the dust-loading test is terminated. (See ASHRAE Standard 52.2 for more detail.)

Particle Size Removal Efficiency (PSE) Test ASHRAE Standard 52.2 prescribes a way to test air-cleaning devices for removal efficiency by particle size while addressing two air cleaner performance characteristics important to users: the ability of the device to remove particles from the airstream and its resistance to airflow. In this method, air cleaner testing is conducted at a specific airflow based on the upper limit of the air cleaner’s application range. Airflows are based on specific face velocities between 0.60 and 3.80 m/s, which yields between 0.22 and 1.4 m3/s for a nominal 610 by 610 mm filter. The test aerosol consists of laboratory-generated potassium chloride particles dispersed in the airstream. Optical particle counters measure and count the particles in 12 geometric logarithmic-scale, equally distributed particle size ranges both upstream and downstream for efficiency determinations. The test encompasses 0.3 to 10 µm polystyrene latex. The clean-filter efficiency test is followed by five dust loadings, using ASHRAE dust, to increase the pressure drop across the filter. Efficiency tests are performed after each dust loading. A set of particle size removal efficiency performance curves is developed from the test and, together with the initial clean performance curve, is the basis of a composite curve representing performance in the range of sizes. Points on the composite curve are averaged and these averages are used to determine the MERV of the air cleaner. A complete test report includes (1) a summary of the test, (2) removal efficiency curves of the clean devices at each of the loading steps, and (3) a composite minimum removal efficiency curve. In 2008, Appendix J on conditioning was added as an optional part of Standard 52.2. If used, this conditioning step replaces the current first (small) dust loading. This step simulates the efficiency drop seen in many charged filters in actual use: the high concentration of 40 to 50 nm aerosol exposes the filter to many particles of a size common in ambient and indoor air. This step does not completely remove the charge on the fibers. When Appendix J is used in a full Standard 52.2 test, the nomenclature used to report results is MERV-A (ASHRAE 2008).

DOP Penetration Test For high-efficiency filters of the type used in cleanrooms and nuclear applications (HEPA filters), the normal test in the United States is the thermal DOP method, outlined in U.S. Military Standard MIL-STD-282 (Revision B; 2015) and U.S. Army document 136-300-175A (1965). DOP is dioctyl phthalate or bis-(2-ethylhexyl) phthalate, which is an oily liquid with a high boiling point. In this method, a smoke cloud of DOP droplets condenses from DOP vapor. The count median diameter for DOP aerosols is about 0.18 µm, and the mass median diameter is about 0.27 µm with a cloud concentration of approximately 20 to 80 mg/m3 under properly controlled conditions. The procedure is sensitive to the mass median diameter, and DOP test results are commonly referred to as efficiency at 0.30 µm particle size, believed to be the most penetrating particle size (MPPS) for the filtration materials currently used. This penetration diameter is velocity dependent.

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29.5

The DOP smoke cloud is fed to the filter, which is held in a special test fixture. Any smoke that penetrates the body of the filter or leaks through gasket cracks passes into the region downstream from the filter, where it is thoroughly mixed. Air leaving the fixture thus contains the average concentration of penetrating smoke. This concentration, as well as the upstream concentration, is measured by a light-scattering photometer. Filter penetration P (%) is given as

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 Downstream concentration P = 100  ----------------------------------------------------------------  Upstream concentration 

Note that field data generally show larger uncertainties than measurements made in a laboratory. For filter/filter system tests, this is because of the variety of different HVAC environments and particle types likely to be encountered, and the difficulties in controlling environmental variables during the test. ISO Standard 29462 includes a procedure for reducing uncertainty by subjecting a reference filter to measurement both in the field and in a laboratory.

Other Performance Tests (1)

Penetration, not efficiency, is usually specified in the test procedure because HEPA filters have efficiencies so near 100% (e.g., 99.97% or 99.99% on 0.30 µm particles). Penetration and efficiency E are related by the equation E = 100 – P. U.S. specifications frequently call for testing HEPA filters at both rated flow and 20% of rated flow. This procedure helps detect gasket leaks and pinholes that would otherwise escape notice. Such defects, however, are not located by the DOP penetration test. Other popular test aerosols include polyalphaolefin (PAO Emery 3004), which is also a liquid aerosol similar in particle size distribution to DOP, and polystyrene latex spheres (PSL), a solid aerosol that can be made to a specified size. The Institute of Environmental Sciences and Technology has published two recommended practices: IEST RP-CC 001.6, HEPA and ULPA Filters, and IEST RP-CC 007.3, Testing ULPA Filters.

Leakage (Scan) Tests For HEPA filters, leakage tests are sometimes desirable to show that no small “pinhole” leaks exist or to locate any that may exist so they may be patched. Essentially, this is the same technique as used in the oil aerosol (DOP) penetration test, except that the downstream concentration is measured by scanning the face of the filter and its gasketed perimeter with a moving isokinetic sampling probe. The exact point of smoke penetration can then be located and repaired. This same test (described in IEST RP-CC 001.6) can be performed after the filter is installed. Smoke produced by a portable generator is not uniform in size, but its average diameter can be approximated as 0.6 µm. Particle diameter is less critical for leak location than for penetration measurement. The leak scan test is typically done for filters with 99.99% or higher efficiencies, in applications for which locating and repairing a discrete leak are more important.

ISO Standard 29462 In 2008, ASHRAE published Guideline 26 to provide a test method for evaluating filters in place in building HVAC systems. This guideline was adopted without technical change by the ISO and published as ISO Standard 29462-2013. Subsequently, ASHRAE withdrew Guideline 26. ISO Standard 29462 presents a method for determining the inplace efficiency of individual particle filters or filter systems installed in building HVAC systems, as long as the filter or system is amenable to testing (e.g., enough space in the HVAC system to install sensors, well-sealed doors; a checklist is provided). Using a particle counter, particles in several size ranges between 0.3 and 5 m that are circulating in the HVAC system are measured several times upstream and downstream of the filter to provide statistically robust data. Then, the removal efficiency is calculated by particle size. Pressure drop across the filter, temperature, and relative humidity are recorded. The test method is theoretically applicable to all filters in HVAC systems. However, it is unlikely to yield statistically significant results for filters with efficiencies lower than MERV 11 because of the size distribution of particles typically found in building HVAC systems.

The European Standardization Institute (Comité Européen de Normalisation, or CEN) developed EN Standard 779, Particulate Air Filters for General Ventilation—Requirements, Testing, Marking, in 1993. Its latest revision, Particulate Air Filters for General Ventilation—Determination of the Filtration Performance, was approved and published in 2012. Efficiency is reported as an average efficiency after loading with ASHRAE dust. In this latest revision, minimum efficiency requirements for discharged filters are 35% for F7, 55% for F8, and 70% for F9. This is 0.4 m efficiency, measured on media samples after discharge in liquid isopropyl alcohol (IPA). The test aerosol as specified in the standard is DEHS. Eurovent working group 4B (Air Filters) developed Eurovent Documents 4/9 (1996), Method of Testing Air Filters Used in General Ventilation for Determination of Fractional Efficiency, and 4/10 (2005), In Situ Determination of Fractional Efficiency of General Ventilation Filters. CEN also developed EN Standard 1822-2009, which requires that HEPA and ULPA filters must be tested. Also, special test standards have been developed in the United States for respirator air filters (NIOSH 1995) and ULPA filters (IEST RP-CC 007.2). In 2004, the International Organization for Standardization (ISO) reactivated technical committee TC 142, Cleaning Equipment for Air and Other Gases. The committee’s scope includes test methods for particle filters of various types and efficiencies, particle filter media, and room air cleaners. ASHRAE has cooperated closely with TC 142, and several ASHRAE documents have been used in the development of TC 142 standards. The unanimously approved ISO 16890:2016 set of standards is a new global filtration standard for laboratory testing and rating. It has replaced EN Standard 779 (wherever specified) and is currently being discussed to replace ASHRAE Standard 52.2 in the United States. This standard uses a similar test procedure to ASHRAE Standard 52.2, but includes a unique rating system that relates measured filter performance to typical urban and rural air pollution conditions. The result is a removal efficiency specified in terms of PM and shown as ePM. ISO Standard 16890 is split into four parts: Part 1. Technical specifications, requirements and classification system based upon particulate matter efficiency (ePM) Part 2. Measurement of fractional efficiency and air flow resistance Part 3. Determination of the gravimetric efficiency and the air flow resistance versus the mass of test dust captured Part 4. Conditioning method to determine the minimum fractional test efficiency ISO Standard 29463, parts 2 to 5, have replaced EN Standard 1822, parts 2 to 5, for the method of test for very high efficiency filters. EN Standard 1822 Part 1 remains for the classification of filters.

Environmental Tests Air cleaners may be subjected to fire, high humidity, a wide range of temperatures, mechanical shock, vibration, and other environmental stress. Several standardized tests exist for evaluating these environmental effects on air cleaners. U.S. Military Standard MIL-STD-282 includes shock tests (shipment rough handling) and filter media water-resistance tests. Several U.S. Atomic Energy

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Commission agencies (now part of the U.S. Department of Energy) specify humidity and temperature-resistance tests (DOE 1997, 2003). UL Standard 900. This destructive test procedure measures the amount of flame, smoke, or both generated by a filter when subjected to a burner in a test duct. Additionally, to pass, a filter must not generate sparks beyond the discharge length of the test duct. This procedure does not consider filter performance regarding particle collection at any point of its test life, and pertains to a clean filter only. As filters accumulate particles, the burning, smoke, and spark characteristics change based on the contaminant in the media; thus, UL Standard 900 applies only to new filters. Once a filter has been put in service, the classification is no longer valid. UL Standard 586. This test was specifically designed to evaluate the performance of a HEPA filter in extreme and rigorous conditions. While being particle challenged, the filter is subjected to extreme cold, heat, and humidity. In addition, there are shock tests simulating shipping and handling stresses to evaluate physical durability, and a flame test on the media. The standard also has some specific construction requirements for various portions of each element. In 1991, UL Standard 586 was required by the Department of Defense for Nuclear material applications. It is an integral part of ASME Standard AG-1, Code on Nuclear Air and Gas Treatment. The UL tests do not evaluate the effect of collected dust on filter flammability; depending on the dust, this effect may be severe. UL Standard 867 applies to electronic air cleaners.

AHRI Standards The Air-Conditioning, Heating, and Refrigeration Institute published AHRI Standards 680 (residential) and 850 (commercial/ industrial) for air filter equipment. Standard 680 applies to both media and electronic air cleaners, and specifies rating conditions for performance, capacity, and restriction. These standards establish (1) definitions and classification; (2) requirements for testing and rating; (3) specification of standard equipment; (4) performance and safety requirements; (5) proper marking; (6) conformance conditions; and (7) literature and advertising requirements. However, certification of air cleaners is not a part of these standards.

8.

TYPES OF AIR CLEANERS

Common air cleaners are broadly grouped as follows: In fibrous media unit filters, the accumulating dust load causes pressure drop to increase up to some maximum recommended or predetermined value before changing filters. During this period, efficiency normally increases. However, at high dust loads, dust may adhere poorly to filter fibers and efficiency drops because of offloading. Filters in this condition should be replaced or reconditioned, as should filters that have reached their final (maximum recommended) pressure drop. This category includes viscous impingement and dry air filters, available in low-efficiency to ultrahigh-efficiency construction. Renewable media filters are fibrous media filters where fresh media is introduced into the airstream as needed to maintain essentially constant resistance and, consequently, constant average efficiency. Electronic air cleaners require a power source and, if maintained properly by regular cleaning, have relatively constant pressure drop and efficiency. If they are not cleaned regularly, the accumulated dust can build up to the point that arcing can occur in the collection section. This may reduce collection efficiency as well as produce unwanted extraneous noise. Dust buildup in the ionizer section can also reduce efficiency. Combination air cleaners combine the other types. For example, an electronic air cleaner may be used as an agglomerator with a fibrous media downstream to catch the agglomerated particles

blown off the plates. Low-efficiency pads, throwaway panels and automatically renewable media roll filters, or low- to medium-efficiency pleated prefilters may be used upstream of a high-efficiency filter to extend the life of the better and more costly final filter. Charged media filters are also available that increase particle deposition on media fibers by an induced electrostatic field. With these filters, pressure loss increases as it does on a non-charged fibrous media filter. The benefits of combining different air cleaning processes vary.

9.

FILTER TYPES AND PERFORMANCE

Panel Filters Viscous impingement panel filters are made up of coarse, highly porous fibers. Filter media are generally coated with an odorless, nonmigrating adhesive or other viscous substance, such as oil, which causes particles that impinge on the fibers to stick to them. Design air velocity through the media usually ranges from 1 to 4 m/ s. These filters are characterized by low pressure drop, low cost, and good efficiency on lint and larger particles (10 µm and larger), but low efficiency on normal atmospheric aerosol. They are commonly made 13 to 100 mm thick. Unit panels are available in standard and special sizes up to about 610 by 610 mm. This type of filter is commonly used in residential furnaces and air conditioning and is often used as a prefilter for higher-efficiency filters. Filter media materials include metallic wools, expanded metals and foils, crimped screens, random matted wire, coarse (15 to 60 µm diameter) glass fibers, coated animal hair, vegetable or synthetic fibers, and synthetic open-cell foams. Although viscous impingement filters usually operate between 1.5 and 3 m/s, they may be operated at higher velocities. The limiting factor, other than increased flow resistance, is the danger of blowing off agglomerates of collected dust and the viscous coating on the filter. The loading rate of a filter depends on the type and concentration of dirt in the air being handled and the operating cycle of the system. Manometers, static pressure differential gages, or pressure transducers are often installed to measure pressure drop across the filter bank. This measurement can identify when the filter requires service. The final allowable pressure differential may vary from one installation to another, but, in general, viscous impingement filters are serviced when their operating resistance reaches 125 Pa. Life-cycle cost (LCC), including energy necessary to overcome the filter resistance, should be calculated to evaluate the overall cost of the filtering system. The decline in filter efficiency caused by dust coating the adhesive, rather than by the increased resistance because of dust load, may be the limiting factor in operating life. The manner of servicing unit filters depends on their construction and use. Disposable viscous impingement panel filters are constructed of inexpensive materials and are discarded after one period of use. The cell sides of this design are usually a combination of cardboard and metal stiffeners. Permanent unit filters are generally constructed of metal to withstand repeated handling. Various cleaning methods have been recommended for permanent filters; the most widely used involves washing the filter with steam or water (frequently with detergent) and then recoating it with its recommended adhesive by dipping or spraying. Unit viscous filters are also sometimes arranged for in-place washing and recoating. The adhesive used on a viscous impingement filter requires careful engineering. Filter efficiency and dust-holding capacity depend on the specific type and quantity of adhesive used; this information is an essential part of test data and filter specifications. Desirable adhesive characteristics, in addition to efficiency and dust-holding capacity, are (1) a low percentage of volatiles to prevent excessive evaporation; (2) viscosity that varies only slightly within the service temperature range; (3) the ability to inhibit growth of bacteria and

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Air Cleaners for Particulate Contaminants mold spores; (4) high capillarity or ability to wet and retain dust particles; (5) high flash point and fire point; and (6) freedom from odorants or irritants. Typical performance of viscous impingement unit filters operating within typical resistance limits is shown as MERV 1 through 6 in Table 3. Dry extended-surface filters use media of random fiber mats or blankets of varying thicknesses, fiber sizes, and densities. Bonded glass fiber, cellulose fibers, wool felt, polymers, synthetics, and other materials have been used commercially. Media in these filters are frequently supported by a wire frame in the form of pockets, or V-shaped or radial pleats. In other designs, the media may be self supporting because of inherent rigidity or because airflow inflates it into extended form (e.g., bag filters). Pleating media provides a high ratio of media area to face area, thus allowing reasonable pressure drop and low media velocities. In some designs, the filter media is replaceable and is held in position in permanent wire baskets. In most designs, the entire cell is discarded after it has accumulated its maximum dust load. Efficiency is usually higher than that of panel filters, and the variety of media available makes it possible to furnish almost any degree of cleaning efficiency desired. The dust-holding capacities of modern dry filter media and filter configurations are generally higher than those of panel filters. Using coarse prefilters upstream of extended-surface filters is sometimes justified economically by the longer life of the main filters. Economic considerations include the prefilter material cost, changeout labor, and increased fan power. Generally, prefilters should be considered only if they can substantially reduce the part of the dust that may plug the protected filter. A prefilter usually has an arrestance of at least 70% (MERV 3), but is commonly rated up to 92% (MERV 6). Temporary prefilters protecting higher-efficiency filters are worthwhile during construction to capture heavy loads of coarse dust. Filters of MERV 16 and greater should always be protected by prefilters. A single filter gage may be installed when a panel prefilter is placed adjacent to a final filter. Because the prefilter is frequently changed on a schedule, the final filter pressure drop can be read without the prefilter in place every time the prefilter is changed. For maximum accuracy and economy of prefilter use, two gages can be used. Some air filter housings are available with pressure taps between the pre- and final filter tracks to accommodate this arrangement. Typical performance of some types of filters in this group, when operated within typical rated resistance limits and over the life of the filters, is shown as MERV 7 through 16 in Table 3. Initial resistance of an extended-surface filter varies with the choice of media and filter geometry. Commercial designs typically have an initial resistance from 25 to 250 Pa. It is customary to replace the media when the final resistance of 125 Pa is reached for low-resistance units and 500 Pa for the highest-resistance units. Dry media providing higher orders of cleaning efficiency have a higher average resistance to airflow. The operating resistance of the fully dust-loaded filter must be considered in design, because that is the maximum resistance against which the fan operates. Variable-airvolume and constant-air-volume system controls prevent abnormally high airflows or possible fan motor overloading from occurring when filters are clean. Flat panel filters with media velocity equal to duct velocity are made only with the lowest-efficiency dry-type media (open-cell foams and textile denier nonwoven media). Initial resistance of this group, at rated airflow, is generally between 10 and 60 Pa. They are usually operated to a final resistance of 120 to 175 Pa. In intermediate-efficiency extended-surface filters, the filter media area is much greater than the face area of the filter; hence, velocity through the filter media is substantially lower than the velocity approaching the filter face. Media velocities range from

29.7 0.03 to 0.5 m/s, although approach velocities run to 4 m/s. Depth in direction of airflow varies from 50 to 900 mm. Intermediate-efficiency filter media include (1) fine glass or synthetic fibers, from nanofiber to 10 µm in diameter, in mats up to 13 mm thick; (2) wet laid paper or thin nonwoven mats of fine glass fibers, cellulose, or cotton wadding; and (3) nonwoven mats of comparatively large-diameter fibers (more than 30 µm) in greater thicknesses (up to 50 mm). Electret filters, which require no power, are composed of electrostatically charged fibers. The charges on the fibers augment collection of smaller particles by interception and diffusion (Brownian motion) with Coulomb forces caused by the charges. Examples of this type of filter include resin wool, electret, and an electrostatically sprayed polymer. The charge on resin wool fibers is produced by friction during the carding process. During production of the electret, a corona discharge injects positive charges on one side of a thin polypropylene film and negative charges on the other side. These thin sheets are then shredded into fibers of rectangular cross section. The third process spins a liquid polymer into fibers in the presence of a strong electric field, which produces the charge separation. Efficiency of charged-fiber filters is determined by both the normal collection mechanisms of a media filter (related to fiber diameter) and the strong local electrostatic effects (related to the amount of electrostatic charge). The effects induce efficient preliminary loading of the filter to enhance the caking process. However, ultrafine-particle dust collected on the media can affect the efficiency of electret filters. Very high-efficiency dry filters, HEPA (high-efficiency particulate air) filters, and ULPA (ultralow-penetration air) filters are made in an extended-surface configuration of deep space folds of submicrometre glass fiber paper. These filters operate at duct velocities from 1.3 to 2.6 m/s, with resistance rising from 120 to more than 500 Pa over their service life. These filters are the standard for cleanroom, nuclear, and toxic particulate applications, and are increasingly used in numerous medical and pharmaceutical applications. Membrane filters are used mainly for air sampling and specialized small-scale applications where their particular characteristics compensate for their fragility, high resistance, and high cost. They are available in many pore diameters and resistances and in flatsheet and pleated forms. Renewable-media filters may be one of two types: (1) movingcurtain viscous impingement filters or (2) moving-curtain drymedia roll filter. Commonly described as automatic roll filters, these are typically lower on the efficiency scale. In one viscous type, random-fiber (nonwoven) media is furnished in roll form. Fresh media is fed manually or automatically across the face of the filter, while the dirty media is rewound onto a roll at the bottom. When the roll is exhausted, the tail of the media is wound onto the take-up roll, and the entire roll is thrown away. A new roll is then installed, and the cycle repeats. Moving-curtain filters may have the media automatically advanced by motor drives on command from a pressure switch, timer, or media light-transmission control. A pressure switch control measures the pressure drop across the media and switches on and off at chosen upper and lower set points. This saves media, but only if the static pressure probes are located properly and unaffected by modulating outdoor and return air dampers. Most pressure drop controls do not work well in practice. Timers and media light-transmission controls help avoid these problems; their duty cycles can usually be adjusted to provide satisfactory operation with acceptable media consumption. Filters of this replaceable roll design generally have a signal indicating when the roll is nearly exhausted. At the same time, the drive motor is deenergized so that the filter cannot run out of media. Normal service requirements involve inserting a clean roll of media

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29.8

2019 ASHRAE Handbook—HVAC Applications (SI) Table 2

Description

Performance of Renewable Media Filters (Steady-State Values) ASHRAE Type ASHRAE Dust-Holding of Media Arrestance, % Capacity, g/m2

Licensed for single user. © 2020 ASHRAE, Inc.

20 to 40 µm glass and Viscous synthetic fibers, impingement 50 to 65 mm thick Permanent metal media cells Viscous or overlapping elements impingement Coarse textile denier nonwoDry ven mat, 12 to 25 mm thick Fine textile denier nonwoven Dry mat, 12 to 25 mm thick

70 to 82

600 to 2000

70 to 80

N/A (permanent media)

60 to 80

150 to 750

80 to 90

100 to 550

Fig. 2 Cross Section of Plate-Type Precipitator Air Cleaner

at the top of the filter and disposing of the loaded dirty roll. Automatic filters of this design are not, however, limited to the vertical position; horizontal arrangements are available for makeup air and air-conditioning units. Adhesives must have qualities similar to those for panel viscous impingement filters, and they must withstand media compression and endure long storage. The second type of automatic viscous impingement filter consists of linked metal mesh media panels installed on a traveling curtain that intermittently passes through an adhesive reservoir. In the reservoir, the panels give up their dust load and, at the same time, take on a new coating of adhesive. The panels thus form a continuous curtain that moves up one face and down the other face. The media curtain, continually cleaned and renewed with fresh adhesive, lasts the life of the filter mechanism. The precipitated captured dirt must be removed periodically from the adhesive reservoir. New installations of this type of filter are rare in North America, but are often found in Europe and Asia. The resistance of both types of viscous impingement automatically renewable filters remains approximately constant as long as proper operation is maintained. A resistance of 100 to 125 Pa at a face velocity of 2.5 m/s is typical of this class. Special automatic dry filters are also available, designed for removing lint in textile mills, laundries, and dry-cleaning establishments and for collecting lint and ink mist in printing press rooms. The medium used is extremely thin and serves only as a base for the build-up of lint, which then acts as a filter medium. The dirt-laden media is discarded when the supply roll is used up. Another form of filter designed specifically for dry lint removal consists of a moving curtain of wire screen, which is vacuum cleaned automatically at a position out of the airstream. Recovery of the collected lint is sometimes possible with these devices. ASHRAE arrestance and dust-holding capacities for typical viscous impingement and dry renewable-media filters are listed in Table 2.

Electronic Air Cleaners Electronic air cleaners use an electrostatic charge to enhance filtration of particulate contaminants such as dust, smoke, and pollen. The electrostatic charge can create higher efficiencies than mechanical means alone. Electronic air cleaners are available in many designs, but fall into two major categories: (1) electronic, plate-type precipitators and (2) electrically enhanced air filtration. Plate Precipitators. Precipitators use electrostatic precipitation to remove and collect particulate contaminants on plates. The air cleaner has an ionization section and a collecting plate section. In the ionization section, small-diameter wires with a positive direct current potential between 6 and 25 kV are suspended equidistant between grounded plates. The high voltage on the wires creates an ionizing field for charging particles. The positive ions created in the field flow across the airstream and strike and adhere to the particles,

Fig. 3

Electrically Enhanced Air Cleaner

imparting a charge to them. The charged particles then pass into the collecting plate section. The collecting plate section consists of a series of parallel plates equally spaced with a positive direct current voltage of 4 to 10 kV applied to alternate plates. Plates that are not charged are at ground potential. As the particles pass into this section, they are attracted to the plates by the electric field on the charges they carry; thus, they are removed from the airstream and collected by the plates. Particle retention is a combination of electrical and intermolecular adhesion forces and may be augmented by special oils or adhesives on the plates. Figure 2 shows a typical electronic air cleaner cell. In lieu of positive direct current, a negative potential also functions on the same principle, but generates more ozone. With voltages of 4 to 25 kV (dc), safety measures are required. A typical arrangement makes the air cleaner inoperative when the doors are removed for cleaning the cells or servicing the power pack. Electronic air cleaners typically operate from a 120 or 240 V (ac) single-phase electrical service. The high voltage supplied to the air cleaner cells is normally created with solid-state power supplies. The electric power consumption ranges from 40 to 85 W per m3/s of air cleaner capacity. Electrically Enhanced Air Filtration. Electrically enhanced air cleaners incorporate an electrostatic field to charge contaminants before capture in a high-efficiency pleated filter. Advantages include high efficiency and reduced maintenance frequency. Figure 3 shows that the air cleaners consist of an ionizing section and a filtration section. The ionizing section has a prefilter to prevent large debris from entering the air filter and to focus the electrostatic field. The air is charged in a high-voltage ionizing section. In the ionization section, the ionizer is connected to a high-voltage power supply

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Air Cleaners for Particulate Contaminants and the particulate is charged. The charged particles are collected in the media filter at earth ground potential. Maintenance. Plate-type air cleaner cells must be cleaned periodically with detergent and hot water. Some designs incorporate automatic wash systems that clean the cells in place; in others, the cells are removed for cleaning. The frequency of cleaning (washing) the cell depends on the contaminant and the concentration. Industrial applications may require cleaning every 8 h, but a residential unit may only require cleaning every one to three months. The timing of the cleaning schedule is important to keep the unit performing at peak efficiency. For some contaminants, special attention must be given to cleaning the ionizing wires. The air cleaner must be maintained based on the recommendations of the manufacturer. Electrically enhanced air cleaners have a longer service life between maintenance than plate-type precipitators. Maintenance consists of replacing the filter and cleaning the ionizing section and prefilter. Performance. Currently AHRI Standard 680 is the industryaccepted test method for electronic air cleaners. This test involves loading the filter with a dust that does not contain a conductive component and allows comparison to media filtration. Application. As with most air filtration devices, duct approaches to and from the air cleaner housing should be arranged so that airflow is distributed uniformly over the face area. Panel prefilters should also be used to help distribute airflow and to trap large particles that might short out or cause excessive arcing in the high-voltage section of the air cleaner cell. Electronic air cleaner design parameters of air velocity, ionizer field strength, cell plate spacing, depth, and plate voltage must match the application requirements (e.g., contaminant type, particle size, volume of air, required efficiency). Many units are designed for installation into central heating and cooling systems for total air filtration. Other self-contained units are furnished complete with air movers for source control of contaminants in specific applications that need an independent air cleaner. Optional features are often available for electronic air cleaners. Afterfilters such as roll filters collect particulates that agglomerate and blow off the cell plates. These are used mainly where heavy contaminant loading occurs and extension of the cleaning cycle is desired. Cell collector plates may be coated with special oils, adhesives, or detergents to improve both particle retention and particle removal during cleaning. High-efficiency dry extended-media area filters are also used as afterfilters in special designs. The electronic air cleaner used in this system improves the service life of the dry filter and collects small particles such as smoke. A negative ionizer uses the principle of particle charging, but does not use a collecting section. Particles enter the ionizer of the unit, receive an electrical charge, and then migrate to a grounded surface closest to the travel path. Space Charge. Particulates that pass through an ionizer and are charged, but not removed, carry the electrical charge into the space. If continued on a large scale, a space charge builds up, which tends to drive these charged particles to walls and interior surfaces. Thus, a low-efficiency electronic air cleaner used in areas of high ambient dirt concentrations (or a malfunctioning unit), can blacken walls faster than if no cleaning device were used (SMACNA 2010). Ozone. All high-voltage devices can produce ozone, which is toxic and damaging not only to human lungs, but to paper, rubber, and other materials. When properly designed and maintained, an electronic air cleaner produces an ozone concentration that only reaches a fraction of the limit acceptable for continuous human exposure and is less than that prevalent in many American cities (EPA 1996). Continuous arcing and brush discharge in an electronic air cleaner may yield an ozone level that is annoying or mildly toxic; this is indicated by a strong ozone odor. Although the nose is sensitive to ozone, only actual measurement of the concentration can determine whether a hazardous condition exists.

29.9 Outdoor air can also be a source of indoor ozone. Weschler et al. (1989) found that ozone levels indoors closely tracked outdoor levels, despite ozone’s reactions with HVAC interior surfaces. Indoor concentrations were typically 20 to 80% of outdoor concentrations depending on the ventilation rate. The U.S. Environmental Protection Agency (EPA) limits the maximum allowable exposure to ozone in outdoor air to 0.070 ppm averaged over 8 h (EPA 2015). ASHRAE Standard 62.1 requires ozone removal systems in buildings where the intake air concentration exceeds the EPA limit.

10.

SELECTION AND MAINTENANCE

To evaluate filters and air cleaners properly for a particular application, consider the following factors: • • • • • •

Types of contaminants present indoors and outdoors Sizes and concentrations of contaminants Air cleanliness levels required in the space Air filter efficiency needed to achieve cleanliness Space available to install and access equipment Life-cycle costing, including • Operating resistance to airflow (static pressure differential) • Disposal or cleaning requirements of spent filters • Initial cost of selected system • Cost of replacement filters or cleaning • Cost of warehousing filter stock and change-out labor

Savings (from reduced housekeeping expenses, protection of valuable property and equipment, dust-free environments for manufacturing processes, improved working conditions, and even health benefits) should be credited against the cost of installing and operating an adequate system. The capacity and physical size of the required unit may emphasize the need for low maintenance cost. Operating cost, predicted life, and efficiency are as important as initial cost because air cleaning is a continuing process. Panel filters do not have efficiencies as high as can be expected from extended-surface filters, but their initial cost and upkeep are generally low. Compared to moving-curtain filters, panel filters of comparable efficiencies require more attention to maintain the resistance within reasonable limits. However, single-stage, face- or sideaccess, low- to medium-efficiency filters of MERV 6 to 10 from a 50 mm pleat to a 300 mm deep cube, bag, or deep pleated cartridge, require less space with lower initial cost, and have better efficiency. The bag and cartridges generally have a similar service life to that of a roll filter. If efficiency of MERV 11 or higher is required, extended-surface filters or electronic air cleaners should be considered. The use of very fine glass fiber mats or other materials in extended-surface filters has made these available at the highest efficiency. Initial cost of an extended-surface filter is lower than for an electronic unit, but higher than for a panel type. Operating and maintenance costs of some extended-surface filters may be higher than for panel types and electronic air cleaners, but efficiencies are always higher than for panel types; the cost/benefit ratio must be considered. Pressure drop of media-type filters is greater than that of electronic types, and slowly increases during their useful life. Advantages include the fact that no mechanical or electrical services are required. Choice should be based on both initial and operating costs (life-cycle costs), as well as on the degree of cleaning efficiency and maintenance requirements. Although electronic air cleaners have higher initial and maintenance costs, they have high initial efficiencies in cleaning atmospheric air, largely because of their ability to remove fine particulate contaminants. System resistance remains unchanged as particles are collected, and efficiency is reduced until the resulting residue is removed from the collection plates to prepare the equipment for further duty. The manufacturer must supply information

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29.10

2019 ASHRAE Handbook—HVAC Applications (SI)

Fig. 4 Typical Filter Locations for HVAC System Table 3 Cross-Reference and Application Guidelines Standard 52.2 Arrestance MERV Value

Example Range of Contaminants Controlled

Example Applications

Sample Air Cleaner Type(s)

N/A N/A N/A N/A

0.3 to 1.0 m size range: bacteria, smoke (ETS), paint pigments, face powder, some virus, droplet nuclei, insecticide dusts, soldering fumes

Day surgery, general surgery, hospital general ventilation, turbo equipment, compressors, welding/soldering air cleaners, prefilters to HEPAs, LEED for existing (EB) and new (NC) commercial buildings, smoking lounges

Box-style wet-laid or lofted fiberglass, boxstyle synthetic media, minipleated synthetic or fiberglass paper, depths from 50 to 300 mm. Pocket filters of fiberglass or synthetic media 300 to 900 mm.

N/A N/A N/A N/A

1.0 to 3.0 m size range: milled flour, lead dust, combustion soot, Legionella, coal dust, some bacteria, process grinding dust

Food processing facilities, air separation plants, commercial buildings, better residential, industrial air cleaning, prefiltration to higherefficiency filters, schools, gymnasiums

Box-style wet-laid or lofted fiberglass, boxstyle synthetic media, minipleated synthetic or fiberglass paper, depths from 50 to 300 mm. Pocket filters either rigid or flexible in synthetic or fiberglass, depths from 300 to 900 mm.

MERV 8 MERV 7 MERV 6 MERV 5

N/A N/A N/A N/A

3.0 to 10 m size range: pollens, earth-origin dust, mold spores, cement dust, powdered milk, snuff, hair spray mist

General HVAC filtration, industrial equipment Wide range of pleated media, ring panels, cubes, pockets in synthetic or fiberglass, filtration, commercial property, schools, prefilter to high-efficiency filters, paint booth disposable panels, depths from 25 to 600 intakes, electrical/phone equipment protection mm.

MERV 4 MERV 3 MERV 2 MERV 1

>70% >70% >65% 10 >15 >5 >10 >10

>10 >10 >2 >2 >2 >2 >2 >2

Pressure Loss Liquid, kPa

50 50 70 80 85 95 90 90

25 to 125 125 to 625 0.01 Low-voltage >0.001

>0.2 1

Fabric filters Baghouses Cartridge filters

>0.08 >0.05

.>1 >0.2

99 99+

500 to 1500 500 to 2000

Wet scrubbers Gravity spray Centrifugal Impingement Packed bed Dynamic

>10 >5 >5 >5 >2

>2 >2 >2 >0.2 >2

70 90 95 90 95

25 to 250 500 to 2000 500 to 2000 125 to 2500 Provides pressure

>2 >2 >0.1

>0.2 >0.2 >0.2

Submerged orifice Jet Venturi

Comparative Capacity Space Energy Superficial Limits, Required Utilities m3/s (Relative) per m3/s (gas) Requirement Velocity,b m/s

Gas, Pa

—

140 to 690 70 to 270 L/s 140 to 690 0.14 to 1.4 L/s 140 to 690 0.14 to 0.7 L/s 35 to 210 0.7 to 70 L/s 35 to 210 0.14 to 0.7 L/s, 2.25 to 15 kW 90 500 to 1500 None No pumping 90 Provides pressure 345 to 610 7 to 14 L/s 95 to 99 2500 to 15 000 70 to 210 0.4 to 1.4 L/s

Source: IGCI (1964). Information updated by ASHRAE Technical Committee 5.4. particle diameter for which the device is effective.

a Minimum

Fig. 1 Typical Louver and Baffle Collectors

Inertial Collectors Louver and Baffle Collectors. Louvers are widely used to control particles larger than about 15 m in diameter. The louvers cause a sudden change in direction of gas flow. By virtue of their inertia, particles move away from high-velocity gases and are either collected in a hopper or trap or withdrawn in a concentrated sidestream. The sidestream is cleaned using a cyclone or high-efficiency collector, or it is simply discharged to the atmosphere. In general, the pressure drop across inertial collectors with louvers or baffles is greater than that for settling chambers, but this loss is balanced by higher collection efficiency and more compact equipment. Inertial collectors are occasionally used to control mist. In some applications, the interior of the collector may be irrigated to prevent reentrainment of dry dust and to remove soluble deposits. Typical louver and baffle collectors are shown in Figure 1.

b Average

1 1.5 3.0 1.5 to 6.0 1.5 to 9.0 6.0 to 20 3.0 to 6.0 10 to 20

1.5 to 3.0 5.0 to 10 10 to 20 10 to 20 10 to 20 10 to 20 10 to 20 —

None None 25 15 25 95 None 25

Large Medium Small Medium Medium Small Small —

0.3 to 2.0 1.0 to 3.5

5 to 940 Large 0.5 to 50 Medium

6.0 to 20

0.005 to 0.10 0.0025 to 0.025

95 Large 20 to 25 Medium

5.0 12 to 26 9.0 to 31 4.0 to 34 30 to 200

0.5 to 1 10 to 20 15 to 30 0.5 to 1.5 15 to 20

9.0 to 21 15 to 30 30 to 300

15 10 to 100 60 to 210

50 50 50 25 25 25 25 50 50

Medium Medium Medium Medium Small Medium Small Small

speed of gases flowing through the equipment’s collection region.

Cyclones and Multicyclones. A cyclone collector transforms a gas stream into a confined vortex, from which inertia moves suspended particles to the wall of the cyclone’s body. The inertial effect of turning the gas stream, as used in the baffle collector, is used continuously in a cyclone to improve collection efficiency. Cyclone collectors are often used as precleaners to reduce the loading of more efficient pollution control devices. Figure 2 shows some typical cyclone collectors. A low-efficiency cyclone operates with a static pressure drop from 250 to 370 Pa between its inlet and outlet and can remove 50% of the particles from 5 to 10 m. High-efficiency cyclones operate with static pressure drops from 0.75 to 2 kPa between their inlet and outlet and can remove 70% of the particulates of approximately 5 m. The efficiency of a cyclone depends on particle density, shape, and size (aerodynamic size Dp , which is the average of the size range). Cyclone efficiency may be estimated from Figure 3. The parameter Dp c , known as the cut size, is defined as the diameter of particles collected with 50% efficiency. The cut size may be estimated using the following equation: Dpc =

1.4 b -------------------------------------N e V i   p –  g 

(5)

where Dpc  b Ne

= = = =

cut size, m absolute gas viscosity, Pa·s cyclone inlet width, m effective number of turns in cyclone; approximately 5 for a highefficiency cyclone and may be from 0.5 to 10 for other cyclones Vi = inlet gas velocity, m/s p = density of particle material, kg/m3 g = density of gas, kg/m3

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Industrial Gas Cleaning and Air Pollution Control

30.5

Licensed for single user. © 2020 ASHRAE, Inc.

Fig. 4 Typical Single-Stage Electrostatic Precipitator

Fig. 2 Typical Cyclone Collectors

Fig. 5 Fig. 3 Cyclone Efficiency (Lapple 1951)

At inlet gas velocity above 25 m/s, internal turbulence limits improvements in the efficiency of a given cyclone. The pressure drop through a cyclone is proportional to the inlet velocity pressure and hence the square of the volumetric flow.

2.2

ELECTROSTATIC PRECIPITATORS

Electrostatic precipitators use the forces acting on charged particles passing through an electric field to separate those particles from the airstream in which they were suspended. In every precipitator, three distinct functions must be accomplished: • Ionization: charging contaminant particles • Collection: subjecting particles to a precipitating force that moves them toward collecting electrodes

Typical Two-Stage Electrostatic Precipitators

• Collector cleaning: removing collected contaminant from precipitator Units in which ionization and collection are accomplished simultaneously in a single structure are called single-stage precipitators (Figure 4). They have widely spaced electrodes (80 to 150 mm) and typically operate with high voltages (20 to 60 kV) but relatively low (rarely as high as 0.43 kV/mm) field gradients.In two-stage precipitators, ionization and collection are performed independently in discrete charging and precipitating structures (Figure 5). Because their ionizing and collecting electrodes are closely spaced (18 to 38 mm), two-stage precipitators normally operate with high field gradients (usually more than 0.39 kV/mm) but low voltages (usually 10 kV or less, and never more than 14 kV) (White 1963). Because of fundamental differences in their ionization processes and practical differences in the way they are usually constructed, high-voltage (single-stage) and low-voltage (two-stage) precipitators are suited for entirely different air-cleaning requirements.

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30.6

2020 ASHRAE Handbook—HVAC Systems and Equipment(SI)

Licensed for single user. © 2020 ASHRAE, Inc.

Table 4 Collectors Used in Industry

Operation Ceramics Raw product handling Fettling Refractory sizing Glaze and vitreous enamel spray Glass melting Frit smelting Fiberglass forming and curing Chemicals Material handling Crushing, grinding Pneumatic conveying Roasters, kilns, coolers Incineration Coal mining and handling Material handling Bunker ventilation Dedusting, air cleaning Drying Combustion fly ash Coal burning: Chain grate Spreader stoker Pulverized coal Fluidized bed Coal slurry Wood waste Municipal refuse Oil Biomass Foundry Shakeout Sand handling Tumbling mills Abrasive cleaning Grain elevator, flour and feed mills Grain handling Grain drying Flour dust Feed mill Metal melting Steel blast furnace Steel open hearth, basic oxygen furnace Steel electric furnace Ferrous cupola Nonferrous reverberatory furnace Nonferrous crucible

Rotating HighEfficiency Centrifugal Mist Centrifugal

Wet Collectors

Self-Cleaning Disposable Media Fabric Filter Filter

Electrostatic Precipitators HighVoltage

LowVoltage

Notes

N/A N/A N/A N/A N/A N/A Rare

N/U N/U N/U N/A Usual Often Usual

N/A N/A N/A N/A N/A N/A N/A

1 2 3 — — — —

Frequent Frequent Usual Rare Rare

N/A N/A N/A N/A N/A

N/U N/U N/U Often Frequent

N/A N/A N/A N/A N/A

4 5 6 7 8

N/U N/U N/U Occasional

Usual Usual Usual N/U

N/A N/A N/A N/A

N/A N/A N/A N/A

N/A N/A N/A N/A

9 10 11 12

N/U N/U N/U — — N/U Occasional N/U Occasional

N/U N/U N/U — — N/U N/U N/U N/U

Frequent Frequent Frequent Frequent Often Occasional Usual Usual Usual

N/A N/A N/A N/A N/A N/A N/A N/A N/A

N/U Rare Usual Frequent Often Often Frequent Often Frequent

N/A N/A N/A N/A N/A N/A N/A N/A N/A

13 14 14 — — 15 — — —

N/A N/A N/A N/A

Rare Usual Frequent Frequent

Seldom N/U N/U N/U

Usual Rare Usual Usual

N/A N/A N/A N/A

N/U N/U N/U N/U

N/A N/A N/A N/A

16 17 18 19

Occasional N/A Often Often

N/A N/A N/A N/A

Rare N/U Occasional Occasional

N/U N/U N/U N/U

Frequent See Note 20 Usual Frequent

N/A N/A N/A N/A

N/A N/A N/A N/A

N/A N/A N/A N/A

20 21 22 23

Frequent N/A

Rare N/A

N/A N/A

Frequent N/A

Frequent Often

N/U Rare

N/A N/A

Frequent Frequent

N/A N/A

24 25

N/A N/A N/A N/A

N/A N/A N/A N/A

N/A N/A N/A N/A

N/A Frequent Rare Rare

Occasional Often Occasional Rare

Usual Frequent Usual Occasional

N/A N/A N/A N/A

Rare Occasional N/U N/U

N/A N/A N/A N/A

26 27 28 29

MediumPressure

HighEnergy

Concentration

Particle Size

Cyclone

Light Light Heavy Moderate Light Light Light

Fine Fine to medium Coarse Medium Fine Fine Fine

Rare Rare Seldom N/U N/A N/A N/A

Seldom Occasional Occasional N/U N/A N/A N/A

N/A N/A N/A N/A N/A N/A N/A

Frequent Frequent Frequent Usual Occasional N/U Occasional

N/U N/U Rare N/U N/U Often N/U

Frequent Frequent Frequent Occasional Occasional Often N/U

Light to moderate Moderate to heavy Very heavy Heavy Light to medium

Fine to medium Fine to coarse Fine to coarse Medium to coarse Fine

Occasional Often Usual Occasional N/U

Frequent Frequent Occasional Usual N/U

N/A N/A N/A N/A N/A

Frequent Frequent Rare Usual N/U

Frequent Occasional Rare Frequent Frequent

Moderate Moderate Heavy Moderate

Medium Fine Medium to coarse Fine

Rare Occasional Occasional Rare

Occasional Frequent Frequent Occasional

N/A N/A N/A N/A

Occasional Occasional Occasional Frequent

Light Moderate Heavy Moderate Light Varied Light Light Moderate

Fine Fine to coarse Fine Fine — Coarse Fine Fine Fine to coarse

N/A Rare N/A Usual — Usual N/U N/U N/U

Rare Rare Frequent — — Usual N/U N/U N/U

N/A N/A N/A N/A N/A N/A N/A N/A N/A

Light to moderate Moderate Moderate Moderate to heavy

Fine Fine to medium Medium to coarse Fine to medium

Rare Rare N/A N/A

Rare Rare N/A Occasional

Light Light Moderate Moderate

Medium Coarse Medium Medium

Usual N/A Rare Often

Heavy Moderate

Varied Fine to coarse

Light Moderate Varied Light

Fine Varied Fine Fine

2020 ASHRAE Handbook—HVAC Systems and Equipment(SI)

30.7

Table 4 Collectors Used in Industry (Continued)

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Operation Metal mining and rock products Material handling Dryers, kilns Cement rock dryer Cement kiln Cement grinding Cement clinker cooler Metal working Production grinding, scratch brushing, abrasive cutoff Portable and swing frame Buffing Tool room Cast-iron machining Steel, brass, aluminum machining

Rotating HighEfficiency Centrifugal Mist Centrifugal

Concentration

Particle Size

Cyclone

Moderate Moderate Moderate Heavy Moderate Moderate

Fine to medium Medium to coarse Fine to medium Fine to medium Fine Coarse

Rare Frequent N/A N/A N/A N/A

Light

Coarse

Occasional Frequent

Light Light Light Moderate Light to moderate

Medium Varied Fine Varied 1 m smoke, med. mist to solids Welding Light to moderate 1 m fume to med. Plasma and laser cutting Moderate Laser welding Moderate Abrasive machining Moderate to heavy Fine to 1 m Milling, turning, cutting tools Light to moderate Fine to 1 m Annealing, heat treating, induction Moderate to heavy 1 m heating, quenching Pharmaceutical and food products Mixers, grinders, weighting, Light Medium blending, bagging, packaging Coating pans Varied Fine to medium Plastics Raw material processing (See comments under Chemicals) Plastic finishing Light to moderate Varied Molding, extruding, curing Light to moderate 1 m smoke Pulp and paper Recovery boilers: Direct contact Heavy Medium Low odor Heavy Medium Lime kilns Heavy Coarse Wood-chip dryers Varied Fine to coarse Rubber products Mixers Moderate Fine Batch-out rolls Light Fine Talc dusting and dedusting Moderate Medium Grinding Moderate Coarse Molding, extruding, curing Light to moderate 1 m smoke Wood particle board and hard board Particle dryers Moderate Fine to coarse Woodworking Woodworking machines Moderate Varied Sanding Moderate Fine Waste conveying, hogs Heavy Varied Source: Kane and Alden (1982). Information updated by ASHRAE Technical Committee 5.4.

Occasional Occasional Frequent Frequent Rare Occasional

Wet Collectors MediumPressure

N/A N/A N/A N/A N/A N/A

Usual Frequent Occasional Rare N/U N/U

N/A N/A N/A N/A N/A Frequent

HighEnergy N/U Occasional Rare N/U N/U N/U

Self-Cleaning Disposable Electrostatic Precipitators Media HighLowFabric Filter Voltage Voltage Notes Filter Considerable N/U N/U Usual Usual Occasional

N/A N/A N/A N/A N/A N/A

N/U Occasional Occasional Usual Rare N/U

N/A N/A N/A N/A N/A N/A

30 31 30 32 33 34

Considerable N/U

Considerable

N/A

N/U

N/A

35

Frequent Frequent Frequent Considerable Occasional

N/U N/U N/U N/U N/U

Considerable Rare Frequent Considerable Occasional

N/A N/A N/A N/A Frequent

N/U N/U N/U N/U N/U

N/A N/A N/A N/A Frequent

— 36 37 38 39

Occasional Occasional Occasional Occasional Occasional Rare

N/U N/U N/U N/U N/U Rare

Frequent Frequent Frequent Rare N/A N/A

Frequent Rare Rare Frequent Frequent Rare

Rare N/A N/A N/A N/A N/A

Occasional N/U N/U Rare Frequent Frequent

40 41 41 39 — —

Rare Frequent Frequent Rare N/A

Frequent Rare Frequent Frequent N/A

N/A N/A N/A N/A N/A N/A

N/A N/A N/A N/U N/U N/U

Rare

Frequent

N/A

Frequent

N/U

Frequent

Occasional

N/U

N/U

42

Rare

Rare

N/A

Frequent

N/U

Frequent

Rare

N/U

N/U

43

Frequent N/A

Frequent N/A

N/A N/A

Frequent Rare

N/U N/U

Frequent N/A

Rare Occasional

N/U N/U

N/U Considerable

44 45 46

N/U N/U N/U N/U

N/U N/U N/U N/U

N/A N/A N/A N/A

N/U N/U N/U N/U

N/U N/U N/U N/U

Occasional Occasional Often Occasional

N/A N/A N/A N/A

Usual Usual Often Often

N/A N/A N/A N/A

— — — —

N/A N/A N/A Often N/A

N/A N/A N/A Often N/A

N/A N/A N/A N/A N/A

Frequent Usual Frequent Frequent Rare

N/U N/U N/U N/U N/U

Usual Frequent Usual Often N/A

Rare N/A Rare Rare Occasional

N/U N/U N/U N/U N/A

N/U Rare N/U N/U Considerable

47 48 49 50 46

Usual

Occasional

N/A

Frequent

Occasional Rare

N/A

Occasional

Rare

51

Usual Frequent Usual

Occasional Occasional Rare

N/A N/A N/A

Rare Occasional Occasional

N/U N/U N/U

N/A Rare N/A

N/U N/U N/U

N/A N/A N/A

52 53 54

N/A N/A N/A Occasional Frequent N/A

Frequent Frequent Occasional

30.8

2020 ASHRAE Handbook—HVAC Systems and Equipment(SI) Notes for Table 4

Definitions N/A: Not applicable because of inefficiency or process incompatibility. N/U: Not widely used. Particle size Fine: 50% in 0.5 to 7 m diameter range Medium: 50% in 7 to 15 m diameter range Coarse: 50% over 15 m diameter range Concentration of particulate matter entering collector (loading) Light: 10 g/m3

26 27

1

Dust released from bin filling, conveying, weighing, mixing, pressing, forming. Refractory products, dry pan, and screening operations more severe. Operations found in vitreous enameling, wall and floor tile, pottery. Grinding wheel or abrasive cutoff operation. Dust abrasive. Operations include conveying, elevating, mixing, screening, weighing, packaging. Category covers so many different materials that recommendation will vary widely. 5 Cyclone and high-efficiency centrifugal collectors often act as primary collectors, followed by fabric filters or wet collectors. 6 Usual setup uses cyclone as product collector followed by fabric filter for high overall collection efficiency. 7 Dust concentration determines need for dry centrifugal collector; plant location, product value determine need for final collectors. High temperatures are usual, and corrosive gases not unusual. Liquid smoke emissions may be controlled by condensing precipitator systems using low-voltage, two-stage electrostatic precipitators. 8 Ionizing wet scrubbers are widely used. 9 Conveying, screening, crushing, unloading. 10 Remote from other dust-producing points. Separate collector generally used. 11 Heavy loading suggests final high-efficiency collector for all except very remote locations. 12 Loadings and particle sizes vary with different drying methods. 13 Boiler blowdown discharge is regulated, generally for temperature and, in some places, for pH limits; check local environmental codes on sanitary discharge. 14 Collection for particulate or sulfur control usually requires a scrubber (dry or wet) and a fabric filter or electrostatic precipitator. 15 Public nuisance from settled wood char indicates collectors are needed. 16 Hot gases and steam usually involved. 17 Steam from hot sand, adhesive clay bond involved. 18 Concentration very heavy at start of cycle. 19 Heaviest load from airless blasting because of high cleaning speed. Abrasive shattering greater with sand than with grit or shot. Amounts removed greater with sand castings, less with forging scale removal, least when welding scale is removed. 20 Operations such as car unloading, conveying, weighing, storing. 21 Special filters are successful. 22 In addition to grain handling, cleaning rolls, sifters, purifiers, conveyors, as well as storing, packaging operations are involved. 23 In addition to grain handling, bins, hammer mills, mixers, feeders, conveyors, bagging operations need control. 24 Primary dry trap and wet scrubbing usual. Electrostatic precipitators are added where maximum cleaning is required. 25 Air pollution control is expensive for open hearth, accelerating the use of substitute melting equipment, such as basic oxygen process and electric-arc furnace.

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2 3 4

Fabric filters have found extensive application for this air pollution control problem. Cupola control varies with plant size, location, melt rate, and air pollution emission regulations. 28 Corrosive gases can be a problem, especially in secondary aluminum. 29 Zinc oxide plume can be troublesome in certain plant locations. 30 Crushing screening, conveying, storing involved. Wet ores often introduce water vapor in exhaust airstream. 31 Dry centrifugal collectors are used as primary collectors, followed by a final cleaner. 32 Collectors usually allow salvage of material and also reduce nuisance from settled dust in plant area. 33 Salvage value of collected material is high. Same equipment used on raw grinding before calcining. 34 Coarse abrasive particles readily removed in primary collector types. 35 Roof discoloration, deposition on autos can occur with cyclones and, less frequently, with dry centrifugal. Heavy-duty air filter sometimes used as final cleaner. 36 Linty particles and sticky buffing compounds can cause trouble in high-efficiency centrifugals and fabric filters. Fire hazard is also often present. 37 Unit collectors extensively used, especially for isolated machine tools. 38 Dust ranges from chips to fine floats, including graphitic carbon. 39 Coolant mist and thermal smoke, often with solid swarf particulate entrained. 40 Submicrometre smoke. Arc welding creates mostly dry metal oxide particulate, sometimes with liquid oil smoke. Resistance welding usually creates only liquid oil smoke, unless done at extremely high currents that vaporize some of the metal being welded. 41 Plasma and laser cutting and welding of clean metals usually creates dry submicrometre smoke, but oily work pieces frequently generate a sticky mix of liquid and solid submicrometre smoke or fume. 42 Materials involved vary widely. Collector selection may depend on salvage value, toxicity, sanitation yardsticks. 43 Controlled temperature and humidity of supply air to coating pans makes recirculation from coating pans desirable. 44 Manufacture of plastic compounds involves operations allied to many in chemical field and vanes with the basic process used. 45 Operations are similar to woodworking, and collector selection involves similar considerations. 46 Submicrometre liquid smoke is frequently emitted when plastic and rubber products are heated. 47 Concentration is heavy during feed operation. Carbon black and other fine additions make collection and dust-free disposal difficult. 48 Often, no collection equipment is used where dispersion from exhaust stack is good and stack location is favorable. 49 Salvage of collected material often dictates type of high-efficiency collector. 50 Fire hazard from some operations must be considered. 51 Granular-bed filters, at times electrostatically augmented, have occasionally been used in this application. 52 Bulky material. Storage for collected material is considerable; bridging from splinters and chips can be a problem. 53 Production sanding produces heavy concentrations of particles too fine to be effectively captured by cyclones or dry centrifugal collectors. 54 Primary collector invariably indicated with concentration and partial size range involved; when used, wet or fabric collectors are used as final collectors.

Table 5 Terminal Settling Velocities of Particles, m/s Particle Density, kg/m3

1

2

5

10

20

50

100

200

500

1000

50 100 200 500 1000 2000 5000 10 000

1.8 E–6 3.7 E–6 7.4 E–6 1.8 E–5 3.7 E–5 7.4 E–5 1.8 E–4 3.7 E–4

7.0 E–6 1.4 E–5 2.8 E–5 7.0 E–5 1.4 E–4 2.8 E–4 7.0 E–4 1.4 E–3

4.0 E–5 8.0 E–5 1.6 E–4 4.0 E–4 8.0 E–4 1.6 E–3 4.0 E–3 8.0 E–3

1.6 E–4 3.2 E–4 6.4 E–4 1.6 E–3 3.2 E–3 6.4 E–3 1.5 E–2 3.1 E–2

7.0 E–4 1.4 E–3 2.8 E–3 7.0 E–3 1.2 E–2 2.5 E–2 6.4 E–2 0.12

4.0 E–3 8.0 E–3 1.6 E–2 3.9 E–2 7.6 E–2 0.14 0.35 0.60

1.6 E–2 3.0 E–2 5.9 E–2 0.14 0.25 0.45 0.97 1.66

0.055 0.11 0.19 0.42 0.70 1.14 2.23 3.52

0.23 0.39 0.65 1.21 1.95 3.10 5.75 8.90

0.51 0.83 1.30 2.50 3.90 6.25 11.0 17.3

Particle or Aggregate Diameter, m

Note: E–6106, etc.

Source: Billings and Wilder (1970).

Single-Stage Designs Figure 6 shows several types of single-stage precipitators. The charging electrodes are located between parallel collecting plates. The gas flows horizontally through the precipitators. High-voltage precipitators collect larger particles better than small particles (they are less efficient at collecting contaminants smaller than 1 to 2 m). Their precipitation efficiency depends in part on the relative electrical resistivity of the pollutant being collected; most are less efficient when collecting either conductive or highly dielectric contaminants.

Single-stage high-voltage precipitators can easily handle heavy loadings of dry dust. Most are configured to operate continuously (using online vibratory or shaker cleaning). They can continuously collect large quantities (hundreds of kilograms per hour) of airborne materials such as foundry shakeout, cement, ceramics, chemical dusts, fly ash, blast furnace dust and fumes, and paper mill recovery boiler emissions. Although they are rarely used to clean exhaust airflows much smaller than 25 m3/s, single-stage precipitators can be constructed to handle airflows as large as 1000 m3/s. Gas velocity through the electrostatic field is ordinarily 0.3 to 2 m/s, with treatment time in

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Industrial Gas Cleaning and Air Pollution Control the range of 2 to 10 s. Only special-purpose single-stage “wet” precipitators are configured for collection of liquid contaminants. Single-stage industrial electrostatic precipitators are distinguished from low-voltage two-stage precipitator designs by several attributes: • Separation between the electrodes is larger to secure acceptable collection and electrode cleaning under the high dust loading and hostile conditions of industrial gas streams. • Construction is heavy-duty for operation to 450°C and +7.5 kPa. • They are generally used for exhaust applications where ozone generation is not of concern. Consequently, they may operate with negative ionization, the polarity that gives the maximum electric field strength between the electrodes. • They are normally custom-engineered and assembled on location for a particular application. Single-stage precipitators can be designed to operate at collection efficiencies above 99.9% for closely specified conditions. Properties of the dust, such as particle size distribution and a deposit’s electrical resistivity, can affect performance significantly, as can variations in gas composition and flow rate.

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Two-Stage Designs Two-stage precipitators are manufactured in two general forms: (1) electronic air cleaners, which are commonly used for light dust loadings in commercial/residential ventilation and air-conditioning service (see Chapter 29); and (2) heavy-duty industrial electrostatic precipitators designed primarily to handle the heavy loadings of submicrometre liquid particulates that are emitted from hot industrial

Fig. 6

Fig. 7

30.9 processes and machining operations. Two-stage industrial precipitators often include sumps and drainage provisions that encourage continuous gravity runoff of collected liquids (Beck 1975; Shabsin 1985). Two-stage precipitators are frequently used in industry to collect submicrometre pollutants that would be difficult (or impossible) to collect in other types of equipment. Two-stage precipitators are frequently sized to handle less than 0.5 m3/s of air and rarely built to handle more than 25 to 35 m3/s of contaminated air in one unit. Gas velocity through the collecting fields usually ranges from 1 to 3.5 m/ s, with treatment times per pass of 0.015 to 0.05 s in the ionizing fields and 0.06 to 0.25 s in the collecting fields. Because gravity drainage of precipitated liquid smoke or fog is a dependable collector-cleaning mechanism, low-voltage two-stage precipitators are most often recommended to collect liquefiable pollutants in which few (or no) solids are entrained. Since the early 1970s, an important application for low-voltage two-stage electrostatic precipitators has been as the gas-cleaning component of condensing precipitator systems. Design of these air pollution control systems (and of similar condensing filtration systems) is based on the following principle: although the hot gases or fumes emitted by many processes are not easily filtered or precipitated (because they are in the vapor phase), the condensation aerosol fogs or smokes that form as those vaporized pollutants cool can be efficiently collected by filtration or precipitation (Figure 7). Many condensation aerosol smokes consist of submicrometre liquid droplets, making them a good match for the collecting capability of

Typical Single-Stage Precipitators

Condensing Precipitator Systems for Control of Hot Organic Smokes

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30.10

2020 ASHRAE Handbook—HVAC Systems and Equipment(SI)

low-voltage two-stage precipitators (Rossnagel 1973; Sauerland 1976; Thiel 1977). Low-voltage precipitators can be very effective at collecting the aerosol particles smaller than 1 m (down to 0.001 to 0.01 m) that are responsible for most plume opacity and for virtually all blue smoke (blue-tinted) emissions. Condensing precipitator systems may therefore be a good choice for eliminating the residual opacity of blue smoke formed by condensation aerosol plumes from hot processes, dryers, ovens, furnaces, or other exhaust air cleaning devices (Beltran 1972; Beltran and Surati 1976; Thiel 1977). When condensation aerosol pollutants are odorous in character, precipitation of the submicrometre droplets of odorant can prevent the long-distance drift of odorous materials, possibly eliminating neighborhood complaints that are associated with submicrometre particulate smokes. Odors that have been successfully controlled by precipitation include asphalt fumes, food frying smokes, meat smokehouse smoke, plasticizer smokes, rubber curing smoke, tar, and textile smokes (Chopyk and Larkin 1982; Thiel 1983).

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2.3

FABRIC FILTERS

Fabric filters are dry dust collectors that use a stationary medium to capture the suspended dust and remove it from the gas stream. This medium, called fabric, can be composed of a wide variety of materials, including natural and synthetic cloth, glass fibers, and even paper. Three types of industrial fabric filters are in common use: pulse jet (and, rarely, reverse air or shaker) cleaned baghouses, pulse jet cleaned cartridge filters, and disposable media filters. Although most commercially available fabric filter systems are currently one of these three types, there are a number of design variations among competing products that can greatly influence the suitability of a particular collector for a specific air-cleaning application. Most baghouses (see Figures 9, 10, 12, and 13) use flexible cloth or other fabric-like filter media in the shape of long cylindrical bags. Bags in pulse jet cleaned systems are rarely more than 160 mm in diameter but can be up to 6 m long. Timed pulses of compressed air flex the bags while blowing collected dust off the media surface. Continuous operation, with no downtime for filter cleaning or dust removal, is common (Figure 12). Cartridge filters, illustrated in Figure 14, use a nearly identical compressed-air media cleaning system. However, their fabric is relatively rigid material, packaged into pleated cylindrical cartridges. Cartridges are self-supported and much easier to handle and replace than are long tubular baghouse bags. Most cartridges are 325 mm or less in diameter and rarely longer than 760 mm. With the pleated media construction, a large filter surface area can be packaged in a relatively small housing to reduce both cost and space required. Most cartridge filter units are not nearly as tall as baghouses having comparable capacity. Pleat depth, spacing and media material construction are critically important variables that determine the suitability and useful lifetime of particular filter cartridges under the conditions of each specific application. Z-flow pack filters are similar to cartridge filters, but with an even greater density of media. The filter media is corrugated, with alternating ends of the corrugated flutes sealed. Dirty air can enter an open-ended flute that is sealed shut at the other end. The air must pass through the media, and exit via an adjacent flute. The dust contaminant is trapped on the media surface. Z-flow packs use cleaning systems nearly identical to those of cartridge systems. Baghouse, cartridge, and Z-flow pack filter systems are practical only when airborne contaminants consist almost exclusively of dry dust. The presence of any entrained liquid in the airstream usually creates a severe maintenance problem because the filter self-cleaning systems (i.e., pulse jet, reverse air, or shaker) become

less effective, so the filters become plugged or “blinded” by collected material and fail after only brief operation. Conservatively selected and carefully applied baghouse, cartridge, and Z-flow filter systems, on the other hand, can easily provide excellent dust collection performance with a filter service life of more than 1 year. They often require very little maintenance, even when handling heavy dust loads in continuous 24 h, 7 day operation. Disposable fabric (or disposable-media) filter collectors are usually simple and economical units that hold enough fabric or similar media to collect modest quantities of almost any particulate pollutant, including liquid, solid, and sticky or waxy materials, regardless of particle size (at least for particles larger than 0.5 to 1 m). When each filter has accumulated as much material as it can practically hold, it is discarded and replaced by a new element. Both envelope bag arrays and pleated rectangular cartridge elements are popular media forms for use in disposable filter collectors. When considering using disposable media filter collectors, serious attention must be given to safe, legal, and ethical disposal of spent and/or contaminated filter elements.

Principle of Operation When contaminated gases pass through a fabric, particles may attach to the fibers and/or other particles and separate from the gas stream. The particles are normally captured near the inlet side of a fabric to form a deposit known as a dust cake. In self-cleaning designs, the dust cake is periodically removed from the fabric to prevent excessive resistance to the gas flow. Finer particles may penetrate more deeply into a fabric and, if not removed during cleaning, may blind it. Surface-loading-type fabrics can help mitigate this effect by not allowing particles to penetrate as deeply into the fabric. Because particles remain on the surface, instead of being embedded into the fabric, the dust cake releases from the fabric more easily during cleaning. Some examples of surface-loading fabrics are polytetrafluoroethylene (PTFE) membrane and nanofiber-coated fabrics. The primary mechanisms for particle collection are direct interception, inertial impaction, electrostatic attraction, and diffusion (Billings and Wilder 1970). Direct interception occurs when the fluid streamline carrying the particle passes within one-half of a particle diameter of a fiber. Regardless of the particle’s size, mass, or inertia, it will be collected if the streamline passes sufficiently close. Inertial impaction occurs when the particle would miss the fiber if it followed the streamline, but its inertia overcomes the resistance offered by the flowing gas, and the particle continues in a direct enough course to strike the fiber. Electrostatic attraction occurs when the particle, the filter, or both possess sufficient electrical charge to cause the particles to precipitate on the fiber. Diffusion makes particles more likely to pass near fibers and thus be collected. Once a particle resides on a fiber, it effectively increases the size of the fiber. Self-cleaning fabric filters have several advantages over other high-efficiency dust collectors such as electrostatic precipitators and wet scrubbers: • They provide high efficiency with lower installed cost. • Particulate matter is collected in the same state in which it was suspended in the gas stream (a significant factor if product recovery is desired). • Process upsets seldom result in the violation of emission standards. The mass rate of particulate matter escaping collection remains low over the life of the filter media and is insensitive to large changes in the mass-loading of dust entering the collector.

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Industrial Gas Cleaning and Air Pollution Control

30.11

Self-cleaning fabric filter dust collectors also have some limitations: • Liquid aerosols, moist and sticky materials, and condensation blind fabrics and reduce or prevent gas flow. These actions can curtail plant operation. • Fabric life may be shortened in the presence of acidic or alkaline components in the gas stream. • Use of fabric filters is generally limited to temperatures below 260°C. • Should a spark or flame accidently reach the collector, fabrics can contribute to the fire/explosion hazard. • When a large volume of gas is to be cleaned, the large number of fabric elements (bags, cartridges, or envelopes) required and the maintenance problem of detecting, locating, and replacing a damaged element should be considered. Monitoring equipment can detect leakage in an individual row of filters or an individual bag.

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Pressure-Volume Relationships The size of a fabric filter is based on empirical data on the amount of fabric media required to clean the desired volumetric flow of gas with an acceptable static pressure drop across the media. The appropriate media and conditions for its use are selected by pilot test or from experience with media in similar full-scale installations. These tests provide a recommended range of approach velocities for a specific application. The approach velocity is stated in practical applications as a gas-to-media ratio or filtration velocity (i.e., the velocity at which gas flows through the filter media). The gas-to-media ratio is the ratio of the volumetric flow of gases through the fabric filter to the area of fabric that participates in the filtration. It is the average approach velocity of the gas to the surface of the filter media. It is difficult to estimate the correlation between pressure drop and gas-to-media ratio for a new installation. However, when flow entering a filter with a dust cake is laminar and uniform, the pressure drop is proportional to the approach velocity: p = KViW = KQWA

Fig. 8 Time Dependence of Pressure Drop Across Fabric Filter cycle. The volumetric flow, particulate concentration, and distribution of particle sizes are assumed to remain constant. In practice, these assumptions are not usually valid. However, the interval between cleaning events is usually long enough that these variations are insignificant for most systems; between cleaning events, the dependence of pressure drop on time is approximately linear. The volumetric flow of gases from a process often varies in response to changes in pressure drop across the fabric filter. The degree to which this variation is significant depends on the operating point for the fan and the process requirements.

Electrostatic Augmentation Electrostatic augmentation involves establishing an electric field between the fabric and another electrode, precharging the dust particles, or both. The effect of electrostatic augmentation is that the interstitial openings in the fabric material function as if they were smaller, and hence smaller particles are retained. Its principal advantage is the more rapid build-up of the dust layer and somewhat higher efficiency for a given pressure drop. Although tested by many, this technique has not been broadly applied.

(6)

where p = pressure drop, Pa K = specific resistance coefficient, pressure drop per unit air velocity and mass of dust per unit area Vi = approach velocity or gas-to-media ratio, m/s W = area density of dust cake, g/m2 Q = volumetric flow of gases, m3/s A = area of cloth that intercepts gases, m2

Equation (6) suggests that increasing the area of the fabric during initial installation has some advantage. A larger fabric area reduces both the gas-to-media ratio and the thickness of the dust cake, resulting in a decreased pressure drop across the collector and reduced cleaning requirements. A lower gas-to-media ratio generally lowers operational cost for the fabric filter system, extends the useful life of the filter elements, and reduces maintenance frequency and expense. In addition, the lower gas-to-media ratio allows for some expansion of the system and, more importantly, additional surge capacity when upset conditions such as unusually high moisture content occur. The specific resistance coefficient K is usually higher for fine dusts. The use of a primary collector to remove the coarse fraction seldom causes a significant change in the pressure drop across the collector. In fact, the coarse dust fraction helps reduce operating pressure because it results in a more porous dust cake, which provides better dust cake release. Figure 8 illustrates the dependence of the pressure drop on time for a single-compartment fabric filter, operating through its cleaning

Fabrics Commercially available fabrics, when applied appropriately, will separate 1 m or larger particles from a gas with an efficiency of 99.9% or better. Particle size is not the major factor influencing efficiency attained from an industrial fabric filter. Most manufacturers of fabric filters will guarantee such efficiencies on applications in which they have prior experience. Lower efficiencies are generally attributed to poor maintenance (torn fabric seams, loose connections, etc.) or the inappropriate selection of lighter/higher-permeability fabrics in an effort to reduce the cost of the collector. Fabric specifications summarize information on such factors as cost, fiber diameter, type of weave, fabric density, tensile strength, dimensional stability, chemical resistance, finish, permeability, and abrasion resistance. Usually, comparisons are difficult, and the supplier must be relied on to select the appropriate material for the service conditions. Table 6 summarizes experience with the exposure of fabrics to industrial atmospheres. Although higher temperatures are acceptable for short periods, reduced fabric life can be expected with continued use above the maximum temperature. The filter is often protected from high temperatures by thermostatically controlled air bleed-in or collector bypass dampers. When the gases are moist or the fabric must collect hygroscopic or sticky materials, synthetic media are recommended. They are also recommended for high-temperature gases. Polypropylene is a frequent selection. One limitation of synthetic media is greater penetration of the media during the cleaning cycle.

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Table 6

Temperature Limits and Characteristics of Fabric Filter Media

Maximum Cost Continuous Relative Operating to Acid Alkali Flex Temp., °C Resistance Resistance Abrasion Cotton Cotton Wool Nylona,b Nomexa,b Acrylic Polypropylene Polyethylene Teflonb Glass fiber Polyestera Cellulose aThese

82 93 93 200 127 82 62 218 260 135 82

Poor Very good Very good Good Poor Fair to good Poor Excellent Excellent Fair Very good Very good Good Fair Good Excellent Excellent Very good Excellent Excellent Very good Excellent Excellent Good Fair to good Fair to good Poor Good Good Very good Poor Good Good

1.00 2.75 2.50 8.00 3.00 1.75 2.00 30.00 5.00 2.50 —

fibers are subject to hydrolysis when they are exposed to hot, wet atmospheres. trademark.

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bDuPont

Woven fabrics are generally porous, and effective filtration depends on prior formation of a dust cake. New cloth collects poorly until particles bridge the openings in the cloth. Once the cake is formed, the initial layers become part of the fabric; they are not destroyed when the bulk of the collected material is dislodged during the cleaning cycle. Cotton and wool fibers in woven media as well as most felted fabrics accumulate an initial dust cake in a few minutes. Synthetic woven fabrics may require a few hours in the same application because of the smoothness of the monofilament threads. Spun threads in the fill direction, when used, reduce the time required to build up the initial dust cake. Felted fabrics contain no straightthrough openings and have a reasonably good efficiency for most particulates, even when clean. After the dust cake builds internally, as well as on the fabric’s surface, shaking does not make it porous. Cartridges manufactured from pleated paper and various synthetic microfibers (usually spun-bonded, not resin-glued, to form a stiff, nonwoven, microporous media) are also fabric filters because they operate in the same general manner as high-efficiency fabrics. As with cloth filters, the efficiency of the filter is increased by the formation of a dust cake on the medium. Media used in cartridge type filters is usually manufactured to have many more pores (through which gas flows while being filtered) than do any of the common baghouse fabrics. Initial filtration efficiency of clean new cartridge filters is usually much greater, particularly for submicrometre-sized dust particles, than that of bare baghouse media (before a significant cake of filtered dust forms) because the pores in cartridge media are much smaller than those in baghouse fabrics. Despite having pores approximately one-tenth the diameter of those in the best baghouse felts, both cellulose (paper) and synthetic (most often polyester) cartridge media have so many pores that their permeability to gas flow is considerably greater than that of commonly used polyester felt baghouse media. Pulse-jet-cleaned cartridge filter dust collectors are usually designed to operate at much lower filtration velocity (typically 0.005 to 0.015 m/s) than pulse jet cleaned tubular media baghouses. Submicrometre dust collection efficiency of cartridge filter media is so high that cartridge-type collectors are often used in applications where cleaned air will be directly recirculated back into the factory to reduce the expense of heating or cooling replacement (makeup) air.

Fig. 9

Bag-Type Shaker Collector

Fig. 10 Envelope-Type Shaker Collector

Types of Self-Cleaning Mechanisms for Fabric Dust Collectors The most common filter cleaning methods are (1) shaking the bags, (2) reversing the flow of gas through the bags, and (3) using an air pulse (pulse jet) to shock the dust cake and break it from the bags. Pulse jet fabric filters are usually cleaned on-line, whereas fabric filters using shakers or reverse air cleaners are usually cleaned off-line. Generally, large installations are compartmented and use off-line cleaning. Other cleaning methods include shakedeflate, a combination of shaker and reverse air cleaning, and acoustically augmented cleaning. Shaker Collectors. When a single compartment is needed that can be cleaned off-line during shift change or breaks, shaker-type fabric filters are usually the least expensive choice. The fabric medium in a shaker-type fabric filter, whether formed into cylindrical tubes or rectangular envelopes, is mechanically agitated to remove the dust cake. Figures 9 and 10 show typical shaker-type fabric filters using bag and envelope media, respectively. When the fabric filter cannot be stopped for cleaning, the collector is divided into a number of independent sections that are sequentially taken off-line for cleaning. Because it is usually difficult to

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Industrial Gas Cleaning and Air Pollution Control maintain a good seal with dampers, relief dampers are often included. The relief dampers introduce a small volume of reverse gas to keep gas flow at the fabric suitable for cake removal. Use of compartments, with their frequent cleaning cycles, does not allow a substantial increase in flow rates over those of a single-compartment unit cleaned periodically. The best situation for fabric reconditioning is when the system is stopped, because even small particles will then fall into the hopper. Figure 11 shows typical pressure diagrams for four- and sixcompartment fabric filters that are continuously cleaned with mechanical shakers. Continuously cleaned units have compartment valves that close for the shaking cycle. This diagram is typical for a multiple-compartment fabric filter where individual compartments are cleaned off-line. The gas-to-media ratio for a shaker-type dust collector is usually in the range of 0.01 to 0.02 m/s; it might be lower where the collector filters particles that are predominantly smaller than 2 m. The abatement of metallurgical fumes is one example where a shaker collector is used to control particles that are less than 1 m in size. The bags are usually 100 to 200 mm in diameter and 3 to 6 m in length, whereas envelope assemblies may be almost any size or shape. For ambient air applications, a woven cotton or polypropylene fabric is usually selected for shaker-type fabric filters. Synthetics are chosen for their resistance to elevated temperature and to chemicals. Reverse-Flow Collectors. Reverse-flow cleaning is generally chosen when the volumetric flow of gases is very large. This method of cleaning inherently requires a compartmented design because the reverse flow needed to collapse the bags entrains dust that must be returned to on-line compartments of the fabric filter. Each compartment is equipped with one main shutoff valve and one reverse gas valve (whether the system is blown through or drawn through). A secondary blower and duct system is required to reverse the gas flow in the compartment to be cleaned. When a compartment is isolated for cleaning, the reverse gas circuit increases the volumetric flow and dust loading through the collector’s active compartments. The fabric medium is reconditioned by reversing the direction of flow through the bags, which partially collapse. The cleaning action is illustrated in Figure 12. After cleaning, reintroduction of gas is delayed to allow dislodged dust to fall into the hopper. Reverse-flow cleaning reduces the number of moving parts in the fabric filter system—a maintenance advantage, especially when large volumetric flows are cleaned. However, the cleaning or reconditioning is less vigorous than other methods, and the residual drag of the reconditioned fabric is higher. Reverse-flow cleaning is particularly suited for fabrics, like glass cloth, that require gentle cleaning. Reverse-flow bags are usually 200 to 300 mm in diameter and 6 to 10 m long and generally operate at flow velocities in the 0.01 to 0.02 m/s range. As a consequence, reverse-air dust collectors tend to be substantially larger than pulse-jet-cleaned designs of similar capacity.

Fig. 11 Pressure Drop Across Shaker Collector Versus Time

30.13 For ambient air applications, a woven cotton or polypropylene fabric is the usual selection for reverse-flow cleaning. For higher temperatures, woven polyester, glass fiber, or trademarked fabrics are often selected. Pulse Jet Collectors. Efforts to decrease fabric filter sizes by increasing flow rates through the fabric have concentrated on implementing frequent or continuous cleaning cycles without taking major portions of the filter surface out of service. In the pulse jet design shown in Figure 13, a compressed air jet operating for a fraction of a second causes a rapid vibration or ripple in the fabric, which dislodges the accumulated dust cake. Simultaneously, outflow of both compressed cleaning air and entrained air from the top clean air plenum helps to sweep pulsed-off dust away from the filter surface. The pulse jet design is predominantly used because (1) it is easier to maintain than the reverse-jet mechanism and

Fig. 12 Draw-Through Reverse-Flow Cleaning of Fabric Filter

Fig. 13 Typical Pulse-Jet Fabric Filter

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2020 ASHRAE Handbook—HVAC Systems and Equipment(SI)

(2) collectors can be smaller and less costly because the greater useful cleaning energy makes operation with higher filtration velocities practical. The tubular bags are supported by wire cages during normal operation. The reverse-flow pulse breaks up the dust layer on the outside of the bag, and dislodged material eventually falls to a hopper. Filtration velocities for reverse-jet or pulse jet designs range from 0.025 to 0.075 m/s for favorable dusts but require greater fabric area for many materials that produce a low-permeability dust cake. Felted fabrics are generally used for these designs because the jet cleaning opens the pores of woven cloth and produces excessive leakage through the filter. Most pulse jet designs require high-pressure, dry, compressed (up to 700 kPa) air, the cost of which should be considered when aircleaning systems are designed. When collecting light or fluffy dust of low bulk density (e.g., arc welding fumes, plasma or laser cutting fumes, finish sanding of wood, fine plastic dusts), serious attention must be given to the direction and velocity at which dust-laden air travels as it moves from the collector inlet to approach the filter media surface. Dustyair velocity is called can velocity and is defined as the actual velocity of airflow approaching the filter surfaces. Can velocity is computed by subtracting the total area occupied by filter elements (bags or cartridges, measured perpendicularly to the direction of gas flow) from the overall cross-sectional area of the collector’s dusty-air housing to compute the actual area through which dusty gas flows. The total gas flow being cleaned is then divided by that flow area to yield can velocity: Q Vc = --------------------A h – NA f

(7)

ter surfaces. The result is that on-line pulse cleaning cannot function, and the collector must be operated in the downtime pulse mode, with filter cleaning done only when there is no airflow through the collector. Can velocity is sometimes overlooked when attempting to increase upflow collector capacity with improved fabrics or cartridges. Regardless of the theoretical gas-to-media ratio at which a filter operates, if released dust cannot fall through the rising airflow into the hopper, the collector will not be able to clean itself. Collector designs in which dusty gas flows downward around the filters are much less susceptible to problems caused by high can velocity because the downward gas flow sweeps pulsed-off dust down toward (and into) the bottom dust discharge hopper, from which it can easily be removed. This chapter cannot adequately cover all the collector design variables and experience-related factors that must be considered when deciding which baghouse or cartridge self-cleaning dust collector design is best suited for each particular application. Engineers making dust collector selections are encouraged to discuss all aspects of each application in detail with all vendors being considered. It is necessary to judge • The relative expertise of each prospective vendor • Which dust collector design is most desirable • How much media surface is needed in each design for each specified gas flow rate • Which filter media is best suited to the particular application • In the case of pulse-jet-cleaned cartridge collectors, what pleat spacing and pleat depth will give optimum or acceptable dust cake removal performance under the particular application conditions

where Vc Q Ah N Af

= = = = =

can velocity, m/s gas flow being cleaned, m3/s cross-sectional area of collector dusty-gas housing, m2 number of filter bags or cartridges in collector cross-sectional area, perpendicular to gas flow, of each filter bag or cartridge, m2

The maximum can velocity in upflow collectors (i.e., those in which dusty gas enters through a plenum or hopper beneath the filter elements) is at the bottom end of the filter elements, where the entire gas flow must pass between and around the filters. Unless the maximum can velocity is low enough that pulsed-off dust can fall through the upwardly flowing gas, dust will simply redeposit on fil-

2.4

Principle of Operation A typical granular-bed filter is shown in Figure 15. Particulateladen gas travels horizontally through the louvers and a granular medium, while the bed material flows downward. The gases typically travel with a superficial velocity near 0.5 to 0.75 m/s. The filter medium moves continuously downward by gravity to prevent a filter cake from forming on the face of the filter and to prevent a high pressure drop. To provide complete cleaning of the louver’s face, the louvers are designed so that some of the medium falls through each louver opening, thus preventing any bridging or build-up of particulate material. Electrostatic augmentation gives the granular-bed filter many of the characteristics of a two-stage electrostatic precipitator. The obvious disadvantage of a granular-bed filter is in removal of the collected dust, which requires liquid backwash or circulation and cleaning of the filter material.

2.5

Fig. 14 Pulse Jet Cartridge Filters (Upflow Design with Vertical Filters)

GRANULAR-BED FILTERS

Usually, granular-bed filters use a fixed bed of granular material that is periodically cleaned off-line. Continuously moving beds have been developed. Most commercial systems incorporate electrostatic augmentation to enhance fine particle control and to achieve good performance with a moving bed. Reentrainment in moving granular-bed filters still significantly influences overall bed efficiency (Wade et al. 1978).

PARTICULATE SCRUBBERS (WET COLLECTORS)

Wet-type dust collectors use liquid (usually, but not necessarily, water) to capture and separate particulate matter (dust, mist, and fumes) from a gas stream. Some scrubbers operate by spraying the scrubbing liquid into the contaminated air. Others bubble air through the scrubbing liquid. In addition, many hybrid designs exist.

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Industrial Gas Cleaning and Air Pollution Control Particle sizes, which can be controlled by a wet scrubber, range from 0.3 to 50 m or larger. Wet collectors can be classified into three categories: (1) low-energy (up to 2 kJ/m3, 0.25 to 1.5 kPa); (2) medium-energy (2 to 6 kJ/m3, 1.5 to 4.5 kPa); and (3) highenergy (> 6 kJ/m3, > 4.5 kPa). Typical wet-scrubber performance is summarized in Table 3. Wet collectors may be used for collection of most particulates from industrial process gas streams where economics allow for collection of the material in a wet state. Advantages of wet collectors include • • • • •

Constant operating pressure No secondary dust sources Small spare parts requirement Ability to collect both gases and particulates Ability to handle high-temperature and high-humidity gas streams, as well as to reduce the possibility of fire or explosion • Reasonably small space requirements for scrubbers • Ability to continuously collect sticky and hygroscopic solids without becoming fouled

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Disadvantages include • High susceptibility to corrosion (corrosion-resistant construction is expensive) • High humidity in discharge gas stream, which may give rise to visible and sometimes objectionable fog plumes, particularly during winter • Large pressure drops and high power requirement for most designs that can efficiently collect fine (particularly submicrometre) particles • Possible difficulty or high cost of disposal of waste water or clarification waste • Rapidly decreasing fractional efficiency for most scrubbers for particles less than 1 m in size • Freeze protection required in many applications in colder environments

30.15 Principle of Operation The more important mechanisms involved in the capture and removal of particulate matter in scrubbers are inertial impaction, Brownian diffusion, and condensation. Inertial impaction occurs when a dust particle and a liquid droplet collide, resulting in the capture of the particle. The resulting liquid/dust particle is relatively large and may be easily removed from the carrier gas stream by gravitation or impingement on separators. Brownian diffusion occurs when the dust particles are extremely small and have motion independent of the carrier gas stream. These small particles collide with one another, making larger particles, or collide with a liquid droplet and are captured. Condensation occurs when the gas or air is cooled below its dew point. When moisture is condensed from the gas stream, fogging occurs, and the dust particles serve as condensation nuclei. The dust particles become larger as a result of the condensed liquid, and the probability of their removal by impaction is increased. Wet collectors perform two individual operations. The first occurs in the contact zone, where the dirty gas comes in contact with the liquid; the second is in the separation zone, where the liquid that has captured the particulate matter is removed from the gas stream. All well-designed wet collectors use one or more of the following principles: • High liquid-to-gas ratio • Intimate contact between the liquid and dust particles, which may be accomplished by forming large numbers of small liquid droplets or by breaking up the gas flow into many small bubbles that are driven through a bath of scrubbing liquid, to increase the chances that contaminants will be wetted and collected • Abrupt transition from dry to wet zones to avoid particle build-up where dry gas enters the collectors For a given type of wet collector, the greater the power applied to the system, the higher the collection efficiency will be (Lapple and Kamack 1955). This is the contacting power theory. Figure 16 compares the fractional efficiencies of several wet collectors, and Figure 17 shows the relationship between the pressure drop across a venturi scrubber and the abatement of particulate matter.

Spray Towers and Impingement Scrubbers Spray towers and impingement scrubbers are available in many different arrangements. The gas stream may be subjected to a single spray or a series of sprays, or the gas may be forced to impinge on a series of irrigated baffles. Except for packed towers, these types of scrubbers are in the low-energy category; thus, they have relatively low particulate removal efficiency. A typical spray tower and an impingement scrubber are illustrated in Figures 18 and 19, respectively. The efficiency of a spray tower can be improved by adding highpressure sprays. A spray tower efficiency of 50 to 75% can be improved to 95 to 99% (for dust particles with size near 2 m) by pressures in the range of 200 to 700 kPa (gage).

Centrifugal-Type Collectors These collectors are characterized by a tangential entry of the gas stream into the collector. They are classed with medium-energy scrubbers. The impingement scrubber shown in Figure 19 is an example of a centrifugal-type wet collector.

Orifice-Type Collectors

Fig. 15 Typical Granular-Bed Filter

Orifice-type collectors are also classified in the medium-energy category. Usually, the gas stream is made to impinge on the surface of the scrubbing liquid and is forced through constrictions where the gas velocity is increased and where the liquid/gaseous/partic-

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ulate interaction occurs. Water usage for orifice collectors is limited to evaporation loss and removal of collected pollutants. A typical orifice-type wet collector is illustrated in Figure 20.

Venturi Scrubber A high-energy venturi scrubber passes the gas through a venturishaped orifice where the gas is accelerated to 60 m/s or more. Depending on the design, the scrubbing liquid is added at, or ahead of, the throat. The rapid acceleration of the gas shears the liquid into

a fine mist, increasing the chance of liquid-particle impaction. Yung has developed a mathematical model for the performance and design of venturi scrubbers (Semrau 1977). Subsequent validation experiments (Rudnick et al. 1986) demonstrated that this model yields a more representative prediction of venturi scrubber performance than other performance models do. In typical applications, the pressure drop for gases across a venturi is higher than for other types of scrubber. Water circulation is also higher; thus, venturi systems use water reclamation systems. One example of a venturi scrubber is illustrated in Figure 21.

Electrostatically Augmented Scrubbers

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Several gas-cleaning devices combine electrical charging of particulate matter with wet scrubbing. Electrostatic augmentation enhances fine particle control by causing an electrical attraction between the particles and the liquid droplets. Compared to venturi

Fig. 18 Typical Spray Tower

Fig. 16

Fractional Efficiency of Several Wet Collectors

Fig. 17 Efficiency of Venturi Scrubber

Fig. 19 Typical Impingement Scrubber

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Fig. 20

30.17

Typical Orifice-Type Wet Collector

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Fig. 22 Typical Electrostatically Augmented Scrubber

3.

Fig. 21 Typical High-Energy Venturi Scrubber scrubbers, electrostatically augmented scrubbers remove particles smaller than 1 m at a much lower pressure drop. There are three generic designs for electrostatic augmentation: 1. Unipolar charged aerosols pass through a contact chamber containing randomly oriented packing elements of dielectric material. A typical electrostatically augmented scrubber of this design is shown in Figure 22. 2. Unipolar charged aerosols pass through a low-energy venturi scrubber. 3. Unipolar charged aerosols pass into a spray chamber where they are attracted to oppositely charged liquid droplets. Collection efficiencies of 50 to 90% can be achieved in a single particle charging and collection stage, depending on the mass loading of fine particles and the superficial velocity of gases in the collector. Higher collection efficiencies can be obtained by using two or more stages. Removal efficiency of gaseous pollutants depends on the mass transfer and absorption design of the scrubber section. In most applications of electrostatically augmented scrubbers, the dirty gas stream is quenched by adiabatic cooling with liquid sprays; thus, it contains a large amount of water vapor that wets the particulate contaminants. This moisture provides a dominant influence on particle adhesion and the electrical resistivity of deposits in the collector. Electrical equipment for particle charging is similar to that for electrostatic precipitators. The scrubber section is usually equipped with a liquid recycle pump, recycle piping, and a liquid distribution system.

GASEOUS CONTAMINANT CONTROL

Many industrial processes produce large quantities of gaseous or vaporized contaminants that must be separated from gas streams. These contaminants are usually removed through absorption into a liquid or adsorption onto a solid medium. Incineration of exhaust gas (see the section on Incineration of Gases and Vapors) has also been successfully used to remove organic gases and vapors. Low-vaporpressure odorous materials that condense to form submicrometre condensation aerosols after being emitted from hot industrial processes can sometimes be successfully controlled by well-designed condensing filter or condensing precipitator submicrometre particulate collection systems (see the section on Two-Stage Designs under Electrostatic Precipitators).

3.1

SPRAY DRY SCRUBBING

Spray dry scrubbing is used to absorb and neutralize acidic gaseous contaminants in hot industrial gas streams. The system uses an alkali spray to react with the acid gases to form a salt. The process heat evaporates the liquid, resulting in a dry particulate that is removed from the gas stream. Typical industrial applications of spray dry scrubbing are • Control of hydrochloric acid (HCl) emissions from biological hazardous-waste incinerators • Control of sulfuric acid and sulfur trioxide emissions from burning high-sulfur coal • Control of sulfur oxides (SOx), boric acid, and hydrogen fluoride (HF) gases from glass-melting furnaces.

Principle of Operation Spray drying involves four operations: (1) atomization, (2) gas droplet mixing, (3) drying of liquid droplets, and (4) removal and collection of a dry product. These operations are carried out in a tower or a specially designed vessel. In any spray dryer design, good mixing and efficient gas droplet contact are desirable. Dryer height is largely determined by the time required to dry the largest droplets produced by the atomizer. Towers used for acid gas control typically have gas residence times of about 10 s, compared to about 3 s for towers designed for evaporative cooling. The longer residence time is needed because drying by

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itself is not the primary goal for the equipment. Many of the acid/ alkali reactions are accelerated in the liquid state. It is, therefore, desirable to cool the gases to as close as possible to the adiabatic saturation temperature (dew point) without risking condensation in downstream particulate collectors. At these low temperatures, droplets survive against evaporation much longer, thus obtaining a better chance of contacting all acidic contaminants while the alkali compounds are in their most reactive state.

needed for acid gas control. It is not uncommon that pumps and nozzles must be chosen to handle and meter slurries. Spray dryer systems include metering valves, pumps and compressors, and controls to ensure optimal chemical feed and temperature in the gas-cleaning system.

Equipment

Packed scrubbers are used to remove gaseous and particulate contaminants from gas streams. Scrubbing is accomplished by impingement of particulate matter and/or by absorption of soluble gas or vapor molecules on the liquid-wetted surface of the packing. There is no limit to the amount of particulate capture, as long as the properties of the liquid film are unchanged. Gas or vapor removal is more complex than particulate capture. The contaminant becomes a solute and has a vapor pressure above that of the scrubbing liquid. This vapor pressure typically increases with increasing concentration of the solute in the liquid and/or with increasing liquid temperature. Contaminant scrubbing continues as long as the partial pressure of contaminant in the gas is above its vapor pressure with respect to the liquid. The rate of contaminant removal is a function of the difference between the partial pressure and vapor pressure, as well as the rate of diffusion of the contaminant. In most common gas-scrubbing operations, small increases in the superficial velocity of gases through the collector decrease the removal efficiency; therefore, these devices are operated at the highest possible superficial velocities consistent with acceptable contaminant control. Usually, increasing the liquid rate has little effect on efficiency; therefore, liquid flow is kept near the minimum required for satisfactory operation and removal of particulate matter. As the superficial velocity of gases in the collector increases above a certain value, there is a tendency to strip liquid from the surface of the packing and entrain the liquid from the scrubber. If pressure drop is not the limiting factor for operation of the equipment, maximum scrubbing capacity occurs at a gas rate just below the rate that causes excessive liquid reentrainment.

The atomizer must disperse a liquid containing an alkali compound that will react with acidic components of the gas stream. The liquid must be distributed uniformly in the dryer and mixed thoroughly with the hot gases in droplets of a size that will evaporate before striking a dryer surface. In typical spray dryers used for acid gas control, the droplets have diameters ranging from 50 to 200 m. The larger droplets are of most concern because these might survive long enough to impinge on equipment surfaces. In general, a tradeoff must be made between the largest amount of liquid that can be sprayed and the largest droplets that can be tolerated by the equipment. The angular distribution or fan-out of the spray is also important. In spray drying, the angle is often 60 to 80°, although both lower and higher angles are sometimes required. The fan-out may change with distance from the nozzle, especially at high pressures. An important aspect of spray dryer design and operation is the production and control of the gas flow patterns in the drying chamber. Because of the importance of the flow patterns, spray dryers are usually classified on the basis of gas flow direction in the chamber relative to the spray. There are three basic designs: (1) cocurrent, in which the liquid feed is sprayed with the flow of the hot gas; (2) countercurrent, in which the feed is sprayed against the flow of the gas; and (3) mixed flow, in which there is a combined cocurrent and countercurrent flow. There are several types of atomizers. High-speed rotating disks achieve atomization through centrifugal motion. Although disks are bulky and relatively expensive, they are also more flexible than nozzles in compensating for changes in particle size caused by variations in feed characteristics. Disks are also used when high-pressure feed systems are not available. They are frequently used when high volumes of liquid must be spray dried. Disks are not well suited to counterflow or horizontal flow dryers. Nozzles are also commonly used. These may be subdivided into two distinct types: centrifugal pressure nozzles and two-fluid (or pneumatic) nozzles. In the centrifugal pressure nozzle, energy for atomization is supplied solely by the pressure of the feed liquid. Most pressure nozzles are of the swirl type, in which tangential inlets or slots spin the liquid in the nozzle. The pressure nozzle satisfactorily atomizes liquids with viscosities of 300 mPa·s or higher. It is well suited to counterflow spray dryers and to installations requiring multiple atomizers. Capacities up to 1.3 kg/s through a single nozzle are possible. Pressure nozzles have some disadvantages. For example, pressure, capacity, and orifice size are independent, resulting in a certain degree of inflexibility. Moreover, pressure nozzles (particularly those with small passages) are susceptible to erosion in applications involving abrasive materials. In such instances, tungsten carbide or a similarly tough material is mandatory. In two-fluid nozzles, air (or steam) supplies most of the energy required to atomize the liquid. Liquid, admitted under low pressure, may be mixed with the air either internally or externally. Although energy requirements for this atomizer are generally greater than for spinning disks or pressure nozzles, the two-fluid nozzle can produce very fine atomization, particularly with viscous materials. The density and viscosity of the feed materials and how these might change at elevated temperature should be considered. Some alkali compounds do not form a solution at the concentrations

3.2

WET-PACKED SCRUBBERS

Scrubber Packings Packings are designed to present a large surface area that will wet evenly with liquid. They should also have high void ratio so that pressure drop will be low. High-efficiency packings promote turbulent mixing of the gas and liquid. Figure 23 illustrates six types of packings that are randomly dumped into scrubbers. Packings are available in ceramic, metal, and thermoplastic materials. Plastic packings are extensively used in scrubbers because of their low mass and resistance to mechanical damage. They offer a wide range of chemical resistance to acids, alkalies, and many organic compounds; however, plastic packing can be deformed by excessive temperatures or by solvent attack. The relative capacity of tower packings at constant pressure drop can be obtained by calculation from the packing factor F. The gashandling capacity G of a packing is inversely proportional to the square root of F: G = K F

(8)

where G = mass flow rate of gases through scrubber F = packing factor (surface area of packing per unit volume of gently poured material)

The smaller the packing factor of a given packing, the greater will be its gas-handling capacity. Typical packing factors for scrubber packings are summarized in Table 7.

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30.19

Fig. 23 Typical Packings for Scrubbers

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Table 7

Packing Factor F for Various Scrubber Packing Materials Nominal Size, mm

Type of Packing

Material

Saddles, scalloped-edge Saddles, scalloped-edge Saddles Saddles, perforated extended Berl saddles Raschig rings Perforated rings Perforated rings Large perforated rings Toroidal helix

Plastic Ceramic Ceramic Plastic Ceramic Ceramic Metal Plastic Metal Plastic

20

25 40 60 92

145

170 255

110 155 48 52 43 36

40

52

65 95 33 40 26

50 80 or 90 21 30 40 32

16 22 21

45 65 20 24 18 18

37 16 16 15 16

Arrangements of Packed Scrubbers The four generic arrangements for wet-packed scrubbers are illustrated in Figure 24: • • • •

Horizontal cocurrent scrubber Vertical cocurrent scrubber Cross-flow scrubber Countercurrent scrubber

Cocurrent flow scrubbers can be operated with either horizontal or vertical gas and liquid flows. A horizontal cocurrent scrubber depends on the gas velocity to carry the liquid into the packed bed. It operates as a wetted entrainment separator with limited gas and liquid contact time. The superficial velocity of gases in the collector is limited by liquid reentrainment to about 3.3 m/s. A vertical cocurrent scrubber can be operated at very high pressure drop (800 to 2450 Pa per metre of packing depth) because there is no flooding limit for the superficial velocity. The contact time in a cocurrent scrubber is a function of bed depth. The effectiveness of absorption processes is lower in cocurrent scrubbers than in the other arrangements because the liquid containing contaminant is in contact with the exit gas stream. Cross-flow scrubbers use downward-flowing liquid and a horizontally moving gas stream. The effectiveness of absorption processes in cross-flow scrubbers lies between those for cocurrent and countercurrent flow scrubbers. Countercurrent scrubbers use a downward-flowing liquid and an upward-flowing gas. The gas-handling capacity of countercurrent scrubbers is limited by pressure drop or by liquid entrainment.

Fig. 24

Flow Arrangements Through Packed Beds

Contact time can be controlled by the depth of packing used. The effectiveness of absorption processes is maximized because the exiting gas is in contact with fresh scrubbing liquid. The most broadly used arrangement is the countercurrent packed scrubber. This type of scrubber, illustrated in Figure 25, gives the best removal of gaseous contaminants while keeping liquid consumption to a minimum. The effluent liquid has the highest contaminant concentration. Extended-surface packings have been used successfully for the absorption of highly soluble gases such as HCl because the required contact time is minimal. This type of packing consists of a woven mat of fine fibers of a plastic material that is not affected by chemical exposure. Figure 26 shows an example of a scrubber consisting of three wetted stages of extended surface packing in series with the gas flow. A final dry mat is used as an entrainment eliminator. If solids are present in the inlet gas stream, a wetted impingement stage precedes the wetted mats to prevent plugging of the woven mats. Figure 27 shows a scrubber with a vertical arrangement of extended surface packing. This design uses three complete stages in series with the gas flow. The horizontal mat at the bottom of each stage operates as a flooded bed scrubber. The flooded bed is used to minimize water consumption. The two inclined upper mats operate as entrainment eliminators.

Pressure Drop The pressure drop through a particular packing in countercurrent scrubbers can be calculated from the airflow and water flow per unit area. Charts, such as the one shown in Figure 28, are available from manufacturers of each type and size of packing. The pressure drop for any packing can also be estimated by using the data on packing factors in Table 7 and the modified

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2020 ASHRAE Handbook—HVAC Systems and Equipment(SI)

Fig. 27

Fig. 25

Vertical Flow Scrubber with Extended Surface

Typical Countercurrent Packed Scrubber

Fig. 26 Horizontal Flow Scrubber with Extended Surface

Fig. 28 Pressure Drop Versus Gas Rate for Typical Packing

generalized pressure drop correlation shown in Figure 29. This correlation was developed for a gas stream substantially of air, with water as the scrubbing liquid. It should not be used if the properties of the gas or liquid vary significantly from air or water, respectively. Countercurrent scrubbers are generally designed to operate at pressure drops between 200 and 530 Pa per metre of packing depth. Liquid irrigation rates typically vary between 3.4 and 13.6 L/ s per square metre of bed area.

amount of absorbent available for reaction, temperature, and reaction rate for absorption. Practically all commercial packings have been tested for absorption rate (mass transfer coefficient) using standard absorber conditions: carbon dioxide (CO2) in air and a solution of caustic soda (NaOH) in water. This system was selected because the interaction of the variables is well understood. Further, the mass transfer coefficients for this system are low; thus, they can be determined accurately by experiment. The values of mass transfer coefficients (KG a) for various packings under these standard test conditions are given in Table 8. The vast majority of wet absorbers are used to control low concentrations (less than 0.005 mole fraction) of contaminants in air. Dilute aqueous solutions of NaOH are usually chosen as the scrubbing fluid. These conditions simplify the design of scrubbers somewhat. Mass transfer from the gas to the liquid is then explained by

Absorption Efficiency The prediction of the absorption efficiency of a packed bed scrubber is much more complex than estimating its capacity because performance estimates involve the mechanics of absorption. Some of the factors affecting efficiency are superficial velocity of gases in the scrubber, liquid injection rate, packing size, type of packing, amount of contaminant to be removed, distribution and

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30.21 Yi – Yo pln = p ------------------------ln  Y i  Y o 

(10)

where pln = driving or diffusion pressure acting to absorb contaminants on packing p = inlet pressure

The rate of absorption of contaminant (mass transfer coefficient) is related to the depth of packing as follows: KG a = N/HApln

(11)

where N = solute absorbed, mol/s H = depth of packing, m A = cross-sectional area of scrubber, m2

The value of N can be determined from N = G(1 – Yi – Yo)

(12)

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where G = mass flow rate of gases through scrubber, mol/s. The superficial velocity of gases is a function of the unit gas flow rate and the gas density: Fig. 29 Generalized Pressure Drop Curves for Packed Beds Table 8

Nominal Size, mm Material

25

40

50

80 or 90

KG a, mol/(s·m3 ·kPa) Scalloped saddles Saddles Perforated extended saddles Raschig rings Perforated rings Perforated rings Large perforated rings Toroidal helix

Plastic Ceramic Plastic Ceramic Metal Plastic Metal Plastic

(13)

where

Mass Transfer Coefficients (KG a) for Scrubber Packing Materials

Type of Packing

V = TGMvC1A

0.096 0.063 0.086 0.075 0.063 0.063 0.076 0.066 0.053 0.102 0.082 0.071 0.087 0.076 0.064 0.097 0.082 0.074 0.096 0.087

0.039 0.036 0.039 0.040 0.039 0.048

System: CO2 and NaOH; gas rate: 0.67 kg/(m2 ·s); liquid rate: 2.7 kg/(m2 ·s).

the two-film theory: the gaseous contaminant travels by diffusion from the main gas stream through the gas film, then through the liquid film, and finally into the main liquid stream. The relative influences of the gas and liquid films on the absorption rate depend on the contaminant’s solubility in the liquid. Sparingly soluble gases like hydrogen sulfide (H2S) and CO2 are said to be liquidfilm-controlled; highly soluble gases such as HCl and ammonia (NH3) are said to be gas-film-controlled. In liquid-film-controlled systems, the mass transfer coefficient varies with the liquid injection rate but is only slightly affected by the superficial velocity of the gases. In gas-film-controlled systems, the mass transfer coefficient is a function of both the superficial velocity of the gases and the liquid injection rate. In the absence of leakage, the percentage by volume of the contaminant removed from the air can be found from the inlet and outlet concentrations of contaminant in the airstream: % Removed = 100(1 – Yo Yi) (9) where Yi = mole fraction of contaminants entering scrubber (dry gas basis) Yo = mole fraction of contaminants exiting scrubber (dry gas basis)

The driving pressure for absorption (assuming negligible vapor pressure above the liquid) is controlled by the logarithmic mean of inlet and outlet concentrations of the contaminant:

V Mv T C1

= = = =

superficial gas velocity, m/s molar volume, m3/mol exit gas temperature, K = °C + 273 273 K

By combining these equations and assuming ambient pressure, a graphical solution can be derived for both liquid-film- and gas-filmcontrolled systems. Figures 30, 31, and 32 show the height of packing required versus percent removal for various mass transfer coefficients at superficial velocities of 0.6, 1.2, and 1.8 m/s, respectively, with liquid-film-controlled systems. Figures 33, 34, and 35 show the height of packing versus percent removal for various mass transfer coefficients at the same three superficial velocities with gas film-controlled systems. These graphs can be used to determine the height of 50 mm plastic saddles (see Figure 23) required to give the desired percentage of contaminant removal. The height for any other type or size of packing is inversely proportional to the ratio of standard KG a taken from Table 8. Thus, if 4.0 m packing depth were required for 95% removal of contaminants, the same efficiency could be obtained with a 2.9 m depth of 25 mm plastic perforated rings (Figure 23), at the same superficial velocity and liquid injection rate. However, the pressure drop would be higher for the smaller-diameter packing. Figures 30 to 35 are useful when the value of the mass transfer coefficient for the particular contaminant to be removed is known. Table 9 contains mass transfer coefficients for 50 mm plastic scalloped saddles in typical liquid-film-controlled scrubbers. These values can be compared with the mass transfer coefficients in Table 10 for the same packing used in gas-film-controlled scrubbers. When the scrubbing liquid is not water, the mass transfer coefficients in these tables can only be used if the amount of reagent in the solution exceeds by at least 33% the amount needed to completely absorb the gaseous contaminant. When HCl is dissolved in water, there is little vapor pressure of HCl above solutions of less than 8% (by mass) concentration. On the other hand, when NH3 is dissolved in water, there is an appreciable vapor pressure of NH3 above solutions, even at low concentrations. The height of packing needed for NH3 removal, obtained from Figures 33 to 35, is based on the use of dilute acid to maintain the pH of the solution below 7.

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2020 ASHRAE Handbook—HVAC Systems and Equipment(SI)

The following is an example of a typical scrubbing problem. Example 1. Remove 95% of the HF from air at 30°C. The concentration of HF is 0.0006 mole fraction on a dry gas basis. The concentration of HF in the exhaust gas should not exceed 0.000 030 mole fraction.

Total volumetric flow of gas G = 2 m3/s Liquid injection rate = 0.0025 m/s Liquid temperature = 20°C Packed tower diameter = 1.5 m Packing material is 50 mm polypropylene scalloped saddles

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The following design conditions apply:

Fig. 32 Contaminant Control at Superficial Velocity = 1.8 m/s (Liquid-Film-Controlled) Fig. 30 Contaminant Control at Superficial Velocity = 0.6 m/s (Liquid-Film-Controlled)

Fig. 31 Contaminant Control at Superficial Velocity = 1.2 m/s (Liquid-Film-Controlled)

Fig. 33 Contaminant Control at Superficial Velocity = 0.6 m/s (Gas-Film-Controlled)

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30.23 Table 9 Relative KG a for Various Contaminants in Liquid-Film-Controlled Scrubbers Gas Contaminant

Scrubbing Liquid

KG a, mol/(s·m3 ·kPa)

CO2 H2S SO2 HCN HCHO Cl2

4% (by mass) NaOH 4% (by mass) NaOH Water Water Water Water

0.088 0.26 0.13 0.26 0.26 0.2

Note: Data for50 mm plastic scalloped saddles. Temperatures: from 16 to 24°C; liquid rate: 6.8 L/(s·m2); gas rate: 1.1 m/s.

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Table 10 Relative KG a for Various Contaminants in Gas-Film-Controlled Scrubbers Gas Contaminant

Scrubbing Liquid

KG a, mol/(s·m3 ·kPa)

HCl HBr HF NH3 Cl2 SO2 Br2

Water Water Water Water 8% (by mass) NaOH 11% (by mass) Na2CO3 5% (by mass) NaOH

0.82 0.26 0.35 0.76 0.63 0.52 0.22

Note: Data for 250 mm plastic scalloped saddles. Temperatures 16 to 24°C; liquid rate: 0.0068 m/s; gas rate: 1.1 m/s.

Fig. 34 Contaminant Control at Superficial Velocity = 1.2 m/s (Gas-Film-Controlled)

Total liquid flow rate L = 0.0025  1.77 = 0.00442 m3/s Packing factor (from Table 7) F = 21 Figure 29 may be used to find the pressure drop through the packed tower: x-axis = 29.4(L  G) = 0.065 y-axis = (F80)(GA)2 = 0.335 From Figure 29, the pressure drop is about 65 Pa per metre of packing depth. From Table 10, KG a = 0.35 for HF. From Figure 33, the depth of packing required for 95% removal is 4 m. Thus, the total pressure drop is 4  65 = 260 Pa.

General Efficiency Comparisons

Fig. 35 Contaminant Control at Superficial Velocity = 1.8 m/s (Gas-Film-Controlled) Solution: Cross-sectional area of absorber A = (1.5/2)2 = 1.77 m2

Figure 35 indicates that, with KG a = 0.35, 90% removal of HF could be achieved with 3.1 m of packing; this is 23% less packing than needed for 96% removal. Furthermore, with the same superficial velocity, both liquid-film- and gas-film-controlled systems require a 43% increase in absorbent depth to raise the removal efficiency from 80 to 90%. A comparison of Figures 34 and 35 shows that increasing the superficial velocity by 50% in a gas-film-controlled scrubber requires only a 12% increase in bed depth to maintain equal removal efficiencies. In the liquid-film-controlled system (Figures 31 and 32), increasing the superficial velocity by 50% requires an approximately 50% increase in bed depth to maintain equal removal efficiencies. Thus, in a gas-film-controlled system, the superficial velocity can be increased significantly with only a small increase in bed depth required to maintain the efficiency. In practical terms, gas-filmcontrolled scrubbers of fixed depth can handle an overload condition with only a minor loss of removal efficiency. Performance of liquid-film-controlled scrubbers degrades significantly under simi-

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lar overload conditions. This occurs because the mass transfer coefficient is independent of superficial velocity.

Liquid Effects Some liquids tend to foam when they are contaminated with particulates or soluble salts. In these cases, the pressure drop should be kept in the lower half of the normal range: 200 to 330 Pa per metre of packing depth. In the control of gaseous pollution, most systems do not destroy the pollutant but merely remove it from the air. When water is used as the scrubbing liquid, effluent from the scrubber will contain suspended particulate or dissolved solute. Water treatment is often required to alter the pH and/or remove toxic substances before the solutions can be discharged.

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3.3

ADSORPTION OF GASEOUS CONTAMINANTS

The surface of freshly broken or heated solids often contains van der Waals (London dispersion) forces that can physically or chemically adsorb nearby molecules in a gas or liquid. The captured molecules form a thin layer on the surface of the solid that is typically one to three molecules thick. Commercial adsorbents are solids with an enormous internal surface area. This large surface area enables them to capture and hold large numbers of molecules. For example, each gram of a typical activated carbon adsorbent contains over 1000 m2 of internal surface area. Adsorbents are used for removing organic vapors, water vapor, odors, and hazardous pollutants from gas streams. The most common adsorbents used in industrial processes include activated carbons, activated alumina, silica gel, and molecular sieves. Activated carbons are derived from coal, wood, or coconut shells. They are primarily selected to remove organic compounds in preference to water. The other three common gas-phase adsorbents have a great affinity for water and adsorb it to the exclusion of any organic molecules also present in a gas stream. They are used primarily as gas-drying agents. Molecular sieves also find use in several specialized pollution control applications, including removal of mercury vapor, sulfur dioxide (SO2), or nitrogen oxides (NOx) from gas streams. The capacity of a particular activated carbon to adsorb

any organic vapor from mass of the organic compound and the temperature of the gas stream. Compounds with a higherrelative molecular mass are usually more strongly adsorbed than those with lower massThe capacity of activated carbon to adsorb any given organic compound increases with the concentration of that compound. Reducing the temperature also favors adsorption. Typical adsorption capacities of an activated carbon for toluene (relative molecular mass 92) and acetone (relative molecular mass 58) are illustrated in Figure 36 for various temperatures and concentrations. Regeneration. Adsorption is reversible. An increase in temperature causes some or all of an adsorbed vapor to desorb. The temperature of low-pressure steam is sufficient to drive off most of a low-boiling-point organic compound previously adsorbed at ambient temperature. Higher-boiling-point organic compounds may require high-pressure steam or hot inert gas to secure good desorption. Compounds with a very highrelative molecular mass can require reactivation of the carbon adsorber in a furnace at 750°C to drive off all the adsorbed material. Regeneration of the carbon adsorbers can also be accomplished, in some instances, by washing with an aqueous solution of a chemical that will react with the adsorbed organic material, making it water soluble. An example is washing carbon containing adsorbed sulfur compounds with NaOH. The difference between an adsorbent’s capacity under adsorbing and desorbing conditions in any application is its working capacity. Activated carbon for air pollution control is found in canisters under the hoods of most automobiles. The adsorber in these canisters captures gasoline vapors escaping from fuel aspiration devices (when the engine is stopped) and from the fuel tank’s breather vent. Gasoline vapors are desorbed by pulling fresh air through the carbon canister and into the carburetor when the engine is running. Although there is no temperature difference between adsorbing and desorbing conditions in this case, the outdoor airflow desorbs enough gasoline vapors to give the carbon a substantial working capacity. For applications where only traces of a pollutant must be removed from exhaust air, the life of a carbon bed is very long. In these cases, it is often more economical to replace the carbon than to invest in regeneration equipment. Larger quantities can be returned to the carbon manufacturer for high-temperature thermal reactivation. Regeneration in place by steam, hot inert gas, or washing with a solution of alkali is sometimes practiced. Impregnated (chemically reactive) adsorbents are used when physical adsorption alone is too weak to remove a particular gaseous contaminant from an industrial gas stream. Through impregnation, the reactive chemical is spread over the immense internal surface area of an adsorbent. Typical applications of impregnated adsorbent include the following: • Sulfur- or iodine-impregnated carbon removes mercury vapor from air, hydrogen, or other gases by forming mercuric sulfide or iodide. • Metal oxide-impregnated carbons remove hydrogen sulfide. • Amine- or iodine-impregnated carbons and silver exchanged zeolites remove radioactive methyl iodide from nuclear power plant work areas and exhaust gases. • Alkali-impregnated carbons remove acid gases. • Activated alumina impregnated with potassium permanganate removes acrolein and formaldehyde.

Equipment for Adsorption

Fig. 36 Adsorption Isotherms on Activated Carbon

Three types of adsorbers are usually found in industrial applications: (1) fixed beds, (2) moving beds, and (3) fluidized beds (Figure 37). Fixed beds of regenerable or disposable media are most common. Carbon filter elements are a typical example.

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Fig. 37 Fluidized-Bed Adsorption Equipment Moving beds use granular adsorbers placed on inclined trays or on vertical frames similar to those used in granular bed particulate collectors. Moving beds offer continuous contaminant control and regeneration. Often, moving bed adsorbers and regeneration equipment are integrated components in a process. Fluidized beds contain a fine granular adsorber, which is continuously mixed with the contaminated gas by suspension in the process gas stream. The bed may be either “fixed” at lower superficial velocities or highly turbulent and conveying (circulating). Illustrations of these types of fluidized beds are shown in Figure 37.

Solvent Recovery The most common use of adsorption in stationary sources is in recovering solvent vapors from manufacturing and cleaning processes. Typical applications include solvent degreasing, rotogravure printing, dry cleaning, and the manufacture of such products as synthetic fibers, adhesive labels, tapes, coated copying paper, rubber goods, and coated fabrics. Figure 38 illustrates the components of a typical solvent recovery system using two carbon beds. One bed is used as an adsorber while the other is regenerated with low-pressure steam. Desorbed solvent vapor and steam are recovered in a water-cooled condenser. If the solvent is immiscible with water, an automatic decanter separates the solvent for reuse. A distillation column is used for water-miscible solvents. Adsorption time per cycle typically runs from 30 min to several hours. The adsorbing carbon bed is switched to regeneration by (1) an automatic timer shortly before the solvent vapor breaks through from the bed or (2) an organic vapor-sensing control device in the exhaust gas stream immediately after the solvent breaks through from the bed. Low-pressure steam consumption for regeneration is generally about 2.2 kW per kilogram of solvent recovered (Boll 1976), but it can range from 1.3 to over 3 kW per kilogram of solvent recovered, depending on the specific solvent and its concentration in the exhaust gas stream being stripped. Steam with only a slight superheat is normally used, so that it condenses quickly and gives rapid heat transfer. After steaming, the hot, moist carbon bed is usually cooled and partially dried before being placed back on stream. Heat for drying

Fig. 38

Schematic of Two-Unit Fixed Bed Adsorber

is supplied by the cooling of the carbon and adsorber, and sometimes by an external air heater. In most cases, it is desirable to leave some moisture in the bed. When solvent vapors are adsorbed, heat is generated. For most common solvents, the heat of adsorption is 90 to 140 J/mol. When high-concentration vapors are adsorbed in a dry carbon bed, this heat can cause a substantial temperature rise and can even ignite the bed, unless it is controlled. If the bed contains moisture, the water absorbs energy and helps to prevent an undue rise in bed temperature. Certain applications may require heat sensors and automatic sprinklers. Because the adsorptive capacity of activated carbon depends on temperature, it is important that solvent-laden air going to a recovery unit be as cool as is practicable. The exhaust gases from many solvent-emitting processes (such as drying ovens) are at elevated temperatures. Water- or air-cooled heat exchangers must be installed to reduce the temperature of the gas that enters the adsorber. Solvent at very low vapor concentrations can be recovered in an activated carbon system. The size and cost of the recovery unit, however, depend on the volume of air to be handled; it is thus advantageous to minimize the volume of an exhaust stream and keep the solvent vapor concentration as high as possible, consis-

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2020 ASHRAE Handbook—HVAC Systems and Equipment(SI)

tent with safety requirements. Insurance carriers specify that solvent vapor concentrations must not exceed 25% of the lower explosive limit (LEL) when intermittent monitoring is used. With continuous monitoring, concentrations as high as 50% of the LEL are permissible. Solvent recovery systems, with gas-handling capacities up to about 5 m3/s, are available as skid-mounted packages. Several of these packaged units can be used for larger gas flows. Custom systems can be built to handle 100 m3/s or more of gas. Materials of construction may be painted carbon steel, stainless steel, Monel, or even titanium, depending on the nature of the gas mixture. The activated carbon is usually placed in horizontal flat beds or vertical cylindrical beds. The latter design minimizes ground space required for the system. Other alternatives are possible; one manufacturer uses a segmented horizontal rotating cylinder of carbon in which one segment is adsorbing while others are being steamed and cooled. Commercial-scale solvent recovery systems typically recover over 99% of the solvent contained in a gas stream. The efficiency of the collecting hoods at the source of the solvent emission is usually the determining factor in solvent recovery. Dust filters are generally placed ahead of carbon beds to prevent blinding of the adsorber by dust. Occasionally, the carbon is removed for screening to eliminate accumulated dust and fine particles of carbon. If the solvent mixture contains high-boiling-point components, the working capacity of activated carbon can decrease with time. This occurs when high-boiling-point organic compounds are only partially removed by low-pressure steam. In this situation, two alternatives should be considered:

Fig. 39

Moving-Bed Adsorber

1. Periodic removal of the carbon and return to its manufacturer for high-temperature furnace reactivation to virgin carbon activity. 2. Use of more rigorous solvent desorbing conditions in the solvent recovery system. High-temperature steam, hot inert gas, or a combination of electrical heating and application of a vacuum may be used. The last method is selected, for example, to recover lithography ink solvents with high boiling points. Note that this method may not remove all of the high-boiling-point compounds.

Odor Control

Fig. 40 Typical Odor Adsorber

Incineration and scrubbing are usually the most economical methods of controlling high concentrations of odorous compounds from equipment such as cookers in rendering plants. However, many odors that arise from harmlessly low concentrations of vapors are still offensive. The odor threshold (for 100% response) of acrolein in air, for example, is only 0.21 ppm, whereas that for ethyl mercaptan is 0.001 ppm and that for hydrogen sulfide is 0.0005 ppm (AIHA 1989; MCA 1968). Activated carbon beds effectively overcome many odor emission problems. Activated carbon is used to control odors from chemical and pharmaceutical manufacturing operations, foundries, sewage treating plants, oil and chemical storage tanks, lacquer drying ovens, food processing plants, and rendering plants. In some of these applications, activated carbon is the sole odor control method; in others, the carbon adsorber is applied to the exhaust from a scrubber. Odor control systems using activated carbon can be as simple as a steel drum fitted with appropriate gas inlet and outlet ducts, or as complex as a large, vertically moving bed, in which carbon is contained between louvered side panels. A typical moving-bed adsorber is shown in Figure 39. In this arrangement, fresh carbon can be added at the top, and spent or dust-laden carbon is periodically removed from the bottom. Figure 40 shows a fixed-bed odor adsorber. Adsorbers of this general configuration are available as packaged systems, complete

with motor and blower. Air-handling capacities range from 0.2 to 6 m3/s. The life of activated carbon in odor control systems ranges from a few weeks to a year or more, depending on the concentration of the odorous emission.

Applications of Fluidized Bed Adsorbers The injection of alkali compounds into fluidized bed combustors for control of sulfur-containing compounds is one example of the use of fluidized bed adsorbers. Another example is the control of HF emissions from Søderberg aluminum reduction processes by a fixed or circulating fluidized bed of alumina.

3.4

INCINERATION OF GASES AND VAPORS

Incineration is the process by which volatile organic compounds (VOCs), organic aerosols, and most odorous materials in a contaminated gas stream are converted to innocuous carbon dioxide and water vapor using heat energy. Incineration is an effective means for totally eliminating VOCs. The types of incineration commonly used for air pollution control are thermal and catalytic, sometimes with recuperative or regenerative heat recovery.

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Industrial Gas Cleaning and Air Pollution Control To differentiate such air-cleaning systems from liquid and solid waste incinerators, the preferred term to describe such gas and aerosol phase air pollution control systems is now oxidizers.

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Thermal Oxidizers Thermal oxidizers, also known as afterburners or direct flame incinerators, consist of an insulated oxidation chamber in which gas and/or oil burners are typically located. The contaminated gas stream enters the chamber and comes into direct contact with the flame, which provides the heat energy necessary to promote oxidation. Under the proper conditions of time, temperature, and turbulence, the gas stream contaminants are oxidized effectively. The contaminated gas stream enters the combustion chamber near the burner, where turbulence-inducing devices are usually installed. The final contaminant conversion efficiency largely depends on good mixing in the contaminated gas stream and on the temperature of the oxidation chamber. Supplemental fuel is used for start-up, to raise the temperature of the contaminated gas stream enough to initiate contaminant oxidation. Once oxidation begins, the temperature rises further because of the energy released by combustion of the contaminant. The supplementary fuel feed rate is then modulated to maintain the desired oxidizer operating temperature. Most organic gases oxidize to approximately 90% conversion efficiency if a temperature of at least 650°C and a residence time of 0.3 to 0.5 s are achieved in the oxidation chamber. However, oxidation temperatures are typically maintained in the range of 760 to 820°C with residence times of 0.5 to 1 s to ensure conversion efficiencies of 95% or greater. Although the efficiency of thermal oxidizers can exceed 95% destruction of the contaminant, the reaction may form undesirable products of combustion. For example, oxidation of chlorinated hydrocarbons causes the formation of hydrogen chloride, which can have an adverse effect on equipment. These new contaminants then require additional controls. Oxidation systems incorporate primary heat recovery to preheat the incoming contaminated gas stream and, in some cases, provide secondary heat recovery for process or building heating. Primary heat recovery is almost always achieved using air-to-air heat exchangers. Use of a regenerable, ceramic medium for heat recovery has increased due to superior heat recovery efficiency. Secondary heat recovery may incorporate an air-to-air heat exchanger or a waste heat boiler (DOE 1979). Oxidation systems using conventional air-to-air heat exchangers can achieve up to 80% heat recovery efficiency. Regenerative heat exchanger units have claimed as high as 95% heat recovery efficiency and are routinely operated at 85 to 90%. When operated at these high heat recovery efficiencies and with inlet VOC concentrations of 15 to 25% of the LEL, the oxidation process approaches a self-sustaining condition, requiring very little supplementary fuel.

Catalytic Oxidizers Catalytic oxidizers operate under the same principles as thermal oxidizers, except that they use a catalyst to promote oxidation. The catalyst allows oxidation to occur at lower temperatures than in a thermal oxidizer for the same VOC concentration. Therefore, catalytic oxidizers require less supplemental fuel to preheat the contaminated gas stream and have lower overall operating temperatures. A catalytic oxidizer generally consists of a preheat chamber followed by the catalyst bed. Residence time and turbulence are not as important as with thermal oxidizers, but it is essential that the contaminated gas stream be heated uniformly to the required catalytic reaction temperature. The required temperature varies, depending on the catalyst material and configuration. The temperature of the contaminated gas stream is raised in the preheat chamber by a conventional burner. Although the contami-

30.27 nated gas stream contacts the burner flame, the heat input is significantly less than that for a thermal oxidizer, and only a small degree of direct contaminant oxidation occurs. Natural gas is preferred to prevent catalyst contamination, which could occur with sulfurbearing fuel oils. However, No. 2 fuel oil units have been operated successfully. The most effective catalysts contain noble metals such as platinum or palladium. Catalysis occurs at the molecular level. Therefore, an available, active catalyst surface area is important for maintaining high conversion efficiencies. If particulate materials contact the catalyst as either discrete or partially oxidized aerosols, they can ash on the catalyst surface and blind it. This problem is usually accompanied by a secondary pollution problem: odorous emissions caused by the partially oxidized organic compounds. The greatest concern to users of catalytic oxidizers is catalyst poisoning or deactivation. Poisoning is caused by specific gas stream contaminants that chemically combine or alloy with the active catalyst material. Poisons frequently cited include phosphorus, bismuth, arsenic, antimony, lead, tin, and zinc. The first five materials are considered fast-acting poisons and must be excluded from the contaminated gas stream. Even trace quantities of the fastacting poisons can cause rapid catalyst deactivation. The last two materials are slow-acting poisons; catalysts are somewhat tolerant of these materials, particularly at temperatures lower than 540°C. However, even the slow poisons should be excluded from the contaminated gas stream to ensure continuous, reliable performance. Therefore, galvanized steel, another possible source of the slow poisons, should not be used for the duct leading to the oxidizer. Sulfur and halogens are also regarded as catalyst poisons. In most cases, their chemical interaction with the active catalyst material is reversible. That is, catalyst activity can be restored by operating the catalyst without the halogen or sulfur-bearing compound in the gas stream. The potential problem of greater concern with respect to the halogen-bearing compounds is the formation of hydrogen chloride or hydrogen fluoride gas, or hydrochloric or hydrofluoric acid emissions. Some organic compounds, such as polyester amides and imides, are also poisonous. Catalytic oxidizers generally cost less to operate than thermal oxidizers because of their lower fuel consumption. With the exception of regenerative heat recovery techniques, primary and secondary heat recovery can be incorporated into a catalytic oxidation system to further reduce operating costs. Maintenance costs are usually higher for catalytic units, particularly if frequent catalyst cleaning or replacement is necessary. Concern over catalyst life has been the major factor limiting more widespread application of catalytic oxidizers.

Applications of Oxidizers Odor Control. All highly odorous pollutant gases are combustible or chemically changed to less odorous pollutants when they are sufficiently heated. Often, the concentration of odorous materials in the waste gas is extremely low, and the only feasible method of control is oxidation. Odors from rendering plants, mercaptans, and organic sulfides from kraft pulping operations are examples of effluents that can be controlled by incineration. Other forms of oxidation can achieve the same ends (see Chapter 47 of the 2019 ASHRAE Handbook—HVAC Applications). Reduction in Emissions of Reactive Hydrocarbons. Some air pollution control agencies regulate the emission of organic gases and vapors because of their involvement in photochemical smog reactions. Flame afterburning is an effective way of destroying these materials. Reduction in Explosion Hazard. Refineries and chemical plants are among the factories that must dispose of highly combustible or otherwise dangerous organic materials. The safest method of dis-

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posal is usually by burning in flares or in specially designed furnaces. However, special precautions and equipment design must be used in the handling of potentially explosive mixtures.

Adsorption and Oxidation Alternate cycles of adsorption and desorption in an activated carbon bed are used to concentrate solvent or odor vapors before oxidation. This technique greatly reduces the fuel required for burning organic vapor emissions. Fuel savings of 98% compared to direct oxidation are possible. The process is particularly useful in cases where emission levels vary from hour to hour. This technique is common in metal finishing for automotive and office furniture manufacturing. Contaminated gas is passed through a carbon bed until saturation occurs. The gas stream is then switched to another carbon bed, and the exhausted bed is shut down for desorption. A hot inert gas, usually burner flue gas, is introduced to the adsorber to drive off concentrated organic vapors and to convey them to an oxidizer. The volume of this desorbing gas stream is much smaller than the original contaminated gas volume, so that only a small oxidizer, operating intermittently, is required (Grandjacques 1977).

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4.

AUXILIARY EQUIPMENT 4.1

DUCTS

Basic duct design is covered in Chapter 21 of the 2017 ASHRAE Handbook—Fundamentals. This chapter covers only those duct components or problems that warrant special concern when handling gases that contain particulate or gaseous contaminants. Duct systems should be designed to allow thermal expansion or contraction as gases move from the process, through gas-cleaning equipment, and on to the ambient environment. Appropriately designed duct expansion joints must be located in proper relation to sliding and fixed duct supports. Besides withstanding maximum possible temperature, duct must also withstand maximum positive or negative pressure, a partial load of accumulated dust, and reasonable amounts of corrosion. Duct supports should be designed to accommodate these overload conditions as well. Bottom entries into duct junctions create a low-velocity area that can allow contaminants to settle out, thus creating a potential fire hazard. Bottom duct entries should be avoided. Where gases might condense and cause corrosion or sticky deposits, duct should be insulated or fabricated from materials that will survive this environment. A psychrometric analysis of the exhaust gases is useful to determine the dew point. Surface temperatures should then be held above the dew-point temperature by preheating on start-up or by insulating. Slag traps with clean-out doors, inspection doors where direction changes, and dead-end full-sized caps are required for systems having heavy particulate loading. Special attention should be given to high-temperature duct, where the duct might corrode if insulated or become encrusted when molten particulate impacts on cool, uninsulated surfaces. Water-cooled duct or refractory lining is often used where the high operating temperature exceeds the safe limits of low-cost materials. Gas flow through ducts should be considered as a part of overall system design. Good gas flow distribution is essential for measurements of process conditions and can lead to energy savings and increased system life. The minimum speed of gases in a duct should be sufficiently high to convey the heaviest particulate fraction with a degree of safety. Slide gates, balance gates, equipment bypass ducts, and cleanout doors should be incorporated in the duct system to allow for maintenance of key gas-cleaning systems. In some cases, emergency bypass circuitry should be included to vent emissions and protect gas-cleaning equipment from process upsets.

Temperature Controls Control of gas temperature in a gas-cleaning system is often vital to a system’s performance and life. In some cases, gases are cooled to concentrate contaminants, condense gases, and recover energy. In other cases, gas-cleaning equipment, such as fabric filters and scrubbers, can only operate at well-controlled temperatures. Cooling exhaust gases through air-to-air heat transfer has been highly successful in many applications. Controlled evaporative cooling is also used, but it increases the dew point and the danger of acid gas condensation and/or the formation of sticky deposits. However, controlled evaporative cooling to within 28 K of dew point has been used with success. Dilution by injecting ambient air into the duct is expensive because it increases the volumetric flow of gas and, consequently, the size of collector needed to meet gas-cleaning objectives. Water-cooled duct is often used where the gas temperature exceeds the safe limits of the low-cost materials. For dilution cooling, louver-type dampers are often used to inject ambient air and provide fine temperature control. Controls can be used to provide full modulation of the damper or to provide open or closed operation. Emergency bypass damper systems and bypass duct/stacks are used where limiting excessive temperature is critical.

Fans Because static pressure across a gas-cleaning device varies depending on conditions, the fan should operate on the steep portion of the fan pressure-volumetric flow curve. This tends to provide less variation in the volumetric flow. An undersized fan has a steeper characteristic than an oversized fan for the same duty; however, it will be noisier. In the preferred arrangement, the fan is located on the clean gas side of gas-cleaning equipment. Advantages of placing the fan at this point include the following: • A fan on the clean gas side handles clean gases and minimizes abrasive exposure from the collected product. • High-efficiency backward-blade and airfoil designs can be selected because accumulation on the fan wheel is not as great a factor. • Escape of hazardous materials through leaks is minimized. • The collector can be installed inside the plant, even near the process, because any leakage in the duct or collector will be into the system and will not increase the potential for exposure. However, the fan itself should be mounted outdoors, so that the positivepressure duct is outside the work environment. For economic reasons, a fan may be located on the contaminantladen side of the gas-cleaning equipment if the contaminants are relatively nonabrasive, and especially if the equipment can be located outdoors. This arrangement should be avoided because of the potential for leakage of concentrated contaminants to the environment. In some instances, however, the collector housing design, duct design, and energy savings of this arrangement reduce costs. Most scrubbers are operated on the suction side of the fan. This not only eliminates leakage of contaminants into the work area, but also allows for servicing the unit while it is in operation. Additionally, such an arrangement minimizes corrosion of the fan. Stacks on the exhaust streams of scrubbers should be arranged to drain condensate rather than allow it to accumulate and reenter the fan. Fabric filters require special consideration. When new, clean fabric is installed in a collector, the resistance is low, and the fan motor may be overloaded. This overloading may be prevented during startup by using a temporary throttling damper in the main duct, for example, on the clean side of the filter in a pull-through system. Overloading may also be prevented by using a backward-curved blade (nonoverloading fan) on the clean side of the collector.

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Industrial Gas Cleaning and Air Pollution Control 4.2

DUST- AND SLURRY-HANDLING EQUIPMENT

Once the particulate matter is collected, new control problems arise from the need to remove, transport, and dispose of material from the collector. A study of all potential methods for handling the collected material, which might be a hazardous waste, must be an integral part of initial system design.

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Hoppers Dust collector hoppers are intended only to channel collected material to the hopper’s outlet, where it is continuously discharged. When hoppers are used to store dust, the plates and charging electrodes of electrostatic precipitators can be shorted electrically, resulting in failure of entire electrical sections. Fabric collectors, particularly those that have the gas inlet in the hopper, are usually designed on the assumption that the hopper is not used for storage and that collected waste is continuously removed. High dust levels in the hoppers of fabric collectors often result in high dust reentrainment. This reentrainment causes increased operating pressure and the potential for fabric damage. Excessive dust levels not only expose the system to potentially corrosive conditions and fire/explosion hazards, but also place increased structural demands on the system. Aside from misuse of hoppers for storage, common problems with dust-handling equipment include (1) plugging of hoppers, (2) blockage of dust valves with solid objects, and (3) improper or insufficient maintenance. Hopper auxiliaries to be considered include (1) insulation, (2) dust level indicators, (3) rapper plates, (4) vibrators, (5) heaters, and (6) “poke” holes. Hopper Discharge. Dust is often removed continuously from hoppers by means of rotary valves. Alternative equipment includes the double-flap valve, or vacuum system valves. Wet electrostatic precipitators and scrubbers often use sluice valves and drains to ensure that insoluble particulate remains in suspension during discharge.

Dust Conveyors Larger dust collectors are fitted with one or more conveyors to feed dust to a central discharge location or to return it to a process. Drag, screw, and pneumatic conveyors are commonly used with dust collectors. Sequential start-up of conveyor systems is essential. Motion switches to monitor operation of the conveyor are useful.

Dust Disposal Several methods are available for disposal of collected dust. It can be emptied into dumpsters in its as-collected dry form or be pelletized and hauled to a landfill. It can also be converted to a slurry and pumped to a settling pond or to clarification equipment. The advantages and disadvantages of each method are beyond the scope of this chapter, but they should be evaluated for each application.

Slurry Treatment When slurry from wet collectors cannot be returned directly to the process or tailing pond, liquid clarification and treatment systems can be used for recycling the water to prevent stream pollution. Stringent stream pollution regulations make even a small discharge of bleed water a problem. Clarification equipment may include settling tanks, sludge-handling facilities, and, possibly, centrifuges or vacuum filters. Provisions must be provided for handling and disposal of dewatered sludge, so that secondary pollution problems do not develop.

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5.

OPERATION AND MAINTENANCE

A planned program for operation and maintenance of equipment is a necessity. Such programs are becoming mandatory because of the need for operators to prove continuous compliance with emission regulations. Good housekeeping and record-keeping will also help prolong the life of the equipment, support a program for positive relations with regulatory and community groups, and aid problem-solving efforts, should nonroutine maintenance or service be necessary. A typical program includes the following minimum requirements (Stern et al. 1984): • Central location for filing equipment records, warranties, instruction manuals, etc. • Lubrication and cleaning schedules • Planning and scheduling of preventive maintenance (including inspection and major repair) • Storeroom and inventory system for spare parts and supplies • Listing of maintenance personnel (including supplier contacts and consultants) • Costs and budgets for activities associated with operation and maintenance of the equipment • Storage for special tools and equipment

Corrosion Because high-temperature gas cleaning often involves corrosive materials, chemical attack on system components must be anticipated. This is especially true if the temperature in a gas-cleaning system falls below the moisture or acid dew point. Housing insulation should be such that the internal metal surface temperature is 11 to 17 K greater than the moisture and/or acid dew point at all times. In applications with fabric filters where alkali materials are injected into the gas stream to react with acid gases, care must be taken to protect the clean side of the housing downstream of the fabric from corrosion.

Fires and Explosions Industrial gas-cleaning systems often concentrate combustible materials and expose them to environments that are hostile and difficult to control. These environments also make fires difficult to detect and stop. Industrial gas-cleaning systems are, therefore, potential fire or explosion hazards (Billings and Wilder 1970; EEI 1980; Frank 1981). Fires and explosions in industrial process exhaust streams are not generally limited to gas-cleaning equipment. Ignition may take place in the process itself, in the duct, or in exhaust system components other than the gas-cleaning equipment. Once uncontrolled combustion begins, it may propagate throughout the system. Workers around pollution control equipment should never open access doors to gas-cleaning equipment when a fire is believed to be in process; the fire could easily transform into an explosion. The following devices help maintain a safe particulate control system. Explosion Doors. An explosion door or explosion relief valve allows instantaneous pressure relief for equipment when the pressure reaches a predetermined level. Explosion doors are mandatory for certain applications to meet OSHA, insurance, or National Fire Protection Association (NFPA) regulations. Detectors. Temperature-actuated switches or infrared sensors can be used to detect changes in the inlet-to-outlet temperature difference or a localized, elevated temperature that might signal a fire in the gas-cleaning system or a process upset. These detectors can be used to activate bypass dampers, trigger fire alarm/control systems, and/or shut down fans.

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Fire Control Systems. Inert gas and water spray systems can be used to control fires in dust collectors. They are of little value in controlling explosions.

REFERENCES

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ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal.ashrae .org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org/bookstore. ACGIH. 2010. Industrial ventilation: A manual of recommended practice, 27th ed. Committee on Industrial Ventilation, American Conference of Governmental Industrial Hygienists, Cincinnati, OH. AIHA. 1989. Odor thresholds for chemicals with established occupational health standards. American Industrial Hygiene Association, Akron, OH. Beck, A.J. 1975. Heat treat engineering. Heat Treating (January). Beltran, M.R. 1972. Smoke abatement for textile finishers. American Dyestuff Reporter (August). Beltran, M.R. and H. Surati. 1976. Heat recovery vs. evaporative cooling on organic electrostatic precipitators. Proceedings of the American Institute of Plant Engineers/Rossnagel & Associates Sixth Annual Industrial Air Pollution Control Seminar, Cherry Hill, NJ. Bergin, M.H., D.Y.H. Pui, T.H. Kuehn, and W.T. Fay. 1989. Laboratory and field measurements of fractional efficiency of industrial dust collectors. ASHRAE Transactions 95(2):102-112. Billings, C.E., and J. Wilder. 1970. Handbook of fabric filter technology, vol. 1: Fabric filter systems study. NTIS Publication PB 200 648, 2-201. National Technical Information Service, Springfield, VA. Boll, C.H. 1976. Recovering solvents by adsorption. Plant Engineering (January). Buonicore, A.J., and W.T. Davis, eds. 1992. Air pollution engineering manual. Van Nostrand Reinhold, New York. Chopyk, J., and M.C. Larkin. 1982. Smoke and odor subdued with two-stage precipitator. Plant Services (January). DOE. 1979. The coating industry: Energy savings with volatile organic compound emission control. Report TID-28706, U.S. Department of Energy, Washington, D.C. EEI. 1980. Air preheaters and electrostatic precipitators fire prevention and protection (coal fired boilers). Report of the Fire Protection Committee 06-80-07 (September). Edison Electric Institute, Washington, D.C. EPA. 1994. Quality assurance handbook for air pollution measurement systems. Report EPA-600R94038A. Environmental Protection Agency, Washington, D.C. EPA. 2004. Air quality criteria for particulate matter. EPA 600/P-99/002aFbF. U.S. Environmental Protection Agency, Washington, D.C. cfpub2 .epa.gov/ncea/cfm/recordisplay.cfm?deid=87903. Frank, T.E. 1981. Fire and explosion control in bag filter dust collection systems. Proceedings of the Conference on the Hazards of Industrial Explosions from Dusts, New Orleans, LA (October). Grandjacques, B. 1977. Carbon adsorption can provide air pollution control with savings. Pollution Engineering (August). IGCI. 1964. Determination of particulate collection efficiency of gas scrubbers. Publication 1. Industrial Gas Cleaning Institute, Washington, D.C. Kane, J.M., and J.L. Alden. 1982. Design of industrial ventilation systems. Industrial Press, New York. Lapple, C.E. 1951. Processes use many collection types. Chemical Engineering (May):145-151. Lapple, C.E., and H.J. Kamack. 1955. Performance of wet scrubbers. Chemical Engineering Progress (March). MCA. 1968. Odor thresholds for 53 commercial chemicals. Manufacturing Chemists Association, Washington, D.C. (October). NIOSH. 1978. A recommended approach to recirculation of exhaust air. Publication 78-124. National Institute of Occupational Safety and Health, Washington, D.C.

Rossnagel, W.B. 1973. Condensing/precipitator systems on organic emissions. Proceedings of the Third Annual Industrial Air Pollution Control Seminar, Paramus, NJ. Rudnick, S.N., J.L.M. Koehler, K.P. Martin, D. Leith, and D.W. Cooper. 1986. Particle collection efficiency in a venturi scrubber: Comparison of experiments with theory. Environmental Science & Technology 20(3): 237-242. Sauerland, W.A. 1976. Successful application of electrostatic precipitators on asphalt saturator emissions. Proceedings of the American Institute of Plant Engineers/Rossnagel & Associates Sixth Annual Industrial Air Pollution Control Seminar, Cherry Hill, NJ. Semrau, K.T. 1977. Practical process design of particulate scrubbers. Chemical Engineering (September):87-91. Shabsin, J. 1985. Clean plant air PLUS energy conservation. Fastener Technology (Apri1). Sink, M.K. 1991. Handbook: Control technologies for hazardous air pollutants. Report EPA/625/6-91/014. Environmental Protection Agency, Washington, D.C. SIP. 1991. State implementation plans and guidance. Available from regional offices of the U.S. EPA and state environmental authorities. Stern, A.C., R.W. Boubel, D.B. Turner, and D.L. Fox. 1984. Fundamentals of air pollution control, 2nd ed. Chapter 25, Control Devices and Systems. Academic Press, San Diego. Thiel, G.R. 1977. Advances in electrostatic control techniques for organic emissions. Proceedings of the Seventh Annual Industrial Air Pollution/ Contamination Control Seminar, Paramus, NJ. Thiel, G.R. 1983. Cleaning and recycling plant air…Improvement of air cleaner performance and recirculation procedures. Plant Engineering (January 6). Wade, G., J. Wigton, J. Guillory, G. Goldback, and K. Phillips. 1978. Granular bed filter development program. U.S. DOE Report FE-2579-19 (April). White, H.J. 1963. Industrial electrostatic precipitation. Addison-Wesley, Reading, MA.

BIBLIOGRAPHY Crynack, R.B., and J.D. Sherow. 1984. Use of a mobile electrostatic precipitator for pilot studies. Proceedings of the Fifth Symposium on the Transfer and Utilization of Particulate Control Technology 2:3-1. Deutsch, W. 1922. Bewegung und Ladung der Elektrizitätzträger im Zylinderkondensator. Annalen der Physik 68:335. DuBard, J.L., and R.F. Altman. 1984. Analysis of error in precipitator performance estimates. Proceedings of the Fifth Symposium on the Transfer and Utilization of Particulate Control Technology 2:2-1. Faulkner, M.G., and J.L. DuBard. 1984. A mathematical model of electrostatic precipitation, 3rd ed. Publication EPA-600/7-84-069a. Environmental Protection Agency, Washington, D.C. GPO. Annual. Code of federal regulations 40(60). U.S. Government Printing Office, Washington, D.C. Revised annually and published in July. Hall, H.J. 1975. Design and application of high voltage power supplies in electrostatic precipitation. Journal of the Air Pollution Control Association 25(2). HEW. 1967. Air pollution engineering manual. Publication 999-AP-40. Department of Health and Human Services (formerly Department of Health, Education, and Welfare), Washington, D.C. Noll, C.G. 1984. Electrostatic precipitation of particulate emissions from the melting of borosilicate and lead glasses. Glass Technology (April). Noll, C.G. 1984. Demonstration of a two-stage electrostatic precipitator for application to industrial processes. Proceedings of the Second International Conference on Electrostatic Precipitation, Kyoto, Japan (November), pp. 428-434. Oglesby, S., and G.B. Nichols. 1970. A manual of electrostatic precipitator technology. NTIS PB-196-380. National Technical Information Service, Springfield, VA. White, H.J. 1974. Resistivity problems in electrostatic precipitation. Journal of the Air Pollution Control Association 24(4).

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AUTOMATIC FUEL-BURNING SYSTEMS GENERAL CONSIDERATIONS .............................................. 31.1 Terminology ............................................................................. 31.1 System Application................................................................... 31.1 Safety......................................................................................... 31.2 Efficiency and Emission Ratings.............................................. 31.2 GAS-BURNING APPLIANCES ............................................... 31.3 Gas-Fired Combustion Systems............................................... 31.3 Residential Appliances ............................................................. 31.5 Commercial-Industrial Appliances.......................................... 31.6 Applications ............................................................................. 31.7 OIL-BURNING APPLIANCES .............................................. 31.11

Residential Oil Burners.......................................................... Commercial/Industrial Oil Burners ....................................... Dual-Fuel Gas/Oil Burners.................................................... Equipment Selection............................................................... SOLID-FUEL-BURNING APPLIANCES ..................................................................... Capacity Classification of Stokers ......................................... Stoker Types by Fuel-Feed Methods ...................................... CONTROLS............................................................................ Safety Controls and Interlocks ............................................... Operating Controls ................................................................

F

Equipment. Devices other than appliances, such as supply piping, service regulators, sediment traps, and vents in buildings. Flue. General term for passages and conduit through which flue gases pass from the combustion chamber to the outdoors. Flue gas. Products of combustion plus excess air in appliance flues, heat exchangers, and vents. Input rate. Fuel-burning capacity of an appliance in kilowatts as specified by the manufacturer. Appliance input ratings are marked on appliance rating plates. Vent. Passageway used to convey flue gases from appliances or their vent connectors to the outdoors. Vent gas. Products of combustion plus excess air and dilution air in vents.

UEL-BURNING systems provide a means to mix fuel and air in the proper ratio, ignite it, control the position of the flame envelope within the combustion chamber, and control a fuel flow rate for safe combustion-heat energy release for space conditioning, water heating, and other processes. This chapter covers the design and use of automatic fuel-burning systems. The fuel can be gaseous (e.g., natural or liquefied petroleum gas), liquid (primarily the lighter grades of fuel oil or biodiesel), or solid (e.g., coal, or renewable items such as wood or corn). For discussion of some of these fuels, their combustion chemistry, and thermodynamics, see Chapter 28 of the 2017 ASHRAE Handbook—Fundamentals.

1. GENERAL CONSIDERATIONS 1.1

1.2

TERMINOLOGY

The following terminology for combustion systems, equipment, and fuel-fired appliances is consistent with usage in gas-fired appliance standards of the American National Standards Institute (ANSI) and Canadian Standards Association, the National Fire Protection Association’s National Fuel Gas Code (ANSI Standard Z223.1/ NFPA Standard 54), and the Canadian Standards Association’s Natural Gas and Propane Installation Code (CSA Standard B149.1). Air, circulating. Air distributed to habitable spaces for heating, cooling, or ventilation. Air, dilution. Air that enters a draft hood or draft regulator and mixes with flue gas. Air, excess. Air that passes through the combustion chamber in excess of the amount required for complete (stoichiometric) combustion. Air, primary. Air introduced into a burner that mixes with fuel gas before the mixture reaches the burner ports. Air, secondary. Air supplied to the combustion zone downstream of the burner ports. Appliance. Any device that uses a gas, a liquid, or a solid as a fuel or raw material to produce light, heat, power, refrigeration, or air conditioning. Draft. Negative static pressure, measured relative to atmospheric pressure; thus, positive draft is negative static pressure. Draft is the force (buoyancy of hot flue gas or other form of energy) that produces flow and causes pressure drop through an appliance combustion system and/or vent system. See Chapter 34 for additional information. The preparation of this chapter is assigned to TC 6.10, Fuels and Combustion.

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31.11 31.13 31.15 31.15 31.17 31.17 31.18 31.20 31.21 31.22

SYSTEM APPLICATION

The following considerations are important in the design, specification, and/or application of systems for combustion of fossil fuels. Safety. Safety is of prime concern in the design and operation of automatic fuel-burning appliances. For more information, see the sections on Safety and Controls. Appliance standards and installation codes (e.g., ANSI Standard Z21.47/CSA Standard 2.3 for gasfired central furnaces and ANSI Z223.1/NFPA 54, National Fuel Gas Code, in the United States) provide minimum safety requirements. Appliance manufacturers may include additional safety components to address hazards not covered by appliance standards and installation codes. Suitability for Application. The system must meet the requirements of the application, not only in heating capacity, but also in its ability to handle the load profile. It must be suitable for its environment and for the substance to be heated. Combustion System Type. System operation is very much a function of the type of burner(s), means for moving combustion products through the system, proper combustion air supply, and venting of combustion gases to the outdoors. Efficiency. Efficiency can be specified in various ways, depending on the application. Stack loss and heat output are common measures, but for some applications, transient operation must be considered. In very high-efficiency appliances, heat extraction from combustion products may cool vent gas below its dew point, so condensation of water vapor in the combustion products must be handled, and venting design must consider corrosion by combustion products, as well as their lack of buoyancy. Operating Control. Heat load or process requirements may occur in batches or may be transient events. The burner control system must accommodate those requirements, and the combustion system must be able to respond to the controls.

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2020 ASHRAE Handbook—HVAC Systems and Equipment (SI)

Emissions. For safety and air quality reasons, combustion products must not contain excessive levels of noxious materials, notably carbon monoxide, oxides of nitrogen, unburned hydrocarbons, and particulate material such as soot. Fuel Provision. Liquid and solid fuels, liquefied gases, and some gaseous fuels require space for storage. Gaseous and liquid fuels require appropriate piping for the fuel-burning system. Fuel storage and delivery provisions must be of adequate capacity and must be designed to ensure safe operation. Sustainability. Fuel-burning appliances consume fuel resources and produce combustion products that are emitted to the atmosphere. The effect of these inherent characteristics can be minimized by using highly efficient systems with very low emissions of undesirable substances. (See the section on Efficiency and Emission Ratings.) Appliances using environmentally neutral (nonfossil) fuel, such as biofuels processed from vegetable oils and biomass material (e.g., forestry waste), are available. New technologies are also emerging. Through photosynthesis, biomass chlorophyll captures energy from the sun by converting carbon dioxide from the atmosphere and water from the ground into carbohydrates, complex compounds composed of carbon, hydrogen, and oxygen. When these carbohydrates are burned, they are converted back into carbon dioxide and water, and release the sun’s energy. In this way, biomass stores solar energy and is renewable and carbon neutral (UCS 2007). Venting, Combustion Air Supply, and Appliance Installation. Combustion product gases must be handled properly to ensure safety and satisfactory system operation. Adequate air supply must be provided for combustion and ventilation. Appliances must be located to provide safe clearance from combustible material and for convenient service. Standards and Codes. Building codes typically require that fuel-burning appliances be design-certified or listed to comply with nationally recognized standards. Appliance construction, safe operation, installation practices, and emissions requirements are often specified. In some locations, codes require special restraint for seismic or high wind conditions. Cost. The choice of fuel-burning system is often based on it being the least expensive way to provide the heat needed for a process. The basic cost of energy tends to narrow the choices, but the total cost of purchase and ownership should dictate the final decision. Initial cost is the cost of the appliance(s), associated equipment and controls, and installation labor. Operating cost includes the cost of fuel, other utilities, maintenance, depreciation, and various ongoing charges, taxes, fees, etc. Energy cost analysis may indicate that one fuel is best for some loads and seasons, and another fuel is best for other times. Substantial operating cost is incurred if skilled personnel are required for operation and maintenance. Sometimes these costs can be reduced by appliances and control systems that automate operation and allow remote monitoring of system performance and maintenance requirements. Warranties should be considered. See Chapter 38 of the 2019 ASHRAE Handbook—HVAC Applications for a thorough discussion of costs.

1.3

1.4

EFFICIENCY AND EMISSION RATINGS

Heating capacity may be the primary factor in selecting fuelburning appliances, but efficiency and emission ratings are often of equal importance to building owners and governmental regulators.

Steady-State and Cyclic Efficiency Efficiency calculations are discussed in Chapter 28 of the 2017 ASHRAE Handbook—Fundamentals. Boiler and furnace efficiencies are discussed in Chapters 32 and 33 of this volume. Stack Efficiency. Stack efficiency is a widely used rating approach based on measurement of the temperature and composition of gases exhausted by fuel-burning appliances. Knowing the oxygen or carbon dioxide concentration of the flue gases and the fuel’s hydrocarbon content provides a measure of stack mass flow. In conjunction with flue gas temperature, these data allow determination of energy loss in the flue gas exiting the stack. The difference between stack loss and energy input is assumed to be useful energy, and stack efficiency is the ratio of that useful energy to energy input, expressed as a percentage. Generally, the rating is applied to steady-state combustion processes. Flue gas carbon dioxide and oxygen concentrations are affected mostly by the fuel’s hydrocarbon content and by the appliance’s combustion system design. Flue gas temperature is mostly affected by the appliance’s heat exchanger design. Heat Output Efficiency. Some rating standards require actual measurement of heat transferred to the substance being heated. Heat output measurement accounts for all heat losses, not just those in the flue gases. Nonstack heat loss, often called jacket loss, is difficult to measure and may be quite small, but can be accounted for by measuring heat output. The ratio of heat output to energy input, expressed as a percentage, is the heat output efficiency. Load Profile Efficiency. U.S. Federal Trade Commission rules require that some types of residential appliances be rated under protocols that consider load profile. Residential and commercial spaceheating furnaces and boilers, for example, are rated by their annual fuel utilization efficiency (AFUE), which considers steady-state efficiency, heat-up and cooldown transients, and off-season energy consumption by gas pilot burners. Residential storage-type water heaters are rated under a protocol that requires measurement of energy consumption over a 24 h period, during which prescribed amounts of heated water are drawn. A water heater energy factor Ef is calculated from the measurements. The Ef rating accounts for standby losses (i.e., energy loss through the tank and fittings that does not go into the water). Ratings and discussion of AFUE and energy factors can be found in ASHRAE Standard 103 and in product directories by the Gas Appliance Manufacturers Association.

Emissions

SAFETY

All appliance systems must either operate safely or have a way to sense unsafe operation, and safely and promptly shut off the fuel supply before injury or property damage occurs. Safe and unsafe operation sensing and control is generally designed into the combustion control system. Examples of what controls must detect, evaluate, and act on include the following: • • • •

• Heat exchange operation (e.g., circulating air blower operating speed and timing for furnaces) • Flame containment (flame rollout) • Appliance component temperatures • Loss of control power supply

Time to achieve fuel ignition Sufficient combustion air and/or flue gas flow rates Fuel flow rate (e.g., gas orifice pressure) Loss of flame

Regulated Flue Gas Constituents. Appliance safety standards and environmental regulations specify limits for various substances that may be found in combustion flue gases. Substances most often regulated are carbon monoxide, oxides of nitrogen, and soot. Limits for sulfur oxides, unburned hydrocarbons, and other particulate matter may also be specified. Rules vary with location, type of installation, and type of combustion appliance. Regulations that restrict fuel sulfur content are generally intended to reduce sulfur oxide emissions, which may also reduce particulate emissions under certain conditions. Flue Gas Concentration Limits. Standards and codes often specify maximum concentration levels permitted in flue gases.

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Because flue gases may be diluted by air, requirements invariably specify that measured concentration levels be adjusted by calculation to a standard condition. In appliance certification standards, that condition is typically the air-free level (i.e., what the concentration would be if there were no excess air). Carbon monoxide, for example, is often limited to 400 parts per million air-free. Air quality regulations sometimes specify the maximum level for a fixed degree of excess air. For oxides of nitrogen, the level is often specified in parts per million at 3% oxygen. The calculation, in effect, adds or subtracts dilution air to reach a condition at which oxygen concentration is 3% of the exhaust gas volume. Emission per Unit of Useful Heat. In some U.S. jurisdictions, regulations for gas-fired residential furnaces and water heaters require that emission of oxides of nitrogen not exceed a specified level in nanograms per joule of useful heat. Measured flue gas emission levels are compared with the benefit in terms of heat output. Under this rating method, high efficiency is rewarded. Regulations for new installations may differ from those for existing systems and may be more stringent. Mass Released to Atmosphere. Emissions from large fuel-fired appliances are often regulated at the site in terms of mass released to the atmosphere. Limits may be expressed, for example, in terms of kilograms per gigajoule, or megagrams per year.

2.

GAS-BURNING APPLIANCES

2.1

GAS-FIRED COMBUSTION SYSTEMS

Gas-burning combustion systems vary widely, the most significant differences being the type of burner and the means by which combustion products are moved through the system. Gas input rate control also has a substantial effect on combustion system design.

Burners A primary function of a gas burner is mixing fuel gas and combustion air in the proper ratio before their arrival at the flame. In a partially-aerated burner (Bunsen burner), only part of the necessary combustion air is mixed with the gas ahead of the flame. This primary air is typically about 30 to 50% of the stoichiometric air (i.e., that amount of air necessary for complete combustion of the gas). Combustion occurs at the point where adequate secondary air enters the combustion zone and diffuses into the mixture. In most cases, secondary air entry continues downstream of the burner and heat release is distributed accordingly. The total of primary and secondary air typically ranges from 140 to 180% of the stoichiometric air (i.e., 40 to 80% excess air). Most often, partially aerated burners are atmospheric or naturaldraft burners (i.e., they operate without power assist of any kind), which have the advantage of quiet operation. Fuel gas is injected from a pressurized gas supply through an injector (orifice) to form a gas jet, which propels discharged gas into the burner throat, entraining primary air by viscous shear. Primary air may also be drawn into the burner throat by venturi action. Fuel gas and air are mixed in a mixing tube before their arrival at the burner ports where burning occurs. A typical partially aerated burner is illustrated in Figure 1. A premix burner is a power burner in which all or nearly all of the combustion air is mixed with the fuel gas before arrival at the flame. Because the necessary air is present at the flame front, combustion and heat release take place in a compact zone and there is no need for secondary aeration. Combustion quality (i.e., emission performance) tends to be better than that of partially aerated burners because of inherent mixing advantages, resulting in lower peak temperatures and lower residence times at these temperatures. This especially benefits NOx emission levels. Premix burners can normally be operated at lower excess air levels (often 15 to 20%) to increase combustion efficiency. Low excess air increases flame

31.3 temperature, which enhances heat exchange but imposes greater thermal stress on the combustion chamber and its components. Extremely low excess air may result in higher CO and/or NOx emissions, as with any burner design. A fan is almost always necessary to force the mixture of gas and air through a premix burner. Airflow is three or four times that through partially aerated burners, and the associated pressure drop is normally too much to be handled by fuel gas entrainment or stack draft. In general, appliances with premix burners are tuned more finely than those with partially aerated burners, to take advantage of their inherent advantages and to ensure reliable operation. A typical premix burner system is illustrated in Figure 2.

Combustion System Flow In broad terms, flow through the combustion system (i.e., burner, combustion chamber, heat exchanger, and venting) is powered by natural draft and/or by a fan. In a natural-draft system, the low density of hot combustion products creates a buoyant flow through the combustion chamber, heat exchanger, and venting system. The amount of flow through the system is a function of many factors, including chimney height, diameter, flue gas temperature, and friction loss through the appliance and vent system components. For additional information, see Chapter 35. Fan-assisted combustion systems have become common. A fan pushes or pulls combustion air, flue gas products, and fuel gas through the burner, and combustion products through the combustion chamber and heat exchanger (and, in some cases, the venting). Some fan-assisted systems use atmospheric burners, applying the fan power mainly to force flow through an enhanced heat exchange process. Serpentine heat exchangers, for example, typically require fan assist because they have too much pressure drop to operate in natural-draft mode. In systems with significant burner pressure drop, such as that of a premix burner, fan power is required for the burner as well. Fan-assisted systems can operate with or without pressurizing the vent, depending on the flow rate of the combustion system, flue gas conditions (temperature and buoyancy), resistance of the venting system, and location of an induced-draft fan, if used. For more information, see the discussion of venting in the Applications section.

Fig. 1 Partially Aerated (Bunsen) Burner

Fig. 2 Premix Burner

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Forced-draft systems (Figure 3) use a fan to force air into the combustion chamber and heat exchanger under positive pressure (higher than atmospheric). Often, forced-draft systems also operate with positive pressure in at least part of the vent system. Induced-draft systems (Figure 4) use a fan in the appliance near the flue collar or vent outlet to pull flue gas through the combustion chamber and heat exchanger. These systems operate with negative pressure at all points before the fan. Packaged power burners are often used in factory-built heating appliances and in field-assembled installations. These burners include all gas- and air-handling components, along with ignition controls, housing, mechanical means for mounting to the heating appliance, and connection of fuel gas and electric power. They tend to have a gun-type configuration (i.e., gas and air are mixed in an outlet tube with burner head, the latter inserted into the heating appliance’s combustion chamber). Packaged power burners may also include special hardware and controls for gas input rate control, and some may include special features to reduce combustion emissions. Flue gas recirculation (i.e., returning some combustion products to the flame) is sometimes used to reduce NOx production. A typical packaged power burner is shown in Figure 5. Pulse combustion is a process in which combustion system flow is motivated by low-frequency pressure pulses created in the combustion chamber by cyclic/repetitive self-generating ignition of an air/gas mixture. It provides low emission levels and enhanced heat transfer. The oscillating nature of flow through the system provides a beneficial “scrubbing” effect on heat exchanger surfaces. Additional discussion of pulse combustion is provided in Chapter 28 of the 2017 ASHRAE Handbook—Fundamentals.

Ignition Safety standards and codes specify requirements for ignition and proof of flame presence. Good design practices result in ignition

Fig. 3

Forced-Draft Combustion System

Fig. 4 Induced-Draft Combustion System

that is immediate, smooth, and complete. Once flame is established, the ongoing presence of the ignition source or the flame itself must be ensured (i.e., flame supervision by the combustion control must detect loss of pilot and/or main flame and immediately shut off fuel gas flow). Pilot burners have been used very effectively for decades in appliances such as small residential water heaters and in very large field-assembled installations. The pilot flame may be detected by a temperature sensor (thermocouple) or by various electronic sensing systems. Pilot flames may be continuous or ignited only when there is a demand for heat (intermittent pilot operation). Some types of appliances use direct ignition. A spark igniter or a hot surface igniter is applied directly at the main burner ports to ignite the gas/air mixture. Direct-ignition systems also include a means, usually electronic, to sense presence of the flame. Ignition system standards and installation codes usually include requirements for other parameters such as flame failure response time, trial for ignition, and combustion chamber purging. For listed or design-certified appliances, applicable requirements have been test-verified by listing or certifying agencies. Other appliances and those installed in some building occupancy classes may be subject to special ignition and flame safety requirements in building and safety codes or by insurance underwriters. (See the section on Controls for more detail.)

Input Rate Control A wide range of heat input is sometimes required of gas-fired appliances. Space heating, for example, is often done by zones, and must work under a wide range of outdoor conditions. Some waterheating and process applications also require a wide gas input rate range, and transients can be very steep input rate swings. Gas burners with staged or modulating control can be applied to meet these requirements. Staged systems can be operated at discrete input rate levels, from full rated input to preset lower rates, sometimes with airflow remaining at the full-rate level and sometimes with proportional control of combustion air. Efficiency can be enhanced when combustion air is proportionally controlled, because flame temperature can remain high while the amount of heat exchanger surface per unit of input effectively becomes larger. A common staging approach uses a twostage gas pressure regulator to change fuel gas pressure at an injector or metering orifice(s). In other designs, staging is accomplished by operating groups of individual burners or combustion chambers, each under control of its own gas valve and, if necessary, having its own ignition control. An extension of this approach is to use multiple individual heating appliances, controlled such that one or more can be called on as needed to meet the demand. Modulating burner systems vary the input rate continuously, from full rated input to a minimum value. Modulation may be done by a throttling device in the gas burner piping, or with a modulating gas pressure regulator. Modulating systems require special controllers to provide a signal to the gas flow control device that is in some

Fig. 5 Packaged Power Burner

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31.5 using the combustion air mass flow, along with burner temperature or flame ionization signal, are also coming into commercial use. This technology promises to allow the input and air-fuel ratio to be closely controlled independently of factors, such as altitude, vent length, and even fuel gas composition. In addition, because electronically linked systems do not rely on a minimum combustion air pressure drop to control gas flow, much greater turn-down ratios are theoretically possible than with pneumatic systems.

2.2

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Fig. 6

Combustion System with Linked Air and Gas Flow

Fig. 7 Pneumatically Linked Gas/Air Ratio Combustion System way proportional to the demand. As with staged systems, some designs provide commensurate control of combustion air, whereas others control only gas input rate with constant combustion airflow rate. Figure 6 illustrates a system in which combustion air and gas flow rates are throttled and linked. Pneumatically linked gas/air ratio burner systems, in which combustion air is controlled and gas follows proportionately, are a type of modulating burner system. One such system is shown in Figure 7. In these systems, airflow is controlled by dampers or a variable-speed fan. Combustion air is directed through a restriction (typically a venturi), so that the pressure drop across this restriction increases as a function of air mass flow (in accordance with the Bernoulli Equation). A fuel gas regulator is referenced to the upstream side of this restriction and controls gas flow into the combustion air at a point downstream of the restriction. The regulator set point is fixed (often very close to zero). As airflow increases through the restriction, so does the pressure drop across it. As this happens, the regulator must open further, increasing the gas flow, to maintain its set point. At a regulator set point of zero, the air/fuel ratio is theoretically constant at all firing rates. The regulators typically used in these systems are referred to as zero-pressure regulators, zero governors, or negative-pressure regulators. The air restriction may be located upstream or downstream of the blower. An advantage to placing the venturi at the outlet is that the blower does not handle the combustible gas mixture. Placing the venturi at the blower inlet, however, promotes better mixing. Pneumatically linked systems require a minimum pressure drop across the combustion air restriction for the gas regulator to be able to track the airflow accurately. This requirement typically limits the turndown ratio (maximum firing rate divided by minimum firing rate) to approximately 5:1. Electronically linked gas-air ratio burner systems adjust gas flow to match a variable combustion airflow in a manner similar to pneumatically linked systems; however, measurement of airflow and control of gas flow are by electronic means. Typically, such systems measure airflow using a mass flow sensor and adjust gas flow using a throttling valve operated by a stepper motor (in some cases completely eliminating the need for a gas regulator on the appliance). Oxygen trim and other technologies to infer the actual air/fuel ratio

RESIDENTIAL APPLIANCES

Boilers Residential space heating is often done with gas-fired lowpressure steam or hot-water boilers (i.e., steam boilers operating at 100 kPa [gage] or less, or hot-water boilers operating at 1100 kPa [gage] or less with 120°C maximum water temperature). Steam or hot water is distributed to convectors, radiators, floor piping, fan-coils, or other heat transfer devices in the space to be heated. Space temperature control may be by zone, in which case the boiler and distribution system must be able to accommodate reduced-load operation. Burners and combustion systems can be any of the previously described types, and some designs include input rate control. Rules of the U.S. Federal Trade Commission (FTC) and federal law require residential boilers with input rates less than 88 kW to comply with minimum efficiency requirements, following the rating protocol of ASHRAE Standard 103. For hot-water boilers, 82% annual fuel utilization efficiency (AFUE) is required; for steam boilers, 80%. For ratings and technical information on rating protocol, see the AHRI Directory of Certified Product Performance (ahridirectory.org). Manufacturers’ literature also provides technical data and ratings. Some boilers have low mass and essentially instantaneous response, whereas others have higher water volume and mass, which provides a degree of inherent storage capacity to better handle load change. Both steam and hot-water space-heating boilers are available in models having internal coils for service water heating. These combination boilers eliminate the need for a separate water heater, but they must be operated whenever service water may be needed, including times when space heating is not necessary. For a comprehensive discussion of boilers, see Chapter 32.

Forced-Air Furnaces Central gas-fired, forced-air furnaces are the most common residential space-heating systems in the United States and Canada. Forced-air furnaces are available in configurations for upflow, downflow, and horizontal flow air distribution. Most have induceddraft combustion systems with Bunsen-type burners, and are typically of modular design (i.e., burner and heat exchanger modules are used in multiples to provide appliance models with a range of heating capacities). Some are available with staged or modulated input rate, and some have coordinated control of combustion air and circulating airflow. U.S. federal law require furnaces with firing rates less than 66 kW to meet minimum efficiency ratings; see the National Appliance Energy Conservation Act (NAECA) for details. The recognized rating procedure for furnaces is defined in Code of Federal Regulations 10 CFR 430.32, which references ASHRAE Standard 103. Federal Trade Commission (FTC) rules require manufacturers to publish AFUE ratings in their literature. In addition, independently certified AFUE ratings are published by organizations such as AHRI (2014). For detailed discussion of furnaces, see Chapter 33.

Water Heaters In the United States, most residential water heaters are of the storage type (i.e., they have relatively low gas input rates and significant hot-water storage capacity). Typically, a single Bunsentype burner is applied beneath a flue that rises through the stored

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water. Flue gas usually flows by natural draft. ANSI and Canadian Standards Association (CSA) appliance standards limit the input rate of these heaters to 22 kW and require that water heaters manufactured since mid-2003 be designed to be flammable vapor ignition resistant (FVIR) because water heaters are often installed in garages and should not be an ignition source in case gasoline or other volatile or flammable substances may be present. Instantaneous water heaters, often designed for mounting to a wall, are characterized by low water storage capacity and low mass, and have special burners and control systems designed for immediate response to demand for hot water. Burner input rate is typically linked to water flow rate and temperature, often by both hydraulic and electronic mechanisms. Because there is no storage, input rates tend to be higher than for storage heaters, and must be adequate for instantaneous demand. Standby losses of instantaneous water heaters are typically less than those of storage heaters. However, U.S. water use habits favor storage water heaters, which are used more extensively throughout the country. In the United States, the efficiency of residential storage-type water heaters is based on a Department of Energy (DOE) protocol that requires measurement of energy consumption over a 24 h period, during which prescribed amounts of heated water are drawn. An energy factor Ef is calculated from the measurements. The rating accounts for standby losses (i.e., heat lost from stored hot water by conduction and convection through tank walls, flue, and pipe fittings to the environment). Hot-water delivery flow also appears in ratings. For storage heaters, the first-hour rating (i.e., volume of water that can be drawn in the first hour of use) is provided. For instantaneous heaters, the maximum flow rate in litres per second is provided. See AHRI (2007) for ratings and details about the energy factor.

Combination Space- and Water-Heating Appliances Residential appliances that provide both space and water heating are available in a variety of configurations. One configuration consists of a specially designed storage water heater that heats and stores water for washing activities or for use as a heat transfer medium in a space-heating fan-coil unit. Another configuration, common in Europe and Asia, is a wall-hung boiler that provides hot water for use in either mode. Storage or instantaneous heating capacity must be adequate to meet user demand for showers or other peak activity. Control systems normally prioritize hot-water consumption requirements over the need for space heating, which is allowed only when the hot-water demand has been satisfied. Descriptions and technical data for these and other configurations are available from manufacturers. Ratings are provided in a special section of AHRI (2007). The method for testing and rating of combination space and water heating appliances is specified in ASHRAE Standard 124.

Pool Heaters Pool heaters are a special type of water heater designed specifically for handling high flow rates of water at relatively low water temperature. Various burner and combustion system approaches are used in pool heaters. Input rate control is not normally incorporated because swimming pools are of very high mass and do not change temperature rapidly. Consult manufacturer and pool industry technical data for pool heater selection and application factors.

Conversion Burners Conversion burners are complete burner and control units designed for installation in existing boilers and furnaces. Atmospheric conversion burners may have drilled-port, slotted-port, or singleport burner heads. These burners are either upshot or inshot types. Figure 8 shows a typical atmospheric upshot gas conversion burner. Several power burners are available in residential sizes. These are of gun-burner design and are desirable for furnaces or boilers with restricted flue passages or with downdraft passages.

Fig. 8 Typical Single-Port Upshot Gas Conversion Burner Conversion burners for domestic application are available in sizes ranging from 12 to 120 kW input, the maximum rate being set by ANSI Standard Z21.17/CSA 2.7. However, large gas conversion burners for applications such as apartment building heating may have input rates as high as 260 kW or more. Successful and safe performance of a gas conversion burner depends on numerous factors other than those incorporated in the appliance, so installations must be made in strict accordance with current ANSI Standard Z21.8. Draft hoods conforming to current ANSI Standard Z21.12 should also be installed (in place of the dampers used with solid fuel) on all boilers and furnaces converted to burn gas. Because of space limitations, a converted appliance with a breeching over 300 mm in diameter is often fitted with a double-swing barometric regulator instead of a draft hood.

2.3

COMMERCIAL-INDUSTRIAL APPLIANCES

Boilers Boilers for commercial and industrial application can be very large, both in physical size and input rate. Virtually any requirement for space heating or other process can be met by large boilers or multiple boilers. The heated medium can be water or steam. (See Chapter 32 for extensive discussion.)

Space Heaters A wide variety of appliances is available for large air-heating applications. Some of them heat air by means of hot-water coils and are used in conjunction with boilers. Some accomplish space heating by means of a fuel-fired heat exchanger. Others fire directly into the heated space. Forced-air fuel-burning furnaces for commercial and industrial application are essentially like those for residential use, but have larger heating and air-handling capacity. Input rates for single furnaces certified under ANSI Standard Z21.47/CSA Standard 2.3 may be as high as 117 kW. High capacity can also be provided by parallel (twinned) application of two furnaces. Most manufacturers provide kits to facilitate and address the special safety, mechanical, and control issues posed by twinned application. See manufacturer data and Chapter 33 for additional information about forced-air furnaces and their application. Duct furnaces are fuel-fired appliances for placement in fieldassembled systems with separate air-moving means. Combustion products heat air through heat exchangers mounted in the airstream. The combustion components, heat exchangers, and controls are prepackaged in a cabinet suitable for mounting in a duct system. For proper operation of the duct furnace, the airflow rate must be within the range specified by the manufacturer. See manufacturer data and Chapter 33 for additional information. Unit heaters are free-standing appliances for heating large spaces without ductwork. They are often placed overhead, positioned to

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Automatic Fuel-Burning Systems direct heat to specific areas. Typically, they incorporate fuel-fired heat exchangers and a fan or blower to move air through the exchangers and into the space. However, in some designs the fuel-fired exchangers are replaced by air-heating coils using hot water as the heating medium. Chapter 28 provides further information on unit heaters. Direct gas-fired makeup-air heaters do not have heat exchangers. They heat large spaces by firing combustion products directly into the space, accompanied by a large quantity of dilution airflow. They have special burners capable of operation in high airflow and controls that allow operation over a wide fuel input rate range. As their name implies, they are used in applications requiring heated makeup air in conjunction with building exhaust systems. (See Chapter 28 for additional information.) Infrared heaters radiate heat directly to surfaces and objects in a space. Air may be heated by convective heat transfer from the objects. They are suitable for overall heating of a building, but their selection is often based on their ability to radiate heat directly to people in limited areas without intentionally heating items or air in the space, such as in work stations in large open spaces that are not otherwise heated. Indirect infrared heaters use a radiating surface, such as a tube, between the combustion products and the space. Direct infrared heaters use the burner surface, typically a glowing ceramic or metal matrix, as the radiator. Chapter 16 provides additional information. See also manufacturers’ data.

Water Heaters Large-load water-heating applications, such as for large residence buildings, schools and hospitals, restaurants, and industrial processes, require water-heating appliances with correspondingly high fuel input rates, usually in conjunction with substantial hot-water storage. Large tank-type heaters with multiple flues are used in many intermediate-sized applications. They may operate as natural-draft systems, but forced- and induced-draft designs are common. Large applications or those with very high short-term water draw are often handled by means of large unfired storage tanks in conjunction with water-tube or other low-volume, high-input heaters. Fan-assisted combustion systems are increasingly common in those heaters. Premix burners may be used mainly to help meet emissions restrictions. Chapter 51 of the 2019 ASHRAE Handbook—HVAC Applications extensively discusses water-heating issues and appliances. ASHRAE Standards 118.1 and 118.2 provide methods of testing for rating commercial and residential water heaters, respectively.

Pool Heaters Fuel-fired pool heaters are available in very large sizes, with fuel input rates ranging to more than 1000 kW. They are designed to handle low-temperature water at high flow rates, and have sensitive temperature controls to ensure swimmer comfort and energy efficiency. Heating pool water with appliances not designed for that purpose can result in severe problems with combustion product condensation, corrosion, and/or scaling. See manufacturers’ data for complete information.

2.4

APPLICATIONS

Gas-burning appliances cannot perform as intended unless they are properly installed and set up. Once an appliance of appropriate type, size, and features is selected, the location, fuel supply, air for combustion and ventilation, and venting must be considered and specified correctly. Other factors, notably elevation above sea level, must also be considered and handled.

Location Listed appliances are provided with rating plate and installation information, with explicit requirements for location. The required clearance to combustible material is particularly important, to eliminate the hazard of fire caused by overheating. Other requirements

31.7 may be less obvious. Adequate space must be provided for connecting ductwork, piping, and wiring, and for convenient maintenance and service. There must be access to chimneys and vents, and vent terminal locations must comply with specific requirements for safe discharge of combustion products without injury to people or damage to surroundings. Outdoor appliances must be located in consideration of wind effects and similar factors. Local building codes provide basic rules for unlisted appliances and may impose additional requirements on listed appliances.

Gas Supply and Piping Natural Gas. Natural gas is usually provided by the local gas utility. Most North American utilities provide substantial and reliable supply pressure, but it is important to verify adequate supply pressure during maximum simultaneous gas consumption by all appliances sharing the supply, and to design adequate supply piping between the utility supply point and the gas-burning appliance. On listed appliances, rating plates specify the minimum supply pressures at which the appliances will operate safely and as intended. This information is also provided in installation instructions, and is available from manufacturers before purchasing appliances. Gas piping between the utility company meter and the appliance must provide adequate pressure when all concurrent loads operate at their maximum rate. Tables provided in ANSI Standard Z223.1/ NFPA Standard 54 (National Fuel Gas Code), CSA Standard B149.1, local codes, and elsewhere provide procedures for ensuring adequate pressure. For residential and light commercial services, utility companies typically provide gas at 1.75 kPa. Building distribution piping is usually designed for a full-load pressure drop of less than 75 or 125 Pa. Industrial and large-building applications are often supplied with gas at higher pressures; in that case, the distribution piping can be designed for larger pressure drop, but the end result must supply pressure to an individual appliance within the range required by its manufacturer. A pressure regulator may be required at the appliance to reduce pressure to comply with the rating plate pressure requirement. Liquefied Petroleum Gas (LPG). LPG can contain a range of gas components. If it is not commercial propane or butane, the actual composition must be ascertained and accommodated. LPG is stored on site as a liquid at moderate pressure; it is vaporized as gas is drawn. A pressure regulator at the tank reduces pressure for distribution to the appliance through piping, subject to the same considerations as for natural gas. In the United States and Canada, normal supply pressure for residential and light commercial applications is 2.75 kPa. Appliance rating plates and installation instructions typically require that supply pressure be maintained at or near that level. Piping must be designed to ensure adequate pressure when concurrent connected loads operate at their maximum input rates. An important but sometimes overlooked issue is the need to provide heat to vaporize liquefied petroleum in the tank to deliver gas. In most residential applications, heat for vaporization is simply taken from outdoor air through the tank walls. This natural heat source may become inadequate, however, as the air temperature falls, draw rate increases, or tank liquid level falls. Commercial propane and butane have boiling point temperatures of –42°C and 0°C, respectively. As the LPG tank approaches the boiling point temperature, tank pressure falls to the point where the gas cannot be supplied at the required rate. In these cases and in high-demand applications, supplemental heat may be necessary to vaporize LPG. Information on selection and application of vaporization equipment is available from LPG dealers and distributors, from the National Propane Gas Association, and in various codes and standards.

Air for Combustion and Ventilation In application of fuel-burning appliances, inadequate provision of air for combustion and ventilation is a serious mistake. In the worst

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scenario, shortage of combustion air results in incomplete combustion and production of poisonous carbon monoxide, which can kill. Inadequate ventilation of the space in which an appliance is installed can result in high ambient temperatures that stress the appliance itself or other appliances or materials in the vicinity. For those reasons, building codes and manufacturers’ installation instructions include requirements for combustion and ventilation air supply. Requirements vary, with several factors having to do with how easily air can get to the appliance from the outdoors. Infiltration is seldom adequate, and it is usually necessary to provide dedicated means for supply of air for combustion and ventilation. In cold regions, take measures to prevent freezing of water pipes and other equipment by cold air in the appliance space. For more information, see Ackerman et al. (1995) and Dale et al. (1997).

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Draft Control Natural-draft appliances typically include a draft hood or diverter to decouple the combustion system from undesirable draft effects, notably the draft or pull of the chimney. These devices provide a path for dilution air to mix with flue gas before entering the connector between the appliance and the chimney, and to accommodate updraft and downdraft variations that occur in the field because of wind. A barometric draft control is a similar device that uses a damper to control the flow of dilution air into the vent. The damper of a barometric draft control can be manually adjusted to regulate the draft imposed on the appliance. Often, this is accomplished with special weights, in conjunction with measurement with a draft gage. Draft controls should be supplied by the appliance manufacturer as part of the appliance combustion controls. See Chapter 35 for design considerations for vent and chimney draft control.

Venting Safety and technical factors must be considered in venting appliances, including some less obvious considerations. Consequences of incorrect venting are very serious, and can include production of lethal carbon monoxide, spilling of heat and combustion moisture indoors, and deterioration of the vent or chimney caused by condensation of water vapor from vent gas. High-efficiency appliances can produce combustion products at temperatures near or below their dew point. Venting those flue gases requires use of special materials and installation practices, and provision for condensate drainage and disposal. Comprehensive guidance for design of venting systems is provided in Chapter 35. For many North American gas-fired appliances, however, ANSI Standard Z21.47/CSA Standard 2.3 and ANSI Standard 21.13/CSA Standard 4.9 require categorization by the type of vent system necessary for safe and effective operation. Appliances are tested to determine the temperature and pressure of vent gas released into the vent. The categories, which apply only to appliances design-certified as complying with standards having category specifications, are as follows: Category I II III IV

Vent Static Vent Gas Temperature High Enough to Avoid Pressure Excessive Condensate Production in Vent? Nonpositive Nonpositive Positive Positive

Yes No Yes No

Category I and II appliances with a forced- or induced-draft blower to move combustion air and combustion products through the appliance flue create no pressure at the appliance flue exit (entrance to the venting system), and therefore do not augment draft in the vent. Most local building codes include vent sizing tables and requirements that must be used for design of venting systems for category I

appliances. Those tables are adopted from the National Fuel Gas Code, which distinguishes between appliances with draft hoods and appliances having fan-assisted combustion systems without draft hoods. In both cases, vent gases flow into the vent at category I conditions, but there is less dilution air with fan-assisted appliances than with draft-hood-equipped appliances. Vent gas flow and condensation tendencies differ accordingly. The tables specify • Maximum (NAT Max) input rates for single-appliance vent systems and multiple-appliance vent connectors of given sizes for draft hood-equipped appliances • Minimum (Fan Min) and maximum (Fan Max) input rates for single-appliance vent systems and multiple-appliance vent connectors of given sizes for fan-assisted appliances • Maximum (Fan + NAT) input rates for multiple-appliance common vents of given sizes for combinations of draft-hoodequipped and fan-assisted appliance systems • Maximum (Fan + Fan) input rates for multiple-appliance common vents of given sizes for fan-assisted appliance systems Category II appliances are rare because it is difficult to vent low-temperature flue gas by its own buoyancy. Category III and IV appliances, with positive vent pressure, are common. Those appliances must be vented in accordance with the manufacturers’ installation instructions, and require special venting materials. Category II and IV appliances also require venting designs that provide for collection and disposal of condensate. Condensate tends to be corrosive and may require treatment. Appliances designed for installation with piping and terminals for both venting of flue gas and intake of combustion air directly to the appliance are called direct-vent (and sometimes, erroneously, sealed combustion) systems. (Sealed combustion systems take combustion air from outside the space being heated, not necessarily outdoors, and all flue gases are discharged outdoors; this is not a balanced system.) The vent and combustion air intake terminals of direct-vent appliances should be located outdoors, close to each other, so that they form a balanced system that is not adversely affected by winds from various directions. Vent pipe and combustion air intake pipe materials are provided or specified by the appliance manufacturer, and their use is mandatory. A variation in which only vent materials are specified, with combustion air taken directly from inside the conditioned space, is referred to as a direct-exhaust system. Most category III and IV systems are direct-vent or direct-exhaust systems. Unlisted appliances must be vented in accordance with local building codes and the manufacturers’ installation instructions. Clearance from vent piping to combustible material, mechanical support of vent piping, and similar requirements are also included in local codes.

Building Depressurization Appliance operation can be affected by operation of other appliances and equipment in the building that change building pressure with respect to outdoor pressure. Building pressure can be reduced by bathroom and kitchen exhaust fans, cooktop range downdraft exhausters, clothes dryers, fireplaces, other fuel-burning appliances, and other equipment that removes air from the building. If building pressure is significantly lower than outdoor pressure, venting flue gases to the outdoors might be adversely affected and potentially hazardous combustion products may be spilled into the inhabited space, especially from category I and II appliances. Category III and IV appliances are less susceptible to venting and spillage problems, because these appliances produce pressure to force the flue gases through their vents to the outdoors. In addition, direct-vent appliances of all vent categories take their combustion air directly from the outdoors, which makes them even less susceptible to building depressurization. Wind can produce building depressurization, if building infiltration and exfiltration are unfavorably imbalanced.

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Gas Input Rate Gas input rate is the rate of heat energy input to an appliance, measured in kilowatts. The unit heating value of the fuel gas is expressed in megajoules per cubic metre. In North America, higher heating value (HHV) is commonly used to specify the heat available from gas when combusted. HHV includes all of the heat available by burning fuel gas delivered at 15.0°C and 101.325 kPa (i.e., standard conditions for the gas industry in North America) when combustion products are cooled to 15.0°C and water vapor formed during combustion is condensed at 15.0°C. (Note: Standard conditions of 15.0°C and 101.325 kPa [see the section on U.S. Standard Atmosphere in Chapter 1 of the 2017 ASHRAE Handbook—Fundamentals] are slightly different from the standard conditions of 60°F [15.56°C] and 14.735 psia [30.00 in. Hg, or 101.594 kPa] used by the gas industry in North America and in the I-P version of this chapter.) See Chapter 28 of the 2017 ASHRAE Handbook — Fundamentals for more information on HHV. In practical laboratory or field situations, fuel gas is not delivered at standard conditions. Determination of appliance input rate must include compensation for the actual temperature and pressure conditions. In the laboratory, gas input rate is calculated with the following equation:  Ts  P  Q = 0.2778HHV  VFR  ------------------ T  Ps  where Q = gas input rate, kW HHV = gas higher heating value at standard temperature and pressure, MJ/m3 VFR = fuel gas volumetric flow rate at meter temperature and pressure, m3/h Ts = standard temperature, 288.15 K (15.0°C + 273.15 K) P = fuel gas pressure in gas meter, kPa T = absolute temperature of fuel gas in meter, K (fuel gas temperature in °C + 273.15 K) Ps = standard pressure, 101.325 kPa Example 1. Calculate the gas input rate for 38.00 MJ/m3 HHV fuel gas, 24.0°C fuel gas temperature, 97.788 kPa barometer pressure (300.0 m altitude), 2.8 m3/h of fuel gas volumetric flow rate clocked at the meter with 1.7 kPa fuel gas pressure in the gas meter. P = B + Pf where B Pf P T

= = = =

local barometric pressure, kPa fuel gas pressure in gas meter, kPa 97.788 kPa + 1.7 kPa = 99.488 kPa 24.0°C + 273.15 K = 297.15 K

Thus, 0.2778 kWh Q = -----------------------------MJ 3

3

38.00 MJ/m  2.800 m  h   288.15 K  99.488 kPa   -------------------------------------------------------------------------------------------------------------------------------------297.15 K  101.325 kPa = 28.143 kW gas input rate

In the field, input can be measured in two ways: • With a gas meter, usually furnished by the gas supplier at the gas entrance point to a building • By using the appliance burner gas injector (orifice) size and the manifold pressure (pressure drop through the injector [orifice])

31.9 Gas input rate is calculated with the following equation, when fuel gas volumetric flow rate is measured using the gas supplier’s meter. Q = 0.2778HC  VFR where Q 0.2778 HC VFR

= = = =

gas input rate, kW conversion factor, MJ/h to kW local gas heat content, MJ/m3 fuel gas volumetric flow rate, m3/h

Effect of Gas Temperature and Barometric Pressure Changes on Gas Input Rate In the field, gas temperature is typically unknown, but is sometimes assumed to be a standard temperature such as 15.0°C (15.6°C or 60°F in North America); some meters have built-in temperature compensation to that temperature. Some gas suppliers also tabulate data for both local barometric pressure and local metering pressure. Add the meter gage pressure to the barometric pressure to get a correct fuel-gas absolute pressure. This methodology is used in ANSI Z Standard 223.1/NFPA Standard 54 (National Fuel Gas Code), sections 11.1.1, 11.1.2, and A.11.1.1, and Table A.11.1.1. Also see the installation codes in the references for methods for using gas meters and for using injector (orifice) size with pressure drop to measure fuel gas flow rate.

Fuel Gas Interchangeability Gas-burning appliances are normally set up for operation at their rated input with fuel gas of specified properties. Because fuel gases vary greatly, other gases cannot be substituted indiscriminately. In most gas appliances, input rate is controlled by establishing a specified gas pressure difference (often referred to as manifold pressure) across one or more precisely sized orifices. Per the Bernoulli equation, gas velocity through the orifice is proportional to the square root of the pressure difference and inversely proportional to the square root of the density. The input rate also depends on the heat content of the gas and orifice size. Simplified to the basics, it is expressed as follows for particular conditions of temperature and barometric pressure: 2 P Q = K  HHV  D o ------mSG

(1)

where Q = input rate, kW K = constant accounting for measurement units, orifice discharge coefficient, and atmospheric conditions HHV = gas higher heating value, MJ/m3 Do = orifice diameter, mm Pm = manifold pressure, Pa SG = gas specific gravity, dimensionless

If a substitute gas is introduced, only the higher heating value and specific gravity change. The orifice diameter, manifold pressure, and factors accounted for in the constant K do not change. Therefore, the new input is directly proportional to the gas higher heating value and inversely proportional to the square root of the specific gravity:  HHV  Q2 2 ------ =  ---------------- Q1  SG 2 

 HHV  1   ----------------  SG 1 

(2)

where subscripts 1 and 2 indicate the original and substitute gases, respectively. The ratio of higher heating value to the square root of the specific gravity has been named the Wobbe index W. The units of the Wobbe index are the same as those of the heating value because the specific gravity is dimensionless. Substituting for Wobbe index,

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31.10

2020 ASHRAE Handbook—HVAC Systems and Equipment (SI) HHV ------------- = W SG

(3)

Q2 W2 ------ = ------Q1 W1

(4)

W Q2 = Q1 ------2W1

(5)

In other words, when one gas is substituted for another in an appliance and no other changes are made, the input rate changes in proportion to their Wobbe indices. Preservation of input rate does not necessarily ensure proper operation of an appliance, however. Ignition and burning characteristics can differ significantly for gases that have the same Wobbe index. To ensure safe, efficient operation, appliance manufacturers typically limit the range of gases that may be used. Converting an appliance for use with a gas substantially different from the gas it was originally set up for requires changing the gashandling components and/or adjustments. Typical natural gas and commercial propane, for example, have Wobbe indices of about 50 and 76 MJ/m3, respectively. Based on that difference, substituting propane in an unaltered natural gas appliance results in overfiring of the appliance by about 53%, leading to appliance overheating and high production of soot and carbon monoxide. To accommodate the change, it is necessary to change the gas orifice size and, usually, the gas pressure regulator setting to achieve the same gas input rate. Acceptable performance may require additional appliance modifications to avoid other problems (e.g., resonance, poor flame carryover between ganged burners) that manufacturers must identify and resolve. The components necessary for conversion are normally provided by the appliance manufacturer, along with instructions for their installation and checkout of appliance operation. Natural gas is increasingly being transported across oceans as liquefied natural gas (LNG), shipped at high pressure and low temperature in specially designed ships. It is regasified in facilities at the destination, then distributed by utilities in conventional pipelines and service piping. Depending on the characteristics of a utility’s existing gas supply, LNG can differ significantly in its mix of various hydrocarbons, inert gases, etc., and its burning characteristics may be of concern relative to the existing gas. Depending on the extent of the difference, a utility may mix it with other components to improve its compatibility with the existing appliance load. The Wobbe index is useful in evaluating the need for such accommodation.

Altitude When gas-fired appliances are operated at altitudes substantially above sea level, three notable effects occur: • Oxygen available for combustion is reduced in proportion to the atmospheric pressure reduction • With gaseous fuels, the heat of combustion per unit volume of fuel gas (gas heat content) is reduced because of reduced fuel gas density in proportion to the atmospheric pressure reduction • Reduced air density affects the performance and operating temperature of heat exchangers and appliance cooling mechanisms In addition to reducing the gas heat content of fuel gas, reduced fuel gas density also causes increased gas velocity through flow metering orifices. The net effect is for gas input rate to decrease naturally with increases in altitude, but at less than the rate at which atmospheric oxygen decreases. This effect is one reason that derating is required when appliances are operated at altitudes significantly above sea level. Early research by American Gas Association Laboratories with draft hood-equipped appliances established that appliance input rates should be reduced at the rate of 4% per 305 m above sea level, for altitudes higher than 610 m above sea level (Figure 9).

Fig. 9 Altitude Effects on Gas Combustion Appliances

Experience with recently developed appliances having fanassisted combustion systems demonstrated that the 4% rule may not apply in all cases. It is therefore important to consult the manufacturer’s listed appliance installation instructions, which are based on both how the combustion system operates and other factors such as impaired heat transfer. Note also that manufacturers of appliances having tracking-type burner systems may not require derating at altitudes above 610 m. In those systems, fuel gas and combustion airflow are affected in the same proportion by density reduction. In terms of end use, it is important for the appliance specifier to be aware that the heating capacity of appliances is substantially reduced at altitudes significantly above sea level. To ensure adequate delivery of heat, derating of heating capacity must also be considered and quantified. By definition, fuel gas HHV value remains constant for all altitudes because it is based on standard conditions of 101.325 kPa and 15.0°C (288.15 K). Some fuel gas suppliers at high altitudes (e.g., at Denver, Colorado, at 1525 m) may report fuel gas heat content at local barometric pressure instead of standard pressure. Local gas heat content can be calculated using the following equation: B HC = HHV  ----Ps where HC = local gas heat content at local barometric pressure and standard temperature conditions, MJ/m3 HHV = gas higher heating value at standard temperature and pressure of 15.0°C (288.15 K) and 101.325 kPa, respectively, MJ/m3 B = local barometric pressure, kPa (not corrected to sea level: do not use barometric pressure as reported by weather forecasters, because it is corrected to sea level) Ps = standard pressure = 101.325 kPa

For example, at 1525 m, the barometric pressure is 84.316 kPa. If HHV of a fuel gas sample is 37.5 MJ/m3 (at standard temperature and pressure), the local gas heat content would be 31.21 MJ/m3 at 84.316 kPa barometric pressure 1525 m above sea level. HC = 37.5 MJ/m3  84.316 kPa/101.325 kPa = 31.21 MJ/m3 Therefore, the local gas heat content of a sample of fuel gas can be expressed as 31.21 MJ/m3 at local barometric pressure of 84.316 kPa and standard temperature or as 37.5 MJ/m3 (HHV). Both gas heat contents are correct, but the application engineer must understand the difference to use each one correctly. As described earlier, the local heat content HC can be used to determine appliance input rate. When gas heat value (either HHV or HC) is used to determine gas input rate, the gas pressure and temperature in the meter must also be considered. Add the gage pressure of gas in the meter to the

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Automatic Fuel-Burning Systems local barometric pressure to calculate the heat content of the gas at the pressure in the meter. The gas temperature in the meter also affects the heat content of the gas in the meter. The gas heat value is directly proportional to the gas pressure and inversely proportional to its absolute temperature in accordance with the perfect gas laws, as illustrated in the following example calculations for gas input rate with either the HHV or the local heat content. Example 2. Calculate the gas input rate for 37.5 MJ/m3 HHV fuel gas, 2.800 m3/h volumetric flow rate of 24°C fuel gas at 84.316 kPa barometer pressure (1525 m altitude) with 1.700 kPa (gage) fuel gas gage pressure in the gas meter. HHV Method:

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Q = 0.2778HHV × VFRs where Q = fuel gas input rate, kW 0.2778=conversion factor, MJ/h to kW HHV= fuel gas higher heating value at standard temperature and pressure, MJ/m3 VFRs=fuel gas volumetric flow rate adjusted to standard temperature and pressure, m3/h = VFR(Ts × P)/(T × Ps) VFR= fuel gas volumetric flow rate at local temperature and pressure conditions, m3/h Ts = standard temperature, 288.15 K (15.0°C + 273.15 K) P = gas meter absolute pressure, kPa (local barometer pressure + gas pressure in meter relative to barometric pressure) = 84.316 kPa + 1.700 kPa = 86.016 kPa gas meter absolute pressure T = absolute temperature of fuel gas, K (fuel gas temperature in °C + 273.15 K) Ps = standard pressure, 101.325 kPa Substituting given values into the equation for VFRs gives 3

2.800 m  h  288.15 K  86.016 kPa VFRs = ------------------------------------------------------------------------------------------- = 2.305 m3/h  24.00C + 273.15 K 101.325 kPa Then, Q = 0.2778  37.5 MJ/m3  2.305 m3/h = 24.012 kW Local Gas Heat Content Method: The local gas heat content is simply the HHV adjusted to local gas meter pressure and temperature conditions. The gas input rate is simply the observed volumetric gas flow rate times the local gas heat content. Q = 0.2778HC × VFR where Q = gas input rate, kW 0.2778=conversion factor, MJ/h to kW HC = gas heat content at local gas meter pressure and temperature conditions, MJ/m3 = HHV(Ts × P)/(T × Ps) VFR = fuel gas volumetric flow rate, referenced to local gas meter pressure and temperature conditions, m3/h Ts = standard temperature, 288.15 K (15.00°C + 273.15 K) P = gas meter absolute pressure, kPa (local barometer pressure + gas pressure in gas meter relative to barometric pressure) Ps = standard pressure = 101.325 kPa B = local barometric pressure = 84.316 kPa T = absolute temperature of fuel gas, 297.15 K (24.00°C fuel gas temperature + 273.15 K) Substituting given values into the equation for HC gives 37.5  288.15  84.316 + 1.700  HC = --------------------------------------------------------------------------- = 30.87 MJ/m3 297.15  101.325 Then, Q =0.2778 × 30.87 MJ/m3 × 2.800 m3/h = 24.012 kW The gas input rate is exactly the same for both calculation methods.

31.11

3.

OIL-BURNING APPLIANCES

An oil burner is a mechanical device for preparing fuel oil to combine with air under controlled conditions for combustion. Fuel oil is atomized at a controlled flow rate. Air for combustion is generally supplied with a forced-draft fan, although natural or mechanically induced draft can be used. Ignition is typically provided by an electric spark, although gas pilot flames, oil pilot flames, and hotsurface igniters may also be used. Oil burners operate from automatic temperature- or pressure-sensing controls. Oil burners may be classified by application, type of atomizer, or firing rate. They can be divided into two major groups: residential and commercial/industrial. Further distinction is made based on design and operation; different types include pressure atomizing, air or steam atomizing, rotary, vaporizing, and mechanical atomizing. Unvented, portable kerosene heaters are not classified as residential oil burners or as oil heat appliances.

3.1

RESIDENTIAL OIL BURNERS

Residential oil burners ordinarily consume fuel at a rate of 0.5 to 3.7 mL/s, corresponding to input rates of about 20 to 150 kW. However, burners up to 3.7 mL/s (about 300 kW input) sometimes fall in the residential classification because of basic similarities in controls and standards. (Burners with a capacity of 7.4 mL/s and above are classified as commercial/industrial.) No. 2 fuel oil is generally used, although burners in the residential size range can also operate on No. 1 fuel oil. Burners in the 0.5 to 2.5 mL/s range are used not only for boilers and furnaces for space heating, but also for separate tank-type residential water heaters, infrared heaters, space heaters, and other commercial appliances. Central heating appliances include warm-air furnaces and steam or hot water boilers. Oil-burning furnaces and boilers operate essentially the same way as their gas counterparts. NFPA Standard 31 prescribes correct installation practices for oil-burning appliances. Steam or hot-water boilers are available in cast iron and steel. In addition to supplying space heating, many boilers are designed to provide hot water using tankless integral or external heat exchangers. Residential boilers designed to operate as direct-vent appliances are available. Over 95% of residential burners manufactured today are highpressure atomizing gun burners with retention-type heads (Figure 10). This type of burner supplies oil to the atomizing nozzle at pressures that range from 700 to 2100 kPa. A fan supplies air for combustion, and generally an inlet damper regulates the air supply at the burner. A high-voltage electric spark ignites the fuel by either constant ignition (on when the burner motor is on) or interrupted ignition (on only to start combustion). Typically, these burners fire into a combustion chamber in which draft is maintained. Increasingly, however, these burners are being fired into applications (including direct-vent and high-efficiency condensing applications) that have a low level of positive combustion chamber pressure. Modern retention-head oil burners use 3450 rpm motors, and their fans achieve a maximum static pressure of about 750 Pa. Older residential oil burners, which can still be found in the field, used lower-speed motors (1750 rpm) and achieved a maximum static pressure of about 250 to 375 Pa. With higher fan static pressure, burner airflow is less sensitive to variations in appliance draft conditions. Higher-static-pressure burners operate better in applications with positive combustion chamber pressure. In many cases, these burners can be operated without a flue draft regulator. Traditionally, oil burners are operated with a fuel pressure of 700 kPa, and burners’ nozzles are all rated for firing rate and spray angle at this pressure. Increasingly higher pressures (900 to 1050 kPa) are specified by manufacturers for some burners and applications, to achieve better atomization and combustion perfor-

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31.12

2020 ASHRAE Handbook—HVAC Systems and Equipment (SI)

mance (Figure 11). Nozzle fuel firing rate increases in proportion to the square root of the fuel pressure. Oil nozzle line heaters are used in some applications. These are very small electric heaters located adjacent to the nozzle adapter and typically integrated with the burner. These include positivetemperature-coefficient heaters that self-regulate to achieve a fuel temperature in the 50 to 65°C range. Heating fuel in this way improves atomization quality and reduces fuel firing rate. Electric fuel solenoid valves may be installed in the fuel line between the fuel pump and the nozzle and may be mounted directly on the fuel pump. These valves serve to provide sharp fuel flow starts and stops during cyclic operation and augment the function of the fuel pump and pressure-operated flow control valves. In addition, solenoid valves can be used to achieve pre- and post-purge operation in which the burner motor runs to provide purge airflow through the burner and appliance. The post purge can be useful in reducing nozzle temperatures after shutdown and preventing odors indoors, particularly with non-direct-vent applications. Pressure-operated valves integrated with oil burner nozzle assemblies are also available. These valves also provide positive fuel cutoff after burner shutdown and avoid after-drip caused by heat transfer from the still-hot combustion chamber to the nozzle assembly and accumulation of fuel vapor pressure in the nozzle assembly and fuel line. When used, these nozzle valves require the fuel pump discharge pressure to be increased, using the integral pump pressure regulator to compensate for the added pressure drop of the valve.

The following designs are still in operation but are not a significant part of the residential market: • The low-pressure atomizing gun burner differs from the highpressure type in that it uses air at a low pressure to atomize oil. • The pressure atomizing induced-draft burner uses the same type of oil pump, nozzle, and ignition system as the high-pressure atomizing gun burner. • Vaporizing burners are designed for use with No. 1 fuel oil. Fuel is ignited electrically or by manual pilot. • Rotary burners are usually of the vertical wall flame type. Beyond the pressure atomizing burner, considerable developmental effort has been put into advanced technologies such as air atomization, fuel prevaporization, and pulsed fuel flow. Some of these are now used commercially in Europe. Reported benefits include reduced emissions, increased efficiency, and the ability to quickly vary the firing rate. See Locklin and Hazard (1980) for a historical review of technology for residential oil burners. Somewhat more recent developments are described in the Proceedings of the Oil Heat Technology Conferences and Workshops (McDonald 1989-2003). In Europe, there is increasing use of low-NOx , blue flame burners for residential and commercial applications. Peak flame temperatures (and thermal NOx emissions) are reduced by using high flame zone recirculation rates. In a typical boiler, a conventional yellow flame retention head oil burner has a NOx emission of 90 to 110 ppm. A blue flame burner emits about 60 ppm. The blue flame burners require flame sensors for the safety control, which responds to the fluctuating light emitted. Blue flame burners are somewhat more expensive than conventional yellow flame burners. See Butcher et al. (1994) for additional discussion of NOx and small oil burners.

3.2

Fig. 10 High-Pressure Atomizing Gun Oil Burner

COMMERCIAL/INDUSTRIAL OIL BURNERS

Commercial and industrial oil burners are designed for use with distillate or residual grades of fuel oil. With slight modifications, burners designed for residual grades can use distillate fuel oils. The commercial/industrial burners covered here have atomizers, which inject the fuel oil into the combustion space in a fine, conical spray with the apex at the atomizer. The burner also forces combustion air into the oil spray, causing an intimate and turbulent mixing of air and oil. Applied for a predetermined time, an electrical spark, a spark-ignited gas, or an oil igniter ignites the mixture, and sustained combustion takes place. Safety controls are used to shut down the burner upon failure to ignite. All of these burners can almost completely burn the fuel oil without visible smoke when they are operated with excess air as low as 20% (approximately 12% CO2 in the flue gases). Atomizing oil burners are generally classified according to the method used for atomizing the oil, such as pressure atomizing, return-flow pressure atomizing, air atomizing, rotary cup atomizing, steam atomizing, mechanical atomizing, or return-flow mechanical atomizing. Descriptions of these burners are given in the following sections, together with usual capacities and applications. Table 1 lists approximate size range, fuel grade, and usual applications. All burners described are available as gas/oil (dual-fuel) burners.

Pressure-Atomizing Oil Burners

Fig. 11 Details of High-Pressure Atomizing Oil Burner

This type of burner is used in most installations where No. 2 fuel oil is burned. The oil is pumped at pressures of 700 to 2100 kPa through a suitable burner nozzle orifice that breaks it into a fine mist and swirls it into the combustion space as a cone-shaped spray. Combustion air from a fan is forced through the burner air-handling parts surrounding the oil nozzle and is directed into the oil spray. For smaller-capacity burners, ignition is usually started by an electric spark applied near the discharge of the burner nozzle. For

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31.13

Table 1 Classification of Atomizing Oil Burners Type of Oil Burner

Heat Range, MW

Flow Volume, mL/s

Pressure-atomizing

0.02 to 2

0.5 to 50

1 and above

25 and above

0.02 to 0.3

0.5 to 7

No. 2 and heavier

0.2 to 10

5 to 30

3.5 and above 3.5 and above

85 and above 85 and above

No. 2 for small sizes No. 4, 5, or 6 for larger sizes No. 2 and heavier No. 2 and heavier

13 to 50

310 to 1260

No. 2 and heavier

Return-flow pressure-atomizing or modulating pressure-atomizing Air-atomizing Horizontal rotary cup Steam-atomizing (register-type) Mechanical atomizing (register-type)

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Return-flow mechanical atomizing

burner capacities above 20 mL/s, a spark-ignited gas or an oil igniter is used. Pressure-atomizing burners are designated commercially as forced-draft, natural-draft, or induced-draft burners. The forced-draft burner has a fan and motor with capacity to supply all combustion air to the combustion chamber or furnace at a pressure high enough to force the gases through the heat-exchange equipment without the assistance of an induced-draft fan or a chimney draft. Mixing of the fuel and air is such that a minimum of refractory material is required in the combustion space or furnace to support combustion. The natural-draft burner requires a draft in the combustion space. Burner range, or variation in burning rate, is changed by simultaneously varying the oil pressure to the burner nozzle and regulating the airflow by a damper. This range is limited to about 1.6 to 1 for any given nozzle orifice. Burner firing mode controls for various capacity burners differ among manufacturers. Usually, larger burners have controls that provide variable heat inputs. If burner capacity is up to 16 mL/s, a staged control is typically used; if it is up to 25 mL/s, a modulation control is typically used. In both cases, the low burning rate is about 60% of the full-load capacity of the burner. For pressure-atomizing burners, no preheating is required for burning No. 2 oil. No. 4 oil must be preheated to about 38°C for proper burning. When properly adjusted, these burners operate well with less than 20% excess air (approximately 12% CO2); no visible smoke (approximately No. 2 smoke spot number, as determined by ASTM Standard D2156); and only a trace of carbon monoxide in the flue gas in commercial applications. In these applications, the regulation of combustion airflow is typically based on smoke level or flue gas monitoring using analyzers for CO, O2, or CO2. Burners with lower firing rates used to power appliances, residential heating boilers, or warm-air furnaces are usually set up to operate to about 50% excess air (approximately 10% CO2). Good operation of these burners calls for (1) a relatively constant draft (either in the furnace or at the breeching connection, depending on the burner selected), (2) clean burner components, and (3) good-quality fuel oil that complies with the appropriate specifications.

Return-Flow Pressure-Atomizing Oil Burners This burner is a modification of the pressure-atomizing burner; it is also called a modulating-pressure atomizer. It has the advantage of a wide load range for any given atomizer: about 3 to 1 turndown (or variation in load) as compared to 1.6 to 1 for the straight pressure atomizing burner. This wide range is accomplished by means of a return-flow nozzle, which has an atomizing swirl chamber just ahead of the orifice. Good atomization throughout the load range is attained by maintaining a high rate of oil flow and high pressure drop through the

Fuel Grade

Usual Application

No. 2 (less than 25 mL/s)

Boilers Warm-air furnaces Appliances Boilers Warm-air furnaces Boilers Warm-air furnaces Boilers Large warm-air furnaces Boilers Boilers Industrial furnaces Boilers

No. 4 (greater than 25 mL/s) No. 2 and heavier

swirl chamber. The excess oil above the load demand is returned from the swirl chamber to the oil storage tank or to the suction of the oil pump. The burning rate is controlled by varying oil pressure in both the oil inlet and oil return lines. Except for the atomizer, load range, and method of control, the information given for the straight pressureatomizing burner applies to this burner as well.

Air-Atomizing Oil Burners Except for the nozzle, this burner is similar in construction to the pressure-atomizing burner. Atomizing air and oil are supplied to individual parts within the nozzle. The nozzle design allows the oil to break up into small droplet form as a result of the shear forces created by the atomizing air. The atomized oil is carried from the nozzle through the outlet orifice by the airflow into the furnace. The main combustion air from a draft fan is forced through the burner throat and mixes intimately with the oil spray inside the combustion space. The burner igniter is similar to that used on pressure-atomizing burners. This burner is well suited for heavy fuel oils, including No. 6, and has a wide load range, or turndown, without changing nozzles. Turndown of 3 to 1 for the smaller sizes and about 6 or 8 to 1 for the larger sizes may be expected. Load range is varied by simultaneously varying the oil pressure, the atomizing air pressure, and the combustion air entering the burner. Some designs use relatively low atomizing air pressure (35 kPa and lower); other designs use air pressures up to 520 kPa. The burner uses from 16 to 58 L of compressed air per litre of fuel oil (on an air-free basis). Because of its wide load range, this burner operates well on modulating control. No preheating is required for No. 2 fuel oil. The heavier grades of oil must be preheated to maintain proper viscosity for atomization. When properly adjusted, these burners operate well with less than 15 to 25% excess air (approximately 14 to 12% CO2, respectively, at full load); no visible smoke (approximately No. 2 smoke spot number); and only a trace of carbon monoxide in the flue gas.

Horizontal Rotary Cup Oil Burners This burner atomizes the oil by spinning it in a thin film from a horizontal rotating cup and injecting high-velocity primary air into the oil film through an annular nozzle that surrounds the rim of the atomizing cup. The atomizing cup and frequently the primary air fan are mounted on a horizontal main shaft that is motor-driven and rotates at constant speed (58 to 100 r/s) depending on the size and make of the burner. The oil is fed to the atomizing cup at controlled rates from an oil pump that is usually driven from the main shaft through a worm and gear.

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A separately mounted fan forces secondary air through the burner windbox. Secondary air should not be introduced by natural draft. The oil is ignited by a spark-ignited gas or an oil-burning igniter (pilot). The load range or turndown for this burner is about 4 to 1, making it well suited for operation with modulating control. Automatic combustion controls are electrically operated. When properly adjusted, these burners operate well with 20 to 25% excess air (approximately 12.5 to 12% CO2, respectively, at full load), no visible smoke (approximately No. 2 smoke spot number), and only a trace of carbon monoxide in the flue gas. This burner is available from several manufacturers as a package comprising burner, primary air fan, secondary air fan with separate motor, fuel oil pump, motor, motor starter, ignition system (including transformer), automatic combustion controls, flame safety equipment, and control panel. Good operation of these burners requires relatively constant draft in the combustion space. The main assembly of the burner, with motor, main shaft, primary air fan, and oil pump, is arranged for mounting on the boiler front and is hinged so that the assembly can be swung away from the firing position for easy access. Rotary burners require some refractory in the combustion space to help support combustion. This refractory may be in the form of throat cones or combustion chamber liners.

Steam-Atomizing Oil Burners (Register Type) Atomization is accomplished in this burner by the impact and expansion of steam. Oil and steam flow in separate channels through the burner gun to the burner nozzle. There, they mix before discharging through an orifice, or series of orifices, into the combustion chamber. Combustion air, supplied by a forced-draft fan, passes through the directing vanes of the burner register, through the burner throat, and into the combustion space. The vanes give the air a spinning motion, and the burner throat directs it into the cone-shaped oil spray, where intimate mixing of air and oil takes place. Full-load oil pressure at the burner inlet is generally some 700 to 1050 kPa, and the steam pressure is usually kept higher than the oil pressure by about 170 kPa. Load range is accomplished by varying these pressures. Some designs operate with oil pressure ranging from 1050 kPa at full load to 70 kPa at minimum load, resulting in a turndown of about 8 to 1. This wide load range makes the steam atomizing burner suited to modulating control. Some manufacturers provide dual atomizers within a single register so that one can be cleaned without dropping load. Depending on the burner design, steam atomizing burners use from 0.1 to 0.6 kg of steam to atomize a litre of oil. This corresponds to 0.5 to 3.0% of the steam generated by the boiler. Where no steam is available for start-up, compressed air from the plant air supply may be used for atomizing. Some designs allow use of a pressure atomizing nozzle tip for start-up when neither steam nor compressed air is available. This burner is used mainly on water-tube boilers, which generate steam at 1050 kPa or higher and at capacities above 3500 kW input. Oils heavier than grade No. 2 must be preheated to the proper viscosity for good atomization. When properly adjusted, these burners operate well with 15% excess air (14% CO2) at full load, without visible smoke (approximately No. 2 smoke spot number), and with only a trace of carbon monoxide in the flue gas.

Mechanical Atomizing Oil Burners (Register Type) Mechanical atomizing, as generally used, describes a technique synonymous with pressure atomizing. Both terms designate atomization of the oil by forcing it at high pressure through a suitable stationary atomizer. The mechanical atomizing burner has a windbox, which is a chamber into which a fan delivers combustion and excess air for dis-

tribution to the burner. The windbox has an assembly of adjustable internal air vanes called an air register. Usually, the fan is mounted separately and connected to the windbox by a duct. Oil pressure of 600 to 6000 kPa is used, and load range is obtained by varying the pressure between these limits. The operating range or turndown for any given atomizer can be as high as 3 to 1. Because of its limited load range, this type of burner is seldom selected for new installations.

Return-Flow Mechanical Atomizing Oil Burners This burner is a modification of a mechanical atomizing burner; atomization is accomplished by oil pressure alone. Load ranges up to 6 or 8 to 1 are obtained on a single-burner nozzle by varying the oil pressure between 700 and 7000 kPa. The burner was developed for use in large installations such as on ships and in electric generating stations where wide load range is required and water loss from the system makes use of atomizing steam undesirable. It is also used for firing large hot-water boilers. Compressed air is too expensive for atomizing oil in large burners. This is a register burner similar to the mechanical atomizing burner. Wide range is possible by using a return-flow nozzle, which has a swirl chamber just ahead of the orifice or sprayer plate. Good atomization is attained by maintaining a high rate of oil flow and a high pressure drop through the swirl chamber. Excess oil above the load demand is returned from the swirl chamber to the oil storage tank or to the oil pump suction. Control of burning rate is accomplished by varying the oil pressure in both the oil inlet and the oil return lines.

3.3

DUAL-FUEL GAS/OIL BURNERS

Dual-fuel, combination gas/oil burners are forced-draft burners that incorporate, in a single assembly, the features of the commercial/ industrial-grade gas and oil burners described in the preceding sections. These burners have controls to ensure that the burner flame relay or programmer cycles the burner through post- and prepurge before starting again on the other fuel. Burner manufacturers for larger boilers design the special mechanical linkages needed to deliver the correct air/fuel ratios at full fire, low fire, or any intermediate rate. Smaller burners may have straight on/off firing. Larger burners may have low-fire starts on both fuels and use a common flame scanner. Smaller dual-fuel burners usually include pressure atomization of the oil. Air atomization systems are included in large oil burners. Dual-fuel burners often have automatic changeover controls that respond to outdoor temperature. A special temperature control, located outdoors and electrically interlocked with the dual-fuel burner control system, senses outdoor temperature. When the outdoor temperature drops to the outdoor control set point, the control changes fuels automatically after putting the burner through postand prepurge cycles. A manual fuel selection switch can be retained as a manual override on the automatic feature. These control systems require special design and are generally provided by the burner manufacturer. The dual-fuel burner is fitted with a gas train and oil piping that is connected to a two-pipe oil system following the principles of the preceding sections. An oil reserve must be maintained at all times for automatic fuel changeover. Boiler flue chimney connectors are equipped with special doubleswing barometric draft regulators or, if required, sequential furnace draft control to operate an automatic flue damper. Dual-fuel burners and their accessories should be installed by experienced contractors to ensure satisfactory operation.

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Table 2 Guide for Fuel Oil Grades Versus Firing Rate Maximum Heat Input of Unit, kW

Volume Flow Rate, mL/s

Fuel Grade

Up to 1000 1000 to 2000 2000 to 4400 Over 4400

Up to 25 25 to 50 50 to 100 Over 100

2 2, 4, 5 5, 6 6

3.4

EQUIPMENT SELECTION

Economic and practical factors (e.g., the degree of operating supervision required by the installation) generally dictate the selection of fuel oil based on the maximum heat input of the oil-burning appliance. For heating loads and where only one oil-burning appliance is operated at any given time, the relationship is as shown in Table 2 (which is only a guide). In many cases, a detailed analysis of operating parameters results in the burning of lighter grades of fuel oil at capacities far above those indicated. Fig. 12 Typical Oil Storage Tank (No. 6 Oil)

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Fuel Oil Storage Systems All fuel-oil storage tanks should be constructed and installed in accordance with NFPA Standard 31 and with local ordinances. Storage Capacity. Dependable and economical operation of oilburning appliances requires ample and safe storage of fuel oil at the site. Design responsibility should include analysis of specific storage requirements as follows: • Rate of oil consumption. • Dependability of oil deliveries. • Economical delivery lots. The cost of installing larger storage capacity should be balanced against the savings indicated by accommodating larger delivery lots. Truck lots and railcar lots vary with various suppliers, but the quantities are approximated as follows: Small truck lots in metropolitan area 1900 to 7600 L Normal truck lots 7600 to 19 000 L Transport truck lots 19 000 to 34 000 L Rail tanker lots 30 300 to 45 400 L Tank Size and Location. Standard oil storage tanks range in size from 200 to 190 000 L and larger. Tanks are usually built of steel; concrete construction may be used only for heavy oil. Unenclosed tanks located in the lowest story, cellar, or basement should not exceed 2500 L capacity each, and the aggregate capacity of such tanks should not exceed 5000 L unless each 2500 L tank is insulated in an approved fireproof room having a fire resistance rating of at least 2 h. If storage capacity at a given location exceeds about 3800 L, storage tanks should be underground and accessible for truck or rail delivery with gravity flow from the delivering carrier into storage. If the oil is to be burned in a central plant such as a boiler house, the storage tanks should be located, if possible, so that the oil burner pump (or pumps) can pump directly from storage to the burners. For year-round operation, except for storage or supply capacities below 7600 L, at least two tanks should be installed to facilitate tank inspection, cleaning, repairs, and clearing of plugged suction lines. When the main oil storage tank is not close enough to the oilburning appliances for the burner pumps to take suction from storage, a supply tank must be installed near the oil-burning appliances and oil must be pumped periodically from storage to the supply tank by a transport pump at the storage location. Supply tanks should be treated the same as storage tanks regarding location within buildings, tank design, etc. On large installations, it is recommended that standby pumps be installed as a protection against heat loss in case of pump failure. Because all piping connections to underground tanks must be at the top, such tanks should not be more than 3.2 m from top to bottom

to avoid pump suction difficulties. (This dimension may have to be less for installations at high altitudes.) At sea level, the total suction pressure for the oil pump must not exceed 42 kPa. Connections to Storage Tank. All piping connections for tanks over 1.0 m3 capacity should be through the top of the tank. Figure 12 shows a typical arrangement for a cylindrical storage tank with a heating coil as required for No. 5 or 6 fuel oils. The heating coil and oil suction lines should be located near one end of the tank. The maximum allowable steam pressure in such a heating coil is 100 kPa. The heating coil is unnecessary for oils lighter than No. 5 unless a combination of high pour point and low outdoor temperature makes heating necessary. A watertight access port with internal ladder provides access to the inside of the tank. If the tank is equipped with an internal heating coil, a second access port is required to permit withdrawal of the coil. The fill line should be vertical and should discharge near the end of the tank away from the oil suction line. The inlet of the fill line must be outside the building and accessible to the oil delivery vehicle unless an oil transfer pump is used to fill the tank. When possible, the inlet of the fill line should be at or near grade where filling may be accomplished by gravity. For gravity filling, the fill line should be at least 50 mm in diameter for No. 2 oil and 150 mm in diameter for No. 4, 5, or 6 oils. Where filling is done by pump, the fill line for No. 4, 5, or 6 oils may be 100 mm in diameter. An oil return line bringing recirculated oil from the burner line to the tank should discharge near the oil suction line inlet. Each storage tank should be equipped with a vent line sized and arranged in accordance with NFPA Standard 31. Each storage tank must have a device for determining oil level. For tanks inside buildings, the gaging device should be designed and installed so that oil or vapor will not discharge into a building from the fuel supply system. No storage tank should be equipped with a glass gage or any gage that, if broken, would allow oil to escape from the tank. Gaging by a measuring stick is permissible for outdoor or underground tanks.

Fuel-Handling Systems The fuel-handling system consists of the pumps, valves, and fittings for moving fuel oil from the delivery truck or car into the storage tanks and from the storage tanks to the oil burners. Depending on the type and arrangement of the oil-burning appliances and the grade of fuel oil burned, fuel-handling systems vary from simple to quite complicated arrangements. The simplest handling system would apply to a single burner and small storage tank for No. 2 fuel oil, similar to a residential heating

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installation. The storage tank is filled through a hose from the oil delivery truck, and the fuel-handling system consists of a supply pipe between the storage tank and the burner pump. Equipment should be installed on light-oil tanks to indicate visibly or audibly when the tank is full; on heavy-oil tanks, a remote-reading liquidlevel gage should be installed. Figure 13 shows a complex oil supply arrangement for two burners on one oil-burning appliance. For an appliance with a single burner, the change in piping is obvious. For a system with two or more appliances, the oil line downstream of the oil discharge strainer becomes a main supply header, and the branch supply line to each appliance includes a flowmeter, automatic control valve, etc. For light oils requiring no heating, all oil-heating equipment shown in Figure 13 would be omitted. Both a suction and a return line should be used, except for gravity flow in residential installations. Oil pumps (steam or electrically driven) should deliver oil at the maximum rate required by the burners (this includes the maximum firing rate, the oil required for recirculating, plus a 10% margin). The calculated suction pressure at the entrance of any burner pump should not exceed 35 kPa for installations at sea level. At higher elevations, the suction pressure should be reduced in direct proportion to the reduction in barometric pressure. Oil temperature at the pump inlet should not exceed 50°C. Where oil burners with integral oil pumps (and oil heaters) are used and suction lift from the storage tank is within the capacity of the burner pump, each burner may take oil directly from the storage tank through an individual suction line unless No. 6 oil is used. Where two or more tanks are used, the piping arrangement into the top of each tank should be the same as for a single tank so that any tank may be used at any time; any tank can be inspected, cleaned, or repaired while the system is in operation. The length of suction line between storage tank and burner pumps should not exceed 30 m. If the main storage tank(s) are located more than 30 m from the pumps, a supply tank should be installed near the pumps, and a transfer pump should be installed at the storage tanks for delivery of oil to the supply tank.

Central oil distribution systems comprising a central storage facility, distribution pumps or provision for gravity delivery, distribution piping, and individual fuel meters are used for residential communities, notably mobile home parks. Provisions of NFPA Standard 31, Installation of Oil Burning Equipment, should be followed in installing a central oil distribution system.

Fuel Oil Preparation System Fuel oil preparation systems consist of oil heater, oil temperature controls, strainers, and associated valves and piping required to maintain fuel oil at the temperatures necessary to control the oil viscosity, facilitate oil flow and burning, and remove suspended matter. Preparation of fuel oil for handling and burning requires heating the oil if it is No. 5 or 6. This decreases its viscosity so it flows properly through the oil system piping and can be atomized by the oil burner. No. 4 oil occasionally requires heating to facilitate burning. No. 2 oil requires heating only under unusual conditions. For handling residual oil from the delivering carrier into storage tanks, the viscosity should be about 156 mm2/s. For satisfactory pumping, viscosity of oil surrounding the inlet of the suction pipe must be 444 mm2/s or lower; for oil with a high pour point, the temperature of the entire oil content of the tank must be above the pour point. Storage tank heaters are usually made of pipe coils or grids using steam or hot water at or less than 100 kPa as the heating medium. Electric heaters are sometimes used. For control of viscosity for pumping, the heated oil surrounds the oil suction line inlet. For heating oils with high pour points, the heater should extend the entire length of the tank. All heaters have suitable thermostatic controls. In some cases, storage tanks may be heated satisfactorily by returning or recirculating some of the oil to the tank after it has passed through heaters located between the oil pump and oil burner. Heaters to regulate viscosity at the burners are installed between the oil pumps and the burners. When required for small packaged burners, the heaters are either assembled integrally with the individ-

Fig. 13 Industrial Burner Auxiliary Equipment

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Automatic Fuel-Burning Systems ual burners or mounted separately. The heat source may be electricity, steam, or hot water. For larger installations, the heater is mounted separately and is often arranged in combination with central oil pumps, forming a central oil pumping and heating set. The separate or central oil pumping and heating set is recommended for installations that burn heavy oils, have a periodic load demand, and require continuous circulation of hot oil during down periods. Another system of oil heating to maintain pumping viscosity that is occasionally used for small- or medium-sized installations consists of an electrically heated section of oil piping. Low-voltage current is passed through the pipe section, which is isolated by nonconducting flanges. The oil-heating capacity for any given installation should be approximately 10% greater than the maximum oil flow. Maximum oil flow is the maximum oil-burning rate plus the rate of oil recirculation. Controls for oil heaters must be dependable to ensure proper oil atomization and avoid overheating of oil, which results in coke deposits inside the heaters. In steam or electric heating, include an interlock with a solenoid valve or switch to shut off the steam or electricity. When the oil pump is not operating, the oil in the heater can become overheated and deposit carbon. Overheating also can be a problem with oil heaters using high-temperature hot water. Provisions must be made to avoid overheating the oil when the oil pump is not operating. Oil heaters with low- or medium-temperature hot water are not generally subject to coke deposits. Where steam or hot water is used in oil heaters located after the oil pumps, the pressure of the steam or water in the heaters is usually lower than the oil pressure. Consequently, heater leakage between oil and steam causes oil to flow into the water or condensing steam. To prevent oil from entering the boilers, discard the condensed steam or the water from such heaters from the system, or provide special equipment for oil removal. Hot-water oil heaters of double-tube-and-shell construction with inert heat transfer oil and a sight glass between the tubes are available. With this type of heater, oil leaks through an oil-side tube appear in the sight glass, and repairs can be made to the oil-side tube before a water-side tube leaks. This discussion of oil-burning equipment also applies to oilfired boilers and furnaces (see Chapters 32 and 33).

4.

SOLID-FUEL-BURNING APPLIANCES

A mechanical stoker is a device that feeds a solid fuel into a combustion chamber. It supplies air for burning the fuel under automatic control and, in some cases, incorporates automatic ash and refuse removal.

31.17 4.1

CAPACITY CLASSIFICATION OF STOKERS

Stokers are classified according to their coal-feeding rates. Although some residential applications still use stokers, their main application is in commercial and industrial areas. The U.S. Department of Commerce, in cooperation with the Stoker Manufacturers Association, use the following classification: Class 1: Capacity less than 27 kg of coal per hour Class 2: Capacity 27 to less than 45 kg of coal per hour Class 3: Capacity 45 to less than 136 kg of coal per hour Class 4: Capacity 136 to less than 544 kg of coal per hour Class 5: Capacity 544 kg of coal per hour and over Class 1 stokers are used primarily for residential heating and are designed for quiet, automatic operation. These stokers are usually underfeed types and are similar to those shown in Figure 14, except that they are usually screw-feed. Class 1 stokers feed coal to the furnace intermittently, in accordance with temperature or pressure demands. A special control is needed to ensure stoker operation to maintain a fire during periods when no heat is required. Class 2 and 3 stokers are usually of the screw-feed type, without auxiliary plungers or other means of distributing the coal. They are used extensively for heating plants in apartment buildings, hotels, and industrial plants. They are of the underfeed type and are available in both the hopper and bin-feed type. These stokers are also built in a plunger-feed type with an electric motor, steam, or hydraulic cylinder coal-feed drive. Class 2 and 3 stokers are available for burning all types of anthracite, bituminous, and lignite coals. The tuyere and retort design varies according to the fuel and load conditions. Stationary grates are used on bituminous models, and clinkers formed from the ash accumulate on the grates surrounding the retort. Class 2 and 3 anthracite stokers are equipped with moving grates that discharge ash into a pit below the grate. This ash pit may be located on one side or both sides of the grate and, in some installations, is big enough to hold the ash for several weeks of operation. Class 4 stokers vary in details of design, and several methods of feeding coal are practiced. Underfeed stokers are widely used, although overfeed types are used in the larger sizes. Bin-feed and hopper models are available in underfeed and overfeed types. Class 5 stokers are spreader, underfeed, chain or traveling grate, and vibrating grate. Various subcategories reflect the type of grate and method of ash discharge.

4.2

STOKER TYPES BY FUEL-FEED METHODS

Class 5 stokers are classified according to the method of feeding fuel to the furnace: (1) spreader, (2) underfeed, (3) chain or traveling grate, and (4) vibrating grate. The type of stoker used in a given installation depends on the general system design, capacity required, and type of fuel burned. In general, the spreader stoker is the most widely used in the capacity range of 9.5 to 50 kg/s because it

Fig. 14 Horizontal Underfeed Stoker with Single Retort

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2020 ASHRAE Handbook—HVAC Systems and Equipment (SI) Table 3 Characteristics of Various Types of Stokers (Class 5)

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Stoker Type and Subclass Spreader Stationary and dumping grate Traveling grate Vibrating grate

Typical Capacity Range, kg/s

Maximum Burning Rate, kW/m2

Characteristics Capable of burning a wide range of coals; best to follow fluctuating loads; high fly ash carryover; low-load smoke

2.5 to 10 13 to 50 2.5 to 13

1400 2400 1300

Underfeed Single- or double-retort Multiple-retort

2.5 to 4

1300

4 to 63

1900

Chain grate and traveling grate

2.5 to 13

1600

Low maintenance; low fly ash carryover; capable of burning a wide variety of weakly caking coals; smokeless operation over entire range

Vibrating grate

0.2 to 19

1300

Characteristics similar to chain and traveling grate stokers, except that these stokers have no difficulty in burning strongly caking coals

Capable of burning caking coals and a wide range of coals (including anthracite); high maintenance; low fly ash carryover; suitable for continuous-load operation

Grates for spreader stokers may be of several types. All grates are designed with high airflow resistance to avoid formation of blowholes through the thin fuel bed. Early designs were simple stationary grates from which ash was removed manually. Later designs allowed intermittent dumping of the grate either manually or by a power cylinder. Both types of dumping grates are frequently used for small and medium-sized boilers (see Table 3). Also, both types are sectionalized, and there is a separate undergrate air chamber for each grate section and a grate section for each spreader stoker. Consequently, both the air supply and the fuel supply to one section can be temporarily discontinued for cleaning and maintenance without affecting operation of other sections of the stoker. For high-efficiency operation, a continuous ash-discharging grate, such as the traveling grate, is necessary. Introduction of the spreader stoker with the traveling grate increased burning rates by about 70% over the stationary and dumping grate types. Although both reciprocating and vibrating continuous ash discharge grates have been developed, the traveling grate stoker is preferred because of its higher burning rates.

Fig. 15 Spreader Stoker, Traveling Grate Type responds quickly to load changes and can burn a wide variety of coals. Underfeed stokers are mainly used with small industrial boilers of less than 3.8 kg/s. In the intermediate range, the large underfeed stokers, as well as the chain and traveling grate stokers, are being displaced by spreader and vibrating grate stokers. Table 3 summarizes the major features of the different stokers.

Spreader Stokers Spreader stokers use a combination of suspension burning and grate burning. As shown in Figure 15, coal is continually projected into the furnace above an ignited fuel bed. The coal fines are partially burned in suspension. Large particles fall to the grate and are burned in a thin, fast-burning fuel bed. Because this firing method provides extreme responsiveness to load fluctuations and because ignition is almost instantaneous on increased firing rate, the spreader stoker is favored over other stokers in many industrial applications. The spreader stoker is designed to burn about 50% of the fuel in suspension. Thus, it generates much higher particulate loadings than other types of stokers and requires dust collectors to trap particulate material in the flue gas before discharge to the stack. To minimize carbon loss, fly carbon reinjection systems are sometimes used to return particles into the furnace for complete burnout. Because this process increases furnace dust emissions, it can be used only with highly efficient dust collectors.

Fuels and Fuel Bed. All spreader stokers (particularly those with traveling grates) can use fuels with a wide range of burning characteristics, including caking tendencies, because the rapid surface heating of the coal in suspension destroys the caking tendency. High-moisture, free-burning bituminous and lignite coals are commonly burned; coke breeze can be burned in mixture with a high-volatile coal. However, anthracite, because of its low volatile content, is not a suitable fuel for spreader stoker firing. Ideally, the fuel bed of a coal-fired spreader stoker is 50 to 100 mm thick. Burning Rates. The maximum heat release rates range from 1250 kW/m2 (a coal consumption of approximately 5 g/s) on stationary, dumping, and vibrating grate designs to 2400 kW/ m2 on traveling grate spreader stokers. Higher heat release rates are practical with some waste fuels in which a greater portion of fuel can be burned in suspension than is possible with coal.

Underfeed Stokers Underfeed stokers introduce raw coal into a retort beneath the burning fuel bed. They are classified as horizontal or gravity feed. In the horizontal type, coal travels within the furnace in a retort parallel with the floor; in the gravity-feed type, the retort is inclined by 25°. Most horizontal-feed stokers are designed with single or double retorts (and, rarely, triple retorts), whereas gravity-feed stokers are designed with multiple retorts. In the horizontal stoker (see Figure 14), coal is fed to the retort by a screw (for smaller stokers) or a ram (for larger stokers). Once the retort is filled, the coal is forced upward and spills over the retort to form and feed the fuel bed. Air is supplied through tuyeres at each

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Fig. 17 Vibrating Grate Stoker

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Fig. 16

Chain Grate Stoker

side of the retort and through air ports in the side grates. Over-fire air provides additional combustion air to the flame zone directly above the bed to prevent smoking, especially at low loads. Gravity-feed stokers are similar in operating principle. These stokers consist of sloping multiple retorts and have rear ash discharge. Coal is fed into each retort, where it is moved slowly to the rear while simultaneously being forced upward over the retorts. Fuels and Fuel Bed. Either type of underfeed stoker can burn a wide range of coal, although the horizontal type is better suited for free-burning bituminous coal. These stokers can burn caking coal, if there is not an excess amount of fines. The ash-softening temperature is an important factor in selecting coals because the possibility of excessive clinkering increases at lower ash-softening temperatures. Because combustion occurs in the fuel bed, underfeed stokers respond slowly to load change. Fuel-bed thickness is extremely nonuniform, ranging from 200 to over 600 mm. The fuel bed often contains large fissures separating masses of coke. Burning Rates. Single-retort or double-retort horizontal stokers are generally used to service boilers with capacities up to 3.8 kg/s. These stokers are designed for heat release rates of 1300 kW/ m2.

Chain and Traveling Grate Stokers Figure 16 shows a typical chain or traveling grate stoker. These stokers are often used interchangeably because they are fundamentally the same, except for grate construction. The essential difference is that the links of chain grate stokers are assembled so that they move with a scissors-like action at the return bend of the stoker, whereas in most traveling grates there is no relative movement between adjacent grate sections. Accordingly, the chain grate is more suitable for handling coal with clinkering ash characteristics than is the traveling grate stoker. Operation of the two types is similar. Coal, fed from a hopper onto the moving grate, enters the furnace after passing under an adjustable gate that regulates the thickness of the fuel bed. The layer of coal on the grate entering the furnace is heated by radiation from the furnace gases or from a hot refractory arch. As volatile matter is driven off by this rapid radiative heating, ignition occurs. The fuel continues to burn as it moves along the fuel bed, and the layer becomes progressively thinner. At the far end of the grate, where the coal completes combustion, ash is discharged into the pit as the grates pass downward over a return bend. Often, furnace arches (front and/or rear) are included with these stokers to improve combustion by reflecting heat to the fuel bed. The front arch also serves as a bluff body, mixing rich streams of volatile gases with air to reduce unburned hydrocarbons. A chain grate stoker with overfire air jets eliminates the need for a front arch for burning volatiles. As shown in Figure 17, the stoker was zoned, or sectionalized, and equipped with individual zone dampers to

control the pressure and quantity of air delivered to the various sections. Fuels and Fuel Bed. The chain grate and traveling grate stokers can burn a variety of fuels (e.g., peat, lignite, subbituminous coal, free-burning bituminous coal, anthracite coal, coke), as long as the fuel is sized properly. However, strongly caking bituminous coals have a tendency to mat and prevent proper air distribution to the fuel bed. Also, a bed of strongly caking coal may not be responsive to rapidly changing loads. Fuel bed thickness varies with the type and size of the coal burned. For bituminous coal, a 125 to 180 mm bed is common; for small-sized anthracite, the fuel bed is reduced to 75 to 125 mm. Burning Rates. Chain and traveling grate stokers are offered for a maximum continuous burning rate of 1100 to 1600 kW/m2, depending on the type of fuel and its ash and moisture content.

Vibrating Grate Stokers The vibrating grate stoker, as shown in Figure 17, is similar to the chain grate stoker in that both are overfeed, mass-burning, continuous ash discharge stokers. However, in the vibrating stoker, the sloping grate is supported on equally spaced vertical plates that oscillate back and forth in a rectilinear direction, causing the fuel to move from the hopper through an adjustable gate into the active combustion zone. Air is supplied to the stoker through laterally exposed areas beneath the stoker formed by the individual flexing of the grate support plates. Ash is automatically discharged into a shallow or basement ash pit. The grates are water cooled and are connected to the boiler circulating system. The rates of coal feed and fuel bed movement are determined by the frequency and duration of the vibrating cycles and regulated by automatic combustion controls that proportion the air supply to optimize heat release rates. Typically, the grate is vibrated about every 90 s for durations of 2 to 3 s, but this depends on the type of coal and boiler operation. The vibrating grate stoker is increasingly popular because of its simplicity, inherently low fly ash carryover, low maintenance, wide turndown (10 to 1), and adaptability to multiple-fuel firing. Fuels and Fuel Bed. The water-cooled vibrating grate stoker is suitable for burning a wide range of bituminous and lignite coals. The gentle agitation and compaction of the vibratory actions allow coal having a high free-swelling index to be burned and a uniform fuel bed without blowholes and thin spots to be maintained. The uniformity of air distribution and fuel bed conditions produce both good response to load swings and smokeless operation over the entire load range. Fly ash emission is probably greater than from the traveling grate because of the slight intermittent agitation of the fuel bed. The fuel bed is similar to that of a traveling grate stoker. Burning Rates. Burning rates of vibrating grate stokers vary with the type of fuel used. In general, however, the maximum heat release rate should not exceed 1300 kW/m2 (a coal use of approximately 5 g/s) to minimize fly ash carryover.

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31.20

2020 ASHRAE Handbook—HVAC Systems and Equipment (SI) Ignition and Flame Monitoring

Fig. 18

Basic Control Circuit for Fuel-Burning Appliance

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5. CONTROLS This section covers controls required for automatic fuel-burning systems. Chapter 7 of the 2017 ASHRAE Handbook—Fundamentals addresses basic automatic control. Automatic fuel-burning appliances require control systems that supervise combustion and take proper corrective action in the event of a failure in the appliance or related installation equipment. Requirements are similar for gas, oil, and solid fuel (stoker) burners. Controls can be classified as safety or operating. Safety controls monitor potentially hazardous operating conditions such as ignition, combustion, temperature, and pressure, and function as required to ensure safe operation. Operating controls handle appliance operation at the required input rate when heat is required. Figure 18 illustrates the basic elements of a control system for a fuel-fired appliance. In this diagram, the limit control and the ignition safety control, including its sensor, are safety controls.

5.1

SAFETY CONTROLS AND INTERLOCKS

Safety controls protect against hazards related to the combustion process. Personnel and material near the appliance must be protected against explosion, fire, or excessive temperature. Common safety control functions include the following: • Ignition and flame monitoring 32. Ignition proving (e.g., gas pilot thermocouple, flame rectification) 33. Proof of flame • Draft proof (e.g., sail switch, pressure switch) • Limit controls 34. Excessive temperature of heated medium (e.g., air or water) 35. Excessive pressure of heated medium (e.g., steam or hot water) 36. Low water level in steam or hot-water boilers 37. Low fuel-oil pressure 38. Low or excessive fuel-oil temperature when burning heavy oils 39. Low atomizing air pressure or atomizing steam pressure 40. Low fuel-gas pressure (e.g., empty propane tank) • Flue gas spillage switch • Flame rollout detection • Carbon monoxide sensors • Similar devices to prevent hazards from abnormal conditions • Safety-related mechanical malfunction controls (e.g., open furnace blower door, burner out of position or rotary cup burner motor failure) • Field interlocks having safety function

Failure to ignite fuel or failure of an established flame can cause explosion, if unburned fuel is subsequently provided with an ignition source. Ignition and flame safeguard controls, properly applied and used, prevent these occurrences. Some ignition controls monitor existence of an ignition source such as a pilot burner, allowing introduction and continued flow of fuel only when the ignition source is proven. Others use devices that ignite the main flame directly, in conjunction with high-speed flame detection capabilities and rapid shutdown of the fuel supply, if ignition is not immediate or if the flame is lost. Ignition controls often include programming features that govern the entire sequence of operation when a burner is started, operated, and stopped. In residential gas-fired appliances, burners are usually ignited by a small standing gas pilot flame, by an intermittent pilot system, or by a direct ignition system. Standing pilot systems prove existence of the ignition source by sensing its heat with a thermocouple. In intermittent pilot systems, a spark or hot surface igniter lights the pilot burner when there is a call for heat, thus saving pilot burner fuel when no heat is needed. Flame rectification (the ability of a flame to change an alternating current to direct current) is commonly used to prove the existence of the pilot flame (in intermittent pilot systems) or main burner flame (in direct-ignition systems). In direct-ignition systems, a spark or hot surface device ignites the main burner directly, eliminating the need for a pilot burner. If flame is not immediately established during ignition or if an established flame is lost, the gas supply system is quickly shut off. Flame rollout (flame escaping the appliance combustion chamber) protection detects flame in appliance combustion air inlet areas by sensing high temperatures and shuts off the fuel supply. Most residential oil-burners use high-energy sparking for ignition. Sparking may be continuous (present for the entire duration of burner operation) or intermittent (present only long enough to establish presence of the flame). Most residential oil burners use an oil primary control that operates the burner motor and ignition spark transformer in conjunction with a flame detector. Flame proof is typically achieved by sensing visible or ultraviolet light emitted by the flame. A cadmium cell is used to detect visible light: its resistance changes when exposed to light from a burner flame, and that change is interpreted and acted upon by the primary control. Commercial appliances with high fuel input rates normally use some form of a proven ignition source, often a standing or intermittent pilot burner, which itself may be quite large. North American safety standards do not allow use of direct-ignition systems for appliances with input greater than 117 kW. Standing pilots or a proved igniter system must be used. The latter may closely resemble the direct-ignition system, except that the existence of the ignition source must be proven before fuel gas is allowed to flow to the main burner. As in the direct-ignition system, main burner gas flow is shut down if flame is not established quickly or if it goes out after ignition. Flame proving in large appliances is typically by flame rectification or ultraviolet light sensing.

Draft Proving Fuel-burning appliances often include forced- or induced-draft blowers to move combustion air and combustion products through the appliance and venting systems. Sometimes blowers are applied outside the appliance in the venting system. Draft-proving controls supervise proper operation of these components to allow gas flow, ignition, and combustion only if appliance draft is adequate, typically using pressure and/or flow sensors to verify sufficient draft.

Limit Controls In a warm-air furnace or space heater, excessive air temperature must be prevented. In a hot-water boiler, water temperature and pres-

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Automatic Fuel-Burning Systems sure must not exceed the ratings of external piping and terminal equipment. Steam pressure must not exceed ratings of pressure vessels and steam system equipment. Fuel oil must be provided at pressures and temperatures that allow proper operation of oil burners. These requirements and restrictions are the focus of limit controls required by installation codes and appliance standards, and applied by appliance manufacturers and field installers. Design and principle of operation vary greatly, but in all cases limit controls must provide reliable detection of a fault and shut down the fuel-burning appliance.

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Other Safety Controls Requirements for safety devices are a function of the particular appliance and installation circumstances, and must be considered in that context. Appliances listed or design-certified as complying with recognized safety standards generally have appropriate safety devices as required by those standards. Building codes typically require appliances to be listed or design-certified by recognized testing laboratories, and that appliances not listed or design-certified comply with requirements within the code. Field circumstances may impose a need for control interlocks that relate to safety. Mechanical-draft venting systems, for example, must be proven to be in operation before an appliance ignition attempt and during burner operation. An appropriate field-installed interlock to prove mechanical-draft venting system operation before an ignition attempt would be considered a safety control.

Prescriptive Requirements for Safety Controls Industrial safety codes and insurers often require that fuel-fired appliances be provided with specific components and construction features related to safety. Codes and requirements such as ASME Standard CSD-1 for boilers or those of insurance organizations may apply. Prescriptive governmental requirements may also apply to some types of installations. Often, requirements are based on the type of appliance, its capacity or size, or the intended use, such as type of building occupancy.

Reliability of Safety Controls Devices suitable for use as safety controls must meet special quality and reliability requirements. Electrical contacts, for example, must be able to switch the intended load through high cycle counts under extreme environmental conditions. Products designed for safety control application are invariably listed, or design-certified as complying with recognized standards that impose a wide range of construction and testing requirements. The degree of reliability of safety controls must be extremely high, and normally far exceeds that required of operating controls to meet appliance safety standards.

5.2

31.21 temperature or pressure sensor. When the temperature or pressure falls to an on set point, the control initiates combustion, which proceeds until the temperature or pressure increases to an off set point. The difference between the off and on values is called the differential. On/off control is satisfactory for systems with high thermal inertia (i.e., those in which heat input does not rapidly change the controlled temperature or pressure). A staged control operates an appliance at multiple fixed input rates in response to heat demand. The input rates are determined by burner system hardware and settings. A residential gas-fired appliance, for example, may include a gas pressure regulator capable of providing high and low gas pressures to the gas orifice(s). Applying electrical power to a solenoid associated with the pressure regulator enables a second orifice pressure and thus a second gas input rate. In large appliances such as commercial boilers, staging is often accomplished with multiple burner assemblies, operated singly or together as necessary to match the load. As in on/off control, combustion initiation and changes between high and low input rates are in response to demand, as determined by the sensor and control logic. Staged control logic is sometimes provided within an appliance control to enhance load response of the appliance. For example, a heating appliance control can provide staged burner operation without a staged room thermostat by beginning burner operation on burner stage 1 when a single-stage thermostat calls for heat. If burner stage 1 does not satisfy the thermostat within a specified time period, the appliance control turns on burner stage 2 until the thermostat is satisfied, at which time the burner is completely shut off until the thermostat’s next call for heat. Figure 19 illustrates the characteristics of a three-stage control system that provides heat in response to a temperature sensor and is required to maintain temperature at a set point tS. If the controlled temperature is above tS, no heat is provided. When the temperature drops to t1On, the first-stage burner is operated; if the temperature then rises back to tS, the burner is turned off. If the first stage is inadequate, the temperature continues to fall; when it reaches t2On, the second stage is operated. A third stage is available if the second stage is not adequate, operating when the temperature falls to its on temperature t3On. Successful handling of the load results in raising of the controlled temperature and stepped sequential reduction of the heat input as the off temperatures t3Off , t2Off , and t1Off (= tS) are reached. The individual stages operate subject to a stage differential, which is the difference between the temperature at which a stage is shut off and the (lower) temperature at which it is again turned on. Interstage differential and

OPERATING CONTROLS

Operating controls start and stop burner operation in response to demands of the application load, and often incorporate capabilities for load matching and other application-related features. In contrast to safety controls, their purpose is to satisfy the application’s fundamental purpose (e.g., warm air, hot water, steam pressure). Related functions may involve operating ancillary components of the appliance or system, such as blowers or pumps, which must operate at the appropriate speed. Load matching is the most significant differentiating feature of operating controls. The major categories are on/off, staged, and modulated control. Fuel input rate varies for the latter two modes, but combustion airflow rate may also be controlled. An on/off control starts and stops the flow of fuel to the burner to satisfy heat demand. An appliance with on/off control can operate only at its rated input regardless of the rate at which combustion heat may be needed. In a typical system, demand is sensed by a

Fig. 19

Control Characteristics of Three-Stage System

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31.22

2020 ASHRAE Handbook—HVAC Systems and Equipment (SI)

overall differential, as shown in Figure 19, are analogous. The on and off temperatures may or may not overlap, depending on the relative size of the stage differentials. In sophisticated electronic controls, the differential settings are adjustable to meet the particular requirements of the application. Staged control logic often includes features that enhance load response or appliance operation. As shown in Figure 19, load matching might be best if the control initiates the lowest input rate on a call for heat and proceeds to higher inputs only if the demand cannot be met at lower rates. Appliance or application characteristics may dictate other approaches. For example, if appliance ignition is most reliable at high input rate, a staged control might require that ignition begin at a high input rate with an immediate reduction to a lower input rate to meet the actual heat demand. Combustion airflow may or may not be staged. If not, operation at lower input rates, particularly in appliances with fan-assisted combustion, results in lean operation (a high percentage of excess air). There may be little effect, but in general, heat transfer efficiency suffers if the combustion air-to-fuel ratio (A/F) is not maintained at an optimum level on each stage. A modulating control regulates input rate to follow load demands more closely than on/off or staged controls. Fuel input rate may be varied by throttling (mechanical restriction of flow by a valve) or by pressure regulation to control injector (orifice) pressure. Demand is often indicated by a temperature or pressure sensor, and the input rate increases as the temperature or pressure falls from set point values. In large fuel-oil appliances, a fuel oil control valve and burner damper respond over a range of positions within the operating range of the burner. Air/fuel proportioning controls are a variation of modulating controls in which the proper ratio of air to fuel is maintained throughout the burner’s fuel input range. In response to the need for heat, both fuel input rate and combustion airflow rate are modulated, maintaining an optimal air/fuel ratio to maximize efficiency. The controls may change blower speed or (particularly in large commercial appliances) link air dampers to fuel input rate control. Tracking controls are proportioning controls that maintain the proper air/fuel ratio by applying basic Bernoulli-equation fluid flow behavior. Gas-fired appliances are especially suitable for this approach. As combustion air or gas flow rate changes, such as through metering orifices or venturis, pressure changes in proportion to the square of the air or gas velocity. The pressure differential created by forcing air through a venturi, for example, can be used to regulate gas flow through an orifice. When the relationship is established, a change of airflow rate results in a pressure differential change that causes a proportional (tracking) change in gas flow rate. In one approach, airflow rate is varied according to heating demand by changing blower speed or adjusting a damper position, and gas flow rate tracks the change. Variations may use intermediate ratio controllers. Other systems control gas input rate, and combustion air rate tracks. In the ideal case, the air-to-fuel ratio remains constant at all input rates, and efficiency increases substantially when firing rate is reduced. Combustion quality is monitored and controlled in some large installations to ensure efficient operation or satisfactory emission characteristics. Air or fuel can be controlled in response to devices that sense concentrations of various combustion product constituents. Butcher (1990) used the measured intensity of light emitted from flames of fixed-input oil burners to judge basic flame quality. Flame steady-state intensity is compared to a predetermined flame intensity set point. Deviation of flame intensity beyond an optimum intensity range indicates poor flame quality. Operating controls often provide functions not directly related to combustion. These appliance controls include things such as operating air-circulating blowers or pumps according to the fuel combustion system’s needs. A blower or pump can be operated in response

to a temperature or pressure sensor or according to a time protocol imposed with electronic or digital capabilities. Additional algorithms in the appliance control can further enhance appliance operation.

Integrated and Programmed Controls Integrated Controls. Many appliance controls combine safety and operating functions into electronic microprocessors with various sensors and electromechanical devices to handle most of the control functions necessary to operate fuel-burning appliances. Figure 20 illustrates an integrated control approach typical of residential forced-air furnaces or boilers with fan-assisted combustion and a limit control. Other safety devices, such as a flame roll-out switch, typically are present. For a forced-air furnace, the remote thermostat would be a wall thermostat. For a hot-water boiler, the operating control might be a wall thermostat, possibly with a second sensor to control boiler water temperature. Ignition and flame safety algorithms are provided in the microprocessor-based control. The ignition device and flame sensor are connected to the control. Note that the control operates both the circulating blower or pump and a combustion blower. Both are controlled in accordance with an appropriate sequence of operation. A draft-proving device is connected to the control to confirm operation of the combustion blower prior to igniter operation and opening of the gas valve, and during subsequent burner operation. Typically, light-emitting diodes (LEDs) are provided on the control to indicate operating status and to facilitate diagnosis of operating problems. Programmed controls for ignition and flame safeguard have been used in large fuel-fired appliances for many years. More recently, programming has been applied in integrated controls for smaller appliances. These controls handle not only the ignition sequence, but also operating functions, and provide status and diagnostic information. Accomplishing those functions with separate components would be difficult and clumsy. When power is supplied, an integrated control conducts a selfcheck to ensure that the appliance and control system are in safe working order, and then typically indicates readiness by flashing one or more LEDs. The self-check may include verifying that the flame-proving device does not indicate flame before the ignition sequence has started and the draft-proving device does not indicate draft before the combustion blower has started. Self-checking for safe conditions may continue during burner operation and during standby after the thermostat is satisfied. If unacceptable operation is detected, the control will take corrective action or attempt safe shutoff of the appliance. On a call for heat, an operating sequence typically proceeds as follows:

Fig. 20 Integrated Control System for Gas-Fired Appliance

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Automatic Fuel-Burning Systems 1. The call for heat is indicated by an assigned LED flashing sequence. 2. The combustion blower starts and its operation is verified by the draft-proving device. 3. When the combustion blower has run long enough to purge the combustion chamber of residual fuel and combustion products (i.e., flush it with four air changes), the ignition device energizes. 4. The control allows enough time for the ignition device to become operative. In some applications, operation is verified electronically. 5. The gas valve opens. 6. Flame detection checks for main burner ignition. If flame presence is not proven within a very short predetermined time, or if continuous proof of flame fails at any time while the gas valve is open, the gas valve closes. Programming typically requires rapid detection of flame failure, and gas shutoff in accordance with safety standards. 7. If flame presence is detected, the air circulating blower or hotwater pump is started. Depending on the application, blower or pump operation could start with burner operation or after a time delay. 8. When the call for heat is satisfied, the gas valve closes to extinguish the burner flame. 9. The combustion blower operates for a short period to flush the combustion chamber with air (post purge). 10. The circulating blower or pump stops, usually after an applicationspecific delay. 11. The status LED(s) return to the standby mode. This sequence can be varied as necessary to accommodate the requirements of specific appliances and applications. For example, staged or modulated operation may be included. Status LEDs are turned on and off in particular patterns to display coded diagnostic information for many different normal and abnormal conditions (e.g., failure to prove draft, failure to detect flame, overtemperature). Appliance standards and installation codes require that appliance controls lock out operation of the burner system in response to certain failure conditions, when manual intervention by a qualified service agency is necessary.

REFERENCES ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal.ashrae .org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org/bookstore. Ackerman, M.Y., J.D. Dale, D.J. Wilson, and N.P. Fleming. 1995. Design guidelines for combustion air systems in cold climates (RP-735). ASHRAE Research Project, Final Report. AHRI. 2015. Directory of certified product performance. Air Conditioning, Heating and Refrigeration Institute, Arlington, VA. www.ahridirectory.org. ANSI. 2012. Installation of domestic gas conversion burners. Standard Z21.81994 (R2012). American National Standards Institute, Washington, D.C. ANSI. 2000. Draft hoods. Standards Z21.12-1990 (R2000), Z21.12a-1993 (R2000), and Z21.12b-1994 (R2000). Secretariat: CSA-America, Cleveland, OH. American National Standards Institute, Washington, D.C. ANSI. 2014. Gas-fired low pressure steam and hot water boilers. Standard Z21.13-2014/CSA 4.9-2014. Secretariat: CSA-America, Cleveland, OH. American National Standards Institute, Washington, D.C. ANSI. 2014. Domestic gas conversion burners. Standard Z21.17-98 (R2014)/ CSA 2.7-M98. Secretariat: CSA-America, Cleveland, OH. American National Standards Institute, Washington, D.C.

31.23 ANSI. 2012. Gas-fired central furnaces. ANSI Standard Z21.47-2012/CSA 2.3-2012. Secretariat: CSA-America, Cleveland, OH. American National Standards Institute, Washington, D.C. ASHRAE. 2007. Method of testing for annual fuel utilization efficiency of residential central furnaces and boilers. Standard 103-2007. ASHRAE. 2012. Method of testing for rating commercial gas, electric, and oil service water heating equipment. ANSI/ASHRAE Standard 118.12012. ASHRAE. 2015. Method of testing for rating residential water heaters. ANSI/ ASHRAE Standard 118.2-2006 (RA2015). ASHRAE. 2007. Methods of testing for rating combination space-heating and water-heating appliances. Standard 124-2007. ASME. 2012. Controls and safety devices for automatically fired boilers. Standard CSD-1-2012. American Society of Mechanical Engineers, New York. ASTM. 2013. Test method for smoke density in flue gases from burning distillate fuels. ANSI/ASTM Standard D2156-2009 (R2013). American Society for Testing and Materials, West Conshohocken, PA. Butcher, T. 1990. Performance control strategies for oil-fired residential heating systems. BNL Report 52250. Brookhaven National Laboratory, Upton, NY. Butcher, T., L.A. Fisher, B. Kamath, T. Kirchstetter, and J. Batey. 1994. Nitrogen oxides (NOx) and oil burners. BNL Report 52430. Brookhaven National Laboratory, Upton, NY. Code of Federal Regulations. [Annual.] Energy conservation program for consumer products; Energy and water conservation standards and their compliance dates. Code of Federal Regulations 10 CFR 430.32. U.S. Government Publishing Office, Washington, D.C. CSA. 2015. Natural gas and propane installation code. Standard CAN/CSA B149.1-15. Canadian Standards Association, Mississauga, ON. Dale, J.D., D.J. Wilson, M.Y. Ackerman, and N.P. Fleming. 1997. A field study of combustion air systems in cold climates (RP-735). ASHRAE Transactions 103(1):910-920. Paper PH-97-13-2. Locklin, D.W., and H.R. Hazard. 1980. Technology for the development of high-efficiency oil-fired residential heating equipment. BNL Report 51325. Brookhaven National Laboratory, Upton, NY. McDonald, R.J., ed. 1989-2003. Proceedings of the Oil Heat Technology Conferences and Workshops. BNL Reports 52217, 52284, 52340, 52392, 52430, 52475, 52506, 52537, 52670, and 71337. Brookhaven National Laboratory, Upton, NY. U.S. Congress. 1987. National Appliance Energy Conservation Act (NAECA). Public Law 100-12. www.govtrack.us/congress/bills/100/s83/text. NFPA. 2011. Installation of oil-burning equipment. ANSI/NFPA Standard 31-2011. National Fire Protection Association, Quincy, MA. NFPA. 2012. National fuel gas code. ANSI/NFPA Standard 54-2012. American Gas Association, Washington, D.C., and National Fire Protection Association, Quincy, MA. UCS. 2007. Clean energy: How biomass energy works. Union of Concerned Scientists, Cambridge, MA. www.ucsusa.org/clean_energy / renewable_energy_basics/offmen-how-biomass-energy-works.html.

BIBLIOGRAPHY ANSI. 2016. Gas unit heaters and gas-fired duct furnaces. Standard Z83.82016/CSA 2.6-2016 (Historical). Secretariat: CSA-America, Cleveland, OH. American National Standards Institute, Washington, D.C. Segeler, C.G., ed. 1965. Gas engineers handbook, Section 12, Ch. 2. Industrial Press, New York. UL. 2017. Oil burners. ANSI/UL Standard 296-11. Underwriters Laboratories, Northbrook, IL. UL. 1995. Oil-fired boiler assemblies. Standard 726-7. Underwriters Laboratories, Northbrook, IL. UL. 2018. Oil-fired central furnaces. Standard 727-10. Underwriters Laboratories, Northbrook, IL. UL. 2016. Commercial-industrial gas-heating equipment. Standard 795-8. Underwriters Laboratories, Northbrook, IL.

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BOILERS Classifications.......................................................................... 32.1 Selection Parameters ................................................................ 32.5 Efficiency: Input and Output Ratings ....................................... 32.6 Performance Codes and Standards .......................................... 32.7

Sizing ......................................................................................... 32.7 Burner Types ............................................................................. 32.7 Boiler Controls......................................................................... 32.7 Flame Safeguard Controls ........................................................ 32.8

B

boilers. Similarly, operating and safety controls and relief valves are required. Steam boilers are generally available in standard sizes from 17 kW to 30 MW, many of which are used for space heating applications in both new and existing systems. On larger installations, they may also provide steam for auxiliary uses, such as hot water heat exchangers, absorption cooling, laundry, and sterilizers. In addition, many steam boilers provide steam at various temperatures and pressures for a wide variety of industrial processes. Water boilers are generally available in standard sizes from 10 kW to over 30 MW, many of which are in the low-pressure class and are used primarily for space heating applications in both new and existing systems. Some water boilers may be equipped with either internal or external heat exchangers for domestic water service. Every steam or water boiler is rated for a maximum working pressure that is determined by the applicable boiler code under which it is constructed and tested. When installed, it also must be equipped at a minimum with operation and safety controls and pressure/temperature-relief devices mandated by such codes.

OILERS are pressure vessels designed to transfer heat (produced by combustion) to a fluid. The definition has been expanded to include transfer of heat from electrical resistance elements to the fluid or by direct action of electrodes on the fluid. In most boilers, the fluid is usually water in the form of liquid or steam. If the fluid being heated is air, the heat exchange device is called a furnace, not a boiler. The firebox, or combustion chamber, of some boilers is also called a furnace. Excluding special and unusual fluids, materials, and methods, a boiler is a cast-iron, carbon or stainless steel, aluminum, or copper pressure vessel heat exchanger designed to (1) burn fossil fuels (or use electric current) and (2) transfer the released heat to water (in water boilers) or to water and steam (in steam boilers). Boiler heating surface is the area of fluid-backed surface exposed to the products of combustion, or the fire-side surface. Various manufacturers define allowable heat transfer rates in terms of heating surface based on their specific boiler design and material limitations. Boiler designs provide for connections to a piping system, which delivers heated fluid to the point of use and returns the cooled fluid to the boiler. Chapters 6, 11, 12, 13, and 15 cover applications of heating boilers. Chapter 7 discusses cogeneration, which may require boilers.

1.

CLASSIFICATIONS

Boilers may be grouped into classes based on working pressure and temperature, fuel used, material of construction, type of draft (natural or mechanical), and whether they are condensing or noncondensing. They may also be classified according to shape and size, application (e.g., heating, process), and state of output medium (steam or water). Boiler classifications are important to the specifying engineer because they affect performance, first cost, and space requirements. Excluding designed-to-order boilers, significant class descriptions are given in boiler catalogs or are available from the boiler manufacturer. The following basic classifications may be helpful.

Working Pressure and Temperature With few exceptions, boilers are constructed to meet ASME Boiler and Pressure Vessel Code, Section IV (SCIV), Rules for Construction of Heating Boilers (low-pressure boilers), or Section I (SCI), Rules for Construction of Power Boilers (high-pressure boilers). Low-pressure boilers are constructed for maximum working pressures of 103 kPa (gage) steam and up to 1100 kPa (gage) hot water. Hot-water boilers are limited to 120°C operating temperature. Operating and safety controls and relief valves, which limit temperature and pressure, are ancillary devices required to protect the boiler and prevent operation beyond design limits. High-pressure boilers are designed to operate above 103 kPa (gage) steam, or above 1100 kPa (gage) and/or 120°C for water The preparation of this chapter is assigned to TC 6.1, Hydronic and Steam Equipment and Systems.

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Fuel Used Boilers may be designed to burn coal, wood, various grades of fuel oil, waste oil, or various types of fuel gas, or to operate as electric boilers. A boiler designed for one specific fuel type may not be convertible to another type of fuel. Some boilers can be adapted to burn coal, oil, or gas. Several designs accommodate firing oil or gas, and other designs allow firing dual-fuel-burning equipment. Accommodating various fuel-burning equipment is a fundamental concern of boiler manufacturers, who can furnish details to a specifying engineer. The manufacturer is responsible for performance and rating according to the code or standard for the fuel used (see section on Performance Codes and Standards).

Construction Materials Most noncondensing boilers are made with cast iron sections or steel. Some small boilers are made of copper or copper-clad steel. Condensing boilers are typically made of stainless steel or aluminum because copper, cast iron, and carbon steel will corrode because of acidic condensate. Cast-iron sectional boilers generally are designed according to ASME Section IV requirements and range in size from 10 kW to 4.1 MW gross output. They are constructed of individually cast sections, assembled into blocks (assemblies) of sections. Push or screw nipples, gaskets, and/or an external header join the sections pressure-tight and provide passages for the water, steam, and products of combustion. The number of sections assembled determines the boiler size and energy rating. Sections may be vertical or horizontal, the vertical design being more common (Figures 1A and 1C). The boiler may be dry-base (the combustion chamber is beneath the fluid-backed sections), as in Figure 1B; wet-base (the combus-

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Fig. 1

Residential Boilers

tion chamber is surrounded by fluid-backed sections, except for necessary openings), as in Figure 2A; or wet-leg (the combustion chamber top and sides are enclosed by fluid-backed sections), as in Figure 2B. The three types of boilers can be designed to be equally efficient. Testing and rating standards apply equally to all three types. The wet-base design is easiest to adapt for combustible floor installations. Applicable codes usually demand a floor temperature under the boiler no higher than 50 K above room temperature. A steam boiler at 102°C or a water boiler at 116°C may not meet this requirement without appropriate floor insulation. Large cast-iron boilers are also made as water-tube units with external headers (Figure 2C). Steel boilers generally range in size from 15 kW to the largest boilers made. Designs are constructed to either ASME SCI or SCIV (or other applicable code) requirements. They are fabricated into one assembly of a given size and rating, usually by welding. The heat exchange surface past the combustion chamber is usually an assembly of vertical, horizontal, or slanted tubes. Boilers of the fire-tube design contain flue gases in tubes completely submerged in fluid (Figures 1D and 1E show residential units, and Figures 3A to 3D and Figure 4A show commercial units). Water-tube boilers contain fluid inside tubes with tube pattern arrangement providing for the combustion chamber (Figures 4C and 4D). The internal configuration may accommodate one or more flue gas passes. As with cast-iron sectional boilers, dry-base, wet-leg, or wet-base designs may be used. Most small steel boilers are of the dry-base, vertical fire-tube type (Figure 1D). Larger boilers usually incorporate horizontal or slanted tubes; both fire-tube and water-tube designs are used. A popular horizontal fire-tube design for medium and large steel boilers is the scotch marine, which is characterized by a central fluid-backed cylindrical combustion chamber, surrounded by fire-tubes accommodating two

Fig. 2

Cast-Iron Commercial Boilers

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or more flue gas passes, all within an outer shell (Figures 3A to 3D). In another horizontal fire-tube design, the combustion chamber has a similar central fluid-backed combustion chamber surrounded by fire tubes accommodating two or more flue gas passes, all within an outer shell. However, this design uses a dry base and wet leg (or mud leg) (Figure 4A). Copper boilers are usually some variation of the water-tube boiler. Parallel finned copper tube coils with headers, and serpentine copper tube units are most common (Figures 1F and 1G). Some are offered as wall-hung residential boilers. The commercial bent water-tube design is shown in Figure 4B. Natural gas is the usual fuel for copper boilers. Stainless steel boilers usually are designed to operate with condensing flue gases. Most are single-pass, fire-tube design and are generally resistant to thermal shock. ASME limits operating temperatures to 99°C and 1103 kPa (gage) working pressure. Aluminum boilers are also usually designed to operate with condensing flue gases. Typical designs incorporate either cast aluminum boiler sections or integrally finned aluminum tubing. ASME limits operating temperatures to 93°C and working pressure to 345 kPa (gage).

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Type of Draft Draft is the pressure difference that causes air and/or fuel to flow through a boiler or chimney. A natural draft boiler is designed to operate with a negative pressure in the combustion chamber and in the flue connection. The pressure difference is created by the tendency of hot gases to rise up a chimney or by the height of the boiler up to the draft control device. In a mechanical draft boiler, a fan or blower or other machinery creates the required pressure difference. These boilers may be either forced draft or induced draft. In a forced-draft boiler, air is forced into the combustion chamber to maintain a positive pressure in the combustion chamber and/or the

space between the tubing and the jacket (breaching). In an induceddraft boiler, air is drawn into the combustion chamber to maintain a negative pressure in the combustion chamber.

Condensing or Noncondensing Traditionally designed boilers must operate without condensing the flue gas in the boiler. This precaution was necessary to prevent corrosion of cast-iron, steel, or copper parts. Hot-water units were operated at 60°C minimum water temperature to prevent this corrosion and to reduce the likelihood of thermal shock. Because a higher boiler efficiency can be achieved with a lower water temperature, the condensing boiler allows the flue gas water vapor to condense and drain. Full condensing boilers are now available from a large number of manufacturers. These boilers are specifically designed for operation with the low return water temperatures found in hot-water reset, water-source heat pump, two-pipe fan-coil, and reheat systems. Two types of commercial condensing boilers are shown in Figure 5. Figure 6 shows a typical relationship of overall condensing boiler efficiency to return water temperature. The dew point of 55°C shown in the figure varies with the percentage of hydrogen in the fuel and oxygen/carbon dioxide ratio, or excess air, in the flue gases. A condensing boiler is shown in Figure 1H. Condensing boilers can be of the fire-tube, water-tube, cast-iron, and cast-aluminum sectional design. Condensing boilers are generally provided with high-turndown modulating burners and are more efficient than noncondensing boilers at any return water temperature (RWT), including noncondensingtemperature applications. Efficiencies of noncondensing boilers must be limited to avoid potential condensing and corrosion. Further efficiencies can be gained by using lower RWT or higher t as recommended by ASHRAE. For example, a natural gas condensing boiler operating with 15°C RWT in a water-source heat pump application has potential boiler efficiency in excess of 98% (Figure 6).

Fig. 3 Scotch Marine Commercial Boilers

Fig. 4 Commercial Fire-Tube and Water-Tube Boilers

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Fig. 5 Commercial Condensing Boilers

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Fig. 7 Relationship of Dew Point, Carbon Dioxide, and Combustion Efficiency for Natural Gas condensing flue gases. In the past, typical cast iron, carbon steel, and copper were not suitable materials for the condensing section of a boiler. Certain stainless steels and aluminum alloys were suitable. However, advances in design, controls, and manufacturing have allowed materials such as cast iron to be used where they previously could not be; as with all products, consult the manufacturer for proper application. Commercial boiler installations can be adapted to condensing operation by adding a condensing heat exchanger in the flue gas vent. Heat exchangers in the flue gas venting require a condensing medium such as (1) low pressure steam condensate or hot water return, (2) domestic water service, (3) fresh water makeup, or (4) other fluid sources in the 20 to 55°C range. The medium can also be a source of heat recovery in HVAC systems. Fig. 6

Effect of Inlet Water Temperature on Efficiency of Condensing Boilers

For maximum reliability and durability over extended product life, condensing boilers should be constructed for corrosion resistance throughout the fireside combustion chamber and heat exchanger. Noncondensing heat plant efficiency may in some cases be improved with the use of external flue gas-to-water economizers. The condensing medium may include domestic hot-water (DHW) preheat, steam condensate or hot-water return, fresh-water makeup, or other fluid sources in the 20 to 55°C range. The medium can also be used as a source of heat recovery in the HVAC system. Be sure to protect the noncondensing boiler from the low-temperature water return in the event of economizer service or control failure. Figure 7 shows how dew point varies with a change in the percentages of oxygen/carbon dioxide for natural gas. Boilers that operate with a combustion efficiency and oxygen and carbon dioxide concentrations in the flue gas such that the flue gas temperature falls between the dew point and the dew point plus 80 K should be avoided, unless the venting is designed for condensation. This temperature typically occurs with boilers operating between 83 and 87% efficiency, and the flue gas has an oxygen concentration of 7 to 10% and the carbon dioxide is 6 to 8%. Chapter 35 gives further details on chimneys. The condensing portion of these boilers may require special material or operating techniques to resist the corrosive effects of the

Wall-Hung Boilers Wall-hung boilers are a type of small residential gas-fired boiler developed to conserve space in buildings such as apartments and condominiums. These boilers are popular in Europe. The most common designs are mounted on outer walls. Combustion air enters through a pipe from the outdoors, and flue products are vented directly through another pipe to the outdoors. In some cases, the air intake pipe and vent pipe are concentric. Other designs mount adjacent to a chimney for venting and use indoor air for combustion. These units may be condensing or noncondensing. Because these boilers are typically installed in the living space, provisions for proper venting and combustion air supply are very important.

Integrated (Combination) Boilers Integrated boilers are relatively small, residential boilers that combine space heating and water heating in one appliance. They are usually wall mounted, but may also be floor standing. They operate primarily on natural gas and are practical to install and operate. The most common designs have an additional heat exchanger and a storage tank to provide domestic hot water. Some designs (particularly European) do not have a storage tank. Instead, they use a larger heat exchanger and the appropriate burner input to provide instantaneous domestic hot water.

Electric Boilers Electric boilers are a separate class of boiler. Because no combustion occurs, a boiler heating surface and flue gas venting are

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unnecessary. The heating surface is the surface of the electric elements or electrodes immersed in the boiler water. The design of electric boilers is largely determined by the shape and heat release rate of the electric elements used. Electric boiler manufacturers’ literature describes available size, shapes, voltages, ratings, and methods of control.

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2.

SELECTION PARAMETERS

Boiler selection depends on many variables of the individual application, including operating characteristics of actual loads; load distribution; total heating demand on the boiler plant; number of boilers in the plant; operational characteristics of individual boilers; and the whole boiler, burner, and control package. The load’s operational characteristics drive boiler selection. The purpose of boilers is not to make steam or hot water, but rather to supply heating as required by the process. Space heating can be accomplished with water or low-pressure steam. It is often possible to use lower water temperatures and a very efficient condensing boiler. A extremely wide range of products exist for this market. In industrial or institutional buildings, a process may require hightemperature hot water or high-pressure steam; this is a more specialized market and process loads require careful review. Often, the initial warm-up of a process load is very large and the loads are intermittent. Where a space-heating load is (after start-up) relatively constant, process loads can cause sudden increases and decreases on heating demand. High-pressure steam or high-temperature hot water allow for economical distribution on large systems. A substantial part of the first cost of a new system is often distribution piping; there also are consequences on the operational cost to distribute the load. It may be appropriate to use a system with a high energy density for larger systems. Trade-offs of this approach include additional equipment to support a high-temperature and -pressure system, as well as lower system efficiency. The boiler plant should be sized for the maximum system load. This is not merely the sum of connected loads, but should also take in to account piping loss, warm-up loads, possible diversity, etc. The plant also must operate at the minimum load, so the boiler should have sufficient turndown so that the boiler does not cycle excessively. It also must be able to increase or decrease its output with the load. If the load has a short burst of high instantaneous demand, a boiler with a high thermal capacity may be appropriate. If the load increases steadily and then remains for a long period, then a boiler with lower thermal mass but a quick response may be preferred. The designer should not only calculate the maximum but also the low load, and estimate the average load. It is sometimes advantageous to split the load among multiple boilers. Splitting the load allows the plant to meet a seasonal high maximum and improves plant turndown in the shoulder season, and can increase efficiency for loads that tend to vary over the course of the year. Another factor to consider when selecting the number of boilers to be installed is redundancy. Redundant boilers allow for a more flexible maintenance schedule and possible future expansion. Also, consider plant layout and physical restrictions on the plant. An important issue is that, once a boiler reaches a certain size or pressure/temperature, many jurisdictions require an operating engineer to be on site at all times. Once the system characteristics are established, the maximum load and general architectural of the plant can be selected. Important performance characteristics for base boiler selection include efficiency, part-load efficiency, and available fuels. The physical boiler should be selected with burner and controls package. None of these elements operate independently of the others; they must operate as a whole. Each component of the boiler system affects

• Emissions • Energy efficiency • Indoor air quality (IAQ) For IAQ purposes, it may be advantageous to have two boiler systems: a low-temperature hot-water heating system on outdoor reset control (60 to 32°C), and a clean-steam boiler for sterilizers and humidifiers. The codes and standards outlined in the section on Performance Codes and Standards include requirements for minimum efficiency, maximum temperature, burner operating characteristics, and safety control. Test agency certification and labeling, which are published in boiler manufacturers’ catalogs and shown on boiler rating plates, are generally sufficient for determining boiler steady-state operating characteristics. Some boilers are not tested and rated by a recognized agency. Nonrated boilers (rated and warranted only by the manufacturer) are used when jurisdictional codes or standards do not require a rating agency label. Almost without exception, both rated and nonrated boilers are of ASME Code construction and are marked accordingly. Boiler selection should be based on a competent review of the following parameters: All Boilers • Application of terminal unit selection • Applicable code under which the boiler is constructed and tested • Gross boiler heat output • Part- versus full-load efficiency (life-cycle cost) • Total heat transfer surface area • Water content mass or volume • Auxiliary power requirement • Cleaning and service access provisions for fireside and waterside heat transfer surfaces • Space requirement and piping arrangement • Water treatment requirement • Operating personnel capabilities and maintenance/operation requirements • Regulatory requirements for emissions, fuel usage/storage Fuel-Fired Boilers • Combustion chamber (furnace volume) • Internal flow pattern of combustion products • Combustion air and venting requirements • Fuel availability/capability Steam Boilers • Steam quality The codes and standards outlined in the section on Performance Codes and Standards include requirements for minimum efficiency, maximum temperature, burner operating characteristics, and safety control. Test agency certification and labeling, which are published in boiler manufacturers’ catalogs and shown on boiler rating plates, are generally sufficient for determining boiler steady-state operating characteristics. However, for noncondensing commercial and industrial boilers, these ratings typically do not consider part-load or seasonal efficiency, which is less than steady-state efficiency. Condensing boilers, generally provided with modulating burners, provide higher part- and full-load efficiency. Some boilers are not tested and rated by a recognized agency, and, therefore, do not bear the label of an agency. Nonrated boilers (rated and warranted only by the manufacturer) are used when jurisdictional codes or standards do not require a rating agency label. As previously indicated, almost without exception, both rated and nonrated boilers are of ASME Code construction and are marked accordingly.

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2020 ASHRAE Handbook—HVAC Systems and Equipment (SI) can fire at reduced inputs while modulating both fuel and air. This increase in efficiency comes from the increase in the ratio of heat exchanger surface area to heat input as the firing rate is reduced.

4.

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Fig. 8

3.

Boiler Efficiency as Function of Fuel and Air Input

EFFICIENCY: INPUT AND OUTPUT RATINGS

The efficiency of fuel-burning boilers is defined in the following ways: combustion, overall, and seasonal. However, manufacturers are not required to test or publish efficiencies that coincide with these industry definitions. Further explanation of boiler test requirements is found in the section on Performance Codes and Standards. Combustion efficiency is input minus stack (flue gas outlet) loss, divided by input, and generally ranges from 75 to 86% for most noncondensing boilers. Condensing boilers generally operate in the range of 88 to 95% combustion efficiency. Overall (or thermal) efficiency is gross energy output divided by energy input. Gross output is measured in the steam or water leaving the boiler and depends on the characteristics of the individual installation. Overall efficiency of electric boilers is generally 92 to 96%. Overall efficiency is lower than combustion efficiency by the percentage of heat lost from the outer surface of the boiler (radiation loss or jacket loss) and by off-cycle energy losses (for applications where the boiler cycles on and off). Overall efficiency can be precisely determined only under controlled laboratory test conditions, directly measuring the fuel input and the heat absorbed by the water or steam of the boiler. Precise efficiency measurements are generally not performed under field conditions because of the inability to control the required parameters and the high cost involved in performing such an analysis. An approximate combustion efficiency for noncondensing boilers can be determined under any operating condition by measuring flue gas temperature and percentage of CO2 or O2 in the flue gas and by consulting a chart or table for the fuel being used. The approximate combustion efficiency of a condensing boiler must include the energy transferred by condensation in the flue gas. Seasonal efficiency is the actual operating efficiency that the boiler will achieve during the heating season at various loads. Because most heating boilers operate at part load, the part-load efficiency, including heat losses when the boiler is off, has a great effect on the seasonal efficiency. The difference in seasonal efficiency between a boiler with an on/off firing rate and one with modulating firing rate can be appreciable if the airflow through the boiler is modulated along with the fuel input. Figure 8 shows how efficiency increases at part load for a typical boiler equipped with a burner that

PERFORMANCE CODES AND STANDARDS

Commercial heating boilers (i.e., boilers with inputs of 90 kW and larger) at present are only tested for full-load steadystate efficiency according to standards developed by either (1) the Hydronics Institute Division of the Air-Conditioning, Heating, and Refrigeration Institute (AHRI), (2) the American Gas Association (AGA), or (3) Underwriters Laboratories (UL). AHRI Standard 1500 for rating cast-iron sectional, steel, and copper boilers bases performance on controlled test conditions for fuel inputs of 90 kW and larger. The gross output obtained by the test is limited by such factors as flue gas temperature, draft, CO2 in the flue gas, and minimum overall efficiency. This standard applies primarily to oil-fired equipment; however, it is also applied to forced-draft gas-fired or dual-fueled units. Gas boilers are generally design-certified by an accredited testing laboratory based on tests conducted in accordance with ANSI Standard Z21.13 or UL Standard 795. Note that the Z21.13 test procedure may be applied to both condensing and noncondensing boilers. This test uses 27°C RWT, 56 K t, steady-state, full-load operation, and allows the presence of condensate to be ignored. Efficiencies published under this test procedure are generally not achieved in actual operation. Instead of the HI-GAMA, AGA, and UL standards, test procedures for commercial-industrial and packaged fire-tube boilers are often performed based on ASME Performance Test Code 4 (2013). Units are tested for performance under controlled test conditions with minimum required levels of efficiency. Further, the American Boiler Manufacturers Association (ABMA) publishes several guidelines for the care and operation of commercial and industrial boilers and for control parameters. Residential heating boilers (i.e., all gas- and oil-fired boilers with inputs less than 90 kW in the United States) are rated according to standards developed by the U.S. Department of Energy (DOE). The procedure determines both on- and off-cycle losses based on a laboratory test. The test results are applied to a computer program, which simulates an installation and predicts an annual fuel utilization efficiency (AFUE). The steady-state efficiency developed during the test is similar to combustion efficiency and is the basis for determining DOE heating capacity, a term similar to gross output. The AFUE represents the part-load efficiency at the average outdoor temperature and load for a typical boiler installed in the United States. Although this value is useful for comparing different boiler models, it is not meant to represent actual efficiency for a specific installation.

5.

SIZING

Boiler sizing is the selection of boiler output capacity to meet connected load. The boiler gross output is the rate of heat delivered by the boiler to the system under continuous firing at rated input. Net rating (I-B-R rating) is gross output minus a fixed percentage (called the piping and pickup factor) to allow for an estimated average piping heat loss, plus an added load for initially heating up the water in a system (sometimes called pickup). This I-B-R piping and pickup factor is 1.15 for water boilers and ranges from 1.27 to 1.33 for steam boilers, with the smaller number applying as the boilers get larger. The net rating is calculated by dividing the gross output by the appropriate piping and pickup factor. Piping loss is variable. If all piping is in the space defined as load, loss is zero. If piping runs through unheated spaces, heat loss from the piping may be much higher than accounted for by the fixed net rating factor. Pickup is also variable. When the actual

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connected load is less than design load, the pickup factor may be unnecessary. On the coldest day, extra output (boiler and radiation) is needed to pick up the load from a shutdown or low night setback. If night setback is not used, or if no extended shutdown occurs, no pickup load exists. Standby capacity for pickup, if needed, can be in the form of excess capacity in baseload boilers or in a standby boiler. If piping and pickup losses are negligible, the boiler gross output can be considered the design load. If piping loss and pickup load are large or variable, those loads should be calculated and equivalent gross boiler capacity added. Boiler capacity must be matched to the terminal unit and system delivery capacity. That is, if the boiler output is greater than the terminal output, the water temperature rises and the boiler cycles on the high-limit control, delivering an average input that is much lower than the boiler gross output. Significant oversizing of the boiler may result in a much lower overall boiler efficiency.

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6.

BURNER TYPES

Burners for installation on boilers are grouped generally by fuel used and pressure type. Fuel groupings include fuel oil, natural gas, propane, wood, or coal. A dual-fuel burner may use two or more fuels (e.g., No. 2 fuel oil and natural gas). The pressure type refers to whether the burner is atmospheric or a fan is used for pressurization. In atmospheric burners, firing generally natural gas or propane, the fuel is introduced across a drilled orifice manifold where it contacts combustion air and is ignited. The chimney or flue produces a natural draft to remove the products of combustion. In power burners, a fan pushes combustion air into a burner or combustion chamber under positive pressure where it mixes with the fuel and is ignited. The products of combustion are pushed through the combustion chamber and boiler by the fan, then flow through the chimney by natural draft, or induced draft caused by a chimney fan. Burners may also be classified by method of fuel atomization. In pressure atomization, fuel oil at 550 to 2070 kPa (gage) is pumped through a nozzle orifice to create a fine mist. The fuel-rich mist is mixed with combustion air provided by a fan and is ignited at the burner. In steam atomization, generally used on heavy grades of fuel oil, high-pressure steam is mixed with pressurized fuel oil through a nozzle orifice to heat the oil and reduce the oil’s viscosity to create a fine mist. The mist is mixed with combustion air provided by a fan and is ignited at the burner.

7.

BOILER CONTROLS

Boiler controls provide automatic regulation of burner and boiler performance to ensure safe and efficient operation. Operating and combustion controls regulate the rate of fuel input in response to a signal representing load change (demand), so that the average boiler output equals the load within some accepted tolerance. Water level and flame safety controls cut off fuel flow when unsafe conditions develop. The National Fire Protection Association (NFPA) Code 85, Boiler and Combustion Systems Hazard Code, is generally accepted as the governing code for boiler control systems. Other requirements from insurance companies or local governing agencies may also be applicable. Often, the governing agency having jurisdiction may specify specific requirements that the heating system designer or specifying engineer must comply with in the design. It is essential that the designer or engineer determine the applicable codes, and specify the controls and skills needed to complete the control system.

Operating Controls Steam boilers are operated by boiler-mounted, pressure-actuated controls, which vary fuel input to the boiler. Traditional examples of burner controls were on/off, high/low/off, and modulating.

Modulating controls infinitely vary fuel input from 100% down to a selected minimum set point. The ratio of maximum to minimum is the turndown ratio. The minimum input is usually between 5 and 33% (i.e., 20 to 1 down to 3 to 1 ratios); input depends on the size and type of fuel-burning equipment and system. High turndown ratios in noncondensing boilers must be considered carefully to prevent condensation at lower firing rates. Hot-water boilers are operated by temperature-actuated controls that are usually mounted on the boiler. Traditionally, burner controls were the same as for steam boilers (i.e., on/off, high/low/off, and modulating). Modulating controls typically offer more precise water temperature control and higher efficiency than on/off or high/ low controls, if airflow through the boiler is modulated along with fuel input. Boiler reset controls can enhance the efficiency of hot-water boilers. These controls may operate with any of the burner controls mentioned previously. They automatically change the high-limit set point of the boiler to match the variable building load demands caused by changing outdoor temperatures. By keeping boiler water temperature as low as possible, efficiency is enhanced and standby losses are reduced. Microprocessor-Based Control Systems. The introduction of microprocessor-based control systems has changed traditional operating controls on boilers. In the past, smaller boilers were equipped with on/off or high/low/off electromagnetic-relay-based burner operating controls with mercury switches, with larger boilers provided with modulating controls. The low cost and greater efficiency of microprocessor-based control systems has resulted in the availability of such controls on small factory-packaged boilers, and nearly all medium and larger boiler installations. The recent introduction of integrated combustion and burner safeguard microprocessor controllers has accelerated this availability. Traditionally, most burner installations used a single actuator to drive the combustion air damper and fuel ratio valves through common linkages. Such installations were called single-jackshaft controls. The fuel ratio valves often used set screw cams to produce efficient combustion throughout the firing range of the burner. Tuning the burner involved positioning the set screws to adjust the cam, which in turn regulated the fuel ratio at various firing rates. Tuning was often cumbersome, and easily lost when set screws loosened. Single-jackshaft control also invariably meant compromises were made in tuning, generally resulting in inefficient combustion with high excess air ratios at low firing rates. Current technology eliminates the single-jackshaft, using “linkless” burners with individual actuators controlling the combustion air damper and each fuel valve. With the high-speed processing ability of the microprocessors, individual actuators can quickly and accurately respond to changes in load, ensuring efficient combustion throughout the full firing range. When oxygen analyzers are installed to measure the oxygen content of the flue gas, microprocessor combustion controllers can modulate the combustion air damper and fuel valve actuators to ensure optimal combustion efficiency.

Water Level Controls Maintaining proper water level in a boiler is of paramount concern. Should water level drop below a preset limit, damage may occur from overheating of boiler surfaces, resulting in cracking of cast-iron sections, or plastic deformation of steel tubes and tubesheets. Such a condition is known as dry-firing of the boiler. Installation of a low-water cutoff switch to stop fuel flow to the burner is necessary to prevent damage from dry firing. To maintain proper water levels, different methods may be used. Smaller boilers generally use a boiler feedwater controller to cause a feedwater pump to pump water directly to the boiler to maintain proper water level. In larger installations, a feedwater piping loop

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may serve several boilers in parallel. In such an installation, each boiler has a feedwater valve controlled by a feedwater controller mounted on the boiler. The controller modulates the feedwater valve to maintain proper water level in the boiler. Simple feedwater control systems use water level as the control parameter to modulate the feedwater valve. Such systems are considered single-element feedwater systems. In larger boiler installations where steam is generated, the steam flow rate or rate of change of steam pressure may also be monitored with a signal sent to the feedwater controller. If the controller is programmed to modulate the feedwater valve based on both water level and steam flow rate or rate of change of pressure, the feedwater control system is referred to a two-element system. A three-element system uses water level, steam flow rate, and rate of pressure change to modulate the feedwater valve.

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8.

FLAME SAFEGUARD CONTROLS

Flame safeguard controls monitor flame condition and shut fuel flow to the burner in the event of an unsafe condition. The safety circuit of a flame safeguard control system typically includes switch contacts for low-water cutoff, high limits, air proving switches, redundant safety and operating controls, and flame monitors. Flame monitors typically use either infrared or ultraviolet scanners to monitor flame condition and deactivate the burner in the event of nonignition or other unsafe flame condition. Flame safeguard controllers usually use preprogrammed algorithms to operate a burner and cycle it through stages of operation. The first stage is a purge cycle wherein the boiler’s combustion chamber is flushed with combustion air to remove any unspent fuel and products of combustion that remain from the previous cycle. After purging, the pilot flame is ignited and the main fuel introduced into the burner and ignited. In the absence of a pilot fuel, such as with direct electronic spark ignition, the main fuel is introduced and ignited. In either case, the flame monitor determines whether ignition has occurred and whether the resulting flame is proper. If ignition has failed, or flame is not indicated, the flame monitor cuts off fuel flow, causing a postpurge cycle to rid the combustion chamber of unspent fuel and products of combustion. On restart, the burner again starts with a purge cycle and repeats the steps to ignition. When proper flame is established,

the flame safeguard control system allows the operating control or combustion control system to control or modulate the burner firing rate. Traditionally, flame safeguard systems were separate controllers from operating or combustion control systems. With the advent of microprocessor-based control systems, separate microprocessors control the individual functions of flame safeguard and combustion control. Recently introduced burner controllers provide an integrated control with algorithms for both flame safeguard and combustion control.

REFERENCES ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal.ashrae .org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org/bookstore. AHRI. 2015. Performance rating of commercial space heating boilers. ANSI/AHRI Standard 1500-2015. Air-Conditioning, Heating, and Refrigeration Institute, Arlington, VA. ANSI. 2010. Gas-fired low-pressure steam and hot water boilers. Standard Z21.13-2010/CSA 4.9-2010. CSA International, Mississauga, Ontario. ASME. 2015. Boiler and pressure vessel code. American Society of Mechanical Engineers, New York. ASME. 2013. Fired steam generators. Performance Test Code PTC 4-2013. American Society of Mechanical Engineers, New York. NFPA. 2015. Boiler and combustion systems hazard code. NFPA Code 852015. National Fire Protection Association, Quincy, MA. UL. 2011. Commercial-industrial gas heating equipment. Standard 795. Underwriters Laboratories, Northbrook, IL.

BIBLIOGRAPHY ABMA. 1998. Packaged boiler engineering manual. American Boiler Manufacturers Association, Arlington, VA. ASME. 2018. Controls and safety devices for automatically fired boilers, Standard CSD-1-2018. American Society of Mechanical Engineers, New York. Strehlow, R.A. 1984. Combustion fundamentals. McGraw-Hill, New York. Woodruff, E.B., H.B. Lammers, and T.F. Lammers. 1984. Steam-plant operation, 5th ed. McGraw-Hill, New York.

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Related Commercial Resources CHAPTER 33

FURNACES Components ............................................................................. Heat Source Types ................................................................... Commercial Equipment ........................................................... Controls and Operating Characteristics .....................................................................

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URNACES are self-enclosed, permanently installed major appliances that provide heated air through ductwork to the space being heated. In addition, a furnace may provide the indoor fan necessary for circulating heated or cooled air from a split or singlepackage air conditioner or heat pump (see Chapter 10). Furnaces may be used in either residential or commercial applications, and may be grouped according to the following characteristics: • Heat source: electricity, natural gas/propane (fan assisted, condensing or noncondensing), or oil (forced draft with power atomizing burner) • Installation location: within conditioned space (indoors), or outside conditioned space (either outdoors, or inside the structure but not within the conditioned space) • Combustion air source: direct vent (outdoor air) or indoor air • Mounting arrangement and airflow: horizontal, vertical upflow, vertical downflow, or multiposition Furnaces that use electricity as a heat source include one or more resistance-type heating elements that heats the circulating air either directly or through a metal sheath that encloses the resistance element. In gas- or oil-fired furnaces, combustion occurs in the heat exchanger sections or in a combustion chamber. Circulating air passes over the outer surfaces of a heat exchanger so that it does not contact the fuel or the products of combustion, which are passed to the outdoor atmosphere through a vent. In North America, natural gas is the most common fuel supplied for residential heating, and the central-system forced-air furnace (Figure 1) is the most common way of heating with natural gas. This

Equipment Selection................................................................. 33.6 Calculations ............................................................................. 33.8 Technical Data ......................................................................... 33.8 Installation ............................................................................... 33.9 Agency Listings ...................................................................... 33.10

type of furnace is equipped with a blower to circulate air through the furnace enclosure, over the heat exchanger, and through the ductwork distribution system. A furnace such as that in Figure 1 is categorized as follows: • Heat source: natural gas • Mounting arrangement and airflow: vertical upflow • Installation location: varies; usually inside the structure but not necessarily within the conditioned space • Combustion air source: varies; high-efficiency furnaces usually are direct vent

1.

COMPONENTS

A typical furnace consists of the following basic components: (1) a cabinet or casing; (2) heat exchangers; (3) heat sources, including burners and controls; (4) venting components, such as an induced-draft blower; (5) a circulating air blower and motor; and (6) an air filter and other accessories such as a humidifier, electronic air cleaner, air-conditioning coil, or a combination of these elements.

Casing or Cabinet The furnace casing is most commonly formed from painted coldrolled steel. Access panels on the furnace allow access to those sections requiring service. The inside of the casing adjacent to the heat exchanger or electric heat elements is lined with a foil-faced blanket insulation and/or a metal radiation shield to reduce heat losses through the casing and to limit the outer surface temperature of the furnace. On some furnaces, the inside of the blower compartment is lined with insulation to acoustically dampen the blower noise. Commercial furnace cabinets may also include the indoor and outdoor air-conditioning or heat pump components. New scrutiny of the airtightness of system ductwork has drawn attention to the furnace casing as a common source of leakage. ASHRAE Standard 193-2010 provides a method of test to determine the airtightness of forced-air HVAC equipment before field installation. The standard is cited in the U.S. Department of Energy (DOE) ENERGY STAR® Program Requirements Product Specification for Furnaces, Version 4.0 (DOE 2013).

Heat Exchangers

Fig. 1 Induced-Draft Gas Furnace The preparation of this chapter is assigned to TC 6.3, Central Forced Air Heating and Cooling Systems.

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Furnaces with gas-fired burners have heat exchangers that are typically made either of left/right sets of formed parts that are joined together to form a clamshell, finless tubes bent into a compact form, or finned-tube (condensing) heat exchangers. Standard indoor furnace heat exchangers are generally made of alloy steel. Common corrosion-resistant materials include aluminized steel and stainless steel. Furnaces certified for use downstream of a cooling coil must have corrosion-resistant heat exchangers. Some problems of heat exchanger corrosion and failure have been encountered because of exposure to halogen ions in flue gas. These problems were caused by combustion air contaminated by substances such as laundry bleach, cleaning solvents, and halogenated hydrocarbon refrigerants.

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Heat exchangers of oil-fired furnaces are normally heavy-gage steel formed into a welded assembly. Hot flue products flow through the inside of the heat exchanger into the chimney, and conditioned air flows over the outside of the heat exchanger and into the air supply plenum.

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Heat Sources Electric Heat Elements. Elements for electric furnaces are generally either open wire, open ribbon, or wire enclosed in a tube. Current is applied to the element and heats it through resistance of the material. Burners and Internal Controls. Gas burners are most frequently made of stamped sheet metal, although cast iron is also used. Fabricated sheet metal burners may be made from cold-rolled steel coated with high-temperature paint or from a corrosion-resistant material such as stainless or aluminized steel. Burner material must meet the corrosion protection requirements of the specific application. Gas furnace burners may be of either the monoport or multiport type; the type used with a particular furnace depends on compatibility with the heat exchanger. Gas furnace controls include an ignition device, gas valve, fan control, limit switch, and other components specified by the manufacturer. These controls allow gas to flow to the burners when heat is required. The most common ignition systems are (1) direct spark, and (2) hot-surface ignition. The section on Technical Data has further details on the function and performance of individual control components. Oil furnaces are generally equipped with pressure-atomizing burners. The pump pressure and size of the injection nozzle orifice regulate the firing rate of the furnace. Electric ignition lights the burners. Other furnace controls, such as the blower switch and the limit switch, are similar to those used on gas furnaces.

Combustion Venting Components Fan-assisted combustion furnaces use a small blower to induce flue products through the furnace. Induced-draft furnaces may or may not have a relief air opening, but they meet the same safety requirements regardless. Residential furnaces built since 1987 are equipped with a blocked-vent shutoff switch to shut down the furnace in case the vent becomes blocked. Research into common venting of natural-draft appliances (water heaters) and fan-assisted combustion furnaces shows that nonpositive vent pressure systems may operate on a common vent. Refer to manufacturers’ instructions for specific information. Direct-vent furnaces use outdoor air for combustion. Outdoor air is supplied to the furnace combustion chamber by direct connections between the furnace and the outdoor air. If the vent or the combustion air supply becomes blocked, the furnace control system will shut down the furnace. ANSI Standard Z21.47/CSA 2.3 classifies venting systems. Central furnaces are categorized by temperature and pressure attained in the vent and by the steady-state efficiency attained by the furnace. Although ANSI Standard Z21.47/CSA 2.3 uses 83% as the steady-state efficiency dividing central furnace categories, a general rule of thumb is as follows: Category I: nonpositive vent pressure and flue loss of 17% or more Category II: nonpositive vent pressure and flue loss less than 17% Category III: positive vent pressure and flue loss of 17% or more Category IV: positive vent pressure and flue loss less than 17% Furnaces rated in accordance with ANSI Standard Z21.47/CSA 2.3 that are not direct vent are marked to show that they are in one of these four venting categories.

Ducted-system, oil-fired, forced-air furnaces are usually forced draft.

Circulating Blowers and Motors Centrifugal blowers with forward-curved blades of the doubleinlet type are used in most forced-air furnaces. These blowers overcome the resistance of furnace air passageways, filters, and ductwork. They are usually sized to provide the additional air requirement for cooling and the static pressure required for the cooling coil. The blower may be a direct-drive type, with the blower wheel attached directly to the motor shaft, or it may be a belt-drive type, with a pulley and V-belt used to drive the blower wheel. Electric motors used to drive furnace blowers are usually custom designed for each furnace model or model series. Direct-drive motors may be shaded-pole, permanent split-capacitor (PSC), or brushless permanent magnet. Speed may be varied by taps connected to extra windings in the motor. Belt-drive blower motors are normally split-phase or capacitor-start. The speed of belt-drive blowers is controlled by adjusting a variable-pitch drive pulley. Electronically controlled, variable-speed motors using a brushless permanent magnet design are inherently more efficient than shaded-pole or PSC motors. The motor is controlled electronically by a microprocessor and electronic controls, which provides the ability to increase or decrease motor speed and maintain efficiency across a wide range of operating speeds. In the United States, the DOE regulates energy efficiency of residential furnace fans. DOE has initiated a rulemaking to consider new energy conservation or use standards for furnace fans. For the current status of DOE regulations, visit www.energy.gov/eere/ buildings/appliance-and-equipment-standards-program.

Filters and Other Accessories Air Filters. An air filter in a forced-air furnace removes dust from the air that could reduce the effectiveness of the blower and heat exchanger(s), and may also help provide cleaner air for the indoor environment (see ASHRAE Standard 52.2). Filters installed in a forced-air furnace are often disposable. Permanent filters that may be washed or vacuum-cleaned and reinstalled are also used. The filter is always located in the circulating airstream ahead of the blower and heat exchanger. Because the air filter keeps airflow components of the furnace clean, it should be cleaned or replaced regularly to extend the life of the furnace components. See Chapters 10 and 29 for further information on air filters. Humidifiers. These are not included as a standard part of the furnace package. However, one advantage of a forced-air heating system is that it offers the opportunity to control the relative humidity of the heated space at a comfortable level. Chapter 22 addresses various types of humidifiers used with forced-air furnaces. Electronic Air Cleaners. These air cleaners may be much more effective than the air filter provided with the furnace, and they filter out much finer particles, including smoke and pollen. Electronic air cleaners create an electric field of high-voltage direct current in which dust particles are given a charge and collected on a plate having the opposite charge. The collected material is then cleaned periodically from the collector plate by the homeowner. Electronic air cleaners are mounted in the airstream entering the furnace. Chapter 29 has detailed information on filters.

Airflow Variations The components of a forced-air furnace can be arranged in a variety of configurations to suit a residential heating system. The relative positions of the components in the different types of furnaces are as follows: • Upflow furnace. In an upflow furnace (Figure 2), the blower is located beneath the heat exchanger and discharges vertically

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upward. Air enters through the bottom or the side of the blower compartment and leaves at the top. This furnace may be used in closets and utility rooms on the first floor or in basements, with the return air ducted down to the blower compartment entrance. Downflow furnace. In a downflow furnace (Figure 3), the blower is located above the heat exchanger and discharges downward. Air enters at the top and is discharged vertically at the bottom. This furnace is normally used with a perimeter heating system in a house without a basement. It is also used in upstairs furnace closets and utility rooms supplying conditioned air to both levels of a two-story house. • Horizontal furnace. In a horizontal furnace, the blower is located beside the heat exchanger (Figure 4). Air enters at one end, travels horizontally through the blower and over the heat exchanger, and is discharged at the opposite end. This furnace is used for loca-

tions with limited head room such as attics and crawlspaces, or is suspended under a roof or floor or placed above a suspended ceiling. These units are often designed so that the components may be rearranged to allow installation with airflow from left to right or from right to left. • Multiposition (multipoise) furnace. This furnace can be installed in more than one airflow configuration (e.g., upflow or horizontal; downflow or horizontal; or upflow, downflow, or horizontal). In some models, field conversion is necessary to accommodate an alternative installation. • Basement furnace. The basement furnace (Figure 5) is a variation of the upflow furnace and requires less head room. The blower is located beside the heat exchanger at the bottom. Air enters the top of the cabinet, is drawn down through the blower, is discharged over the heat exchanger, and leaves vertically at the top. This type of furnace has become less popular because of the advent of short upflow furnaces. • Gravity furnace. These furnaces are no longer available, and they are not common. This furnace has larger air passages through the casing and over the heat exchanger so that the buoyancy force created by the air being warmed circulates the air through the ducts. Wall furnaces that rely on natural convection (gravity) are discussed in Chapter 34.

Combustion System Variations Gas-fired furnaces use a fan-assisted combustion system. Fanassisted combustion furnaces have a combustion blower, which may be located either upstream or downstream from the heat exchangers (Figure 6). If the blower is located upstream, blowing the combustion air into the heat exchangers, the system is known as a forced-draft system. If the blower is downstream, the arrangement is known as an induced-draft system. Direct-vent furnaces obtain combustion air from outside the structure. Mobile home furnaces must be of the direct-vent type.

Fig. 2 Upflow Category I Furnace with InducedDraft Blower

Fig. 4 Horizontal Category I Furnace with Induced-Draft Blower

Fig. 3 Downflow (Counterflow) Category I Furnace with Induced-Draft Blower

Fig. 5

Basement (Lowboy) Category I Furnace with Induced-Draft Blower

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Fig. 7 Electric Forced-Air Furnace

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sell conversion kits containing both the required parts and instructions to convert furnace operation from one gas to the other.

Oil Furnaces

Fig. 6

Terminology Used to Describe Fan-Assisted Combustion

Indoor/Outdoor Furnace Variations Central system residential furnaces are designed and certified for either indoor or outdoor use. Outdoor furnaces are normally horizontal flow and convertible to downflow. The heating-only outdoor furnace is similar to the more common indoor horizontal furnace. The primary difference is that the outdoor furnace is weatherized; the motors and controls are sealed, and the exposed components are made of corrosion-resistant materials such as galvanized or aluminized steel. A common style of outdoor furnace is the combination package unit. This unit is a combination of an air conditioner and a gas or electric furnace built into a single casing. The design varies, but the most common combination consists of an electric air conditioner coupled with a horizontal gas or electric furnace. The advantage is that much of the interconnecting piping and wiring is included in the unit.

2.

HEAT SOURCE TYPES

Natural Gas and Propane Furnaces Most manufacturers have their furnaces certified for both natural gas and propane. The major difference between natural gas and propane furnaces is the pressure at which the gas is injected from the manifold into the burners. For natural gas, the manifold pressure is usually controlled at 750 to 1000 Pa; for propane, the pressure is usually 2500 to 2700 Pa. Because of the higher injection pressure and the greater heat content per volume of propane, there are certain physical differences between a natural gas furnace and a propane furnace. One difference is that the burner orifices must be smaller for propane furnaces. The gas valve regulator spring is also different. Sometimes it is necessary to change burners, but this is not normally required. Manufacturers

Indoor oil furnaces come in the same configurations as gas furnaces. They are available in upflow, downflow, horizontal, and lowboy configurations for ducted systems. Oil-fired outdoor furnaces and combination units are not common. The major differences between oil and gas furnaces are in the combustion system, heat exchanger, and barometric draft regulator.

Electric Furnaces Electric-powered furnaces come in a variety of configurations and have some similarities to gas- and oil-fired furnaces. However, when a furnace is used with an air conditioner, the cooling coil may be upstream from the blower and heaters. On gas- and oil-fired furnaces, the cooling coil is normally mounted downstream from the blower and heat exchangers so cold air leaving the cooling coil does not contact the heat exchangers, which could cause premature corrosion from condensation. If the cooling coil is upstream of the heat exchangers on a gas- or oil-fired product, the heat exchanger may require a mechanism to remove the condensed moisture. Figure 7 shows a typical arrangement for an electric forced-air furnace. Air enters the bottom of the furnace and passes through the filter, then flows up through the cooling coil section into the blower. The electric heating elements are immediately above the blower so that the high-velocity air discharging from the blower passes directly through the heating elements. The furnace casing, air filter, and blower are similar to equivalent gas furnace components. The heating elements are made in modular form, with 5 kW capacity being typical for each module. Electric furnace controls include electric overload protection, contactor, limit switches, and a fan control switch. The overload protection may be either fuses or circuit breakers. The contactor brings the electric heat modules on. The fan control switch and limit switch functions are similar to those of the gas furnace, but one limit switch is usually used for each heating element. Frequently, electric furnaces are made from modular sections; for example, the coil box, blower section, and electric heat section are made separately and then assembled in the field. Regardless of whether the furnace is made from a single-piece casing or a modular casing, it is generally a multiposition unit. Thus, the same unit may be used for upflow, downflow, or horizontal installations. When an electric heating appliance is sold without a cooling coil, it is known as an electric furnace. The same appliance is called a fan-coil air handler when it has an air-conditioning coil already

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installed. When the unit is used as the indoor section of a split heat pump, it is called a heat pump fan-coil air handler. For detailed information on heat pumps, see Chapters 9 and 48. Electric forced-air furnaces are also used with packaged heat pumps and packaged air conditioners.

3.

COMMERCIAL EQUIPMENT

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The basic differences between residential and commercial furnaces are available options such as economizers, outdoor air dampers, and the type of electrical service required (three-phase). Commercial heating equipment comes in almost as many flow arrangements and design variations as residential equipment. Some are identical to residential equipment, whereas others are unique to commercial applications. Some commercial units function as a part of a ducted system, and others operate as unducted space heaters.

or more zones at the same time that cooling is supplied to other zones. Large package units are normally available only in a curbed configuration (i.e., units are mounted on a rooftop over a curbed opening in the roof). Supply and return air enters through the bottom of the unit. Smaller units may be available for either curbed or uncurbed mounting. In either case, the unit is usually connected to ductwork in the building to distribute the conditioned air.

Unducted Heaters Ductless furnaces, floor furnaces (Figure 8), and wall furnaces are discussed in Chapter 34. Infrared heating equipment is covered in Chapter 16.

4.

CONTROLS AND OPERATING CHARACTERISTICS

Ducted Equipment

External to Furnace

Upflow Gas-Fired Commercial Furnaces. These furnaces are normally incorporated into a system in conjunction with a commercial split-system air-conditioning unit and are available in either propane or natural gas. Oil-fired units may be available on a limited basis. Horizontal Gas-Fired Duct Furnaces. Available for built-up light commercial systems, this type of furnace is not equipped with its own blower but is designed for uniform airflow across the entire furnace. Duct furnaces are normally certified for operation either upstream or downstream of an air conditioner cooling coil. If a combination blower and duct furnace is desired, a package called a blower unit heater is available. Duct furnaces and blower unit heaters are available in natural gas, propane, oil, and electric models. Electric Duct Furnaces. These furnaces are available in a large range of sizes and are suitable for operation in upflow, downflow, or horizontal positions. These units are also used to supply auxiliary heat with the indoor section of a split heat pump. Package Units. The most common commercial furnace is the package unit, sometimes known as a rooftop unit (RTU). These are available as air-conditioning units with propane and natural gas furnaces, electric resistance heaters, or heat pumps. Combination oilheat/electric-cool units are not commonly available. Package units of 50 kW and under are available as single-zone units. The entire unit must be in either heating mode or cooling mode. All air delivered by the unit is at the same temperature. Frequently, the heating function is staged so that the system operates at reduced heat output when the load is small. Large package units in the 50 to 175 kW range are available as single-zone units, as are small units; however, they are also available as multizone units. A multizone unit supplies conditioned air to several different zones of a building in response to individual thermostats controlling those zones. These units can supply heating to one

Externally, the furnace is controlled by a low-voltage room thermostat. Chapter 7 of the 2017 ASHRAE Handbook—Fundamentals discusses thermostats in detail. In North America, a furnace that is not part of a package unit is commonly installed with a split-system air conditioner. The furnace control receives input from the room thermostat, which determines the mode of operation. The furnace blower is used to circulate air through the ducts in the heating, cooling, and circulating modes of operation.

Fig. 8 Standing Floor Furnace

Internal to Furnace Several types of gas valves perform various functions within the furnace. The type of valve available relates closely to the type of ignition device used. Two-stage valves, available on some furnaces, operate at full gas input or at a reduced rate, and are controlled by either a two-stage thermostat or a software algorithm programmed in the furnace control system. They provide less heat at the reduced input and, therefore, may produce less space temperature variation and greater comfort during mild weather conditions when full heat output is not required. Two-stage control is used frequently for zoning applications. Fuel savings with two-stage firing rate systems may not be realized unless both the fuel and the combustion air are controlled. The fan control switch controls the circulating air blower. This switch may be temperature-sensitive and exposed to the circulating airstream in the furnace cabinet, or it may be an electronically operated relay. Blower start-up is typically delayed about 1 min after burner start-up. This delay gives the heat exchangers time to warm up and reduces the flow of cold air when the blower comes on. Blower shutdown is also delayed several minutes after burner shutdown to remove residual heat from the heat exchangers and to improve the annual efficiency of the furnace. Constant blower operation throughout the heating season is sometimes used to improve air circulation; however, this increases fan motor energy consumption, duct conductive losses, and air distribution system air leakage losses. Electronic motors that provide continuous but variable airflow use less energy. Both strategies may be considered when air filtering performance is important. The limit switch prevents overheating in the event of severe reduction in circulating airflow. This temperature-sensitive switch is exposed to the circulating airstream and shuts off the heat source (e.g., gas valve or electric element) if the temperature of air leaving the furnace is excessive. The fan control and limit switches are sometimes incorporated in the same housing and may be operated by the same thermostatic element. In the United States, the blockedvent shutoff switch and flame rollout switch have been required on residential furnaces produced since November 1989; they shut off the gas valve if the vent is blocked or when insufficient combustion air is present.

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Furnaces using fan-assisted combustion feature a pressure switch to verify the flow of combustion air before opening the gas valve. Electronic control systems are available in furnaces to provide sequencing of the inducer prepurge, ignition, circulating air blower operation, and inducer postpurge functions according to an algorithm provided by the manufacturer.

5.

EQUIPMENT SELECTION

Many options are available to consumers, and careful planning is needed when selecting equipment. Some decisions can be very basic, whereas others may require research into the various kinds of equipment and features that are available. Several selection considerations are presented here.

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Distribution System A fundamental question in the selection process is whether the space to be heated uses a circulating forced-air system or a hydronic system. These two types of systems are vastly different with respect to equipment selection. For hydronic systems, refer to Chapter 36. Forced-air systems vary widely. A forced-air distribution system typically has a central air duct, with branches feeding air to numerous supply registers. The duct branches are designed to proportion the air to the different spaces in the building, so that temperatures are best managed by a central thermostat location. Some systems are zoned, with different thermostat sensors; in these systems, dampers with electrical motors are placed at strategic branches of the distribution system and are opened or closed according to each zone’s demand for heat. Choosing a zoned system affects the type of equipment selected. Chapter 10 discusses the overall design configuration and efficiency of forced-air systems.

Equipment Location Furnaces can be installed inside or outside a building. For ideal air distribution, locate the unit in the center of the structure being heated. Furnaces are typically located in a closet, mechanical room, basement, attic, crawlspace, garage, or outdoors. As minimum efficiency levels have increased for residential air-conditioning systems, evaporator coils have increased in height to accommodate additional coil surface area. Compact furnaces have been introduced that allow the evaporator coil and furnace system combination to be installed as a replacement without structural modification in some cases. For installations that cannot be accommodated by the more compact furnace designs, the furnace and evaporator coil are more often located in attic or garage spaces. Installation locations are characterized as either indoors, outdoors (weatherized), or isolated combustion systems (ICS). Indoor furnaces are installed in the heated structure, such as in a basement that is connected to the internal living space, in a utility room, or in a closet. In these locations, air that directly surrounds the furnace is in communication with the air of the heated space. Heat that is lost from the cabinet and adjacent ducts by conduction or air leakage is largely recaptured, helping to preserve furnace efficiency. Some furnaces use combustion air from inside the building. In these applications, room air is used for combustion and exhausted to the outdoors. Additionally, dilution air is also drawn from the room. Combustion/dilution makeup air is provided to the combustion appliance zone by infiltration or by a duct designed specifically to provide makeup air, as required by installation codes. Furnaces located outside the conditioned space (e.g., crawlspace, attic, garage, outdoors) regain little or none of the conductive or air leakage energy losses. Outdoor installations require furnaces to be qualified as weatherized. Outdoor installation locations are on rooftops, platforms, or on a pad adjacent to the heated structure. Ducts for supply and return air connections may also be exposed to

the elements and should be weatherized appropriately or moved inside to the conditioned space. Isolated combustion systems (ICS) are within the structure being heated, but the air that directly surrounds the furnace does not communicate with the heated space. Air needed for combustion and ventilation is admitted through grille openings or ducts (NFPA Standard 54). Heat given off from the casing is not considered usable heat, and is subtracted from the furnace efficiency. Typical ICS locations include garages, attics, crawlspaces, and closets that are directly ventilated to the outdoors. These furnaces are protected from the elements (but not temperature) by the surrounding structure.

Forced-Air System Primary Use Forced-air systems have many benefits, the most significant of which is that they can be used for both heating and cooling without needing separate duct systems and separate air-handling units. A primary function of the furnace is to circulate air through the forcedair distribution system. Heating the air is an inherent function of a furnace. Cooling is typically a modular add-on, although some furnaces are included as a part of a packaged furnace-and-air-conditioner combination appliance. Forced-air systems also make it possible to add humidifiers and air filters or purifiers. In some cases, forced-air systems can also be connected to an additional appliance to bring clean air from the outdoors through air-to-air energy recovery heat exchangers. Proposed central-fan-integrated ventilation systems would use the furnace blower to pull in and distribute a controlled amount of outdoor air. Air distribution system design must take into account all the system’s intended functions. The air-handling capacity must be designed to meet the demand of the highest airflow and static pressure needs. Typically, the airflow needed for heating is less than that needed for cooling. Manufacturers provide information on furnace performance capabilities, including how much airflow it can deliver at different static pressures. Furnace specification data typically describe the gas input for heating and the airflow (or cooling capacity) for which the unit is designed.

Fuel Selection The type of fuel selected for heating is based on relative fuel cost, number of heating degree-days, and availability of utilities in the area. The most common fuel is natural gas because of its clean burning characteristics, and because of the continuous supply of this fuel through underground distribution networks to most urban settings. Propane and oil fuels are also commonly used. These fuels require on-site storage and periodic fuel deliveries. Electric heat is also continuously available through electrical power grids and is common especially where natural gas is not provided, or where the heating demand is small relative to the cooling demand. Furnaces are clearly marked for the type of fuel to be used. In some cases, a manufacturer-approved conversion kit may be necessary to convert a furnace from one fuel type to another. If the fuel type is changed after the original installation, the conversion must be done by a qualified service person per the manufacturer’s instructions and using the manufacturer’s specified conversion kit. After conversion, the unit must be properly inspected by the local code authority.

Combustion Air and Venting All fuel-burning furnaces must be properly vented to the outdoors. Metal vents, masonry chimneys, and plastic vents are commonly used for furnaces. Manufacturers provide installation instructions for venting their furnaces, and Chapter 35 has a detailed discussion on venting. Air for combustion enters the combustion zone through louvers or pipes. Outdoor air usually has lower levels of pollutants than are

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typically found in air from indoors, garages, utility rooms, and basements. If the furnace is not exposed to pollutants, the heat exchanger material will have a long life.

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Equipment Sizing The furnace’s heating capacity (i.e., the maximum heating rate the furnace can provide) is provided on the appliance rating plate; it is also available through the Air-Conditioning, Heating, and Refrigeration Institute directories (AHRI 2016) and manufacturers’ product literature. The heating load for the intended space must be determined; variables that must be considered when calculating space load include heat gains or losses through walls, floors, and ceiling; infiltration; fenestration; ventilation; internal loads; and humidification. Chapters 17 to 19 in the 2017 ASHRAE Handbook—Fundamentals provide the information necessary to determine residential and nonresidential heating loads. Other factors should be considered when determining furnace capacity. Thermostat setback recovery may require additional heating capacity. On the other hand, supplemental heat sources or off-peak storage devices may offset some of the peak demand capacity. Increasing furnace capacity may increase space temperature swing, and thus reduce comfort. Two-stage or step-modulating equipment could help by using the unit’s maximum capacity to meet the setback recovery needs, and providing a lower stage of heating capacity at other times. If cooling is included, the cooling load must be considered, especially in climates that have substantial cooling loads but minimal heating loads. When a large cooling load necessitates a large airflow rate, heating capacity may be greater than necessary. Two-stage or modulating heating may be a suitable alternative to reduce the potentially large temperature swings that can occur with excessively oversized furnaces.

Types of Furnaces Fuel-burning furnaces are typically subdivided into two primary categories: noncondensing and condensing. Condensing furnaces typically have high efficiencies, ranging from 89 to 98%, because they have a specially designed secondary heat exchanger that extracts the heat of vaporization of water vapor in the exhaust. Such heat exchangers may collect soot in small passages if the burners are not set up properly for the heating value of the gas used. The dew-point temperatures of flue gases of condensing furnaces are significantly above the vent temperature, so plastic or other corrosion-resistant venting material is required. Condensing furnaces must be plumbed for condensate disposal. Provisions must be taken to prevent the condensate trap and drain line from freezing if installed in a location that is likely to be below freezing at some point in the year. Noncondensing furnaces have generally less than 82% steadystate efficiency. This type of furnace has higher flue gas temperatures and requires either metal, masonry, or a combination of the two for venting materials. Because there is no water management, noncondensing furnaces do not need freeze protection.

Consumer Considerations Safety and Reliability. Gas furnaces sold in North America are tested and certified to ANSI Standard Z21.47/CSA 2.3 requirements. Oil furnaces are tested in accordance with UL Standard 727, and oil burners in accordance with UL Standard 296. These standards are intended to ensure that consumer safety and product reliability are maintained in appliance design. Because of open-flame combustion, the following safety items need to be considered: (1) the surrounding atmosphere should be free of dust or chemical concentrations; (2) a path for combustion air must be provided for both sealed and open combustion chambers; and (3) the gas piping

and vent pipes must be installed according to the NFPA/AGA National Fuel Gas Code (NFPA Standard 54), local codes, and the manufacturer’s instructions. For electric furnaces, safety primarily concerns proper wiring techniques. Wiring should comply with the National Electrical Code® (NEC) (NFPA Standard 70) and applicable local codes. Efficiency, Operating, and Life-Cycle Costs. Annual operating costs of furnaces must take into account both the cost of the heating fuel as well as the electrical efficiency of the blower motor. Lifecycle cost determination includes initial cost, maintenance, energy consumption, design life, and price escalation of the fuel. Annual fuel utilization efficiency (AFUE) and energy consumption data to help calculate the annual cost for heating a building are available in the AHRI (2016) directory. AFUE and fuel cost are primary drivers in the operating cost. Electric furnaces are listed as nearly 100% AFUE (site-based) because all of the electrical energy is converted into heat, and the only inefficiency is from cabinet conduction and air leakage losses. Cabinet leakage can affect energy consumption and indoor air quality when furnaces are installed outside the conditioned living space, especially when the home and distribution system are otherwise of tight construction. Air leakage should be taken into account during system design and location. Care should also be taken to install equipment according to the manufacturer’s instructions, and that gaps are not overlooked by the installer where service entries or attachments are made to the cabinet. The ENERGY STAR® qualification specifications for furnaces require that cabinet leakage be less than or equal to 2% of the maximum airflow, measured according to ANSI/ASHRAE Standard 193-2010. Design Life. Typically, heat exchangers made of aluminized steel have a design life of approximately 15 years. Some furnaces now have a 20 year warranty, or even a “lifetime” warranty on the heat exchanger. The design life of electric furnaces depends on the durability of the contactors and heating elements. The typical design life is approximately 15 years. Comfort. Consumer opinions of comfort vary quite a bit. Thermal comfort is affected by supply air temperature, air velocity leaving the supply registers, and proximity of the supply airflow stream to occupants. Complaints of draftiness are common when delivered supply air temperatures are low and register velocities are high. A common solution is to reduce the blower speed to get more temperature rise, while staying within the rated temperature rise range listed on the rating plate. However, reducing blower speed can lead to distribution problems. System design, including register selection and placement, should take these issues into account to avoid comfort problems. Large temperature swings may also cause discomfort. Factors that affect temperature swing include oversizing, thermostat cycling characteristics related to the number of cycles per hour, and thermostatic control. Two-stage or step-modulated heating can improve comfort by reducing the wide, variable temperature swings. These control schemes reduce furnace capacity through gas and blower modulation, which reduces the amount of oversizing for the current demand. Comfort can also be affected by the indoor air quality. Adding air filters with high minimum efficiency reporting value (MERV; see ASHRAE Standard 52.2) ratings reduces airborne particulate matter and allergens. Winter months typically cause drier indoor air conditions, which can be offset by adding duct-integrated humidifiers. Both filters and humidifiers affect duct resistance, which in turn affects electrical energy consumption, and therefore must be considered in system design. Sound level can be classified as a comfort consideration. Chapter 49 of the 2019 ASHRAE Handbook—HVAC Applications outlines procedures for determining acceptable noise levels.

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Many options are available to increase comfort or economy. To manage comfort and cost of operation, a fuel-burning furnace can be combined with a heat pump; this takes advantage of the heat pump’s relatively high efficiency during mild weather, and switches to fuel heating when outdoor temperatures drop and it becomes more difficult for heat pumps to meet the demand. To reduce peak demand for energy, off-peak storage devices may be used to decrease the required capacity of the furnace. The storage device can supply the additional capacity required during the morning recovery of a night setback cycle or reduce the daily peak loads to help in load shedding. Detailed calculations can determine the contribution of storage devices.

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Selecting Furnaces for Commercial Buildings The procedure for design and selection of a commercial furnace is similar to that for a residential furnace. First, the design capacity of the heating system must be determined, considering heat loss from the structure, recovery load, internal heat sources, humidification, off-peak storage, waste heat recovery, and back-up capacity. Because most commercial buildings use setback during weekends, evenings, or other long periods of inactivity, the recovery load is important, as are internal loads and waste heat recovery. The furnace should be sized according to the load per ACCA (2004). Efficiency of commercial units is about the same as for noncondensing residential units. Two-stage gas valves are frequently used with commercial furnaces, but the efficiency of a two-stage system may be lower than for a single-stage system. At a reduced firing rate, excess combustion airflow through the burners increases, decreasing the steady-state operating efficiency of the furnace. Multistage furnaces with multistage thermostats and controls may be used to more appropriately match load conditions. The design life of commercial heating and cooling equipment is about 20 years. Most gas furnace heat exchangers are either coated steel or stainless steel. Because most commercial furnaces are made for outdoor application, the cabinets are made from corrosionresistant coated steel (e.g., galvanized or aluminized). Blowers can be direct- or belt-driven and can deliver air at higher static pressure. The noise level of commercial heating equipment is important in some applications, such as schools, office buildings, churches, and theaters. Unit heaters, for example, are used primarily in industrial applications where noise is less important. Duct design can greatly affect noise levels. In many jurisdictions, safety requirements are the same for light commercial systems and residential systems. Above 117 kW gas input, ANSI Standard Z21.47/CSA 2.3 requirements for gas controls are more stringent.

6.

CALCULATIONS

Performance Criteria. Residential furnaces and boilers manufactured and distributed in commerce, as defined by 42 U.S.C. 6291 (16), must meet the energy conservation standards specified in the Code of Federal Regulations, 10 CFR 430.32(e)(1)(i) and (e)(2)(i). This information is also available in the Electronic Code of Federal Regulations, and is published in the AHRI (2016) directories. To calculate the furnace’s rated heating capacity, the steady-state efficiency must be determined. In the United States, manufacturers may publish only the AFUE as an efficiency measure. The unit capacity is proportional to the steady-state efficiency when compared to the fuel input rate. U.S. federal law requires manufacturers of furnaces to use AFUE as determined using the isolated combustion system method to rate efficiency. Since January 1, 1992, all furnaces produced have a minimum AFUE (ICS) level of 78%. Table 1 gives efficiency values for different furnaces.

Table 1 Historical and Typical Values of Efficiency AFUE, % Type of Gas Furnace

Indoor

ICSa

64.5 69.0 78.0

63.9b 68.5b 68.5b

80.0

78.0

82.0 66.0 80.0

80.0 64.5b 78.0

80.0 90.0

78.0 88.0

Indoor

ICSa

71.0 80.0 81.0 82.0 91.0

69.0b 78.0 79.0 80.0 89.0

1. Natural-draft with standing pilot 2. Natural-draft with intermittent ignition 3. Natural-draft with intermittent ignition and auto vent damper 4. Fan-assisted combustion with standing pilot or intermittent ignition 5. Same as 4, except with improved heat transfer 6. Direct vent, natural-draft with standing pilot, preheat 7. Direct vent, fan-assisted combustion, and intermittent ignition 8. Fan-assisted combustion (induced-draft) 9. Condensing Type of Oil Furnace 1. 2. 3. 4. 5.

Standard: pre-1992 Standard: post-1992 Same as 2, with improved heat transfer Same as 3, with automatic vent damper Condensing

a Isolated

combustion system (estimate). design (pilot lights and natural draft systems are now obsolete).

b Pre-1992

7.

TECHNICAL DATA

Detailed technical data on furnaces are available from manufacturers, wholesalers, and dealers. The data are generally tabulated in product specification bulletins printed by the manufacturer for each furnace line. These bulletins usually include performance information, electrical data, blower and air delivery data, control system information, optional equipment information, and dimensions.

Natural Gas Furnaces Capacity Ratings. ANSI Standard Z21.47/CSA 2.3 requires that the heating capacity be marked on the rating plates of commercial furnaces in the United States. The heating capacity of residential furnaces, less than 65 kW input, is required by the Federal Trade Commission and can be found in furnace directories provided by AHRI at www.ahridirectory.org. Capacity is calculated by multiplying the input by the steady-state efficiency. Residential gas furnaces with heating capacities ranging from 10 to 50 kW are readily available. Some smaller furnaces are manufactured for special-purpose installations such as mobile homes. Smaller-capacity furnaces are becoming common because new homes are better insulated and have lower heat loads than older homes. Larger furnaces are also available, but these are generally considered for commercial use. Because of the overwhelming popularity of the upflow furnace, or multiposition including upflow, it is available in the greatest number of models and sizes. Downflow furnaces, dedicated horizontal furnaces, and various combinations are also available but are generally limited in model type and size. Residential gas furnaces can be installed as heating-only systems or as part of a heat/cool system. The difference is that, in the heat/ cool system, the furnace operates as the air-handling section of a split-system air conditioner. Heating-only systems typically operate with enough airflow to yield a 22 to 40 K air temperature rise through the furnace. Condensing furnaces may be designed for a lower temperature rise (as low as 20 K). Furnaces have blowers capable of multiple speeds. When the furnace is used as the air handler for a cooling system, the blower is typically capable of delivering about 55 L/s per kilowatt of air conditioning. Furnaces are generally available to accommodate 5 to 18 kW air conditioners. The blower speed for each mode of opera-

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tion should be selected to provide the required airflow for both heating and cooling operation. Controls of furnaces used in a heat/cool system can be installed to operated the multispeed blower motor at the most appropriate speed for either heating or cooling when airflow requirements vary for each mode. Efficiency Ratings. Currently, gas furnaces have steady-state efficiencies that vary from about 78 to 98%. Natural-draft and fanassisted combustion furnaces typically range from 78 to 80% efficiency, whereas condensing furnaces have over 90% steady-state efficiency.

Propane Furnaces

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Most residential natural gas furnaces are also available in a propane version with identical ratings. The technical data for these two furnaces are identical, except for gas control and burner and pilot orifice sizes. Orifice sizes on propane furnaces are much smaller because propane has a higher density and may be supplied at a higher manifold pressure. The heating value and relative density of typical gases are listed as follows: Gas Type

Heating Value, MJ/m3

Relative Density (Air = 1.0)

Natural Propane Butane

38.4 93.1 118.3

0.60 1.53 2.00

As in natural gas furnaces, the ignition systems have a required pilot gas shutoff feature in case the pilot ignition fails. Pilot gas leakage is more critical with propane or butane gas because both are heavier than air and can accumulate to create an explosive mixture in the furnace or furnace enclosure. Besides natural and propane, a furnace may be certified for manufactured gas, mixed gas, or propane/air mixtures; however, furnaces with these certifications are not commonly available. Mobile home furnaces are certified as convertible from natural gas to propane.

Oil Furnaces Oil furnaces are similar to gas furnaces in size, shape, and function, but the heat exchanger, burner, and combustion control are significantly different. Input ratings are based on oil flow rate (L/s), and heating capacity is calculated by the same method as that for gas furnaces. The typical heating value of oil is 39 MJ/L. Fewer models and sizes are available for oil than are available for gas, but residential furnaces in the range of 19 to 44 kW heating capacity are common. Air delivery ratings are similar to gas furnaces. The efficiency of an oil furnace can drop during normal operation if the burner is not maintained and kept clean. In this case, the oil does not atomize sufficiently to allow complete combustion, and energy is lost up the chimney in the form of unburned hydrocarbons. Because most oil furnaces use power burners and electric ignition, the annual efficiency is relatively high. Oil furnaces are available in upflow, downflow, and horizontal models. The thermostat, fan control switch, and limit switch are similar to those of a gas furnace. Oil flow is controlled by a pump and burner nozzle, which sprays the oil/air mixture into a single-chamber drum-type heat exchanger. The heat exchangers are normally heavygage cold-rolled steel. Humidifiers, electronic air cleaners, and night setback thermostats are available as accessories.

Electric Furnaces Residential electric resistance furnaces are available in heating capacities of 5 to 35 kW. Electric resistance furnaces are typically part of a heat/cool system and provide the appropriate airflow for both heating and cooling modes. The only losses associated with an electric resistance furnace are the conductive and air leakage losses in the cabinet. If the furnace is

located fully within the heated space, then the seasonal efficiency would be 100%. Although the efficiency of an electric furnace is high, electricity is generally a relatively expensive form of energy. The operating cost may be reduced substantially by using an electric heat pump in place of a straight electric resistance furnace. Heat pump systems are discussed in Chapter 9. Electric furnaces are available in upflow, downflow, or horizontal models. Internal controls include overload fuses or circuit breakers, overheat limit switches, a fan control switch, and contactors to bring on the heating elements at timed intervals.

Commercial Furnaces Furnaces with capacities above 44 kW are classified as commercial furnaces. The 1992 U.S. Energy Policy and Conservation Act (EPCA) prescribes minimum efficiency requirements for commercial furnaces based on ASHRAE Standard 90.1. Some efficiency improvement components, such as intermittent ignition devices, are common in commercial furnaces.

8.

INSTALLATION

Installation requirements call for a forced-air heating system to meet three basic criteria: (1) the system must be safe, (2) it must provide comfort for the occupants of the conditioned space, and (3) it must be energy efficient. Location of equipment and ducts, materials selected for the distribution system, and installation practices all affect the total system efficiency. Conduction and air leakage losses can result in substantial energy and system performance degradation and deserve special attention. For maximum safety, comfort, and efficiency, proper treatment of distribution system air leakage is necessary. A few significant considerations for installing furnaces are discussed here; for additional issues, see Chapter 10. Generally, the following three categories of installation guidelines must be followed to ensure the safe operation of a heating system: (1) the equipment manufacturer’s installation instructions, (2) local installation code requirements, and (3) national installation code requirements. Local code requirements may or may not be available, but the other two are always available. Depending on the type of fuel being used, one of the following national code requirements apply in the United States: • NFPA 54 National Fuel Gas Code (also AGA Z223.1) • NFPA 70 National Electrical Code® • NFPA 31 Standard for the Installation of Oil-Burning Equipment Comparable Canadian standards are • CAN/CSA-B149.1 Natural Gas and Propane Installation Code • CSA C22.1 Canadian Electrical Code • CAN/CSA B139 Installation Code for Oil Burning Equipment An additional source is the International Fuel Gas Code (IFGC) (ICC 2015). These regulations provide complete information about construction materials, gas line sizes, flue pipe sizes, wiring sizes, and so forth. Proper design of the air distribution system is necessary for both comfort and safety. Chapter 21 of the 2017 ASHRAE Handbook— Fundamentals, Chapter 1 of the 2019 ASHRAE Handbook—HVAC Applications, and Chapter 10 of this volume provide information on the design of ductwork for forced-air heating systems. Forcedair furnaces provide design airflow at a static pressure as low as 30 Pa for a residential unit to above 250 Pa for a commercial unit. The air distribution system must handle the required volumetric flow rate within the pressure limits of the equipment. If the system is a combined heating/cooling installation, the air distribution system

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must meet the cooling requirement because more air is required for cooling than for heating. It is also important to include the pressure drop of the cooling coil. The Air-Conditioning, Heating, and Refrigeration Institute (AHRI) maximum allowable pressure drop for residential cooling coils is 75 Pa. Condensing furnaces generate a large amount of relatively lowtemperature water. It is important that the exhaust vent be sloped so that water runs back into the furnace. All furnace condensate should go into a drain and not be discharged where it might freeze. If there is not a gravity path to a drain, an adequate-capacity condensate pump is required.

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9.

AGENCY LISTINGS

Construction and performance of furnaces are regulated by several agencies. AHRI, in cooperation with its industry members, sponsors a certification program relating to gas- and oil-fired residential furnaces and boilers. This program uses an independent laboratory to verify commercial and residential furnace and boiler manufacturers’ certified AFUEs and heating capacities, as determined by testing in accordance with the U.S. DOE’s Uniform Test Method for Measuring the Energy Consumption of Furnaces and Boilers (Title II, 10 CFR 430, Subpart B, Appendix N). Gas and oil furnaces with input ratings less than 66 kW and gas and oil boilers with input ratings less than 88 kW are currently included in the program. Also included in the program are the online consumers’ directories, which identify certified products and list the input rating, certified heating capacity, and AFUE for each furnace. Participating manufacturers are entitled to use the AHRI certification symbol. ANSI Standard Z21.47/CSA 2.3 (CSA America is secretariat) gives minimum construction, safety, and performance requirements for gas furnaces. The CSA maintains laboratories to certify furnaces and operates a factory inspection service. Furnaces tested and found to be in compliance are listed in the CSA directory and carry the seals of certification. Underwriters Laboratories (UL) and other approved laboratories can also test and certify equipment in accordance with ANSI Standard Z21.47/CSA 2.3. Gas furnaces may be certified for standard, alcove, closet, or outdoor installation. Standard installation requires clearance between the furnace and combustible material of at least 150 mm. Furnaces certified for alcove or closet installation can be installed with reduced clearance, as listed. Furnaces certified for either sidewall venting or outdoor installation must operate properly in a 50 km/h wind. Construction materials must be able to withstand natural elements without degradation of performance and structure. Horizontal furnaces are normally certified for installation on combustible floors and for attic installation and are so marked, in which case they may be installed with point or line contact between the jacket and combustible constructions. Upflow and downflow furnaces are normally certified for alcove or closet installation. Gas furnaces may be listed to burn natural gas, mixed gas, manufactured gas, propane, or propane/air mixtures. A furnace must be equipped and certified for the specific gas to be used because different burners and controls, as well as orifice changes, may be required. Sometimes oil burners and control packages are sold separately; however, they are normally sold as part of the furnace package. Pressure-type or rotary burners should bear the Underwriters Laboratory label showing compliance with ANSI/UL Standard 296. In addition, the complete furnace should bear markings indicating compliance with UL Standard 727. Vaporizing burner furnaces should also be listed under UL Standard 727. UL Standard 1995 gives requirements for the listing and labeling of electric furnaces and heat pumps.

The following list summarizes important standards issued by Underwriters Laboratories, the Canadian Gas Association, and the Canadian Standards Association that apply to space-heating equipment: ANSI/ASHRAE 103

Method of Testing for Annual Fuel Utilization Efficiency of Residential Central Furnaces and Boilers ANSI Z21.66/CGA 6.14 Automatic Vent Damper Devices for Use with Gas-Fired Appliances ANSI Z83.4/CSA 3.7 Non-Recirculating Direct Gas-Fired Industrial Air Heaters ANSI Z83.18 Recirculating Direct Gas-Fired Industrial Air Heaters ANSI Z83.19/CSA 2.35 Gas-Fired High-Intensity Infrared Heaters ANSI Z83.20/CSA 2.34 Gas-fired Low-Intensity Infrared Heaters ANSI Z83.8/CGA 2.6 Gas Unit Heaters, Gas-Packaged Heaters, Gas Utility Heaters, and Gas-Fired Duct Furnaces ANSI Z21.47/CSA 2.3 Gas-Fired Central Furnaces ANSI/UL 296 Oil Burners ANSI/UL 307A Liquid Fuel-Burning Heating Appliances for Manufactured Homes and Recreational Vehicles ASHRAE 90.1 Energy Standard for Buildings Except LowRise Residential Buildings ICC International Fuel Gas Code® NFPA 70 National Electrical Code® UL 307B Gas-Burning Heating Appliances for Manufactured Homes and Recreational Vehicles UL 727 Oil-Fired Central Furnaces UL 1995/CSA C22.2 No. 236 Heating and Cooling Equipment CGA 3.2 Industrial and Commercial Gas-Fired Package Furnaces CSA B140.4 Oil-Fired Warm Air Furnaces

REFERENCES ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal.ashrae .org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org/bookstore. AHRI. 2016. Directory of certified product performance. Air-Conditioning, Heating, and Refrigeration Institute, Arlington, VA. www.ahridirectory .org. ASHRAE. 2007. Method of testing for annual fuel utilization efficiency of residential central furnaces and boilers. ANSI/ASHRAE Standard 1032007. ACCA. 2004. Residential equipment selection. Manual S. Air Conditioning Contractors of America, Arlington, VA. CFR. (Annual.) FTC appliance labeling. 16 CFR 305. Code of Federal Regulations, U.S. Government Printing Office, Washington, D.C. CFR. (Annual.) Uniform test method for measuring the energy consumption of furnaces and boilers. Title II, 10 CFR 430, Subpart B, Appendix N. Code of Federal Regulations, U.S. Government Printing Office, Washington, D.C. DOE. 2013. ENERGY STAR® program requirements product specification for furnaces, version 4.0. U.S. Department of Energy, Washington, D.C. www.energystar.gov/ia/partners/product_specs/program_reqs/Furnaces _Version_4.0_Program_Requirements.pdf Kelly, G.E., J.G. Chi, and M. Kuklewicz. 1978. Recommended testing and calculation procedures for determining the seasonal performance of residential central furnaces and boilers. Available from National Technical Information Service, Springfield, VA (Order No. PB289484). NFPA/AGA. 2015. National fuel gas code. ANSI/NFPA Standard 54-2015. National Fire Protection Association, Quincy, MA. ANSI/AGA Z223.12015. American Gas Association, Arlington, VA.

BIBLIOGRAPHY ASHRAE. 2017. Method of testing general ventilation air-cleaning devices for removal efficiency by particle size. ANSI/ASHRAE Standard 52.2. ASHRAE. 2014. Method of test for determining the airtightness of HVAC equipment. ANSI/ASHRAE Standard 193-2010 (RA 2014).

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Related Commercial Resources CHAPTER 34

RESIDENTIAL IN-SPACE HEATING EQUIPMENT GAS IN-SPACE HEATERS ...................................................... 34.1 Controls ................................................................................... 34.2 Vent Connectors ....................................................................... 34.3 Sizing Units .............................................................................. 34.3 OIL AND KEROSENE IN-SPACE HEATERS ............................................................................. 34.3

ELECTRIC IN-SPACE HEATERS ........................................... 34.3 Radiant Heating Systems .......................................................... 34.4 SOLID-FUEL IN-SPACE HEATERS ....................................... 34.4 Fireplaces ................................................................................. 34.5 Stoves ........................................................................................ 34.5 GENERAL INSTALLATION PRACTICES ............................... 34.6

I

ers require an outdoor air intake. The size of the fresh air opening required is marked on the heater. To ensure adequate fresh air supply, unvented gas-heating equipment must, according to voluntary standards, include a device that shuts the heater off if the oxygen in the room becomes inadequate. Unvented room heaters may not be installed in hotels, motels, or rooms of institutions such as hospitals or nursing homes. Catalytic room heaters are fitted with fibrous material impregnated with a catalytic substance that accelerates the oxidation of a gaseous fuel to produce heat without flames. The design distributes the fuel throughout the fibrous material so that oxidation occurs on the surface area in the presence of a catalyst and room air. Catalytic heaters transfer heat by low-temperature radiation and by convection. The surface temperature is below a red heat and is generally below 650°C at the maximum fuel input rate. The flameless combustion of catalytic heaters is an inherent safety feature not offered by conventional flame-type gas-fueled burners. Catalytic heaters have also been used in agriculture and for industrial applications in combustible atmospheres. Unvented household catalytic heaters are used in Europe. Most of these are portable and mounted on casters in a casing that includes a cylinder of liquefied petroleum gas (LPG) so that they may be rolled from one room to another. LPG cylinders holding more than 0.9 kg of fuel are not permitted for indoor use in the United States. As a result, catalytic room heaters sold in the United States are generally permanently installed and fixed as wall-mounted units. Local codes and the National Fuel Gas Code (NFPA 54/ANSI Z223.1) should be reviewed for accepted combustion air requirements.

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N-SPACE heating equipment differs from central heating in that fuel is converted to heat in the space to be heated. In-space heaters may be either permanently installed or portable and may transfer heat by a combination of radiation, natural convection, and forced convection. The energy source may be liquid, solid, gaseous, or electric.

1.

GAS IN-SPACE HEATERS

Room Heaters A vented circulator room heater is a self-contained, freestanding, nonrecessed gas-burning appliance that furnishes warm air directly to the space in which it is installed, without ducting (Figure 1). It converts the energy in the fuel gas to convected and radiant heat without mixing flue gases and circulating heated air by transferring heat from flue gases to a heat exchanger surface. A vented radiant circulator is equipped with high-temperature glass panels and radiating surfaces to increase radiant heat transfer. Separation of flue gases from circulating air must be maintained. Vented radiant circulators range from 3 to 22 kW. Gravity-vented radiant circulators may also have a circulating air fan, but they perform satisfactorily with or without the fan. Fan-type vented radiant circulators are equipped with an integral circulating air fan, which is necessary for satisfactory performance. Vented room heaters are connected to a vent, chimney, or singlewall metal pipe venting system engineered and constructed to develop a positive flow to the outdoor atmosphere. Room heaters should not be used in a room that has limited air exchange with adjacent spaces because combustion air is drawn from the space. Unvented radiant or convection heaters range in size from 3 to 12 kW and can be freestanding units or wall-mounted, nonrecessed units of either the radiant or closed-front type. Unvented room heat-

Fig. 1 Room Heater The preparation of this chapter is assigned to TC 6.5, Radiant Heating and Cooling.

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Wall Furnaces A wall furnace is a self-contained vented appliance with grilles that are designed to be a permanent part of the structure of a building (Figure 2). It furnishes heated air that is circulated by natural or forced convection. A wall furnace can have boots, which may not extend more than 250 mm beyond the horizontal limits of the casing through walls of normal thickness, to provide heat to adjacent rooms. Wall furnaces range from 3 to 26 kW. Wall furnaces are classified as conventional or direct vent. Conventional vent units require approved B-1 vent pipes and are installed to comply with the National Fuel Gas Code. Some wall furnaces are counterflow units that use fans to reverse the natural flow of air across the heat exchanger. Air enters at the top of the furnace and discharges at or near the floor. Counterflow systems reduce heat stratification in a room. As with any vented unit, a minimum of inlet air for proper combustion must be supplied. Vented-recessed wall furnaces are recessed into the wall, with only the decorative grillwork extending into the room. This leaves more usable area in the room being heated. Dual-wall furnaces are two units that fit between the studs of adjacent rooms, thereby using a common vent. Both vented-recessed and dual-wall furnaces are usually natural convection units. Cool room air enters at the bottom and is warmed

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2020 ASHRAE Handbook—HVAC Systems and Equipment (SI) Table 1

Efficiency Requirements in the United States for Gas-Fired Direct Heating Equipment

Input, kW

Minimum AFUE, %

Wall Furnace (with fan) 12.3 75 >12.3 76 Wall Furnace (gravity type) 7.9 65 >7.9 to 13.5 66 >13.5 67

Input, kW

Minimum AFUE, %

Floor Furnace 10.8 >10.8

57 58

Room Heaters 5.9 >5.9 to 7.9 >7.9 to 13 >13.5

61 66 67 68

16, 2013 (Table 1).The AFUE is measured using the U.S. Department of Energy test method (DOE 2015) and must be met by manufacturers of direct heating equipment (i.e., gas-fired room heaters, wall furnaces, and floor furnaces). Fig. 2 Wall Furnace

1.1

CONTROLS

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Valves Gas in-space heaters are controlled by four types of valves:

Fig. 3 Floor Furnace as it passes over the heat exchanger, entering the room through the grillwork at the top of the heater. This process continues as long as the thermostat calls for the burners to be on. Accessory fans assist in the movement of air across the heat exchanger and help minimize air stratification. Direct-vent wall furnaces are constructed so that combustion air comes from the outdoors, and all flue gases discharge into the outdoor atmosphere. These appliances include grilles or the equivalent and are designed to be attached to the structure permanently. Direct-vent wall heaters are normally mounted on walls with outdoor exposure. Direct-vent wall furnaces can be used in extremely tight (wellinsulated) rooms because combustion air is drawn from outside the room. There are no infiltration losses for dilution or combustion air. Most direct-vent heaters are designed for natural convection, although some may be equipped with fans. Direct-vent furnaces are available with inputs of 2 to 19 kW.

Floor Furnaces Floor furnaces are self-contained units suspended from the floor of the heated space (Figure 3). Combustion air is taken from the outdoors, and flue gases are also vented outdoors. Cold air returns at the periphery of the floor register, and warm air comes up to the room through the center of the register.

U.S. Minimum Efficiency Requirements The National Appliance Energy Conservation Act (NAECA) of 1987 mandates that the U.S. Department of Energy establish minimum annual fuel utilization efficiency (AFUE) requirements for gas-fired direct heating equipment. The first minimums were established on January 1, 1990, and were most recently updated on April

The full on/off, single-stage valve is controlled by a wall thermostat. Models are available that are powered by a 24 V supply or from energy supplied by the heat of the pilot light on the thermocouple (self-generating). The two-stage control valve (with hydraulic thermostat) fires either at full input (100% of rating) or at some reduced step, which can be as low as 20% of the heating rate. The amount of time at the reduced firing rate depends on the heating load and the relative oversizing of the heater. The step-modulating control valve (with a hydraulic thermostat) steps on to a low fire and then either cycles off and on at the low fire (if the heating load is light) or gradually increases its heat output to meet any higher heating load that cannot be met with the low firing rate. This control allows an infinite number of fuel firing rates between low and high fire. The manual control valve is controlled by the user rather than by a thermostat. The user adjusts the fuel flow and thus the level of fire to suit heating requirements.

Thermostats Temperature controls for gas in-space heaters are of the following two types. • Wall thermostats are available in 24 V and millivolt systems. The 24 V unit requires an external power source and a 24 V transformer. Wall thermostats respond to temperature changes and turn the automatic valve to either full-on or full-off. The millivolt unit requires no external power because the power is generated by multiple thermocouples and may be either 250 or 750 mV, depending on the distance to the thermostat. This thermostat also turns the automatic valve to either full-on or full-off. • Built-in hydraulic thermostats are available in two types: (1) a snap-action unit with a liquid-filled capillary tube that responds to changes in temperature and turns the valve to either full-on or full-off; and (2) a modulating thermostat, which is similar to the snap-action unit, except that the valve comes on and shuts off at a preset minimum input. Temperature alters the input anywhere from full-on to the minimum input. When the heating requirements are satisfied, the unit shuts off.

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Residential In-Space Heating Equipment

34.3

Table 2 Gas Input Required for In-Space Supplemental Heaters Gas Consumption per Unit House Volume, W/m3 Steady State Average EffiAFUE, ciency, % %

Heater Type Vented Unvented Direct vent

54.6 90.5 76.0

73.1 90.5 78.2

Outdoor Air Temperature, °C Energy-Efficient Houseb

Older Bungalowa –15

0

10

–15

0

10

67.1 61.7 60.7

36.9 33.9 33.4

16.8 15.4 15.2

28.6 26.3 21.1

15.7 14.4 11.6

7.1 6.6 5.3

bungalow total heated volume = 193.3 m3 and U  1.7 to 2.8 W/(m2 ·K). energy-efficient house total heated volume = 333.7 m3 and U  1.1 to 1.7 W/(m2 ·K).

aTested

bTested

1.2

VENT CONNECTORS

Any vented gas-fired appliance must be installed correctly to vent combustion products. A detailed description of proper venting techniques is found in the National Fuel Gas Code and Chapter 35.

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1.3 SIZING UNITS The size of the unit selected depends on the size of the room, the number and direction of exposures, the amount of insulation in the ceilings and walls, and the geographical location. Heat loss requirements can be calculated from procedures described in Chapter 17 of the 2017 ASHRAE Handbook—Fundamentals. DeWerth and Loria (1989) studied the use of gas-fired, in-space supplemental heaters in two test houses. They proposed a heater sizing guide, which is summarized in Table 2. The energy consumption in Table 2 is for unvented, vented, and direct vent heaters installed in (1) a bungalow built in the 1950s with average insulation, and (2) a townhouse built in 1984 with above-average insulation and tightness.

2.

OIL AND KEROSENE IN-SPACE HEATERS

Vaporizing Oil Pot Heaters These heaters have an oil-vaporizing bowl (or other receptacle) that admits liquid fuel and air in controllable quantities; the fuel is vaporized by the heat of combustion and mixed with the air in appropriate proportions. Combustion air may be induced by natural draft or forced into the vaporizing bowl by a fan. Indoor air is generally used for combustion and draft dilution. Window-installed units have the burner section outdoors. Both natural- and forcedconvection heating units are available. A small blower is sold as an option on some models. The heat exchanger, usually cylindrical, is made of steel (Figure 4). These heaters are available as room units (both radiant and circulation), floor furnaces, and recessed wall heaters. They may also be installed in a window, depending on the cabinet construction. The heater is always vented to the outdoors. An 11 to 19 L fuel tank may be attached to the heater, or a larger outdoor tank can be used. Vaporizing pot burners are equipped with a single constant-level and metering valve. Fuel flows by gravity to the burner through the adjustable metering valve. Control can be manual, with an off pilot and variable settings up to maximum, or it can be thermostatically controlled, with the burner operating at a selected firing rate between pilot and high.

Fig. 4 Oil-Fueled Heater with Vaporizing Pot-Type Burner

Portable Kerosene Heaters Because kerosene heaters are not normally vented, precautions must be taken to provide sufficient ventilation. Kerosene heaters are of four basic types: radiant, natural-convection, direct-fired, forcedconvection, and catalytic. The radiant kerosene heater has a reflector, while the natural convection heater is cylindrical in shape. Fuel vaporizes from the surface of a wick, which is immersed in an integral fuel tank of up to 7.6 L capacity similar to that of a kerosene lamp. Fuel-burning rates range from about 1.5 to 6.5 kW. Radiant heaters usually have a removable fuel tank to facilitate refueling. The direct-fired, forced-convection portable kerosene heater has a vaporizing burner and a heat-circulating fan. These heaters are available with thermostatic control and variable heat output. The catalytic type uses a metal catalyst to oxidize the fuel. It is started by lighting kerosene at the surface; however, after a few moments, the catalyst surface heats to the point that flameless oxidation of the fuel begins.

3. ELECTRIC IN-SPACE HEATERS Wall, Floor, Toe Space, and Ceiling Heaters Heaters for recessed or surface wall mounting are made with open wire or enclosed, metal-sheathed elements. An inner liner or reflector is usually placed between the elements and the casing to promote circulation and minimize the rear casing temperature. Heat is distributed by both convection and radiation; the proportion of each depends on unit construction. Ratings are usually 1000 to 5000 W at 120, 208, 240, or 277 V. Models with air circulation fans are available. Other types can be recessed into the floor. Electric convectors should be placed so that air moves freely across the elements.

Powered Atomizing Heaters

Baseboard Heaters

Wall furnaces, floor furnaces, and freestanding room heaters are also available with a powered gun-type burner using No. 1 or No. 2 fuel oil. For more information, refer to Chapter 31.

These heaters consist of a metal cabinet containing one or more horizontal, enclosed, metal-sheathed elements. The cabinet is less than 150 mm in overall depth and can be installed 460 mm above the

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34.4

2020 ASHRAE Handbook—HVAC Systems and Equipment (SI)

floor; the ratio of the overall length to the overall height is more than two to one. Units are available from 610 to 3700 mm in length, with ratings from 300 to 1300 W/m, and they fit together to make up any desired continuous length or rating. Electric hydronic baseboard heaters containing immersion heating elements and an antifreeze solution are made with ratings of 300 to 2000 W. The placement of any type of electric baseboard heater follows the same principles that apply to baseboard installations (see Chapter 36) because baseboard heating is primarily perimeter heating.

3.1

RADIANT HEATING SYSTEMS

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Heating Panels and Heating Panel Sets These systems have electric resistance wire or etched or graphite elements embedded between two layers of insulation. High-density thermal insulation behind the element minimizes heat loss, and the outer shell is formed steel with baked enamel finish. Heating panels provide supplementary heating by convection and radiation. They can be recessed into or surface mounted on hard surfaces or fit in standard T-bar suspended ceilings. Units are usually rated between 250 and 1000 W in sizes varying from 610 by 610 mm to 610 by 2440 mm in standard voltages of 120, 208, 240, and 277 V.

Embedded Cable and Storage Heating Systems Ceiling and floor electric radiant heating systems that incorporate embedded cables are covered in Chapter 6. Electric storage systems, including room storage heaters and floor slab systems, are covered in Chapter 50.

Cord-Connected Portable Heaters Portable electric heaters are often used in areas that are not accessible to central heat. They are also used to maintain an occupied room at a comfortable level independent of the rest of the residence. Portable electric heaters for connection to 120 V, 15 A outlets are available with outputs of 600 to 1500 W, the most common being 1320 and 1500 W. Many heaters are available with a selector switch for three wattages (e.g., 1100-1250-1500 W). Heavy-duty heaters are usually connected to 240 V, 20 A outlets with outputs up to 4000 W, whereas those for connection to 240 V, 30 A outlets have outputs up to 5600 W. All electric heaters of the same wattage produce the same amount of heat. Portable electric heaters transfer heat by one of two predominant methods: radiation and convection. Radiant heaters provide heat for people or objects. An element in front of a reflector radiates heat outward in a direct line. Conventional radiant heaters have ribbon or wire elements. Quartz radiant heaters have coil wire elements

encased in quartz tubes. The temperature of a radiant wire element usually ranges between 650 and 870°C. Convection heaters warm the air in rooms or zones. Air flows directly over the hot elements and mixes with room air. Convection heaters are available with or without fans. The temperature of a convection element is usually less than 500°C. An adjustable, built-in bimetal thermostat usually controls the power to portable electric heaters. Fan-forced heaters usually provide better temperature control because the fan, in addition to cooling the case, forces room air past the thermostat. One built-in control uses a thermistor to signal a solid logic circuit that adjusts wattage and fan speed. Most quartz heaters use an adjustable control that operates the heater for a percentage of total cycle time from 0 (off) to 100% (full-on).

Controls Low-voltage and line-voltage thermostats with on-off operation are used to control in-space electric heaters. Low-voltage thermostats, operating at 30 V or less, control relays or contactors that carry the rated voltage and current load of the heaters. Because the control current load is small (usually less than 1 A), the small switch can be controlled by a highly responsive sensing element. Line-voltage thermostats carry the full load of the heaters at rated voltage directly through their switch contacts. Most switches carry a listing by Underwriters Laboratories (UL) at 22 A (resistive), 277 V rating. Most electric in-space heating systems are controlled by remote wall-mounted thermostats, but many are available with integral or built-in line-voltage thermostats. Most low-voltage and line-voltage thermostats use small internal heaters, either fixed or adjustable in heat output, that provide heat anticipation by energizing when the thermostat contacts close. The cycling rate of the thermostat is increased by the use of anticipation heaters, resulting in more accurate control of the space temperature. Droop is an apparent shift or lowering of the control point and is associated with line-voltage thermostats. In these thermostats, switch heating caused by large currents can add materially to the amount of droop. Most line-voltage thermostats in residential use control room heaters of 3 kW (12.5 A at 240 V) or less. At this moderate load and with properly sized anticipation heaters, the droop experienced is acceptable. Cycling rates and droop characteristics have a significant effect on thermostat performance.

4.

SOLID-FUEL IN-SPACE HEATERS

Most wood-burning and coal-burning devices, except central wood-burning furnaces and boilers, are classified as solid-fuel inspace heaters (see Table 3). An in-space heater can be either a fireplace or a stove.

Table 3 Solid-Fuel In-Space Heaters Type*

Approximate Efficiency,* %

Simple fireplaces, masonry or prefabricated

–10 to +10

High-efficiency fireplaces

Features

Advantages

Disadvantages

Open front. Radiates heat in one direction only.

Aesthetic.

Low efficiency. Heats only small areas.

25 to 45

Freestanding or built-in with glass doors, grates, ducts, and blowers.

Aesthetic. More efficient. Heats larger areas. Long service life. Maximum safety.

Medium efficiency.

Box stoves

20 to 40

Radiates heat in all directions.

Low initial cost. Heats large areas.

Fire hard to control. Short life. Wastes fuel.

Airtight stoves

40 to 55

Radiates heat in all directions. Sealed Good efficiency. Long burn times, Can create creosote problems. seams, effective draft control. high heat output. Longer service life.

High-efficiency catalytic wood heaters

65 to 75

Radiates heat in all directions. Sealed Highest efficiency. Long burn times, Creosote problems. High seams, effective draft control. high heat output. Long life. purchase price.

*Product categories are general; product efficiencies are approximate.

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Residential In-Space Heating Equipment 4.1

FIREPLACES

Simple Fireplaces Simple fireplaces, especially all-masonry and noncirculating metal built-in fireplaces, produce little useful heat. They lend atmosphere and a sense of coziness to a room. Freestanding fireplaces are slightly better heat producers. Simple fireplaces have an average efficiency of about 10%. In extreme cases, the chimney draws more heated air than the fire produces. The addition of glass doors to the front of a fireplace has both a positive and a negative effect. The glass doors restrict the free flow of indoor heated air up the chimney, but at the same time they restrict the radiation of the heat from the fire into the room.

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Factory-Built Fireplaces A factory-built fireplace consists of a fire chamber, chimney, roof assembly, and other parts that are factory made and intended to be installed as a unit in the field. These fireplaces have fireboxes of refractory-lined metal rather than masonry. Factory-built fireplaces come in both radiant and heat-circulating designs. Typical configurations are open-front designs, but corner-opening, three-sided units with openings either on the front or side; four-sided units; and seethrough fireplaces are also available. Radiant Design. The radiant system transmits heat energy from the firebox opening by direct radiation to the space in front of it. These fireplaces may also incorporate such features as an outdoor air supply and glass doors. Radiant-design factory-built fireplaces are primarily used for aesthetic wood burning and typically have efficiencies similar to those of masonry fireplaces (0 to 10%). Heat-Circulating Design. This unit transfers heat by circulating air around the fire chamber and releasing to the space to be heated. The air intake is generally below the firebox or low on the sides adjacent to the opening, and the heated air exits through grilles or louvers located above the firebox or high on the sides adjacent to it. In some designs, ducts direct heated air to spaces other than the area near the front of the fireplace. Some circulating units rely on natural convection, while others have electric fans or blowers to move air. These energy-saving features typically boosts efficiency 25 to 60%.

Freestanding Fireplaces Freestanding fireplaces are open-combustion wood-burning appliances that are not built into a wall or chase. One type of freestanding fireplace is a fire pit in which the fire is open all around; smoke rises into a hood and then into a chimney. Another type is a prefabricated metal unit that has an opening on one side. Because they radiate heat to all sides, freestanding fireplaces are typically more efficient than radiant fireplaces.

4.2

STOVES

Conventional Wood Stoves Wood stoves are chimney-connected, solid-fuel-burning room heaters designed to be operated with the fire chamber closed. They deliver heat directly to the space in which they are located. They are not designed to accept ducts and/or pipes for heat distribution to other spaces. Wood stoves are controlled-combustion appliances. Combustion air enters the firebox through a controllable air inlet; the air supply and thus the combustion rate are controlled by the user. Conventional controlled-combustion wood stoves manufactured before the mid-1980s typically have overall efficiencies ranging from 40 to 55%. Most controlled-combustion appliances are constructed of steel, cast iron, or a combination of the two metals; others are constructed of soapstone or masonry. Soapstone and masonry have lower thermal conductivities but greater specific heats (the amount of heat that can be stored in a given mass). Other materials such as special refractories and ceramics are used in low-emission appliances.

34.5 Wood stoves are classified as either radiant or convection (sometimes called circulating) heaters, depending on the way they heat interior spaces. Radiant wood stoves are generally constructed with single exterior walls, which absorb radiant heat from the fire. This appliance heats primarily by infrared radiation; it heats room air only to the extent that air passes over the hot surface of the appliance. Convection wood stoves have double vertical walls with an air space between the walls. The double walls are open at the top and bottom of the appliance to permit room air to circulate through the air space. The more buoyant hot air rises and draws in cooler room air at the bottom of the appliance. This air is then heated as it passes over the surface of the inner radiant wall. Some radiant heat from the inner wall is absorbed by the outer wall, but the constant introduction of room temperature air at the bottom of the appliance keeps the outer wall moderately cool. This characteristic generally allows convection wood stoves to be placed closer to combustible materials than radiant wood stoves. Fans in some wood stoves augment the movement of heated air. Convection wood stoves generally provide more even heat distribution than do radiant types.

Advanced-Design Wood Stoves Strict air pollution standards have prompted the development of new stove designs. These clean-burning wood stoves use either catalytic or noncatalytic technology to achieve very high combustion efficiency and to reduce creosote and particulate and carbon monoxide emission levels. Catalytic combustors are currently available as an integral part of many new wood-burning appliances and are also available as addon or retrofit units for most existing appliances. The catalyst may be platinum, palladium, rhodium, or a combination of these elements. It is bonded to a ceramic or stainless steel substrate. A catalytic combustor’s function in a wood-burning appliance is to substantially lower the ignition temperatures of unburned gases, solids, and/ or liquid droplets (from approximately 540°C to 260°C). As these unburned combustibles leave the main combustion chamber and pass through the catalytic combustor, they ignite and burn rather than enter the atmosphere. For the combustor to efficiently burn the gases, the proper amount of oxygen and a sufficient temperature to maintain ignition are required; further, the gases must have sufficient residence time in the combustor. A properly operating catalytic combustor has a temperature in the range of 540 to 930°C. Catalyst-equipped wood stoves have a default efficiency, as determined by the U.S. Environmental Protection Agency (EPA), of 72%, although many stoves are considerably more efficient. This EPA default efficiency is the value one standard deviation below the mean of the efficiencies from a database of stoves. Another approach to increasing combustion efficiency and meeting emissions requirements is the use of technologically advanced internal appliance designs and materials. Generally, noncatalytic, low-emission wood-burning appliances incorporate high-temperature refractory materials and have smaller fireboxes than conventional appliances. The fire chamber is designed to increase temperature, turbulence, and residence time in the primary combustion zone. Secondary air is introduced to promote continued burning of the gases, solids, and liquid vapors in a secondary combustion zone. Many stoves add a third and fourth burn area within the firebox. The location and design of the air inlets is critical because proper air circulation patterns are the key to approaching complete combustion. Noncatalytic wood stoves have an EPA default efficiency of 63%; however, many models approach 80%.

Fireplace Inserts Fireplace inserts are closed-combustion wood-burning room heaters that are designed to be installed in an existing masonry fireplace.

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2020 ASHRAE Handbook—HVAC Systems and Equipment (SI)

They combine elements of both radiant and convection wood stove designs. They have large radiant surfaces that face the room and circulating jackets on the sides that capture heat that would otherwise go up the chimney. Inserts may use either catalytic or noncatalytic technology to achieve clean burning.

Pellet-Burning Stoves

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Pellet-burning stoves burn small pellets made from wood byproducts rather than burning logs. An electric auger feeds the pellets from a hopper into the fire chamber, where air is blown through, creating very high temperatures in the firebox. The fire burns at such a high temperature that the smoke is literally burned up, resulting in a very clean burn, and no chimney is needed. Instead, the waste gases are exhausted to the outdoors through a vent. An air intake is operated by an electric motor; another small electric fan blows the heated air from the area around the fire chamber into the room. A microprocessor controls the operation, allowing the pellet-burning stove to be controlled by a thermostat. Pellet-burning stoves typically have the lowest emissions of all wood-burning appliances and have an EPA default efficiency of 78%. Because of the high air/fuel ratios used by pellet-burning stoves, these stoves are excluded from EPA wood stove emissions regulations.

5.

GENERAL INSTALLATION PRACTICES

The criteria to ensure safe operation are normally covered by local codes and ordinances or, in rare instances, by state and federal requirements. Most codes, ordinances, or regulations refer to the following building codes and standards for in-space heating: Building Codes BOCA/National Building Code CABO One- and Two-Family Dwelling Code International Building Code National Building Code of Canada Standard Building Code Uniform Building Code

BOCA CABO ICC NBCC SBCCI ICBO

Mechanical Codes National Mechanical Code Uniform Mechanical Code International Mechanical Code Standard Mechanical Code

BOCA IAPMO ICC SBCCI

Electrical Codes National Electrical Code Canadian Electrical Code

NFPA 70 CSA C22.1

Chimneys Chimneys, Fireplaces, Vents and Solid Fuel-Burning Appliances Chimneys, Factory-Built Residential Type and Building Heating Appliance Solid-Fuel Appliances Factory-Built Fireplaces Room Heaters, Solid-Fuel Type

NFPA 211 UL 103

Table 4 Chimney Connector Wall Thickness* Diameter

Gage

Less than 152 mm 152 mm to 254 mm 254 mm to 406 mm 406 mm or greater

26 24 22 16

Minimum Thickness, mm 0.48 0.58 0.74 1.42

*Do not use thinner connector pipe. Replace connectors as necessary. Leave at least 460 mm clearance between the connector and a wall or ceiling, unless the connector is listed for a smaller clearance or an approved clearance reduction system is used.

Safety with Solid Fuels The evacuation of combustion gases is a prime concern in the installation of solid-fuel-burning equipment. NFPA Standard 211, Chimneys, Fireplaces, Vents and Solid Fuel-Burning Appliances, lists requirements that should be followed. Because safety requirements for connector pipes (stovepipes) are not always readily available, these requirements are summarized as follows: • Connector pipe is usually black (or blue) steel single-wall pipe; thicknesses are shown in Table 4. Stainless steel is a corrosionresistant alternative that does not have to meet the thicknesses listed in Table 4. • Connectors should be installed with the crimped (male) end of the pipe toward the stove, so that creosote and water drip back into the stove. • The pipe should be as short as is practical, with a minimum of turns and horizontal runs. Horizontal runs should be pitched 20 mm per metre up toward the chimney. • Chimney connectors should not pass through ceilings, closets, alcoves, or concealed spaces. • When passing through a combustible interior or exterior wall, connectors must be routed through a listed wall pass-through that has been installed in accordance with the conditions of the listing, or they must follow one of the home-constructed systems recognized in NFPA Standard 211 or local building codes. Adequate clearance and protection of combustible materials is extremely important. In general, listed devices are easier to install and less expensive than home-constructed systems. Creosote forms in all wood-burning systems. The rate of formation is a function of the quantity and type of fuel burned, the appliance in which it is burned, and the manner in which the appliance is operated. Thin deposits in the connector pipe and chimney do not interfere with operation, but thick deposits (greater than 6 mm) may ignite. Inspection and cleaning of chimneys connected to wood-burning appliances should be performed on a regular basis (at least annually). Only the solid fuel that is listed for the appliance should be burned. Coal should be burned only in fireplaces or stoves designed specifically for coal burning. The chimney used in coal-fired applications must also be designed and approved for coal and wood. Solid-fuel appliances should be installed in strict conformance with the clearance requirements established as part of their safety listing. When clearance reduction systems are used, stoves must remain at least 300 mm and connector pipe at least 150 mm from combustibles, unless smaller clearances are established as part of the listing.

Utility-Furnished Energy UL 127 UL 1482

Chapter 51 has further information, including the names and addresses of these agencies. Safety and performance criteria are furnished by the manufacturer.

Those systems that rely on energy furnished by a utility are usually required to comply with local utility service rules and regulations. The utility usually provides information on the installation and operation of the equipment using their energy. Bottled gas (LPG) equipment is generally listed and tested under the same standards as natural gas. LPG equipment may be identical to natural gas equipment, but it always has a different orifice and sometimes has a dif-

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Residential In-Space Heating Equipment ferent burner and controls. The listings and examinations are usually the same for natural, mixed, manufactured, and liquid petroleum gas.

Products of Combustion The combustion chamber of equipment that generates products of combustion must be connected by closed piping to the outdoors. Gas-fired equipment may be vented through masonry stacks, chimneys, specifically designed venting, or, in some cases, venting incorporating forced- or induced-draft fans. Chapter 35 covers chimneys, gas vents, and fireplace systems in more detail.

Agency Testing The standards of several agencies contain guidelines for the construction and performance of in-space heaters. The following list summarizes the standards that apply to residential in-space heating; they are coordinated or sponsored by ASHRAE, the American National Standards Institute (ANSI), Underwriters Laboratories (UL), the American Gas Association (AGA), and the Canadian Gas Association (CGA). Some CGA standards have a CAN1 prefix.

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ANSI Z21.11.1 ANSI Z21.11.2 ANSI Z21.44 ANSI Z21.48 ANSI Z21.49 ANSI Z21.60/ CSA 2.26-M96 ANSI Z21.76 ANSI Z21.50/ CSA 2.22-M98 ANSI Z21.86/ CSA 2.32-M98 ANSI Z21.88/ CSA 2.33-M98 CAN1-2.1-M86

Gas-Fired Room Heaters, Vented Gas-Fired Room Heaters, Unvented Gas-Fired Gravity and Fan-Type Direct-Vent Wall Furnaces Gas-Fired Gravity and Fan-Type Floor Furnaces Gas-Fired Gravity and Fan-Type Vented Wall Furnaces Decorative Gas Appliances for Installation in Solid-Fuel Burning Fireplaces Gas-Fired Unvented Catalytic Room Heaters for Use with Liquefied Petroleum (LP) Gases Vented Gas Fireplaces Vented Gas-Fired Space Heating Appliances Vented Gas Fireplace Heaters Gas-Fired Vented Room Heaters

34.7 CAN/CGA-2.5Gas-Fired Gravity and Fan Type Vented Wall M86 Furnaces CAN1/CGA-2.19- Gas-Fired Gravity and Fan Type Direct Vent M81 Wall Furnaces NFPA 211 Chimneys, Fireplaces, Vents and Solid FuelBurning Appliances NFPA/AGA 54 National Fuel Gas Code ANSI/UL 127 Factory-Built Fireplaces UL 574 Electric Oil Heaters UL 647 Unvented Kerosene-Fired Heaters and Portable Heaters ANSI/UL 729 Oil-Fired Floor Furnaces ANSI/UL 730 Oil-Fired Wall Furnaces UL 737 Fireplace Stoves ANSI/UL 896 Oil-Burning Stoves ANSI/UL 1042 Electric Baseboard Heating Equipment ANSI/UL 1482 Heaters, Room Solid-Fuel Type ASHRAE 62.1 Ventilation for Acceptable Indoor Air Quality

REFERENCES ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal.ashrae .org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org/bookstore. DeWerth, D.W., and R.L. Loria. 1989. In-space heater energy use for supplemental and whole house heating. ASHRAE Transactions 95(1). DOE. 2015. Uniform test method for measuring the energy consumption of vented home heating equipment. Federal Register 80:3, 782 (January).

BIBLIOGRAPHY AHRI. 2018. Directory of direct heating equipment. Air-conditioning, Heating, and Refrigeration Institute, Arlington, VA.www.ahridirectory.org MacKay, S., L.D. Baker, J.W. Bartok, and J.P. Lassoie. 1985. Burning wood and coal. Natural Resources, Agriculture, and Engineering Service, Cornell University, Ithaca, NY. Wood Heating Education and Research Foundation. 1984. Solid fuel safety study manual for Level I solid fuel safety technicians. Washington, D.C.

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Related Commercial Resources CHAPTER 35

CHIMNEY, VENT, AND FIREPLACE SYSTEMS Terminology ............................................................................. 35.1 Draft Operating Principles ...................................................... 35.1 Chimney Functions .................................................................. 35.2 Steady-State Chimney Design Equations................................. 35.3 Steady-State Chimney Design Graphical Solutions ........................................................... 35.11 Vent and Chimney Capacity Calculation Examples .............. 35.12 Gas Appliance Venting .......................................................... 35.17 Oil-Fired Appliance Venting ................................................. 35.19

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A

PROPERLY designed chimney or vent system provides and controls draft to convey flue gas from an appliance to the outdoors. This chapter describes the design of chimneys and vent systems that discharge flue gas from appliances and fireplace systems. Sustainability. Good chimney and vent design is not only a safety issue, but also can enhance a building’s sustainability. This chapter explains how to design vent systems to optimize and minimize the materials used to construct fuel-burning appliance vents and chimneys for low cost and long reliability, reducing the need for vent or chimney replacement, thus saving natural resources. Also, systems designed to bring outdoor air directly into the appliance space for combustion and vent gas dilution, instead of relying on air infiltration into the building, reduce heat load and conserve fuel.

1.

TERMINOLOGY

In this chapter, appliance refers to any furnace, boiler, or incinerator (including the burner). Unless the context indicates otherwise, the term chimney includes specialized vent products such as masonry, metal, and factory-built chimneys; single-wall metal pipe; type B gas vents; special gas vents; or masonry chimney liner systems (NFPA Standard 211). Draft is negative static pressure, measured relative to atmospheric pressure; thus, positive draft is negative static pressure. Flue gas is the mixture of gases discharged from the appliance and conveyed by the chimney or vent system. Appliances can be grouped by draft conditions at the appliance flue gas outlet as follows (Stone 1971): 1. Those that require draft applied at the appliance flue gas outlet to induce air into the appliance 2. Those that operate without draft applied at the appliance flue gas outlet (e.g., a gas appliance with a draft hood in which the combustion process is isolated from chimney draft variations) 3. Those that produce positive pressure at the appliance flue gas outlet collar so that no chimney draft is needed; appliances that produce some positive outlet pressure but also need some chimney draft In the first two configurations, hot flue gas buoyancy, induceddraft chimney fans, or a combination of both produces draft. The third configuration may not require chimney draft, but it should be considered in the design if a chimney is used. If the chimney system is undersized, draft inducers in the connector or chimney may supply draft needs. If the connector or chimney pressure requires control for proper operation, draft control devices must be used. The preparation of this chapter is assigned to TC 6.10, Fuels and Combustion.

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Fireplace Chimneys ............................................................... Air Supply to Fuel-Burning Appliances ................................. Vent and Chimney Materials.................................................. Vent and Chimney Accessories .............................................. Draft Fans .............................................................................. Terminations: Caps and Wind Effects.................................... Codes and Standards.............................................................. Symbols ..................................................................................

2.

35.21 35.26 35.26 35.28 35.29 35.30 35.33 35.33

DRAFT OPERATING PRINCIPLES

Available draft Da is the draft supplied by the vent system, available at the appliance flue gas outlet. It can be shown as Da = Dt – p – Dp + Db

(1)

where Da Dt p Dp Db

= = = = =

available draft, Pa theoretical draft, Pa flow losses, Pa depressurization, Pa boost (increase in static pressure by fan), Pa

This equation can account for a nonneutral (nonzero) pressure difference between the space surrounding the appliance or fireplace and the atmosphere. If the surrounding space is at a lower pressure than the atmosphere (space depressurized), the pressure difference Dp should also be subtracted from Dt when calculating available draft Da , and vice versa. This equation applies to all three appliance draft conditions at the vent system inlets; for example, in the second condition with zero draft requirement at the appliance outlet, available draft required is zero, so theoretical draft of the chimney equals the flow resistance, if no depressurization or boost is present. Satisfactory operation of the appliance and vent system depends on Da staying within limits. If Da is too small and the appliance design relies on Da to draw combustion air into the burner system, incomplete combustion, flame rollout, and/or flue gas spillage can occur at the appliance. Even where combustion air is mechanically supplied to the burner, inadequate Da can result in portions of the appliance and/or vent system operating under positive pressure. If those sections are not designed to be pressurized, flue gas can leak into the building. Excessive Da can cause excessive combustion air to be drawn into the burner system, which not only reduces combustion efficiency, but can also cause flame lifting or other detriments to combustion performance. Many vent systems limit Da using a draft hood or draft regulator (integral to the appliance or field supplied); these devices limit draft by admitting dilution air into the vent system. The resulting increased flow through the vent system increases P and reduces Da. Such draft control devices are unsuitable for use in parts of vent systems that may operate with a negative Da (positive pressure). Various methods can be used identify the Da limits within which an appliance can properly operate. Often, the appliance manufacturer publishes a Da or range of Das where the appliance is designed to operate. In other cases, acceptable values for Da are not explicitly specified, but the manufacturer or installation code prescribes a vent system design in terms of length, chimney height, cross-sectional area, etc, that is known to provide the required Da. An example of

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35.2

2020 ASHRAE Handbook—HVAC Systems and Equipment (SI)

the latter approach is that for Category I gas appliances (see the section on Gas Appliance Venting). Theoretical draft Dt is the natural draft produced by the buoyancy of hot gases in the chimney relative to cooler gases in the surrounding atmosphere. It depends on chimney height, local barometric pressure, and the mean chimney flue gas temperature difference tm, which is the difference in temperature between the flue gas and atmospheric gases. Therefore, cooling by heat transfer through the chimney wall is a key variable in chimney design. Precise evaluation of theoretical draft is not necessary for most design calculations because of the availability of chimney design charts, computer programs, and capacity tables in the references, building codes, and vent and appliance manufacturers’ data sheets. Chimney temperatures and acceptable combustible material temperatures must be known in order to determine safe clearances between the chimney and combustible materials. Safe clearances for some chimney systems, such as type B gas vents, are determined by standard tests and/or specified in building codes. Losses from Flow p represent the friction losses imposed on the flue gas by flow resistance through the chimney. Depressurization Dp is negative pressure in the space surrounding the appliance with respect to the atmosphere into which the chimney discharges. Dp can be caused by other appliances and fans operating in the building that remove or add air to the space surrounding the appliance, by building stack effect, by outdoor atmospheric effects such as wind impacting the chimney exit or the side of the building facing or leeward to the wind, and other building phenomena. Boost Db is the pressure boost from a mechanical-draft fan. A chimney with an forced-draft fan at the inlet of the chimney would have positive boost (increased static pressure). A chimney with an induced-draft fan at the outlet of the chimney would have negative boost (decreased static pressure). The following sections cover the basis of chimney design for average steady-state appliance operating conditions, where significant condensation of flue products is not expected. For other appliance operating conditions, a rigorous cyclic evaluation of the flue gas and material surface temperatures in the chimney vent system can be obtained using the VENT-II computer program (GTI 2009). For oil-fired appliances, chimney flue gas and material surface temperature evaluations can be obtained using the OHVAP computer program (Krajewski 1996).

3.

CHIMNEY FUNCTIONS

The proper chimney can be selected by evaluating factors such as draft, configuration, size, and operating conditions of the appliance; construction of surroundings; appliance usage classification; residential, low, medium, or high heat (NFPA Standard 211); and building height. The chimney designer should know the applicable codes and standards to ensure acceptable construction. In addition to chimney draft, the following factors must be considered for safe and reliable operation: adequate air supply for combustion; building depressurization effects; draft control devices; chimney materials (corrosion and temperature resistance); flue gas temperatures, composition, and dew point; wind eddy zones; and particulate dispersion. Chimney materials must resist oxidation and condensation at both high and low fire levels at all design temperatures.

Start-Up The equations and design charts in this chapter may be used to determine vent or chimney size for average vent system operating conditions where significant condensation of flue products is not expected. The equations and charts do not consider modulation, cycling, or time to achieve equilibrium flow conditions from a cold

start. Whereas mechanical draft systems can start gas flow, gravity systems rely on the buoyancy of hot flue gases as the sole force to displace the cold air in the chimney. Priming follows Newton’s laws of motion. The time to fill a system with hot flue gases, displace the cold air, and start flow is reasonably predictable and is usually a minute or less; however, unfavorable thermal differentials, building/ chimney interaction, mechanical equipment (e.g., exhaust fans), or wind forces that oppose the normal flow of vent gases can overwhelm the buoyancy force. Then, rapid priming cannot be obtained solely from correct system design. The VENT-II computer program contains detailed analysis of gas vent and chimney priming and other cold-start considerations and allows for appliance cycling and pressure differentials that affect performance (GTI 2009).

Air Intakes All rooms or spaces containing fuel-burning appliances must have a constant supply of combustion air from outdoors (either directly or indirectly) at adequate static pressure to ensure proper combustion. In addition, air (either directly or indirectly) is required to replace the air entering chimney systems through draft hoods and barometric draft regulators and to ventilate closely confined boiler and furnace rooms. The ANSI-approved U.S. National Fuel Gas Code (NFPA Standard 54/AGA Z223.1) and the Canadian Natural Gas and Propane Installation Code (CSA Standard B149.1), along with appliance manufacturers, provide requirements for air openings. Any design must consider flow resistance of the air supply, including registerlouver resistance, air duct resistance, and air inlet terminations. Compliance with these codes or the appliance manufacturers’ instructions accounts for the air supply flow resistance.

Vent Size Small residential and commercial natural-draft gas appliances need vent diameters of 75 to 300 mm. U.S. and Canadian codes recommend sizes or input capacities for most acceptable gas appliance venting materials. These sizes also apply to gas appliances with integral automatic vent dampers, as well as to appliances with fieldinstalled automatic vent dampers. Field-installed automatic vent dampers should be listed for use with a specific appliance by a recognized testing agency and installed by qualified installers.

Draft Control Pressure, temperature, and other draft controls have replaced draft hoods in many residential furnaces and boilers to attain higher steady-state and seasonal efficiencies. Appliances that use pulse combustion or forced- or induced-draft fans, as well as those designed for sealed combustion or direct venting, do not have draft hoods but may require special venting and special vent terminals. If fan-assisted burners deliver fuel and air to the combustion chamber and also overcome the appliance flow resistance, draft hoods or other control devices may be installed, depending on the design of the appliance. Many appliances do not use draft hoods; in such cases, the listed appliance manufacturer’s vent system design requirements should be followed. The section on Vent and Chimney Accessories has information on draft hoods, barometric regulators, draft fans, and other draft control devices. Frequently, a chimney must produce excess flow or draft. For example, dangerously high flue gas outlet temperatures from an incinerator (or normal noncondensing appliance flue gas temperatures) may be reduced by diluting the flue gas with air in the chimney (or PVC vent pipe) by applying excess draft.

Pollution Control Where control of pollutant emissions is impossible, the chimney should be tall enough to ensure dispersion over a wide area to prevent objectionable ground-level concentrations. The chimney can

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Chimney, Vent, and Fireplace Systems also serve as a passageway to carry flue gas to pollution control equipment. This passageway must meet the building code requirements for a chimney, even at the exit of pollution control equipment, because of possible exposure to heat and corrosion. A bypass chimney should also be provided to allow continued exhaust in the event of pollution control equipment failure, repair, or maintenance.

Equipment Location Chimney materials may allow installation of appliances at intermediate or all levels of a high-rise building without imposing mass penalties. Some gas vent systems allow individual apartment-byapartment heating systems.

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Wind Effects Wind and eddy currents affect discharge of gases from vents and chimneys. A vent or chimney must expel flue gas beyond the cavity or eddy zone surrounding a building to prevent reentry through openings and fresh air intakes. A chimney and its termination can stabilize the effects of wind on appliances and their equipment rooms. In many locations, the equipment room air supply is not at neutral pressure under all wind conditions. Locating the chimney outlet well into the undisturbed wind stream and away from the cavity and wake zones around a building both counteracts wind effects on the air supply pressure and prevents reentry through openings and contamination of fresh air intakes. Chimney outlets below parapet or eaves level, nearly flush with the wall or roof surface, or in known regions of stagnant air may be subjected to downdrafts and are undesirable. Caps for downdraft and rain protection must be installed according to either their listings and the cap manufacturer’s instructions or the applicable building code. Wind effects can be minimized by locating the chimney terminal and the combustion air inlet terminal close together in the same pressure zone while taking care to minimize recirculation of combustion products into the combustion air inlet. See Chapter 24 of the 2017 ASHRAE Handbook—Fundamentals, for more information on wind effects.

Safety Factors Safety factors allow for uncertainties of vent and chimney operation. For example, flue gas must not spill from a draft hood or barometric regulator, even when the chimney has very low available draft. The Table 2 design condition for natural gas vents (i.e., vent gas at 165 K rise and 5.3% CO2 concentration) allows gas vents to operate with reasonable safety above or below the suggested temperature and CO2 limits. Safety factors may also be added to the system friction coefficient to account for a possible extra fitting, soot accumulation, and air supply resistance. The specific gravity of flue gas can vary depending on the fuel burned. Natural gas flue gas, for example, has a density as much as 5% less than air, whereas coke flue gas may be as much as 8% greater. However, these density changes are insignificant relative to other uncertainties, so no compensation is needed.

4.

STEADY-STATE CHIMNEY DESIGN EQUATIONS

Chimney design balances forces that produce flow against those that retard flow (e.g., friction). Theoretical draft is the pressure that produces flow in gravity or natural-draft chimneys. It is defined as the static pressure resulting from the difference in densities between a stagnant column of hot flue gas and an equal column of ambient air. In the design or balancing process, theoretical draft may not equal friction loss because the appliance is frequently built to operate with some specific pressure (positive or negative) at the appliance flue gas exit. This exit pressure is added to the available draft,

35.3 which depends on chimney conditions, appliance operating characteristics, fuel, and type of draft control. Flow losses caused by friction may be estimated by several formulas for flow in pipes or ducts, such as the equivalent length method or the loss coefficient (velocity head) method. Chapter 13 of the 2017 ASHRAE Handbook—Fundamentals covers computation of flow losses. This chapter emphasizes the loss coefficient method because fittings usually cause the greater portion of system pressure drop in chimney systems, and conservative loss coefficients (which are almost independent of piping size) provide an adequate basis for design. A computer program called VENT-II (GTI 2009) dynamically predicts cyclic flows, temperatures, and pressures in venting systems. Similarly, Krajewski (1996) developed a computer program entitled OHVAP: Oil-Heat Vent Analysis Program. For large gravity chimneys, available draft Da may be calculated from Equation (1).The following equations may be derived and expressed in a form that is more readily applied to the problems of chimney design, size, and capacity by considering the following factors, which are the steps used to solve the problems in the section on Vent and Chimney Capacity Calculation Examples: 1. 2. 3. 4. 5. 6. 7. 8.

Mass flow of combustion products in chimney Chimney gas temperature and density Theoretical draft System pressure loss caused by flow Available draft Chimney gas velocity System resistance coefficient Final input-volume relationships

For application to system design, the chimney gas velocity step is eliminated; however, actual velocity can be found readily, if needed.

Mass Flow of Combustion Products in Chimneys and Vents Mass flow in a chimney or venting system may differ from that in the appliance, depending on the type of draft control or number of appliances operating in a multiple-appliance system. Mass flow is preferred to volumetric flow because it remains constant in any continuous portion of the system, regardless of changes in temperature or pressure. For the chimney gases resulting from any combustion process, mass flow can be expressed as w = IM (2) where w = mass flow rate, kg/s I = appliance heat input, MJ/s (MW) M = ratio of mass flow to heat input, kg of combustion products per MJ of fuel burned (M depends on fuel composition and percentage excess air [or CO2] in the chimney)

Volumetric flow rate (m3/min) Q can be found as follows: w Q = 60 -----m

(3)

where m is gas density (kg/m3). In chimney system design, the composition and flow rate of the flue gas must be assumed to determine the ratio M of mass flow to input. When combustion conditions are given in terms of excess air, Figure 1 can be used to estimate CO2. The mass flow equations in Table 1 illustrate the influence of fuel composition; however, additional guidance is needed for system design. The information provided with many heat-producing appliances is limited to whether they have been tested, certified, listed, or approved to comply with applicable standards. From this information and from the type of fuel and draft control, certain inferences can be drawn regarding the flue gas.

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35.4

2020 ASHRAE Handbook—HVAC Systems and Equipment (SI) Table 1 Mass Flow Equations for Common Fuels Ratio M of Mass Flow to

Fuel Natural gas LPG (propane, butane, or mixture) No. 2 oil (light) No. 6 oil (heavy) Bituminous coal (soft) Type 0 waste or wood

M=

Inputa

kg Total Combustion Productsb MJ Fuel Input

 10.72  0.303  0.159 + -----------------  % CO 2    12.61  0.304  0.144 + -----------------  % CO 2    14.4  0.31  0.12 + -----------------  % CO 2    15.8  0.31  0.12 + -----------------  % CO 2  

 18.2  0.327  0.11 + -----------------  % CO 2    19.7  0.30  0.16 + -----------------  % CO 2  

aPercent

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CO2 is determined in combustion products with water condensed (dry basis). bTotal combustion products include combustion products and excess air.

vary with configuration and appliance design and are not necessarily the same as boiler or appliance outlet conditions. Mass flow in incinerator chimneys must account for the probable heat value of the waste, its moisture content, and use of additional fuel to initiate or sustain combustion. Classifications of waste and corresponding values of M in Table 3 are based on recommendations from the Incinerator Institute of America. Combustion data given for types 0, 1, and 2 waste do not include any additional fuel. Where constant burner operation accompanies waste combustion, the additional quantity of products should be considered in the chimney design. The system designer should obtain exact outlet conditions for the maximum rate operation of the specific appliance. This information can reduce chimney construction costs. For appliances with higher seasonal or steady-state efficiencies, however, special attention should be given to the manufacturer’s venting recommendations because the flue gas may differ in composition and temperature from conventional values.

Mean Chimney Gas Temperature and Density Chimney gas temperature depends on the fuel, appliance, draft control, chimney size, and configuration. Density of gas within the chimney and theoretical draft both depend on gas temperature. Although the gases flowing in a chimney system lose heat continuously from entrance to exit, a single mean gas temperature must be used in either the design equation or the chart. Mean chimney gas density is virtually the same as air density at the same temperature. Thus, density may be found as  Ts B  B m = a ------  ------ = 0.00348 -----Tm T m B o

(4)

where m Ts a B Tm Bo

Fig. 1 Graphical Evaluation of Rate of Vent Gas Flow from Percent CO2 and Fuel Rate Table 2 suggests typical values for the vent or chimney systems for gaseous and liquid fuels when specific outlet conditions for the appliance are not known. If a gas appliance with draft hood is used, Table 2 recommends that dilution air through the draft hood reduce the CO2 percentage to 5.3%. For appliances using draft regulators, the dilution and temperature reduction is a function of the draft regulator gate opening, which depends on excess draft. If the chimney system produces the exact draft necessary for the appliance, little dilution takes place. For manifolded gas appliances that have draft hoods, dilution through draft hoods of inoperative appliances must be considered in precise system design. However, with forced-draft appliances having wind box or inlet air controls, dilution through inoperative appliances may be unimportant, especially if pressure at the outlet of inoperative appliances is neutral (atmospheric level). Figure 2 can be used to estimate mass and volumetric flow. Flow conditions in the chimney connector, manifold, vent, and chimney

= = = = = =

chimney gas density, kg/m3 standard temperature = 288.15 K air density at Ts and Bo =1.225 kg/m3 local barometric pressure, Pa mean chimney gas temperature at average system conditions, K standard pressure = 101 325 Pa

The density a in Equation (4) is a compromise value for typical humidity. The subscript m for density and temperature requires that these properties be calculated at mean gas temperature or vertical midpoint of a system (inlet conditions can be used where temperature drop is not significant). Using a reasonably high ambient (e.g., 15°C) for design ensures improved operation of the chimney when ambient temperatures drop because temperature differentials and draft increase. A design requires assuming an initial or inlet chimney gas temperature. In the absence of specific data, Table 4 provides a conservative temperature. For appliances that can operate over a range of temperatures, size should be calculated at both extremes to ensure an adequate chimney. The drop in vent gas temperature from appliance to exit reduces capacity, particularly in sizes of 300 mm or less. In gravity type B gas vents, which may be as small as 75 mm in diameter, and in other systems used for venting gas appliances, capacity is best determined from NFPA Standard 54/AGA Z223.1 or CSA Standard B149.1. In these codes, the tables compensate for the particular characteristics of the chimney material involved, except for very high single-wall metal pipe. Between 300 and 460 mm diameters, the effect of heat loss diminishes greatly because there is greater gas flow relative to system surface area. For 500 mm and greater diameters, cooling has little effect on final size or capacity.

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35.5

Table 2 Typical Chimney and Vent Design Conditionsa Mass Flow/Input Ratio (M), Flue Gas Flow Rate/Unit Heat Input,c Density,c m3/s per (MJ/s) or m3/MJ kg Total Flue Gasb per MJ Fuel Input kg/m3 at Flue Gas Temperature

Fuel

Appliance

% CO2

Temperature Rise, K

Natural gas Propane gas Natural gas Low-efficiency High-efficiency No. 2 oil Oil

Draft hood Draft hood

5.3 6.0

165 165

0.661 0.683

0.777 0.777

0.851 0.878

8.0 7.0 9.0 13.5

220 135 280 165

0.454 0.512 0.533 0.368

0.693 0.832 0.620 0.777

0.655 0.615 0.860 0.473

9.0

740

0.705

0.343

2.056

Waste, Type 0

No draft hood No draft hood Residential Forced-draft, over 120 kW or 0.120 MJ/s Incinerator

aValues

are for appliances with flue losses of 17% or more. For appliances with lower flue losses (high-efficiency types), see appliance installation instructions or ask manufacturer for operating data.

bTotal

flue gas includes combustion products and excess air. cAt sea level and 15.56°C ambient temperature.

Table 3 Mass Flow for Incinerator Chimneys

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Combustion Products Auxiliary a kg M, kg Type Heat Value Fuel per L/s per kg per kg Products of of Waste,a Unit Waste, Waste at MJ/kg Wastec per MJ Waste MJ/kg 760°Cb 0 1 2 3 4 aAuxiliary

19.8 15.1 10.0 5.8 2.3

0 0 0 3.5 7.0

11.17 8.74 6.18 5.12 4.31

13.76 10.80 7.68 6.25 5.33

0.697 0.714 0.770 1.075 2.292

fuel may be used with any type of waste, depending on incinerator design.

bSpecialized units may produce higher or lower outlet gas temperatures, which must be

considered in chimney sizing, using Equation (12) or Equation (21). cMultiply these values by kilograms of waste burned per second to establish mass flow.

Table 4 Mean Chimney Gas Temperature for Various Appliances Mean Temperature tm* in Chimney, °C

Appliance Type Natural gas-fired heating appliance with draft hood (low-efficiency) LP gas-fired heating appliance with draft hood (low-efficiency) Gas-fired heating appliance, no draft hood Low-efficiency Mid-efficiency

180

240 150

Oil-fired heating appliance (low-efficiency)

290

180

Conventional incinerator

760

Controlled air incinerator

980 to 1320

Pathological incinerator

980 to 1540

Turbine exhaust

480 to 760

Diesel exhaust

480 to 760

Ceramic kiln

980 to 1320

*Subtract 15°C ambient to obtain temperature rise.

A straight vertical vent or chimney directly off the appliance requires little compensation for cooling effects, even with smaller sizes. However, a horizontal connector running from the appliance to the base of the vent or chimney has enough heat loss to diminish draft and capacity. Figure 3 is a plot of temperature correction Cu, which is a function of connector size, length, and material for either conventional single-wall metal connectors or double-wall metal connectors. For a more accurate method of calculating temperature loss in vents and chimneys, including effects from condensation, see Rutz et al. (1991).

Fig. 2 Flue Gas Mass and Volumetric Flow To use Figure 3, estimate connector size and length and read the temperature multiplier. For example, a 5 m length of 175 mm diameter single-wall connector has a multiplier of 0.59. If inlet temperature rise te above ambient is 165 K, operating mean temperature rise tm will be 0.59  165 = 97 K. This factor adequately corrects the temperature at the midpoint of the vertical vent for heights up to 30 m. The correction procedure includes the assumption that the overall heat transfer coefficient of a vertical chimney is approximately 3.4 W/(m2 ·K) or less (the value for double-wall metal). This procedure does not correct for cooling in very high stacks constructed entirely of single-wall metal, especially those exposed to cold ambient temperatures. For severe exposures or excessive heat loss, a trial

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35.6

2020 ASHRAE Handbook—HVAC Systems and Equipment (SI) Table 5 Overall Heat Transfer Coefficients of Various Chimneys and Vents

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U,

W/(m2 · K)*

Material

Observed

Design

Industrial steel stacks Clay or iron sewer pipe Asbestos-cement gas vent Black or painted steel stove pipe

— 7.4 to 7.9

7.4 7.4

4.09 to 8.06

6.8

—

6.8

Single-wall galvanized steel

1.76 to 7.84

5.7

Single-wall unpainted pure aluminum Brick chimney, tile-lined

—

5.7

2.8 to 5.7

5.7

2.1 to 5.9

3.4

1.9 to 4.0

2.3

1.9 to 4.0

1.7

Double-wall gas vent, 6.4 mm air space Double-wall gas vent, 12.7 mm air space Insulated prefabricated chimney

Remarks Under wet wind Used as single-wall material Tested per UL Standard 441 Comparable to weathered galvanized steel Depends on surface condition and exposure No. 1100 or other bright-surface aluminum alloy For gas appliances in residential construction per NFPA Standard 211 Galvanized steel outer pipe, pure aluminum inner pipe; tested per UL Standard 441 Solid insulation meets UL Standard 103 when chimney is fully insulated

*U-factors based on inside area of chimney.

 L m t m t q - ------ = --------e = exp  Ud i --------------qm  t m qm 

(6)

 L m t m t q - ------ = --------e = exp  Ud f --------------qm  t m qm 

(6)

where A cp di exp x g H k

= = = = = = =

Lm = q qm te tm T Tm To Ts U m

= = = = = = = = = =

area of passage cross section, m2 specific heat inside diameter, m ex gravitational constant = 9.806 65 m/s2 height of chimney above grade or inlet, m fixed, dimensionless system flow resistance coefficient for pipe and fittings length from inlet to location (in the vertical) of mean gas temperature, m heat flow rate at vent inlet, J/s (W) heat flow rate at midpoint of vent, J/s (W) (T – To) = temperature difference entering system, K (Tm – To) = chimney gas mean temperature rise, K flue gas temperature at vent inlet, °C chimney mean flue gas temperature, K design ambient temperature, 288 K 288.15 K heat transfer coefficient, W/(m2 ·K) mean flue gas density from Equation (4)

Assuming reasonable constancy of U , the overall heat transfer coefficient of the venting system material, Equations (5) and (6) provide a solution for maximum vent gas capacity. They can also be used to develop cooling curves or calculate the length of pipe where internal moisture condenses. Kinkead (1962) details methods of solution and application to both individual and combined gas vents.

Theoretical Draft The theoretical draft of a gravity chimney or vent is the difference in weight between a given column of warm (light) chimney gas and an equal column of cold (heavy) ambient air. Chimney gas density or temperature, chimney height, and barometric pressure determine theoretical draft; flow is not a factor. The equation for theoretical draft assumes chimney gas density is the same as that of air at the same temperature and pressure; thus, 1 1 Dt = 0.03413BH  ----– ------  T T  o m

(7)

where

Fig. 3

Temperature Multiplier Cu for Compensation of Heat Losses in Connector

calculation assuming a conservative operating temperature shows whether capacity problems will be encountered. For more precise heat loss calculations, Table 5 suggests overall heat transfer coefficients for various constructions installed in typical environments at usual flue gas flow velocities (Segeler 1965). For masonry, any additional thickness beyond the single course of brick plus tile liner used in residential chimneys decreases the coefficient. Equations (5) and (6) describe flow and heat transfer, respectively, in a venting system with Da = 0.

B Dt H Tm To

= = = = =

local barometric pressure, Pa theoretical draft, Pa height of chimney above grade or inlet, m mean flue gas temperature at average conditions in system, K ambient temperature, K

Figure 4Theoretical draft thus increases directly with height and with the difference in density between the hot and cold columns, and is always positive (unless chimney gases are colder than ambient air). Theoretical draft should be estimated and included in system calculations, even for appliances producing considerable positive outlet static pressure, to achieve the economy of minimum chimney size. Equation (7) may be used directly to calculate exact values for theoretical draft at any altitude. For ease of application, Table 6 lists approximate theoretical draft for typical gas temperature rises above 15°C ambient.

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Chimney, Vent, and Fireplace Systems

35.7 instructions for setting the draft regulator. Available draft requirements for larger packaged boilers or appliances assembled from components may be negative, zero (neutral), or positive. Compensation of theoretical draft for altitude or barometric pressure is usually necessary for appliances and chimneys functioning at elevations greater than 600 m. Depending on the design, one of the following approaches to pressure or altitude compensation is necessary for chimney sizing.

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1. Equation (19): Use local theoretical draft with actual energy input, or use sea-level theoretical draft with energy input multiplied by ratio of sea level to local barometric pressure (Table 7 factor). 2. Equation (21): Use local theoretical draft and barometric pressure with volumetric flow at the local density.

Fig. 4

Theoretical Draft Nomograph

Table 6 Approximate Theoretical Draft of Chimneys Vent Gas Temperature Rise, K

Dt per Metre, Pa

60 100 150 200 300 400 600 900 1500

2.0 3.0 4.0 5.0 6.0 7.0 8.0 9.0 10.0

System Pressure Loss Caused by Flow In any chimney system, flow losses, expressed as pressure drop p in pascals, are the difference between theoretical and available draft:

Notes: Ambient temperature = 15°C = 288 K Chimney gas density = air density Sea level barometric pressure = 101 325 Pa Equation (7) may be used to calculate exact values for Dt at any altitude.

Table 7 Altitude, m Sea level 600 1000 2000 3000 4000

Altitude Correction

Barometric Pressure B, kPa 101.325 94.41 89.87 79.50 70.11 61.64

The altitude correction multiplier for input (Table 7) is the only method of correcting to other elevations. Equation (7) for theoretical draft is the basis for Figure 4, which can be used up to about 540°C and 2300 m elevation. For example, using this figure, ambient air temperature at 4.4°C and mean chimney gas temperature at 176°C provide D100 draft of 408 Pa per 100 m of height at 87.3 kPa (1219 m above sea level). For D100 draft at 84.42 kPa barometric pressure or 1493 m above sea level, follow the intersection to the pivot line and read the new D100 draft of 399 Pa. Gas appliances with draft hoods, for example, are usually derated 4% per 300 m of elevation above sea level when they are operated at 610 m altitude or above (see NFPA Standard 54/AGA Z223.1 or CSA Standard B149.1). The altitude correction factor derates the design input so that the vent size at altitude for derated gas appliances is effectively the same as at sea level. For other appliances where burner adjustments or internal changes might be used to adjust for reduced density at altitude, the same factors produce an adequately compensated chimney size. For example, an appliance operating at 1830 m elevation at 2.930 MJ/s (2930 kW) input, but requiring the same draft as at sea level, should have a chimney selected on the basis of 1.244 times the operating input, or 3.645 MJ/s.

p = Dt – Da

where p is always positive. In any chimney or vent system, flow losses resulting from velocity and resistance can be determined from the Bernoulli equation: k m V 2 p = --------------2

Factor* 1.00 1.08 1.13 1.27 1.45 1.64

(8)

(9)

where k = dimensionless system resistance coefficient of piping and fittings V = system gas velocity at mean conditions, m/s m = mean flue gas density, kg/m3

*Multiply operating input by factor to obtain design input.

Pressure losses are thus directly proportional to the resistance factor and to the square of the velocity.

Appliances with fixed fuel beds, such as hand-fired coal stoves and furnaces, require positive available draft (negative gage pressure). Small oil heaters with pot-type burners, as well as residential furnaces with pressure-atomizing oil burners, need positive available draft, which can usually be set by following the manufacturer’s

Available Draft Starting with Equation (1), the difference between theoretical draft and flow losses without depressurization or boost is Da = Dt – p

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35.8

2020 ASHRAE Handbook—HVAC Systems and Equipment (SI) Table 8

Pressure Equations for p

Required Appliance Outlet Pressure or Available Draft Da

p Equationa

Negative, needs positive draft Zero, vent with draft hood or balanced forced draft Positive, causes negative draft

Gravity Only

Gravity plus Inducerb

 p = Dt  D a p = Dt

p = Dt  Da + Db p = Dt + Db

 p = Dt +Da

p = Dt + Da + Db

aEquations

use absolute pressure for Da. bD = static pressure boost of inducer at flue gas temperature and rated flow. b

Q = 4  0.68 = 2.72 m3/s

Available draft Da can therefore be defined as

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2 1 1  k m V – ------  – ---------------Da = 0.03413BH  ----2  To Tm 

(10)

Available draft Da can be negative, zero, or positive. The pressure difference p, or theoretical minus available draft, overcomes the flow losses. Table 8 lists the pressure components for three draft configurations. Table 8 applies to still air (no wind) conditions and a neutral (zero) pressure difference between the space surrounding the appliance or fireplace and the atmosphere. Columns with and without boost are provided. The effect of a nonneutral pressure difference on capacity or draft may be included by imposing a static pressure (either positive or negative). One way to circumvent a space-to-atmosphere pressure difference is to use a sealed-combustion system or direct-vent system (i.e., all combustion air is taken directly from the outdoor atmosphere, and all flue gas is discharged to the outdoor atmosphere with no system openings such as draft hoods). The effect of wind on capacity or draft may be included by imposing a static pressure (either positive or negative) or by changing the vent terminal resistance loss. However, a properly designed and located vent terminal should cause little change in p at typical wind velocities. Although small static draft pressures can be measured at the entrance and exit of gas appliance draft hoods, Da at the appliance is effectively zero. Therefore, all theoretical draft energy produces chimney flow velocity and overcomes chimney flow resistance losses.

Chimney Gas Velocity Velocity in a chimney or vent varies inversely with flue gas density m and directly with mass flow rate. The equation for flue gas velocity at mean gas temperature in the chimney is  4w V = 1000 ------------------------2  m d i

(11)

where V di m w

= = = =

exhaust systems, and other appliances with high outlet pressures or velocities is needed to estimate piping loss coefficients. Equations (9), (11), (12), and (20) can be applied to find velocity. The velocity may also be calculated by dividing volumetric flow rate Q by chimney area A. A similar calculation may be performed when the energy input is known. For example, a 500 mm diameter chimney serving a 4 kW propane gas appliance, at 8.5% CO2 and 200 K above ambient in the chimney, produces (from Figure 2) 0.49 kg/MJ mass flow rate and 0.68 m3/MJ volume flow rate at sea level. Chimney gas flow rate is

flue gas velocity, m/s inside diameter, mm mean flue gas density, kg/m3 mass flow rate, kg/s

(13)

Dividing by area to obtain velocity, V = 2.72/0.1963 = 13.86 m/s. Chimney gas velocity affects the piping friction factor kL and roughness correction factor. The section on Resistance Coefficients has further information, and Example 2 illustrates how these factors are used in the velocity equations. Chimney systems can operate over a wide range of velocities, depending on modulation characteristics of the burner system or the number of appliances in operation. Typical velocity in vents and chimneys ranges from 1.5 to 15 m/s. A chimney design developed for maximum input and maximum velocity should be satisfactory at reduced input because theoretical draft is roughly proportional to flue gas temperature rise, whereas flow losses are proportional to the square of the velocity. Thus, as input is reduced, flow losses decrease more rapidly than system motive pressures. Effluent dispersal may occasionally require a minimum upward chimney outlet velocity, such as 15 m/s. A tapered exit cone can best meet this requirement. For example, to increase the outlet velocity from the 500 mm diameter chimney (A = 0.1963 m3 area) from 12.8 to 15 m/s, the cone must have a discharge area of 0.1963  12.815 = 0.168 m2 and a 460 mm diameter. An exit cone avoids excessive flow losses because the entire system operates at the lower velocity, and a resistance factor is only added for the cone. In this case, the added resistance for a gradual taper approximates the following (see Table 9):  500  d i1  k =  -------  – 1 =  --------- – 1 = 0.396 d  460  i2  4

4

(14)

Noise in chimneys may be caused by turbulent flow at high velocity or by combustion-induced oscillations or resonance. Noise is seldom encountered in gas vent systems or in systems producing positive available draft, but it may be a problem with forced-draft appliances. Turbulent flow noise can be avoided by designing for lower velocity, which may entail increasing the chimney size above the minimum recommended by the appliance manufacturer. Chapter 49 of the 2019 ASHRAE Handbook—HVAC Applications has more information on noise control.

System Resistance Coefficient

To express velocity as a function of input and chimney gas composition, w in Equation (11) is replaced by using Equation (2): 1000  4IM 1273IM V = ----------------------------- = ------------------2 2  m d i m di

(12)

The input capacity or diameter of a chimney may usually be found without determining flow velocity. Internal or exit velocity must occasionally be known, to ensure effluent dispersal or avoid flow noise. Also, the flow velocity of incinerator chimneys, turbine

The velocity head method of determining resistance losses assigns a fixed numerical coefficient (independent of velocity) or k factor to every fitting or turn in the flow circuit, as well as to piping. The resistance coefficient k that appears in Equations (19) and (21) summarizes the friction loss of the entire chimney system, including piping, fittings, and configuration or interconnection factors. Capacity of the chimney varies inversely with the square root of k, whereas diameter varies as the fourth root of k. The insensitivity of diameter and input to small variations in k simplifies design. Analyzing details such as pressure regain, increasers and reducers, and gas cooling junction effects is unnecessary if slightly high resistance coefficients are assigned to any draft diverters, elbows, tees, terminations, and, particularly, piping.

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Chimney, Vent, and Fireplace Systems Table 9

Inlet acceleration (k1) Gas vent with draft hood Barometric regulator Direct connection Round elbow (k2) 90° 45° Tee or 90° connector (k3) Y connector Cap, top (k4) Open straight Low-resistance (UL) Other Spark screen

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Converging exit cone Tapered reducer (di1 to di2) Increaser Piping (kL)

Resistance Loss Coefficients Suggested Design Value, Dimensionless*

Component

35.9

1.5 0.5 0.0

Estimated Span and Notes 1.0 to 3.0 0.0 to 0.5 Also dependent on blocking damper position

0.75 0.3 1.25 0.75

0.5 to 1.5 — 1.0 to 4.0 0.5 to 1.5

0.0 0.5 — 0.5

— 0.0 to 1.5 1.5 to 4.5 —

Fig. 5

(Lapple 1949) k3 k4 F L di

1 – (di2 di1)4 System designed using di2 See Chapter 2, 2017 ASHRAE Handbook—Fundamentals.

0.033L/df

= = = = =

tee loss coefficient, n3 = number of tees cap, top, or exit cone loss coefficient friction factor length of all piping in chimney system, m inside diameter, m

For combined gas vents using appliances with draft hoods, the summation k must be multiplied by a diversity factor of 1.5 (see Table 9 note and Example 4). This multiplier does not apply to forced- or induced-draft vents or chimneys. When size is unknown, the following k values may be used to run a first trial estimate:

(di1 di2)4 – 1 System designed using di1

—

Friction Factor for Commercial Iron and Steel Pipe

Numerical coefficient (friction factor F) from 0.017 to 0.042; see Figure 12, Chapter 2, 2017 ASHRAE Handbook— Fundamentals for size, roughness, and velocity effects.

*Initial assumption, when size is unknown: k = 5.0 for entire system, for first trial; k = 7.5 for combined gas vents only Note: For combined gravity gas vents serving two or more appliances (draft hoods), multiply total k (components + piping; see Equations [15], [16], and [17]) by 1.5 to obtain gravity system design coefficient. (This rule does not apply to forced- or induced-draft vents or chimneys.)

For any chimney with fittings, the total flow resistance is a constant plus variable piping resistance (a function of centerline length divided by diameter). Table 9 suggests moderately conservative resistance coefficients for common fittings. Elbow resistance may be lowered by long-radius turns; however, corrugated 90° elbows may have resistance values at the high end of the scale. Table 9 shows resistance as a function of inlet diameter di1 and outlet diameter di2. System resistance k may be expressed as follows: k = kf + kL

(15)

kf = k1 + n2k2 + n3k3 + k4 + …

(16)

kL = FL ------di

(17)

with

and

where kf = fixed fitting loss coefficient kL = piping resistance loss function (Figure 13, Chapter 3 of the 2017 ASHRAE Handbook—Fundamentals, adjusted for units) k1 = inlet acceleration coefficient k2 = elbow loss coefficient, n2 = number of elbows

k = 5.0 for the entire system k = 7.5 for combined gas vents only

The resistance coefficient method adapts well to systems in which the fittings cause significant losses. Even for extensive systems, an initial assumption of k = 5.0 gives a tolerably accurate vent or chimney diameter in the first trial solution. Using this diameter with the piping resistance function (Equation [17]) in a second trial normally yields the final answer. The minimum system resistance coefficient in a gas vent with a draft hood is always 1.0 because all gases must accelerate through the draft hood from almost zero velocity to vent velocity. For a system connected directly to the outlet of a boiler or other appliance where the capacity is stated as full-rated heat input against a positive static pressure at the chimney connection, minimum system resistance is zero, and no value is added for existing velocity head in the system. For simplified design, a value of 0.4 for F in Equation (17) applies for all sizes of vents or chimneys and for all velocities and temperatures. As diameter increases, this function becomes increasingly conservative, which is desirable because larger chimneys are more likely to be made of rough masonry construction or other materials with higher pressure losses. The 0.033 constant also introduces an increasing factor of safety for flow losses at greater lengths and heights. Figure 5 is a plot of friction factor F versus velocity and diameter for commercial iron and steel pipe at a flue gas temperature of 150°C above ambient (Lapple 1949). The figure shows, for example, that a 1.2 m diameter chimney with a flue gas velocity of 24 m/ s may have a friction factor as low as 0.016. In most cases, kL = 0.025 L/di gives reasonable design results for chimney sizes 460 mm and larger because systems of this size usually operate at flue gas velocities greater than 3 m/s. At 540°C or over, the factors in Figure 5 should be multiplied by 1.2. Because Figure 5 is for commercial iron and steel pipe, an additional correction for greater or less surface roughness may be imposed. For example, the factor for a very rough 300 mm diameter pipe may be doubled at a velocity as low as 10 m/s. For most chimney designs, a friction factor F of 0.033 gives a conservative solution for diameter or input for all sizes, types, and

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35.10

2020 ASHRAE Handbook—HVAC Systems and Equipment (SI)

operating conditions of prefabricated and metal chimneys; alternatively, F = 0.025 is reasonable if the diameter is 460 mm or more. Because neither input nor diameter is particularly sensitive to the total friction factor, the overall value of k requires little correction. Masonry chimneys, including those lined with clay flue tile, may have rough surfaces, tile shape variations that cause misalignment, and joints at frequent intervals with possible mortar protrusions. In addition, the inside cross-sectional area of liner shapes may be less than expected because of local manufacturing variations, as well as differences between claimed and actual size. To account for these characteristics, the estimate for kL should be on the high side, regardless of chimney size or velocity. Computations should be made by assuming smooth surfaces and then adding a final size increase to compensate for shape factor and friction loss. Performance or capacity of metal and prefabricated chimneys is generally superior to that of site-constructed masonry.

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Configuration and Manifolding Effects The most common configuration is the individual vent, stack, or chimney, in which one continuous system carries products from appliance to terminus. Other configurations include the combined vent serving a pair of appliances, the manifold serving several, and branched systems with two or more lateral manifolds connected to a common vertical system. As the number of appliances served by a common vertical vent or chimney increases, the precision of design decreases because of diversity factors (variation in the number of appliances in operation) and the need to allow for maximum and minimum input operation (Stone 1957). For example, the vertical common vent for interconnected gas appliances must be larger than for a single appliance of the same input to allow for operating diversity and draft hood dilution effects. Connector rise, headroom, and configuration in the equipment room must be designed carefully to avoid draft hood spillage and related oxygen depletion problems. For typical combined vents, the diversity effect must be introduced into the equations by multiplying system resistance loss coefficient k by 1.5 (see Table 9 note and Example 4). This multiplier compensates for junction effect and part-load operation. Manifolds for appliances with barometric draft regulators can be designed without allowing for dilution air through inoperative appliances. In this case, because draft regulators remain closed until regulation is needed, dilution under part load is negligible. In addition, airflow through any inoperative appliance is negligible because the combustion air inlet dampers are closed and the multiple-pass heat exchanger has a high internal flow resistance. Manifold systems of oil-burning appliances, for example, have a lower flow velocity and, hence, lower losses. As a result, they produce reasonable draft at part load or with only one of several appliances in operation. Therefore, diversity of operation has little effect on chimney design. Some installers set each draft regulator at a slightly different setting to avoid oscillations or hunting possibly caused by burner or flow pulsations. Calculation of the resistance coefficient of any portion of a manifold begins with the appliance most distant from the vertical portion. All coefficients are then summed from its outlet to the vent terminus. The resistance of a series of tee joints to flow passing horizontally straight through them (not making a turn) is the same as that of an equal length of piping (as if all other appliances were off). This assumption holds whether the manifold is tapered (to accommodate increasing input) or of a constant size large enough for the accumulated input. Coefficients are assigned only to inlet and exit conditions, to fittings causing turns, and to the piping running from the affected appliance to the chimney exit. Initially, piping shape (round, square, or rectangular) and function (for connectors, vertical piping, or both) are irrelevant.

Some high-pressure, high-velocity packaged boilers require special manifold design to avoid turbulent flow noise. Figure 6 shows a typical configuration for this application. The loss coefficients listed in Table 9 for standard tees and elbows are higher than necessary for long-radius elbows or Y entries. Occasionally, on appliances with high chimneys augmenting boiler outlet pressure, it may appear feasible to reduce the diameter of the vertical portion to below the size recommended by the manufacturer. However, any reduction may cause turbulent noise, even though all normal design parameters have been considered. With the simplifying assumption that the maximum velocity of the flue gas (which exists in the smaller of the two portions) exists throughout the entire system, and Equations (18) and (19) can be used to calculate the size of a vertical portion smaller in area than the manifold or of a chimney connector smaller than the vertical. This assumption leads to a conservative design because true losses in the larger area are lower than assumed. Further, if the size change is small, either as a contraction or expansion, the added loss coefficient for this transition fitting (see Table 9) is compensated for by reduced losses in the enlarged part of the system. These comments on size changes apply more to individual than to combined systems because it is undesirable to reduce the vertical area of the combined type, and, more frequently, it is desirable to enlarge it. If an existing vertical chimney is slightly undersized for the connected load, the complete chart method must be applied to determine whether a pressure boost is needed, because size is no longer a variable. Sectional gas appliances with two or more draft hoods do not pose any special problems if all sections fire simultaneously. In this case, the designer can treat them as a single appliance. The appliance installation instructions either specify the size of manifold for interconnecting all draft hoods or require a combined area equal to the sum of all attached draft hood outlet areas. Once the manifold has been designed and constructed, it can be connected to a properly sized chimney connector, vent, or chimney. If the connector and chimney size is computed as less than manifold size (as may be the case with a tall chimney), the operating resistance of the manifold will be lower than the sum of the assigned component coefficients because of reduced velocity. The general rule for conservative system design, in which manifold, chimney connector, vent, or chimney are different sizes, can be stated as follows: Always assign full resistance coefficient values to all portions carrying combined flow, and determine system capacity from the smallest diameter carrying the combined flow. In addition, horizontal chimney connectors or vent connectors should pitch upward toward the stack at a minimum of 20 mm per metre of connector length.

Input, Diameter, and Temperature Relationships To obtain a design equation in which all terms are readily defined, measured, or predetermined, the gas velocity and density terms must be eliminated. Using Equation (4) to replace m and Equation (12) to replace V in Equation (9) gives  T m B o   4  10 6 IM  2 - p = ---k-  ---------------   -------------------------2 2  o Ts B    d i

(18)

Rearranging to solve for I and including the values of ,To, o, and Bo gives d i  pB  I = 6.55  10–8 -----  ----------  M  kT m  2

(19)

Solving for input using Equation (19) is a one-step process, given the diameter and configuration of the chimney. More frequently,

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Chimney, Vent, and Fireplace Systems

35.11

however, input, available draft, and height are given and the diameter di must be found. Because system resistance is a function of the chimney diameter, a trial resistance value must be assumed to calculate a trial diameter. This method allows for a second (and usually accurate) solution for the final required diameter.

Volumetric Flow in Chimney or System Volumetric flow Q may be calculated in a chimney system for which Equation (19) can be solved by solving Equation (9) for velocity at mean density (or temperature) conditions: V=

2 p ---------k m

(20)

This equation can be expressed in terms of the same variables as Equation (19) by using the density value m of Equation (4) in Equation (20) and then substituting Equation (20) for velocity V. Area is expressed in terms of di. Multiplying area and velocity and adjusting for units,

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Q = 18.8 

2  pTm  10–6 d i  ------------- kB





12

(21)

where Q = volumetric flow rate, m3/s. The volumetric flow obtained from Equation (21) is at mean gas temperature Tm and local barometric pressure B. Equations (19) and (21) do not account for the effects of heat transfer or cooling on flow, draft, or capacity. Equation (21) is useful in the design of forced-draft and induceddraft systems because draft fans are usually specified in terms of volumetric flow rate at some standard ambient or selected gas temperature. An induced-draft fan is necessary for chimneys that are undersized, that are too low, or that must be operated with draft in the manifold under all conditions.

5.

STEADY-STATE CHIMNEY DESIGN GRAPHICAL SOLUTIONS

Design ambient temperature is 15.6°C (288.9 K), and all temperatures given are in terms of rise above this ambient. Theoretical draft may be corrected for altitude or reduced air density by multiplying the operating input by the factor in Table 7. The resistance coefficient k summarizes the friction loss of the entire chimney system, including piping, fittings, and configuration or interconnection factors. Figure 2 can be used to estimate mass and volumetric flow. Table 6 lists approximate theoretical draft for typical gas temperature rises above 15.6°C ambient. The equations are based on the fuel combustion products and temperatures in the chimney system. The first trial solution for diameter, using Equations (19) and (21), need not consider the cooling temperature multiplier, even for small sizes. A first approximate size can be used for the temperature multiplier for all subsequent trials because capacity is insensitive to small changes in temperature. The equations do not contain the same number or order of steps as the derivation; for example, a step disappears when theoretical and available drafts are combined into p. Similarly, the examples selected vary in their sequence of solution, depending on which parameters are known and on the need for differing answers, such as diameter for a given input, diameter versus height, or the amount of pressure boost Db from a forced-draft fan.The following sample calculation illustrates the direct solution for input, velocity, and volume. Example 1. Find the input capacity (MJ/s) of a vertical, double-wall type B gas vent, 600 mm in diameter, 30 m high at sea level. This vent is used with draft hood natural-gas-burning appliances.

Fig. 6 Typical Connector Design

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35.12

2020 ASHRAE Handbook—HVAC Systems and Equipment (SI) From Equation (12), 6

4  10  3.00   0.661  V = ---------------------------------------------------- = 9.01 m/s 2  3.14   0.778   600  9. Volume flow can now be found because velocity is known. The flow area of 600 mm diameter is 0.283 m2, so Q = 0.283  9.01 = 2.55 m3/s No heat loss correction is needed to find the new flue gas temperature because the size is greater than 500 mm, and this vent is vertical with no horizontal connector.

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6.

Fig. 7

Gas Vent with Lateral

Solution: 1. Mass flow from Table 2. M = 0.661 kg/MJ for natural gas, if no other data are given. 2. Temperature from Table 4. Temperature rise = 165 K and Tm = 180 + 273 = 453 K for natural gas. 3. Theoretical draft from Table 6 or Equation (7). For 30 m height at 165 K rise, Dt = 131 Pa. 4. Available draft for draft hood appliances: Da = 0. 5. Flow losses from Table 8. p = Dt = 131 Pa; flow losses for a gravity gas vent equal theoretical draft at mean gas temperature. 6. Resistance coefficients from Table 9. For a vertical vent, Draft hood Vent cap 30 m piping

k1 = 1.5 k4 = 1.0 kL = 0.033(30/0.6) = 1.65

System total

k = 4.15

7. Solution for input. Altitude: Sea level, B = 101 325 Pa from Table 7. di = 600 mm. These values are substituted into Equation (19) as follows:  600  2   131   101 325  I = 6.55  10–8 ----------------  --------------------------------------- 0.661   4.15   453  

12

I = 3.00 MJ/s input capacity 8. A solution for velocity requires a prior solution for input to apply to Equation (12). First, using Equation (4),  101 325 m = 0.00348  ------------------- = 0.778 kg/m3  453 

VENT AND CHIMNEY CAPACITY CALCULATION EXAMPLES

Figures 7 to 10 show chimney capacity for individually vented appliances computed by the methods presented. These capacity curves may be used to estimate input or diameter for Equation (19) or Equation (21). These capacity curves apply primarily to individually vented appliances with a lateral chimney connector; systems with two or more appliances or additional fittings require a more detailed analysis. Figures 7 to 10 assume the length of the horizontal connector is (1) at least 3 m and (2) no longer than 50% of the height or 15 m, whichever is less. For chimney heights of 3 to 6 m, a fixed 3 m long connector is assumed. Between 6 and 30 m, the connector is 50% of the height. If the chimney height exceeds 30 m, the connector is fixed at 15 m long. For a chimney of similar configuration but with a shorter connector, the size indicated in the figures is slightly larger than necessary. In deriving the data for Figures 7 to 10, additional conservative assumptions were used, including the temperature correction Cu for double-wall laterals (see Figure 3) and a constant friction factor (0.033) for all sizes. The loss coefficient k4 for a low resistance cap is included in Figures 7, 8, and 9. If no cap is installed, these figures indicate a larger size than needed. Figure 7 applies to a gas vent with draft hood and a lateral that runs to the vertical section. Maximum static draft is developed at the base of the vertical, but friction reduces the observed value to less than the theoretical draft. Areas of positive pressure may exist at the elbow above the draft hood and at the inlet to the cap. The height of the system is the vertical distance from the draft hood outlet to the vent cap. Figure 8 applies to a typical boiler system requiring both negative combustion chamber pressure and negative static outlet pressure. The chimney static pressure is below atmospheric pressure, except for the minor outlet reversal caused by cap resistance. Height of this system is the difference in elevation between the point of draft measurement (or control) and the exit. (Chimney draft should not be based on the height above the boiler room floor.) Figure 9 shows the use of a negative static pressure connector serving a forced-draft boiler. This system minimizes flue gas leakage in the equipment room. The draft is balanced or neutral, which is similar to a gas vent, with zero draft at the appliance outlet and pressure loss p equal to theoretical draft. Figure 10 applies to a forced-draft boiler capable of operating against a positive static outlet pressure of up to 125 Pa. The chimney system has no negative pressure, so outlet pressure may be combined with theoretical draft to get minimum chimney size. For chimney heights or system lengths less than 125 Pa, the effect of adding 125 Pa positive pressure to theoretical draft causes all curves to fall into a compressed zone. An appliance that can produce 125 Pa positive forced draft (negative draft) is adequate for venting any simple arrangement with up to 30 m of flow path and no wind back pressure, for which additional forced draft is required.

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Chimney, Vent, and Fireplace Systems

Fig. 8 Draft-Regulated Appliance with 25 Pa Available Draft Required The following examples illustrate the use of the corresponding equations. Example 2. Gravity incinerator chimney (see Figure 11). Located at 2400 m elevation, the appliance burns 75 g/s of type 0 waste with 100% excess air at 760°C outlet temperature. Ambient temperature To is 15°C. Outlet pressure is zero at low fire, and +75 Pa at high fire. The chimney will be a prefabricated medium-heat type with an 18 m connector and a roughness factor of 1.2. The incinerator outlet is 460 mm in diameter, and it normally uses a 6 m vertical chimney. Find the diameter of the chimney and the connector and the height required to overcome flow and fitting losses. Solution: 1. Find mass flow from Table 3 as 13.76 kg combustion products per kilogram of waste, or w = 75  13.76 = 1032 g/s total. 2. Find mean chimney gas temperature. Based on 18 m length of 460 mm diameter double-wall chimney, Cu = 0.83 (see Figure 3). Temperature rise te = 760 – 15 = 745 K; thus, tm = 745(0.83) = 618 K rise above 15°C ambient. Tm = 618 + 15 + 273 = 906 K. Use this temperature in Equation (4) to find flue gas density at 2400 m elevation (from Table 7, B = 75.7 kPa):  75 700  m = 0.00348  -----------------  = 0.291 kg/m3  906  3. Find the required height by finding theoretical draft per metre from Equation (7) (Table 6 applies only to sea level).  1 Dt 1  ----- =  0.0314   75 700   -------- – ---------  H  288 906  = 6.12 Pa per metre of height

35.13

Fig. 9 Forced-Draft Appliance with Neutral (Zero) Draft (Negative Pressure Lateral) 4. Find allowable pressure loss p in the incinerator chimney for a positive-pressure appliance having an outlet pressure of +25 Pa. From Table 8, p = Dt + Da, where Dt = 6.12H, Da = 0.25, and p = 6.12H + 25 Pa. 5. Calculate flow velocity at mean temperature from Equation (11) to balance flow losses against diameter/height combinations:  1000  4   1032  V = -------------------------------------------------- = 21.4 m/s 2  3.14   0.291   460  This velocity exceeds the capability of a gravity chimney of moderate height and may require a draft inducer if a 460 mm chimney must be used. Verify velocity by calculating resistance and flow losses by the following steps. 6. From Table 9, resistance coefficients for fittings are 1 Tee 1 Elbow Spark screen

k3 = 1.25 k2 = 0.75 k4 = 0.50

Fitting total

kf = 2.50

The piping resistance, adjusted for length, diameter, and a roughness factor of 1.2, must be added to the total fitting resistance. From Figure 5, find the friction factor F at 460 mm diameter and 21.4 m/s as 0.0183. Assuming 6 m of height with an 18 m lateral, piping friction loss is  1.2   0.0183   18 + 6  1.2FL kL = --------------- = ------------------------------------------------------ = 1.15 di 460  1000 and total k = 2.50 + 1.15 = 3.65 Use Equation (9) to find p, to determine whether this chimney height and diameter are suitable.

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35.14

2020 ASHRAE Handbook—HVAC Systems and Equipment (SI)

Fig. 11 Illustration for Example 2 1. As before, w = 1032 g/s. 2. At 610 mm diameter, no temperature correction is needed for the 18 m connector. Thus, Tm = 760 + 273 = 1033 K (see Table 4), and density is  75 700  m = 0.00348  ----------------  = 0.255 kg/m3  1033 

Fig. 10 Forced-Draft Appliance with Positive Outlet Pressure (Negative Draft) 2

 3.65   0.291   21.4  p = ---------------------------------------------------- = 243 Pa flow loss 2 For these operating conditions, theoretical draft plus available draft yields p = (6.12  6) + 25 = 62 Pa driving force Flow losses of 243 Pa exceed the 62 Pa driving force; thus, the selected diameter, height, or both are incorrect, and this chimney will not work. This can also be shown by comparing draft per metre with flow losses per metre for the 460 mm diameter configuration: Flow losses per metre of 460 mm chimney = 243/(18 + 6) = 10 Pa Draft per metre of height = 6.12 Pa Regardless of how high the chimney is, losses caused by a 21.4 m/s velocity build up faster than draft. 7. A draft inducer could be selected to make up the difference between losses of 243 Pa and the 62 Pa driving force. Operating requirements are w 1032 Q = ------------------------ = ------------------------------------ = 3.55 m3/s  1000 m   1000  0.291  p = 243 – 62 = 181 Pa at 3.55 m3/s and 618 K rise If the inducer selected (see Figure 25C) injects single or multiple air jets into the gas stream, it should be placed only at the chimney top or outlet. This location requires no compensation for additional air introduced by an enlargement downstream from the inducer. Because 460 mm is too small, assume that a 610 mm diameter may work at a 6 m height and recalculate with the new diameter.

3. Theoretical draft per metre of chimney height is Dt ----- = 0.03413 (75 700) H

 1 1   --------- – ------------  = 6.47 Pa/m  288 1033 

4. Velocity is  1000  4   1032  V = -------------------------------------------------- = 13.86 m/s 2  3.14   0.255   610  5. From Figure 5, the friction factor is 0.0188, which, when multiplied by a roughness factor of 1.2 for the piping used, becomes 0.0226. For the entire system with kf = 2.50 and 24 m of piping, find k = 2.5 + 0.0226(24/0.610) = 3.39. From Equation (9), 2

 3.39   0.255   13.86  p = ------------------------------------------------------- = 83.0 Pa flow losses 2 Dt + Da = 6  6.47 + 25 = 64 Pa, so driving force is less than losses. The small difference indicates that, although 6 m of height is insufficient, additional height may solve the problem. The added height must make up for 19 Pa additional draft. As a first approximation, Added height = Additional draft/draft per metre = 19/6.47 = 2.94 m Total height = 6 + 2.94 = 8.94 m This is less than the actual height needed because resistance changes have not been included. For an exact solution for height, the driving force can be equated to flow losses as a function of H. Substituting H = +18 for L in Equation (15) to find k for Equation (9),

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Chimney, Vent, and Fireplace Systems

35.15 Using a positive outlet pressure of 125 Pa as available draft, p = 65.6 + 125 = 190.6 Pa. Assume k = 5.0. Assuming 80% efficiency, input is 1.25 MJ/s. Substitute in Equation (19): d i 190.6  101 325 1  2 I = 1.25 – 6.55  10–8 -------------  ---------------------------------------   0.368  5.0  453 2

Solving, di = 276 mm as a first approximation. From Table 9, find correct k using next largest diameter, or 300 mm: Inlet acceleration (direct connection) 90° Elbow Tee 21.25 m piping

k1 k2 k3 kL

System total

Fig. 12 Illustration for Example 3

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k = kf + kL = kf + (FLdi ) = 2.5 + (0.0226  18/0.610) = 3.17 k m V 2 Driving force = p = --------------2

 1  2 3.17 +  0.0226   -------------  H  0.255   13.86   0.610  25 + 6.47H = -------------------------------------------------------------------------------------------------------------2 H = 9.464 m at 610 mm diameter. Checking by substitution, total driving force = 85.95 Pa and total losses, based on a system with 27.415 m of piping, equal 85.95 Pa. The value of H = 9.464 m is the minimum necessary for proper system operation. Because of the great variation in fuels and firing rate with incinerators, greater height should be used to ensure adequate draft and combustion control. An acceptable height would be from 12 to 15 m.

= 0.0 = 0.75 = 1.25 = 0.033(21.25) = 2.34 k = 4.34

Note: Assume tee over boiler B has k = 0 in subsystem A. Corrected temperature rise (see Figure 3 for 6 m single-wall connector) = 0.75  165 = 124 K, and Tm = 124 + 15 + 273 = 412 K. This corrected temperature rise changes Dt to 3.61 Pa/m (Table 6 or Equation [7]), or 55.1 Pa for 15.25 m. Thus, p becomes 55.1 + 125 = 180.1 Pa. di 180.1  101 325 1  2 125 = 6.55  10–8 -------------  ---------------------------------------  0.368 4.34  412  2

So di = 264 mm, or a 300 mm diameter is adequate. For size of manifold and vertical, starting with the tee over boiler B, assume 400 mm diameter (see also Figure 10). System k = 3.88 for the 16.75 m of piping from B to outlet: Inlet acceleration k1 = 0.0 Two tees (boiler B subsystem) k3 = 2.5 16.75 m piping kL = 0.033(16.75/0.4) = 1.38 System total

k = 3.88

Temperature and p will be as corrected (a conservative assumption) in the second step for boiler A. Having assumed a size, find input. 400 180.1  101 325 1  2 = 3.04 MJ/s I = 6.55  10–8 -------------  ---------------------------------------  0.368  3.88  412  2

Example 3. Two forced-draft boilers (see Figure 12). This example shows how multiple-appliance chimneys can be separated into subsystems. Each boiler is rated 1 MJ/s (1 MW) on No. 2 oil. The manufacturer states flue gas operation at 13.5% CO2 and 165 K temperature rise against 125 Pa positive static pressure at the outlet. The 15.25 m high chimney has a 6 m single-wall manifold and is at sea level. Find the size of connectors, manifold, and vertical. Solution: First, find the capacity or size of the piping and fittings from boiler A to the tee over boiler B. Then, size the boiler B tee and all subsequent portions to carry the combined flow of A and B. Also, check the subsystem for boiler B; however, because its shorter length compensates for greater fitting resistance, its connector may be the same size as for boiler A. Find the size for combined flow of A and B either by assuming k = 5.0 or by estimating that the size will be twice that found for boiler A operating by itself. Estimate system resistance for the combined portion by including those fittings in the B connector with those in the combined portion. Data needed for the solution of Equation (19) for boiler A for No. 2 oil at 13.5% CO2 include the following:  14.4  M = 0.31  0.12 + ----------  = 0.368 kg/MJ 13.5  

(from Table 1)

For a temperature rise of 165 K and an ambient of 15°C, Tm = 165 + 15 + 273 = 453 K Interpolating from Table 6, theoretical draft = 4.3 Pa per metre of height; for 15.25 m height, Dt = (4.3)(15.25) = 65.6 Pa.

A 400 mm diameter manifold and vertical is more than adequate. Solving for the diameter at the combined input, di = 364 mm; thus, a 400 mm chimney must be used. Note: Regardless of calculations, do not use connectors smaller than the appliance outlet size in any combined system. Applying the temperature correction for a single-wall connector has little effect on the result because positive forced draft is the predominant motive force for this system. Example 4. Six gas boilers manifolded at 1825 m elevation (see Figure 13). Each boiler is fired at 0.5 MJ/s, with draft hoods and an 25 m long manifold connecting into a 125 m high vertical. Each boiler is controlled individually. Find the size of the constant-diameter manifold, vertical, and connectors with a 0.6 m rise. All are double-wall. Solution: Simultaneous operation determines both the vertical and manifold sizes. Assume the same appliance operating conditions as in Example 1: CO2 = 5.3%, natural gas, flue gas temperature rise = 165 K. Initially assume k = 5.0 is multiplied by 1.5 for combined vent (see note at bottom of Table 9); thus, design k = 7.5. For a gas vent at 125 m height, p = Dt = H × Dt /m = 125  4.3 = 537 Pa rise 165 K in Table 6; Dt = 4.3 Pa/m. At 1825 m elevation, operating input must be multiplied by an altitude correction (Table 7) of 1.25. Total design input is 0.5(6)(1.25) = 3.75 MW. From Table 2, M = 0.661 kg/MJ at operating conditions. Tm = 165 + 15 + 273 = 453 K, and B = 101 325 Pa because the 1.25 input multiplier corrects back to sea level. From Equation (19),

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35.16

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d i  537  101 325 -  ----------------------------------- 0.5  6  1.25 = 3.75 = 6.55  10–8 -----------0.661  7.5  453  2

12

Because the diameter is greater than 500 mm, no temperature correction is needed. Recompute k for 25 + 125 m of 600 mm diameter (Table 9): Draft hood inlet acceleration k1 = Two tees (connector and base of chimney) 2k3 = Low-resistance top k4 = 150 m piping kL = 0.033(150/ 0.6) = System total

1.5 2.5 0.5 8.25

k =12.75

Combined gas vent design k = multiple vent factor 1.5  12.75 = 19.12. Substitute again in Equation (19):  d i   537  101 325 3.75 = 6.55  10–8  -------------  -----------------------------------  0.661  19.12  453  2

12

di = 691 mm; thus use 700 mm

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2

12

di = 674 mm (virtually the) same as above)

di = 547 mm

kf = 4.5 kL = 0.033(150/0.700) = 7.1 k = 11.6 (1.5)(11.6) = 17.4 Substitute in Equation (19) to obtain the third trial: 3.75 = 6.55 

 d i   431  81 300 0.5  6 = 6.55  10–8  -------------  --------------------------------  0.661  17.4  453 

12 2 d i  537  101 325 –8 --------------------------------------------10

  0.661  17.4  453 

di = 675 mm The third trial is less than the second and again shows the manifold and vertical chimney diameter to be between 600 and 700 mm. For connector size, see the National Fuel Gas Code (NFPA Standard 54/AGA Standard Z223.1) for double-wall connectors of combined vents. The height limit of the (I-P only) table is 30.48 m (100 ft); do not extrapolate and read the capacity of 457 mm (18 in.) connector as 510 MJ/s (MW) (1 740 000 Btu/h) at 610 mm (2 ft) rise. Use 457 mm connector or draft hood outlet, whichever is larger. No altitude correction is needed for connector size; the draft hood outlet size considers this effect. Note: Equation (19) can also be solved at local elevation for exact operating conditions. At 1825 m, the local barometric pressure is 81.3 kPa (interpolated from Table 7), and assumed theoretical draft must be corrected in proportion to the reduction in pressure: Dt = 537(81.3/101 325) = 431 Pa. Operating input of 6  0.5 = 3 MJ/s is used to find di, again taking final k = 17.4:

Fig. 13 Illustration for Example 4

This example illustrates the equivalence of the two solutions. Equation (19) gives the correct solution using either method 1, with only the fuel input corrected back to sea-level condition, or method 2, correcting p for local barometric pressure and using operating input at altitude. Example 5. Pressure boost for undersized chimney (not illustrated). A natural gas boiler at sea level (no draft hood) is connected to an existing 300 mm diameter chimney. Input is 1.15 MJ/s with flue gas at 10% CO2 and 165 K temperature rise above ambient. System resistance loss coefficient k = 5.0 with 6 m high chimney. The appliance operates with neutral outlet static draft, so, Da = 0 Pa. a. How much draft boost is needed at operating temperature? b. What fan rating is required at 15°C ambient temperature? c. Where in the system should the fan be located? Solution: Combustion data: from Figure 2, 10% CO2 at 165 K temperature rise indicates 0.48 m3/s per megawatt (m3/MJ). Total flow rate Q = 0.48  1.15 = 0.552 m3/s at chimney gas temperature. Then, Equation (21) can be solved for the only unknown,  p:  p  165 + 15 + 273  0.552 = 18.8  10–6(300)2 -------------------------------------------------5.0  101 325

12

 p = 119 Pa needed at 165 K temperature rise. For 6 m of height at 165 K temperature rise (Table 6 or Equation [7]), (per Table 6)

Dt = 4.3  6 = 25.8 Pa

(per Equation [7]) Dt = 0.03413 BH [(1/To) – (1/Tm)] = 0.03413  101 325  6{[1/288] – [1/(288 + 165)]} = 26.2 Pa a. Pressure boost Db supplied by the fan must equal  p minus theoretical draft (Table 8) when available draft is zero. Db =  p – Dt = 119 – 26 = 93 Pa at operating temperature. b. Draft fans are usually rated for standard ambient (15°C) air. Pressure is inversely proportional to absolute flue gas temperature. Thus, for ambient air, Tm 165 + 15 + 273 = 93  ------------------------------------ = 146 Pa Db = 93 ----- 15 + 273  To

This pressure is needed to produce 93 Pa at operating temperature. In specifying power ratings for draft fan motors, a safe policy is to select one that operates at the required flow rate at ambient temperature and pressure (see Example 6). c. A fan can be located anywhere from boiler outlet to chimney outlet. Regardless of location, the amount of boost is the same; however, chimney pressure relative to atmosphere will change. At boiler outlet, the fan pressurizes the entire connector and chimney. Thus, the system should be gastight to avoid leaks. At the chimney outlet, the system is below atmospheric pressure; any leaks flow into the system and seldom cause problems. With an ordinary sheet metal connector attached to a gastight vertical chimney, the fan may be placed close to the vertical chimney inlet. Thus, the connector leaks safely inward, while the vertical chimney is under pressure. Example 6. Draft inducer selection (see Figure 14). A third gas boiler must be added to a two-boiler system at sea level with an 450 mm diameter, 4.6 m horizontal, and 23 m of total height of connector and chimney system. Outlet conditions for natural gas draft hood appliances are 5.3% CO2 at 165 K temperature rise. Boilers are controlled individually, each with 0.5 MJ/s, for 1.5 MJ/s total input. The system is currently undersized for gravity full-load operation. Find capacity, pressure, size, and power rating of a draft inducer fan installed at the outlet.

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Chimney, Vent, and Fireplace Systems

35.17

Solution: Using Equation (21) requires evaluating two operating conditions: (1) full input at 165 K rise and (2) no input with ambient air. Because the boilers are controlled individually, the system may operate at nearly ambient temperature (55 K flue gas temperature rise or less) when only one boiler operates at part load. Use the system resistance k for boiler 3 as the system value for simultaneous operation. It needs no compensating increased draft, as with gravity multiple venting, because a fan induces flow at all flue gas temperatures. From Table 9, the resistance summation is Inlet acceleration (draft hoods) k1 Tee above boiler k3 Tee at base of vertical k3 27.6 m of 450 mm diameter pipe kL = 0.033(27.6/0.45) System total

= = = =

1.50 1.25 1.25 2.02

k = 6.02

At full load, Tm = 165 + 15 + 273 = 453 K; B = 101 325 Pa. Flow rate Q must be found for operating conditions of 0.661 kg/MJ (Table 2) at density m and full input 1.5 MJ/s. From Equation (4), m = 0.00348B/Tm = 0.00348(101 325/453) =0.778 kg/m3

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Volumetric flow rate is 0.661  1.5 Q = --------------------------- = 1.27 m3/s 0.778 For 165 K flue gas temperature rise, Dt = 4.3  23 = 98.9 Pa theoretical draft in the system (Table 6). Solving Equation (21) for p, 2

 6.02   101 325   1.27  p = ----------------------------------------------------------------- = 150 Pa –6 2 4  18.8  10   450   453  Thus, a fan is needed because p exceeds Dt. Required static pressure boost (Table 8) is Db = 150 – 98.9 = 51.1 Pa at 165 K flue gas temperature rise or a flue gas density of 0.778 kg/m3

temperature rise against 51.1 Pa pressure requires the static pressure boost with standard air to be Db = 51.1(453/288) = 80.4 Pa Thus, a fan that delivers 1.27 m3/s of flue gas at 80.4 Pa at 15°C is required. Figure 15 shows the operating curves of a typical fan that meets this requirement. The exact volume flow rate and pressure developed against a system k of 6.02 can be found for this fan by plotting airflow rate versus p from Equation (21) on the fan curve. The solution, at point C, occurs at 0.92 m3/s, where both the system p and fan static pressure equal 130 Pa. Although some fan manufacturers’ ratings are given at standard air conditions, the motors selected will be overloaded at temperatures below 165 K air temperature rise. Figure 15 shows that power required for two conditions with ambient air is as follows: 1. 0.92 m3/s at 130 Pa static pressure requires 0.42 kW 2. 1.27 m3/s at 80.4 Pa static pressure requires 0.40 kW Thus, the minimum size motor will be 0.40 kW brake power and run at 1590 rpm. Manufacturers’ literature must be analyzed carefully to discover whether the sizing and selection method is consistent with appliance and chimney operating conditions. Final selection requires both a thorough analysis of fan and system interrelationships and consultation with the fan manufacturer to verify the fan and motor capacity and power ratings.

7.

GAS APPLIANCE VENTING

In much of the United States, gas-burning appliances requiring venting of combustion products are installed and vented in accordance with the National Fuel Gas Code (NFPA Standard 54/AGA Z223.1), and in Canada with the Natural Gas and Propane Installation Code (CAN/CSA Standard B149.1). This standard includes capacity data and definitions for commonly used gas vent systems.

Fans are rated for ambient or standard air (15 to 20°C) conditions. Flue gas pressure is directly proportional to density or inversely proportional to absolute temperature. Moving 1.27 m3/s of flue gas at 165 K

Fig. 14 Illustration for Example 6

Fig. 15 Typical Fan Operating Data and System Curves

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In addition, these codes group vented gas-fired appliances by draft and flue gas conditions as follows: • Category I appliances operate with nonpositive vent static pressure and a vent gas temperature that avoids excessive condensate production in the vent. • Category II appliances operate with nonpositive vent static pressure and a vent gas temperature that may cause condensate production in the vent. • Category III appliances operate with positive vent static pressure and a vent gas temperature that avoids excessive condensate production in the vent. • Category IV appliances operate with positive vent static pressure and a vent gas temperature that may cause condensate production in the vent. Category I appliances are further divided into two broad types: natural-draft appliances rely solely on flue gas buoyancy to draw combustion air through the appliance/vent system and are equipped with draft hood (integral or external to the appliance itself) used to limit draft at the outlet of the appliance. Fan-assisted appliances use a blower to move combustion air through the appliance, but rely on flue gas buoyancy alone to move flue products through the vent system (vent pressure at the outlet of the appliance is zero or negative). From a vent design standpoint, the difference between these two types of Category I appliances is important because the vent systems used with natural-draft Category I appliances must handle not only the products of combustion, but also dilution air drawn into the system through the draft hood. Dilution air increases the total gas volume that must be handled by the vent system, and therefore reduces the capacity of a given size vent system. However, a given vent system handling less dilution air tends to be more prone to forming flue gas condensate, particularly under transient conditions, which can cause corrosion or drainage problems when present in excessive amounts. Both the National Fuel Gas Code (ANSI/NFPA Standard 54/ ANSI/AGA Z223.1) and the Natural Gas and Propane Installation Code (CAN/CSA Standard B149.1) include tables and other information that can be used to design vent systems for one or more Category I appliances. These tables include maximum capacities, representing the largest appliance input rating that could safely be vented using a given vent diameter, height, lateral, and material. For reasons discussed previously, these capacities are larger for fanassisted appliances than they are for natural draft appliances. At the same time, these tables specify a minimum capacity for a given size chimney used to vent fan assisted appliances to minimize condensate formation. Minimizing condensate formation, as well as the time required to “prime” the chimney (establish draft when the chimney is initially cold) are the reasons for limitations placed by the above codes on the use of single wall vent connectors, and masonry chimneys. Category I guidelines in these codes were developed using the VENT-II computer program (GTI 2009) to perform transient analysis with the appliance(s) cycling. Gas-fired central furnace cycle times of 3.87 min on, 13.3 min off, 17.17 min total, were determined based on a design outdoor ambient of –14.7°C with an oversize factor of 1.7. One criterion used in developing these guidelines is that the vent connector dry before the end of the appliance on cycle, whereas the vertical portion of the vent is required to dry out before the end of the total cycle. Methods for determining the category of a given gas appliance vary. ANSI Standards Z21.13, Z21.47, and Z21.10 include category determination tests for boilers, furnaces, and water heaters and include a requirement that the resulting category be marked on the appliance’s data plate. Appliances not certified to one of these standards may not be categorized at all or may be categorized in a way inconsistent with ANSI test procedures. In such cases, caution

should be exercised in attempting to use the venting guidelines mentioned in this section. No universal venting guidelines exist for Category II, III, or IV appliances. In many cases, guidelines for such an appliance are defined by the appliance manufacturer in terms of an acceptable range of vent length, diameter, material, and terminal design. In such cases, the appliance manufacturer may also specify the means of condensate disposal and vent terminal location. Criteria used to develop these guidelines primarily include maintaining draft (or pressure) at the appliance outlet within acceptable limits, avoiding corrosion damage caused by condensate, and withstanding wind loads. ANSI Standards Z21.13, Z21.47, and Z21.10 include performance tests to establish these guidelines and require the manufacturer to publish them in the installation manual. In some cases the appliance manufacturer may specify an acceptable range of draft (or pressure) at the appliance outlet, rather than defining the vent system geometry.

Vent Connectors Vent connectors connect gas appliances to the gas vent, chimney, or single-wall metal pipe, except when the appliance flue outlet or draft hood outlet is connected directly to one of these vent systems. Materials for vent connectors for conversion burners or other appliances without draft hoods must have resistance to corrosion and heat not less than that of galvanized 0.701 mm thick sheet steel. Where a draft hood is used, the connector must have resistance to corrosion and heat not less than that of galvanized 0.475 mm thick sheet steel or type B vent material.

Masonry Chimneys for Gas Appliances A masonry chimney serving a gas-burning appliance should have a tile liner and should comply with applicable building codes such as NFPA Standard 211. The National Fuel Gas Code has other provisions pertaining to masonry chimneys. An additional chimney liner may be needed to avoid slow priming and/or condensation, particularly for an exposed masonry chimney with high mass and low flue gas temperature. A low-temperature chimney liner may be a single-wall passage of pure aluminum or stainless steel or a double-wall type B vent.

Type B and Type L Factory-Built Venting Systems Factory-prefabricated vents are listed by Underwriters Laboratories for use with various types of fuel-burning appliances. These vents should be installed according to the manufacturer’s instructions and the appliance’s listing. Type B gas vents are listed for vented gas-burning appliances. They should not be used for incinerators, appliances readily converted to the use of solid or liquid fuel, or combination gas-oil burning appliances. They may be used in multiple gas appliance venting systems. Type BW gas vents are listed for vented wall furnaces certified as complying with the pertinent ANSI standard. Type L venting systems are listed by Underwriters Laboratories in 75 to 150 mm sizes and may be used for those oil- and gas-burning appliances (primarily residential) certified or listed as suitable for this type of venting. Under the terms of the listing, a single-wall connector may be used in open accessible areas between the appliance’s outlet and the type L material in a manner analogous to type B. Type L piping material is recognized in the National Fuel Gas Code and NFPA Standard 211 for certain connector uses between appliances such as domestic incinerators and chimneys.

Gas Appliances Without Draft Hoods The equations may be used to calculate chimney size for nonresidential gas appliances with the draft configurations listed as 1 and 3 at the beginning of this chapter. Draft conditions 1 and 3 (see draft

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Chimney, Vent, and Fireplace Systems conditions 1, 2, and 3 under Terminology and Table 8) for residential gas appliances, such as boilers and furnaces, may require special vent systems. The appliance test and certification standards include evaluation of manufacturers’ appliance installation instructions (including the vent system) and of operating and application conditions that affect venting. The instructions must be followed strictly.

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Conversion to Gas Installation of conversion burner equipment requires evaluating for proper chimney draft and capacity by the methods in this chapter or by conformance to local regulations. The physical condition and suitability of an existing chimney must be checked before it is converted from a solid or liquid fuel to gas. For masonry chimneys, local experience may indicate how well the construction will withstand the lower temperature and higher moisture content of natural or liquefied petroleum gas combustion products. The section on Masonry Chimneys for Gas Appliances has more details. The chimney should be relined, if required, with corrosionresistant masonry or metal to prevent deterioration. The liner must extend beyond the top of the chimney. The chimney drop-leg (bottom of the chimney) must be at least 100 mm below the bottom of the connection to the chimney. The chimney should be inspected and, if needed, cleaned. The chimney should also have a cleanout at the base.

8.

OIL-FIRED APPLIANCE VENTING

Oil-fired appliances requiring venting of combustion products must be installed and vented in accordance with ANSI/NFPA Standard 31. The standard offers recommendations for metal relining of masonry chimneys. Recommendations for minimum chimney areas for oil-fired natural-draft appliances are offered in Tables 3 and 4 in the HYDI (1989). Implementation of the U.S. National Appliance Energy Conservation Act (NAECA) of 1987 brought attention to heating appliance efficiency and the effect of NAECA on existing chimney systems. Oil-fired appliances maintained a steady growth in efficiency since the advent of the retention-head oil burner and its broad application in both new appliances and the replacement of older burners in existing appliances. Higher appliance efficiencies brought about lower flue gas temperatures. Reduced firing rates became more common as heating appliances were more closely matched to the building heating load. Burner excess air levels also dropped, which resulted in lower mass flows through the chimney and additional reductions in the flue gas temperature. However, the improvements in overall appliance efficiency have not been accompanied by upgrades in existing chimney systems, and upgraded systems are not commonly applied in new construction. An upgrade in a vent system probably involves application of corrosion-resistant materials and/or the reduction in heat loss from the vent system to maintain draft and reduce condensation on interior surfaces of the vent system.

Condensation and Corrosion Condensation and corrosion within the vent system are of growing concern as manufacturers of oil-fired appliances strive to improve equipment efficiencies. The conditions for condensation of the corrosive components of the flue gas produced in oil-fired appliances involve a complex interaction of the water formed in the combustion process and the sulfur trioxide formed from small quantities of sulfur in the fuel oil. The sulfur is typically less than 0.5% by mass for no. 2 fuel oil. The dew point of the two-component system (sulfuric acid and water) in the flue gas resulting from the combustion of this fuel is about 107 to 116°C. This is similar to the effect on dew point of fuel gas sulfur (see Figure 5 in Chapter 28 of the 2017

35.19 ASHRAE Handbook—Fundamentals). In determining the proper curve for fuel oil in that figure, a value of 111 may be used. This value applies to no. 2 fuel oil with 0.5% sulfur by mass. The effect of post-combustion air dilution on the dew point characteristics of flue gas from oil-fired appliances is highly dependent on the presence of sulfur in the fuel. Verhoff and Banchero (1974) developed an equation relating flue gas dew point to the partial pressures of both water and sulfuric acid present in the flue gas. Predictions obtained using this equation are in good agreement with experimental data. Applying this equation to the flue gas from the combustion of no. 2 fuel oil without sulfur and with varying sulfur contents reveals that a broad range of dew point temperatures is possible. For a fuel oil with zero sulfur and 20% excess combustion air, dew point of the combustion products is calculated as approximately 45.6°C, typical for the presence of water formed in combustion. With the addition of post-combustion dilution air to a level equivalent to increasing the excess air in the flue gas to 100%, the apparent water dew point is reduced to 37.2°C. For fuel oil with 0.25% sulfur and 20% excess combustion air, the dew point is elevated to 109.4°C because of formation of dilute sulfuric acid in the flue gas. With the addition of post-combustion dilution air, the dew point is 113.9°C at 100% excess air and 112.2°C at 200% excess air. The implication of these calculations is that, for combustion of fuel oil, the flue gas dew-point temperature can range from a low at the apparent water dew point (calculated with no fuel sulfur) up to some elevated dew point caused by the presence of sulfur in the fuel. With no sulfur in the fuel, adding post-combustion dilution air to the flue gas has some effect on depressing the apparent water dew-point temperature. Adding post-combustion dilution air has no significant effect on the elevated flue gas dew point when sulfur is present in the fuel. It is difficult to meet the flue-gas-side material surface temperatures required to exceed the elevated dew point at all points in the venting system of an oil-fired appliance. Even if the dew-point temperature is not reached at these surfaces, however, a surface temperature approaching 93.3°C allows condensation of sulfuric acid at higher concentrations and results in lower rates of corrosion. The condensed acid concentration is critical to the applicability of plain carbon and stainless steels in vent connectors and chimney liners. The flue-gas-side surface temperatures of conventional connectors and masonry chimney tile liners are often at or below the dew point for some portion of the burner on period. During cooldown (burner off period), these surface temperatures can drop to below the apparent water dew point and, in some cases, to ambient conditions. This is of great concern because, while system surfaces are below the apparent water dew point during burner operation, the condensed sulfuric acid is formed in concentrations well below limits acceptable for steel connectors. This can be seen from an interpretation of a sulfuric acid/water phase diagram presented by Land (1977). An estimate of the condensed liquid sulfuric acid concentration at a condensing surface temperature of 48.9°C, for example, shows a concentration for the condensed liquid acid of about 10 to 20%. According to Fontana and Greene (1967), the relative corrosion of plain carbon steel rises rapidly at sulfuric acid concentrations below 65%. According to Land (1977), for the condensed liquid acid concentration to rise above 65%, the condensing surface temperature must be above 93.3°C. However, according to Fontana and Greene, at acid concentrations above 65%, corrosion rates increase at metal surface temperatures above 79.4°C. This presents the designer with a restrictive operating range for steel surfaces (i.e., between 79.4 and 93.3°C). This is a compromise that does not completely satisfy either the acid concentration or the metal tem-

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perature criterion, but should minimize the corrosion rates induced by each. Another important phenomenon is that when vent system surfaces are at or below the apparent water dew point, a large amount of water condensation occurs on these surfaces. This condensate contains, in addition to sulfuric acid, quantities of sulfurous, nitrous, carbonic, and hydrochloric acids. Under these conditions, the corrosion rate of commonly used vent materials is severe. Koebel and Elsener (1989) found that corrosion rates increase by a factor of 10 when material temperatures on the flue gas side are allowed to fall below the apparent water dew point. The applicability of ordinary stainless steels is very limited and generally follows that of plain carbon steel. Nickel-molybdenum and nickel-molybdenum-chromium alloys show good corrosion resistance over a wide range of sulfuric acid concentrations for surface temperatures up to 104.4°C. This is also true for high-silicon cast iron, used in heat exchangers for oil-fired appliances, and which might find application as a liner system for masonry and metal chimneys.

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Connector and Chimney Corrosion Water and acid condensation can each result in corrosion of the connector wall and deterioration of the chimney material. Although there is little documentation of specific failures, concern in the industry is growing. The volume of anecdotal information regarding corrosive failures is significant and well supported by findings from the heating industry. For oil-fired appliances, the rate of acid corrosion in the connector and chimney is a function of combustion factors and operational factors. The sulfur content of the fuel and the percent excess combustion air are the major combustion-related factors. The frequency and duration of equipment on and off periods, draft control dilution air, and rate of heat loss from the vent system are the major operational factors. In terms of combustion, Butcher and Celebi (1993) found a direct correlation of acid deposition (condensation) rate and subsequent corrosion to fuel sulfur content and excess combustion air. In general, reductions in fuel sulfur and excess combustion air reduce the amount of sulfuric acid produced in combustion and delivered to condensing surfaces in the appliance heat exchanger and carried over into the vent system. From an operational standpoint, long equipment on and short off periods and low vent system heat loss result in shorter warm-up transients and higher end-point temperatures for surfaces exposed to the flue gas. Within the limits of frictional loss, reduced vent sizes increase flue gas velocities and vent surface temperatures.

Vent Connectors An oil-fired appliance is commonly connected to a chimney through a connector pipe. Generally, a draft control device in the form of a barometric damper is included as a component part of the connector assembly. With the advent of new power burners having high static pressure capability, draft control devices in the vent system have become less important, although many local codes still require their use. The portion of the connector assembly between the appliance flue collar and the draft control is called the flue connector; the portion between the draft control and the chimney is called the stack or chimney connector. Chimney connectors are usually of single-wall galvanized steel. The required wall thickness for these connectors varies as a function of pipe diameter. For example, in accordance with Table 9.2.2.3 in NFPA Standard 211, the material thickness for galvanized steel pipe connectors between 150 and 250 mm in diameter is set at 0.58 mm. In accordance with the 2015 National Fuel Gas Code, Paragraph 12.11.2.3(2), vent, chimney, stack, and flue connectors should not be covered with insulation except listed insulated connectors

installed according to the terms of their listing. Because single-wall connectors must remain uninsulated for inspection, substantial cooling of the connector wall and the flue gas can occur, especially with long connector runs through spaces with low ambient temperatures. Close examination of the connector joints, seams, and surfaces is essential whenever the heating appliance is serviced. If the connector is left unrepaired, corrosion damage can cause a complete separation failure of the connector and leakage of flue gas into the occupied space. Where corrosion in the connector has proven to be a chronic problem, consider replacing the connector with a type L vent pipe or its listed equivalent. One product configuration consists of connector pipe with a double wall (stainless steel inner and galvanized steel outer with 6.4 mm gap). The insulated gap of this type of doublewall connector elevates inner wall surface temperature and reduces the overall connector heat loss.

Masonry Chimneys for Oil-Fired Appliances A masonry chimney serving an oil-fired appliance should have a tile liner and should comply with applicable building codes such as NFPA Standard 211. An additional listed chimney liner may be needed to improve thermal response (warm-up) of the inner chimney surface, thereby reducing transient low draft during start-up and acid/water condensation during cyclic operation. This is particularly true for exposed exterior high-mass chimneys but does not exclude cold interior chimneys serving oil-fired appliances that produce relatively low flue gas temperatures. Application of insulation around tile liners within masonry chimneys is common in Europe and may be worth considering in chimney replacement or new construction. A computational analysis by Krajewski (1996) using the U.S. Department of Energy’s Oil Heat Vent Analysis Program (OHVAP; version 3.1) to analyze a series of masonry chimney systems with various firing rates and exit temperatures revealed that current applications of modern oil-fired heating appliances may have problems with acid/water condensation during winter operation. For residential oil-fired heating appliance firing rates below 1.3 to 1.6 mL/s with flue-loss steady-state efficiencies of 82% or higher, exterior masonry chimneys may need special treatment to reduce condensation. For conservative design, listed chimney liners and listed type L connectors may be required for some exterior chimneys serving equipment operating under these conditions.

Replacement of Appliances The physical condition and suitability of an existing chimney must be checked before installation of a new oil-fired appliance. The chimney should be inspected and, if necessary, cleaned. In accordance with NFPA Standard 211, the chimney drop-leg (bottom of the chimney flue) must be at least 200 mm below the bottom of the appliance connection to the chimney. The liner must be continuous, properly aligned, and intact and must extend beyond the top of the chimney. The chimney should also have a properly installed, reasonably airtight clean-out at the base. For masonry chimneys, local experience may indicate how well the construction has withstood the lower temperatures produced by a modern oil-fired appliance. Visual examination reveals evidence of potential or existing chimney damage. Exterior indicators such as missing or loose mortar/bricks, white deposits (efflorescence) on brickwork, a leaning chimney, or water stains on interior building walls should be investigated further. Interior chimney examination with a mirror or video camera can reveal damaged or missing liner material. Any debris collected in the chimney base, drop-leg, or connector should be removed and examined. If any doubt exists about the chimney’s condition, examination by an experienced professional is recommended. Kam et al. (1993) offer specific guidance

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Chimney, Vent, and Fireplace Systems on the examination and evaluation of existing masonry chimneys in the field.

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9.

FIREPLACE CHIMNEYS

This section is condensed from an ASHRAE Journal article; for more details, please see Stone (1969). Fireplaces with natural-draft chimneys follow the same gravity fluid flow law as gas vents and thermal flow ventilation systems. All thermal or buoyant energy is converted into flow, and no draft exists over the fire or at the fireplace inlet. Formulas have been developed to study a wide range of fireplace applications, but the material in this section covers general cases only. Mass flow of hot flue gases through a vertical pipe is a function of rate of heat release and the chimney area, height, and system pressure loss coefficient k. The flow induced in a vertical pipe has a limiting value. A fireplace may be considered as a gravity duct inlet fitting with a characteristic entrance-loss coefficient and an internal heat source. A fireplace functions properly (does not smoke) when adequate intake or face velocity across those critical portions of the frontal opening nullifies external drafts and internal convection effects. In a fireplace-chimney system, the equations assume that all potential buoyant energy is converted into flow as controlled by various losses. This system is analogous to a gas venting system with a draft hood, thereby allowing use of similar concepts as a starting point for size or capacity analysis. The amount of available draft ahead of the fireplace opening is insignificant and need not be considered. Because chimney efficiency, by one definition, equals available draft divided by theoretical draft, the numerical efficiency value approaches zero. Thus, the flow conversion basis is preferable for design over the efficiency approach. System parameters for preventing flue gas spillage from a draft hood or similar collection fitting, can be computed with considerable certainty when heat input is constant or cannot exceed a predictable limiting value. Fireplaces, however, can be fired at extremes ranging from smoldering embers to flash fires of newspapers or dry kindling. Normal opening width and length allow greater access of combustion air to fires than a typical chimney can carry away; thus, combustion overloading occasionally leads to some smoking. At low rates of combustion, airflow velocities into the fireplace face are less than the velocity of natural convection currents induced at side walls by heat stored in the brick, allowing wisps of smoke to stray away from the combustion products’ main flow path. Smoking tendencies are compounded at low firing rates by indoor/ outdoor pressure differentials caused by winds, thermal forces, or fans, because the accompanying thermal force (buoyancy) of low combustion products temperature is insufficient to overcome strong wind or fan effects. In the following analysis, note that fireplaces are primarily aircollecting hoods, diluting a small amount of combustion products with large amounts of air. Maximum mass flow of air into any given fireplace chimney is limited, and diminishes past a certain maximum. Thus, as combustion rate increases, combustion product temperatures rise to the point where masonry cracks; metals overexpand, warp, and oxidize; and steady smoking can occur because heated flue gases evolve beyond the limited capacity of the chimney. An inoperative fireplace is completely at the mercy of indoor/ outdoor pressure differences caused by winds, building stack effects, and operation of forced-air heating systems or mechanical ventilation. Thus, smoking during start-up can have complex causes seldom related to the chimney. Increasingly in new homes and especially in high-rise multiple family construction, fireplaces of normal design cannot cope with mechanically induced reverse flow or shortages of combustion air. It is mandatory in these circumstances to treat and design a fireplace as a constantly operating

35.21 mechanical exhaust system, with induced-draft blowers (mechanical-draft systems) that can overpower other mechanized air-consuming systems, and can develop sufficient flow to avoid smoking and excessive flue temperatures. The gravity-flow capacity equation (Kinkead 1962) of a fireplace-chimney system equates mass flow with the resultant system driving forces and losses.  1  2 w = Ac  2gH [m (o – m)]12 -----------  k 

(22)

where w Ac g H k m o

= = = = = = =

flue gas flow rate, kg/s chimney flue cross-sectional area, m2 gravitational constant, 9.8 m/s2 height of chimney above lintel, m system equivalent resistance coefficient, dimensionless flue gas density at mean temperature, kg/m3 air density of ambient temperature, kg/m3

From Equation (22), it is possible to develop a relationship for average frontal velocity, maximum chimney capacity, and variation of gas temperature with changes in fuel input rate. Using resistance coefficients in these compact systems is preferable to the usual method of equivalent lengths. The summation term k in a vertical chimney is the total of four individual resistance factors: • Acceleration ka of ambient air to flue gas velocity, a constant value that must always be included in the total; ka = 1.0 • Inlet loss coefficient ki for fireplace configuration, including smoke shelf Cone-type fireplaces ki = 0.5 Masonry (damper throat = 2 × flue area) ki = 1.0 Masonry (damper throat = flue area) ki = 2.5 • Chimney flue pipe friction kc at a typical Reynolds number of 10 000 and roughness of 0.001, where rh = hydraulic radius, m H  kc = 0.0083 ----  r h 

• Termination coefficient kt For open top pipe or tile, same size as chimney Disk or cone cap at D/2 above outlet Manufactured caps

kt = 0 kt = 0.5 kt = 0 to 4.0

For a 3.6 m high open top chimney, 300 mm in diameter on a typical fireplace, the system resistance is ka = ki = kc = 0.0083(12/0.25) = kt =

1.0 2.5 0.4 0.0

k = 3.9 summation

Note that in a short chimney, the wall friction coefficient kc is only 0.4 and has a minor effect on system flow. Greater or lesser chimney roughness, or a change from low to high heat loss materials will have little bearing on fireplace effectiveness in short chimneys. In some situations, it may be necessary, for completeness, to include a k term for air supply resistance. To determine the frontal velocity VF of ambient air, the term w is replaced using the substitution w = o AFVF (23) where AF = fireplace frontal opening area, m2

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Fig. 17 Effect of Chimney Gas (Combustion Products) Temperature on Fireplace Frontal Opening Velocity Fig. 16 Eddy Formation VF = fireplace mean frontal velocity, m/s

Accordingly, mean frontal velocity VF becomes A  2gH1  2  m   o –  m  VF = -----c-  ----------- -----------------------------o AF  k 

1 2

(24)

For present purposes, the relative molecular mass or density of dilute combustion products is practically the same as that of air, and both can be expressed with adequate accuracy in terms of absolute gas temperature by the same relationship: B o = 3.48 -----oTo

(25)

B m = 3.48 ------o Tm

(26)

where Bo = existing barometric pressure, kPa To = ambient air temperature, K Tm = mean combustion products temperature, K

Substitution of the density/temperature relationships (Equations [25] and [26]) into Equation (24) allows further simplification, leading to the general frontal velocity expression: A  2gH1  2  T o  T m – T o   1  2 VF = -----c-  ----------- ---------------------------------AF  k  Tm

(27)

Here, frontal velocity is a function of the product of three terms: • Dimensionless area ratio Ac/AF • Height/resistance term (2gHk)12 • Dimensionless temperature effects [To(Tm – To)]12Tm

For the 3.6 m high example chimney, assume ambient temperature (for calculation purposes) is 21°C, or To = 294 K indoors and outdoors, with no air supply resistance. Equation (27) expresses variation in VF with gas temperature as shown in Figure 17. Assume also Ac /AF = 0.10, so that frontal opening is ten times chimney area. The mean airflow velocity into a fireplace frontal opening is nearly constant from 165 K flue gas temperature rise up to any higher temperature. Local velocities vary within the opening. Depending on design, air typically enters horizontally along the hearth and then flows into the fire and upward, clinging to the back wall (see Figure 16). A recirculating eddy forms just inside the upper front half of the opening, induced by the high velocity of flow along the back. Restrictions or poor construction in the throat area between the lintel and damper also increase the eddy. Because the eddy moves smoke out of the zone of maximum velocity, the tendency of this smoke to escape must be counteracted by some minimum inward air movement over the entire front of the fireplace, particularly under the lintel. Construction of a fireplace, its internal configuration, damper location, height and location of lintel, slope of back and sides, and so on all affect minimum frontal velocity to prevent smoking with ordinary fires. It seems desirable to maintain a smooth, gradual tapering transition between the hearth or flame region up into the damper location. A sudden transition, unless it is well above the lintel, induces velocity components that tend to increase eddying. With a shallow chamber between lintel and damper zone, there is insufficient volume for convection currents, and some flue gases may be diverted horizontally before being captured by the main flow. Masons following the dimensional parameters recommended by Ramsey and Sleeper (1956) or in damper literature can avoid these design flaws. With prefabricated fireplace systems, which often tend to be unconventional, careful testing is essential to ensure safe, smoke-free performance at minimum chimney height. A minimum mean frontal inlet velocity of 0.244 m/s, in conjunction with a chimney flue gas temperature of at least 165 to 280 K

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Fig. 18 Permissible Fireplace Frontal Opening Area for Design Conditions (0.24 m/s mean frontal velocity with 0.3 m inside diameter round flue) above ambient, should control smoking in a well-constructed conventional masonry fireplace. As noted in Figure 17, this velocity can be achieved even in low chimneys by system resistance of 2.9 or less, in conjunction with rates of combustion producing flue gas temperatures above a certain minimum level. Figure 17 also illustrates why increases in flue gas temperature rise greater than 165 K have no perceptible effect on fireplace smoking, because combustion air mass flow and face velocity actually decrease at flue gas temperature rises higher than 280 K. For practical purposes, chimney flue gas temperature has little influence on fireplace performance, if flue gas temperature is at least 165 K above ambient. Fireplace performance analysis thus can be continued assuming a constant flue gas temperature of 280 K rise above ambient: [T o  T m – T o ] 1  2 ------------------------------------------ = 0.5 Tm where Tm = 572 K and To = 294 K. Assuming 21°C ambient (To = 294 K), and factoring out 2g, Equation (27) becomes roughly A  H 1  2 VF = 2.21 -----c- ----- AF  k 

(28)

This expression states that relative fireplace performance is purely a matter of geometry. It allows evaluation of the opening size to maintain minimum frontal velocity as a function of height, as well as quick analysis of effect of frontal area on velocity. Figure 18 shows that 0.24 m/s face velocity requires a relatively small AF Ac area ratio at 2.4 m chimney height, whereas the fireplace frontal opening area can be nearly twice as large with a 12 m high chimney. To compute this curve, system resistance is assumed as k = 2.5 + 0.033HD D = 0.3 m diameter This expression of resistance assumes the fully open free area of the damper throat to be twice chimney flue area. A corollary application of Equation (28) assumes fixed chimney size and height, and explores variation in frontal velocity with changes in area ratio. Figure 19 shows that a 0.3 m diameter, 4.6 m

Fig. 19 Effect of Area Ratio on Frontal Velocity (for chimney height of 4.6 m with 0.3 m inside diameter round flue) high chimney cannot produce adequate velocities for frontal area ratios greater than 11. These curves point out the possibility of further simplification to yield a fireplace-chimney design equation for a constant face velocity of 0.24 m/s.  H 1  2 AF = 9.06Ac ----- k 

(29)

This equation allows permissible frontal area AF to be determined as a function of chimney area, height, and system resistance. It can also be used to determine Ac, if AF is known. However, in this latter case, Equations (30) and (31) are preferable, because chimney size (Dc, Ac, and Rh) appears twice in the right-hand side of each equation. For circular flues, 

AF =

1  2

2 - 7.12 D c --------------------------------H

H

2.5 + 0.033 ------  Dc  

(30)

For other flue shapes,

 1  2 H ----------------------------------- AF = 9.06Ac  H 2.5 + 0.0083 ------  Rh  

(31)

where Dc is chimney inside diameter, m, Rh is inside hydraulic radius, m, frontal velocity VF is 0.244 m/s at maximum combustion air mass flow rate, and damper opening free area is twice chimney flue area combustion air AF . These relationships clearly reveal the origin of rules of thumb specifying opening area as 8, 10, or 12 times chimney area. Some design guides go beyond and classify chimneys by height groups so that short ones serve smaller fireplace openings. Using a mean face velocity of 0.244 m/s yields ratios and heights that fall well within the limits of such rules, as well as providing a unifying concept for computing design charts.

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The previous relationships primarily apply to masonry singleface fireplaces of conventional construction, but are applicable to other types with considerable validity if face or opening area is properly treated. Corner or double-face designs and many freestanding types embody conventional smokeshelf construction with similar resistance coefficients. The preceding equations can be applied to illustrate the effect of excessive firing rates on chimney flue gas temperature. Masonry fireplaces are highly inefficient as heating devices, and tests show that, over a wide range of controlled fuel inputs using a drilled-port nonaerated gas burner, 75 to 80% of the gross heating value goes up the chimney. With constant flue heat loss of 80%, 70% of heat loss is sensible heat, which produces the rise in flue gas temperature. This heat input/temperature relationship may be developed by expressing the system flow relationship in terms of heat input: q w = ----------------------------cp  Tm – To 

(32)

where

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q = heat content of chimney flue gases, kW cp = specific heat of chimney gases, assumed to be approximately 1.046 kJ/(kg·K)

Fig. 20

Variation of Chimney Flue Gas Temperature with Heat Input Rate of Combustion Products

Equating Equations (32) and (22), and eliminating w,  2gH1  2 q ----------------------------- = Ac  ----------- [m (o – m)]12 cp  Tm – To   k 

(33)

Substituting Equations (25) and (26) for density and solving for q, 3.48B o A c c p  2gH1  2  T m – T o  3  2 ------------------------------Tm  k   To 

-  ----------- q = ---------------------------0.5

(34)

For any given system, all terms except (Tm – To)3/2/Tm may be considered fixed; therefore, gas temperature rise for a fireplace can be obtained as a concise function of heat input: 1.5

 Tm – To  q = C  Tm

(35)

where C is a constant. The flow-temperature function of a carefully controlled experimental fireplace (Equation [34] or [35]) in which C as fixed is plotted in Figure 20 as chimney flue gas temperature versus heat input rate. Data taken on a fireplace-chimney system in this way may readily be evaluated to determine system resistance coefficient k. Figure 21 shows fireplace and chimney dimensions for the specific conditions of circular flues at 0.244 m/s frontal velocity. This chart solves readily for maximum frontal opening for a given chimney, as well as for chimney size and height with a predetermined opening. For example, a 1050 mm high by 750 mm wide opening for a 4 m high chimney (measured from the highest point of front opening) requires a 350 mm flue. Figure 21 assumes no wind or air supply difficulties. For other face velocities, AF is found by multiplying frontal area (center scale) by velocity ratio 0.244/VF . To confirm the example results, calculate from Equation (30) as follows: 12 H 2 AF = 7.12 D c ---------------------------------------------- 2.5 + 0.033H   D c

where H is 44 m and Dc is 0.35 m.

Fig. 21 Chimney Sizing Chart for Fireplaces Mean face velocity = 0.244 m/s (Stone 2005)

AF = 0.788 m2 = 0.75 m  1.05 m Although derived specifically for circular flues, AF applies with negligible sacrifice in performance to chimney flue cross sections such as squared or rounded ovals, because flue area is a much more important factor than friction caused by changes in hydraulic radius. For example, in a 6.1 m high chimney, assuming a square flue section equal in area to a 203 mm circle, frontal area is reduced from 0.386 m2 with the round, to 0.380 m2 with the square, or about 2%, a difference that is hardly observable. For some typical constructions, Figure 22 suggests methods of estimating frontal area. These relationships apply to steady-state conditions, which are obtained only after warm-up. Igniting a rapidly flammable charge in a cold system creates pulses of expanding hot combustion products,

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Chimney, Vent, and Fireplace Systems which frequently escape from the fireplace. In a typical chimney, priming time (time to accelerate from no flow to full upward velocity) is around 5 to 10 s. Thus, initial or intermittent smoking caused by momentarily excessive combustion rates occurs because of system inability to increase flue gas velocity in pace with combustion surges. Flue or chimney material is of little relevance to fireplace-chimney operation. Materials to which these equations and charts apply include the very hazardous uninsulated single-wall metal, through conventional masonry, as well as the various constructions of lightweight, insulated, factory-built low-heat-space-appliance chimneys. (Safety standards classify fireplaces as low-heat appliances.) Because most fireplace chimneys are short and vertical, neither heat loss nor wall roughness has any important effect on flow. The governing factors in chimney selection for fireplaces are mainly safety, installation, convenience, and esthetics. Indoor/outdoor pressure differences caused by winds, kitchen or bath exhaust fans, building stack effects, and operation of forced-air heating systems or mechanical ventilation affect the operation of a fireplace. Thus, smoking during start-up can be caused by factors unrelated to the chimney. Frequently, in new homes (especially in high-rise multiple-family construction), fireplaces of normal design cannot cope with mechanically induced reverse flow or shortages of combustion air. In these circumstances, a fireplace should include an induced-draft blower able to overpower other mechanized airconsuming systems. An inducer for this purpose is best located at the chimney outlet and should produce 0.244 to 0.305 m/s fireplace face velocity of ambient air in an individual flue or 3.05 to 3.66 m/ s chimney velocity. In conventional fireplaces, the greater the frontal velocity, the more freedom from smoking. The damper free area, together with its resultant resistance coefficient, are thus major factors in obtaining good masonry fireplace performance, especially with short chimneys. Tests show that damper free area need not exceed twice

Fig. 22 Estimation of Fireplace Frontal Opening Area

35.25 the required flue area, because little further resistance reduction occurs past this limit. If the damper selected has a free area equal to or less than the required area, it will be definitely restrictive, despite complete adequacy of other factors. Manufacturers’ literature seldom includes damper free area or opening dimensions, and the dimensions may vary further after installation because of interferences with lintels and other parts. It is expedient to select dampers of adequate free area for best results. Partially closing a damper during a vigorous fire illustrates this point; what is not so obvious is that greater damper openings may be needed in some cases to control smoke by achieving adequate frontal velocities. Many free-standing fireplaces are built without the usual smokeshelf/throat damper configuration. The same parameters and relationships apply to free-standing fireplaces as to masonry fireplaces; however, in many designs it is difficult to assign a true frontal area for velocity analysis. Where freestanding fireplaces include back outlets, or require a horizontal run to reach the chimney, compensation is necessary for the flow resistance or pressure losses caused by turns. As a rough approximation, increase the system resistance coefficient k about 1.5 for two 90° elbows, or a back outlet with lateral run into a tee. Losses caused by lateral connectors and tees generally mean a one-size increase (e.g., moving from 150 to 200) is sufficient. Collar sizes generally determine correct chimney size, and instructions are furnished to cover special situations. More sophisticated prefabricated fireplace designs are provided with matching correctly sized chimneys, and are intended for installation in conventional woodframe residences. The equations and design charts presented here assume no wind or air supply difficulties. Lack of replacement air, competing ventilation exhaust fans, and negative interior pressures caused by winds are all obvious causes of smoking or poor fireplace priming. Even when fully primed and hot, thermal forces in a fireplace chimney can be overpowered by a combination of adverse influences. In modern high-rise residential apartments, where an effort has been made to provide all amenities, fireplaces may have to cope simultaneously with all these troubles. Continuous induced draft for the chimney alleviates most of these problems by maintaining a chimney prime at all times. An inducer for this purpose should be able to produce 0.24 to 0.30 m/s fireplace frontal velocity of ambient air in any individual flue. Where multiple flues are installed in a chase, a single, larger inducer serving the chase can be sized for the combined fireplace frontal opening area. Even where flues are of different heights and sizes, a draft inducer selection, assuming all flues to be some compromise median height and size, produces far greater user satisfaction than reliance on gravity alone. In single-family dwellings, fireplace problems are more often caused by reduced interior pressures from wind effects than by poor chimney terminal location or characteristics. Efforts to cure smoking, slow priming, or blowback of ashes usually involve one of the myriad forms of stationary or rotating caps, cowls, or chimney pots. This questionable expedient has contradictory effects. Usually, the added still-air resistance of a cap reduces fireplace frontal velocity, which limits combustion airflow rate, and thus may tend to increase smoking. On the other hand, a cap reduces air dilution of smaller fires, raises chimney temperature, and improves stability of flame, thus tending to mitigate wind impulses that cause momentary flow reduction. The usual fireplace damper can also be used to restrict flow, and thus raise temperature. Remedies for fireplace malfunctions may be analyzed using Equation (28). For example, it is apparent that any change of parameters on the right-hand side that might decrease VF can increase smoking tendencies. If frontal area AF or k increases, there will be a corresponding decrease in VF . Similarly, if chimney area Ac or chimney height H are

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reduced, then VF decreases. Further, because frontal velocity varies as the square root of the term H/k, it is more effective to reduce frontal area, thereby increasing Ac /AF , than to increase H or reduce k. Logical expedients for increasing VF frontal velocity, and thus improving performance of fireplaces and chimneys, include the following: • Increase chimney height (using the same flue area) and extend the last tile 150 mm upward, or more. • Decrease frontal opening by lowering the lintel, or raising the hearth. (Glass doors may help by increasing VF.) • Increase free area through damper. (Check that it opens fully without interferences.) CSA Standard P.4.1-02 can be used for measuring annual efficiency in Canada.

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10.

AIR SUPPLY TO FUEL-BURNING APPLIANCES

Failure to supply outdoor air for combustion may result in erratic or even dangerous operating conditions. A correctly designed gas appliance with a draft hood can function with short vents (1.5 m high) using an outdoor air supply opening as small in area as the vent outlet collar. Such an orifice, when equal to vent area, has a resistance coefficient in the range of 2 to 3. If the air supply opening is as much as twice the vent area, however, the coefficient drops to 0.5 or less. The following rules may be used as a guide: • Residential heating appliances installed in unconfined spaces in buildings of conventional construction do not ordinarily require ventilation other than normal air infiltration. In any residence or building that has been built or altered to conserve energy or minimize infiltration, the heating appliance area should be considered a confined space. The air supply should be installed in accordance with NFPA Standard 54/AGA Z223.1, CSA Standard B149.1, or the following recommendations. • Residential heating appliances installed in a confined space having unusually tight construction require two permanent openings to an unconfined space or to the outdoors. An unconfined space has a volume of at least 4.8 m3 per kilowatt (m3 per kJ/s) of the total input rating of all appliances installed in that space. Free opening areas must be greater than 550 mm2 per kilowatt input with vertical ducts or 1100 mm2 per kilowatt with horizontal ducts to the outdoors. The two openings communicating directly with sufficient unconfined space must be greater than 2200 mm2 per kilowatt. Upper openings should be within 300 mm of the ceiling; lower openings should be within 300 mm of the floor. • Complete combustion of natural and propane gas or fuel oil requires approximately 0.27 m3 of air, at standard conditions, for each megajoule of fuel burned, but excess air is usually required for proper burner operation. • The size of these air openings may be modified if special engineering ensures an adequate supply of air for combustion, dilution, and ventilation or if local ordinances apply to boiler and machinery rooms. • In calculating free area of air inlets, consider the blocking effect of louvers, grilles, or screens protecting openings. Screens should not be smaller than 6.4 mm mesh. If the free area through a particular louver or grille is known, use that in calculating the size opening required to provide the free area specified. If the free area is not known, assume that wood louvers have 20 to 25% free area and metal louvers and grilles have 60 to 75% free area. • Mechanical ventilation systems serving the fuel-burning appliance room or adjacent spaces should not be allowed to create negative appliance room air pressure. The appliance room may require tight self-closing doors and provisions to supply air to

spaces under negative pressure so fuel-burning appliances and venting operate properly. • Fireplaces may require special consideration. For example, a residential attic fan can be hazardous if it is inadvertently turned on while a fireplace is in use. • In buildings where large quantities of combustion and ventilation or process air are exhausted, a sufficient supply of fresh uncontaminated makeup air, warmed if necessary to the proper temperature, should be provided. It is good practice to provide about 5 to 10% more makeup air than the amount exhausted.

11.

VENT AND CHIMNEY MATERIALS

Factors to be considered when selecting chimney materials include (1) the temperature of flue gases; (2) their composition and propensity for condensation of water vapor from combustion products (dew point); (3) presence of sulfur, halogens, and other fuel and air contaminants that lead to corrosion of the chimney vent system; and (4) the appliance’s operating cycle (condensate dwell time). Materials for vents and chimneys in the 100 to 1220 mm size range include single-wall metal, various multiwall air- and massinsulated types, and precast and site-constructed masonry. Each has different characteristics (e.g., frequency of joints, roughness, heat loss), but the type of materials used for systems 360 mm and larger is relatively unimportant in determining draft or capacity. This does not preclude selecting a safe product or method of construction that minimizes heat loss and fire hazard in the building. National codes and standards classify heat-producing appliances as low, medium, and high heat, with appropriate reference to chimney and vent constructions permitted with each. These classifications are mainly based on size, process use, or combustion temperature. Often, the appliance classification gives little information about outlet gas temperature or venting needed. The designer should, wherever possible, obtain gas outlet temperature conditions and properties that apply to the specific appliance, rather than going by code classification only. Where building codes allow engineered chimney systems, chimney material selection based on gas outlet temperature can save space as well as reduce structural and material costs. For example, in some jurisdictions, approved gas-burning appliances with draft hoods operating at inputs over 117 kW (kJ/s) may be placed in a heat-producing classification that prohibits use of type B gas vents. An increase in input may not cause an increase in outlet temperature or in venting hazards, and most building codes recommend correct matching appliance and vent. Single-wall uninsulated steel stacks can be protected from condensation and corrosion internally with refractory firebrick liners or by spraying calcium aluminate cement over a suitable interior expanded metal mesh or other reinforcement. Another form of protection applies proprietary silica or other prepared refractory coatings to pins or a support mesh on the steel. The material must then be suitably cured for moisture and heat resistance. Moisture condensation on interior surfaces of connectors, vents, stacks, and chimneys is a more serious cause of deterioration than heat. Chimney wall temperature and flue gas velocity, temperature, and dew point affect condensation. Contaminants such as sulfur, chlorides, and fluorides in the fuel and combustion air raise the flue gas dew point. Studies by Beaumont et al. (1970), Mueller (1968), Pray et al. (1942-53), and Yeaw and Schnidman (1943) indicate the variety of analytical methods as well as difficulties in predicting the causes and probability of actual condensation. Combustion products from any fuel containing hydrogen condense onto cold surfaces or condense in bulk if the main flow of flue gas is cooled sufficiently. Because flue gas loses heat through walls, condensation occurs first on interior wall surfaces cooled to the flue gas dew point, forming a dew and then a liquid film and, with further

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Chimney, Vent, and Fireplace Systems cooling, flowing down into zones where condensation would not normally occur. Start-up of cold interior chimney surfaces is accompanied by transient dew formation, which evaporates on heating above the dew point. This phenomenon causes little corrosion when very low sulfur fuels are used. Proper selection of chimney dimensions and materials minimizes condensation and thus corrosion. Experience shows a correlation between the sulfur content of the fuel and the deterioration of interior chimney surfaces. Figure 5 in Chapter 28 of the 2017 ASHRAE Handbook—Fundamentals illustrates one case, which applies to any fuel gas. The figure shows that the flue gas dew point increases at 40% excess air from 53°C with zero sulfur to 71°C with 16.75 mg of sulfur per megajoule of fuel having a heat value of 20.49 MJ/m3. The figure can also be used to approximate the effect of fuel oil sulfur content on flue gas dew point. For example, fuel oil with a sulfur content of 0.5% (by mass) contains about of sulfur per gallon or 4.31 kg of sulfur per cubic metre of fuel. If the fuel heat value is 559 kJ/m3, the ratio that defines the curves in the figure gives a curve value of 111. Estimates for lower percentages of sulfur (0.25, 0.05) can be formed as factors of the value 111. Because the corrosion mechanism is not completely understood, judicious use of resistant materials, suitably insulated or jacketed to reduce heat loss, is preferable to low-cost single-wall construction. Refractory materials and mortars should be acid-resistant, and steels should be resistant to sulfuric, hydrochloric, and hydrofluoric acids; pitting; and oxidation. Where low flue gas temperatures are expected together with low ambients, an air space jacket or mineral fiber lagging, suitably protected against water entry, helps maintain surface and flue gas temperatures above the dew point. Using lowsulfur fuel, which is required in many localities, reduces both corrosion and air pollution. Type 1100 aluminum alloy or any other non-copper-bearing aluminum alloy of 99% purity or better provides satisfactory performance in prefabricated metal gas vent products. For chimney service, flue gas temperatures from appliances burning oil or solid fuels may exceed the melting point of aluminum; therefore, steel is required. Stainless steels such as type 430 or 304 give good service in residential construction and are referenced in UL-listed prefabricated chimneys. Where more corrosive substances (e.g., high-sulfur fuel or chlorides from solid fuel, contaminated air, refuse) are anticipated, type AL 29-4CR® or equivalent stainless steel offers a good match of corrosion resistance and mechanical properties. As an alternative to stainless steel, porcelain enamel offers good resistance to corrosion if two coats of acid-resistant enamel are used on all surfaces. A single coat, which always has imperfections, allows base metal corrosion, spalling, and early failure. Prefabricated chimneys and venting products are available that use light corrosion-resistant materials, both in metal and masonry. The standardized, prefabricated, double-wall metal type B gas vent has an aluminum inner pipe and a coated steel outer casing, either galvanized or aluminized. Standard air space from 6.35 to 12.7 mm is adequate for applicable tests and a wide variety of exposures. Air-insulated all-metal chimneys are available for low-heat use in residential construction. Thermosiphon air circulation or multiple reflective shielding with three or more walls keeps these units cool. Insulated, double-wall residential chimneys are also available. The annulus between metal inner and outer walls is filled with insulation and retained by coupler end structures for rapid assembly. Prefabricated, air-insulated, double-wall metal chimneys for multifamily residential and larger buildings, classed as building heating appliance chimneys, are available (Figure 23). Refractorylined prefabricated chimneys (medium-heat type) are also available for this use. Commercial and industrial incinerators, as well as heating appliances, may be vented by prefabricated metal-jacketed cast refractory

35.27

Fig. 23 Building Heating Appliance, Medium-Heat Chimney Table 10 Underwriters Laboratories Test Standards Steady-State Appliance Flue Gas Temperature Rise, °F

Fuel

517

All

127 Fireplaces, factory-built

517

Solid or gas

311 Roof jacks for manufactured homes and recreational vehicles

517

Oil, gas

378 Draft equipment (such as regulators and inducers)

—

All

441 Gas vents (type B, BW)

267

Gas only

641 Low-temperature venting systems (type L)

278

Oil, gas

959 Chimneys, factory-built, medium-heat

961

All

1738 Venting systems for gasburning appliances, categories II, III, and IV

78 to 267

Gas only

No. Subject 103 Chimneys, factory-built, residential type (includes building heating appliances)

chimneys, which are listed in the medium-heat category and are suitable for intermittent flue gas temperatures to 1100°C. All prefabricated chimneys and vents carrying a listing by a recognized testing

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laboratory have been evaluated for class of service regarding temperature, strength, clearance to adjacent combustible materials, and suitability of construction in accordance with applicable national standards. Underwriters Laboratories (UL) standards, listed in Table 10, describe the construction and temperature testing of various classes of prefabricated vent and chimney materials. Standards for some related parts and appliances are also included in Table 10 because a listed factory-built fireplace, for example, must be used with a specified type of factory-built chimney. The temperature given for the steady-state operation of chimneys is the lowest in the test sequence. Factory-built chimneys under UL Standard 103 are also required to demonstrate adequate safety during a 1 h test at 740 K rise and to withstand a 10 min simulated soot burnout at either 906 or 1128 K rise. These product tests determine minimum clearance to combustible surfaces or enclosures, based on allowable temperature rise on combustibles. They also ensure that the supports, spacers, and parts of the product that contact combustible materials remain at safe temperatures during operation. Product markings and installation instructions of listed materials are required to be consistent with test results, refer to types of appliances that may be used, and explain structural and other limitations.

12.

VENT AND CHIMNEY ACCESSORIES

Vent or chimney system design must consider the existence of or need for accessories such as draft diverters, draft regulators, induced-draft fans, blocking dampers, expansion joints, and vent or chimney terminals. Draft regulators include barometric draft regulators and furnace sequence draft controls, which monitor automatic flue dampers during operation. The design, materials, and flow losses of chimney and vent connectors are covered in previous sections.

Draft Hoods The draft hood isolates the appliance from venting disturbances (updrafts, downdrafts, or blocked vent) and allows combustion to start without venting action. Suggested general dimensions of draft hoods are given in ANSI Standard Z21.12, which describes certification test methods for draft hoods. In general, the pipe size of the inlet and outlet flues of the draft hood should be the same as that of the appliance outlet connection. The vent connection at the draft hood outlet should have a cross-sectional area at least as large as that of the draft hood inlet. Draft hood selection comes under the following two categories: • Those supplied with a design-certified gas appliance: certification of a gas appliance design under pertinent national standards includes its draft hood (or draft diverter). Consequently, the draft hood should not be altered or replaced without consulting the manufacturer and local code authorities. • Those supplied separately for gas appliances: listed draft hoods for existing vent or chimney connectors should be installed by experienced installers in accordance with accepted practice standards. Every design-certified gas appliance requiring a draft hood must be accompanied by a draft hood or provided with a draft diverter as an integral part of the appliance. The draft hood is a vent inlet fitting as well as a safety device for the appliance, and assumptions can be made regarding its interaction with a vent. First, when the hood is operating without spillage, the heat content of flue gases (enthalpy relative to dilution air temperature) leaving the draft hood is almost the same as that entering. Second, safe operation is obtained with 40 to 50% dilution air. It is unnecessary to assume 100% dilution air for gas venting conditions. Third, during certification tests, the draft hood must function without spillage, using a vent with not over

1.5 m of effective height and one or two elbows. Therefore, if vent heights appreciably greater than 1.5 m are used, an individual vent of the same size as the draft hood outlet may be much larger than necessary. When vent size is reduced, as with tall vents, draft hood resistance is less than design value relative to the vent; the vent tables in the National Fuel Gas Code (NFPA Standard 54/AGA Z223.1) give adequate guidance for such size reductions. Despite its importance as a vent inlet fitting, the draft hood designed for a typical gas appliance primarily represents a compromise of the many design criteria and tests solely applicable to that appliance. This allows considerable variation in resistance loss; thus, catalog data on draft hood resistance loss coefficients do not exist. The span of draft hood loss coefficients, including inlet acceleration, varies from the theoretical minimum of 1.0 for certain lowloss bell or conical shapes to 3 or 4, where the draft hood relief opening is located within a hot-air discharge (as with wall furnaces) and high resistance is needed to limit sensible heat loss into the vent. Draft hoods must not be used on appliances having draft configuration 1 or 3 (see conditions 1, 2, and 3 under Terminology and Table 8) that is operated with either power burners or forced venting, unless the appliances have fan-assisted burners that overcome some or most of the appliance flow resistance and create a pressure inversion ahead of the draft hood or barometric regulator. Gas appliances with draft hoods must have excess chimney draft capacity to draw in adequate draft hood dilution air. Failure to provide adequate combustion air can cause oxygen depletion and spillage of flue gases and flame rollout from the combustion air inlet at the burner(s).

Draft Regulators Appliances requiring draft at the appliance flue gas outlet generally use barometric regulators for combustion stability. A balanced hinged gate in these devices bleeds air into the chimney automatically when pressure decreases. This action simultaneously increases vent gas flow and reduces temperature. Well-designed barometric regulators provide constant flue gas static pressure over a span of impressed vent gas draft (i.e., the draft that would exist without regulation) of about 50 Pa. A regulator can maintain 15 Pa draft for impressed drafts from 15 to 65 Pa. If the chimney system is very high or otherwise capable of generating available draft in excess of the pressure span capability of a single regulator, additional or oversize regulators may be used. Figure 24 shows proper locations for regulators in a chimney manifold. Barometric regulators are available with double-acting dampers, which also swing out to relieve momentary internal pressures or divert continuing downdrafts. In the case of downdrafts, temperature safety switches actuated by hot gases escaping at the regulator sense and limit malfunctions.

Vent Dampers Electrically, mechanically, and thermally actuated automatic vent dampers can reduce energy consumption and improve seasonal efficiency of gas- and oil-burning appliances. Vent dampers reduce loss of heated air through gas appliance draft hoods and loss of specific heat from the appliance after the burner stops firing. These dampers may be retrofit devices or integral components of some appliances. Electrically and mechanically actuated dampers must open before main burner gas ignition and must not close during burner operation. These safety interlocks, which electrically interconnect with existing control circuitry, may include an additional main control valve or special gas-pressure-actuated controls. Vent dampers that are thermally actuated with bimetallic elements and have spillage-sensing interlocks with burner controls are

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Chimney, Vent, and Fireplace Systems available for draft hood-type gas appliances. These dampers open in response to gas temperature after burner ignition. Because thermally actuated dampers may exhibit some flow resistance, even at equilibrium operating conditions, carefully follow instructions regarding allowable heat input and minimum required vent or chimney height. Take special care to ensure that safety interlocks with appliance controls are installed according to instructions. Spillage-free gas venting after the damper has been installed must be verified with all damper types. Energy savings of a vent damper can vary widely. Dampers reduce energy consumption under one or a combination of the following conditions: Heating appliance is oversized. Chimney is too high or oversized. Appliance is located in heated space. Two or more appliances are on the same chimney (a damper must be installed on each appliance connected to that chimney). • Appliance is located in building zone at higher pressure than outdoors. This positive pressure can cause steady flow losses through the chimney. Licensed for single user. © 2020 ASHRAE, Inc.

• • • •

35.29 flow resistance for connectors, vents, and chimneys, comprising typical values for draft hoods, elbows, tees, caps, and piping with no allowance for added devices placed directly in the flue gas stream. Thus, heat exchangers or flue gas extractors should offer no flow resistance or negligible resistance coefficients when they are installed. A heat exchanger that is reasonably efficient and offers some flow resistance may adversely affect the system by reducing both flow rate and flue gas temperature. This may cause moisture condensation in the chimney, draft hood spillage, or both. Increasing heat transfer efficiency increases the probability of the simultaneous occurrence of both effects. An accessory heat exchanger in a solidfuel system, especially a wood stove or heater, may collect creosote or cause its formation downstream. Retrofitting heat exchangers in gas appliance venting systems requires careful evaluation of heat recovered versus both installed cost and the potential for chimney safety and operating problems. Every heat exchanger installation should undergo the same spillage tests given a damper installation. In addition, the flue gas temperature should be checked to ensure it is high enough to avoid condensation between the exchanger outlet and the chimney outlet.

Energy savings may not justify the cost of installing a vent damper if one or more of the following conditions exist: • All combustion and ventilation air is supplied from outdoors to direct-vent appliances or to appliances located in an isolated, unheated room. • Appliance is in an unheated basement that is isolated from the heated space. • A one-story flat-roof house has a short vent, which is unlikely to carry away a significant amount of heated air. For vents or chimneys serving two or more appliances, dampers (if used) should be installed on all attached appliances for maximum effectiveness. If only one damper is installed in such systems, loss of heated air through an open draft hood may negate a large portion of the potential energy savings.

Heat Exchangers or Flue Gas Heat Extractors Sensible heat available in flue gas of properly adjusted furnaces burning oil or gas is about 10 to 15% of the rated input. Small accessory heat exchangers that fit in the connector between the appliance outlet and the chimney can recover some of this heat for localized use; however, they may cause some adverse effects. All gas vent and chimney size or capacity tables assume the gas temperature or heat available to create theoretical draft is not reduced by a heat transfer device. In addition, the tables assume

13.

DRAFT FANS

The selection of draft fans, blowers, or inducers must consider (1) types and combinations of appliances, (2) types of venting material, (3) building and safety codes, (4) control circuits, (5) gas temperature, (6) permissible location, (7) noise, and (8) power cost. Besides specially designed fans and blowers, some conventional fans can be used if the wheel and housing materials are heat- and corrosion-resistant and if blower and motor bearings are protected from adverse effects of the flue gas stream. Small draft inducers for residential gas appliance and unit heater use are available with direct-drive blower wheels and an integral device to sense flow (Figure 25A). The control circuit for these applications must provide adequate vent gas flow both before and while fuel flows to the main burner. Other types of small inducers are either saddle-mounted blower wheels (Figure 25B) or venturi ejectors that induce flow by jet action (Figure 25C). An essential safety requirement for inducers serving draft hood gas appliances does not permit appliance interconnections on the discharge or outlet side of the inducer. This requirement prevents backflow through an inoperative appliance. With prefabricated sheet metal venting products such as type B gas vents, the vent draft inducer should be located at or downstream from the point the vent exits the building. This placement keeps the indoor system below atmospheric pressure and prevents flue gas

Fig. 24 Use of Barometric Draft Regulators

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2020 ASHRAE Handbook—HVAC Systems and Equipment (SI)

from escaping through seams and joints. If the inducer cannot be placed on the roof or outdoor wall, metal joints must be reliably sealed in all pressurized parts of the system. Pressure capability of residential draft inducers is usually less than 250 Pa at rated flow. Larger inducers of the fan, blower, or ejector type have greater pressure capability and may be used to reduce system size as well as supplement available draft. Figure 25D shows a specialized axial-flow fan capable of higher pressures. This unit is structurally self-supporting and can be mounted in any position in the connector or stack because the motor is in a well, separated from the flue gas stream. A right-angle fan, as shown in Figure 25E, is supported by an external bracket and adapts to several inlet and exit combinations. The unit uses the developed draft and an insulated tube to cool the extended shaft and bearings. Pressure, volume, and power curves should be obtained to match an inducer to the application. For example, in an individual chimney system (without draft hood) in which a directly connected inducer only handles combustion products, calculation of the power required for continuous operation need only consider volume at operating flue gas temperature. An inducer serving multiple, separately controlled draft hood gas appliances must be powered for ambient temperature operation at full flow volume in the system. At any input, the inducer for a draft hood gas appliance must handle about 50% more standard chimney gas volume than a directly connected inducer. At constant volume with a given size inducer or fan, these demands follow the Fan Laws (see Chapter 21) applicable to power venting as follows: • Pressure difference developed is directly proportional to gas density. • Pressure difference developed is inversely proportional to absolute gas temperature. • Pressure developed diminishes in direct proportion to drop in absolute atmospheric pressure, as with altitude.

• Required power is directly proportional to gas density. • Required power is inversely proportional to absolute gas temperature. Centrifugal and propeller draft inducers in vents and chimneys are applied and installed the same as in any heat-carrying duct system. Venturi ejector draft boosters involve some added consideration. An advantage of the ejector is that motor, bearings, and blower blades are outside the contaminated flue gas stream, thus eliminating a major source of deterioration. This advantage causes some loss of efficiency and can lead to reduced capacity because undersized systems, having considerable resistance downstream, may be unable to handle the added volume of the injected airstream without loss of performance. Ejectors are best suited for use at the chimney or vent exit or where there is an adequately sized chimney or vent to carry the combined discharge. If total pressure defines outlet conditions or is used for fan selection, the relative amounts of static pressure and velocity pressure must be factored out; otherwise, the velocity head method of calculation does not apply. To factor total pressure into its two components, either the discharge velocity in an outlet of known area or the flow rate must be known. For example, if an appliance or blower produces a total pressure of 67 Pa at 7.6 m/s discharge velocity and the velocity pressure calculates to 35 Pa at ambient standard conditions, then static pressure is 62 – 35 = 27 Pa. Because this static pressure is part of the system driving force, it combines with theoretical draft to overcome losses in the system. Draft inducer fans can be operated either continuously or on demand. In either case, a safety switch that senses flue gas flow or pressure is needed to interrupt burner controls if adequate draft fails. Demand operation links the thermostat with the draft control motor. Once flow starts, as sensed with a flow or pressure switch, the burner is allowed to start. A single draft inducer, operating continuously, can be installed in the common vent of a system serving several separately controlled appliances. This simplifies the circuitry because only one control is needed to sense loss of draft. However, the single fan increases appliance standby loss (especially in boilers) and heat losses via ambient air drawn through inoperative appliances.

14.

Fig. 25 Draft Inducers

TERMINATIONS: CAPS AND WIND EFFECTS

The vent or chimney height and method of termination is governed by a variety of considerations, including fire hazard; wind effects; entry of rain, debris, and birds; and operating considerations such as draft and capacity. For example, the 0.9 m height required for residential chimneys above a roof is necessary so that small sparks will burn out before they fall on the roof shingles. Many vent and chimney malfunctions are attributed to interactions of the chimney termination or its cap with winds acting on the roof or with adjoining buildings, trees, or mountains. Because winds fluctuate, no simple method of analysis or reduction to practice exists for this complex situation. Figures 26 to 28 show some of the complexities of wind flow contours around simple structural shapes. Figure 26 shows three zones with differing degrees of flue gas dispersion around a rectangular building: the cavity or eddy zone, the wake zone, and the undisturbed flow zone (Clarke 1967). In addition, a fourth flow zone of intense turbulence is located downwind of the cavity. Chimney flue gases discharged into the wind at a point close to the roof surface in the cavity zone may be recirculated locally. Higher in the cavity zone, wind eddies can carry more dilute flue gas to the lee side of the building. Flue gases discharged into the wind in the wake zone do not recirculate into the immediate vicinity, but may soon descend to ground level. Above the wake zone, dispersal into the undisturbed wind flow carries and dilutes the flue gases over a wider area. The boundaries of these zones vary

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with building configuration and wind direction and turbulence; they are strongly influenced by surroundings. The possibility of air pollutants reentering the cavity zone because of plume spread or of air pollution intercepting downwind cavities associated with adjacent structures or downwind buildings should be considered. Thus, the design criterion of elevating the stack discharge above the cavity is not valid for all cases. Consult a meteorologist experienced with dispersion processes near buildings for complex cases and for cases involving air contaminants. As chimney height increases through the three zones, draft performance improves dispersion, while additional problems of gas cooling, condensation, and structural wind load are created. As building height increases (Figure 27), the eddy forming the cavity zone no longer descends to ground level. For a low, wide building (Figure 28), wind blowing parallel to the long roof dimension can reattach to the surface; thus, the eddy zone becomes flush with the roof surface (Evans 1957). For satisfactory dispersion with low,

Fig. 26 Wind Eddy and Wake Zones for One- or Two-Story Buildings and Their Effect on Chimney Gas Discharge (Clarke 1967)

Fig. 27

Height of Eddy Currents Around Single High-Rise Buildings

35.31 wide buildings, chimney height must still be determined as if H = W (Figure 26). Chien et al. (1951) and Evans (1957) studied pitched roofs in relation to wind flow and surface pressures. Because the typical residence has a pitched roof and probably uses natural gas or a low-sulfur fossil fuel, dispersion is not important because combustion products are relatively free of pollutants. For example, NFPA Standard 54/AGA Z223.1 requires a minimum distance between the gas vent termination and any air intake, but it does not require penetration above the cavity zone. Flow of wind over a chimney termination can impede or assist draft. In regions of stagnation on the windward side of a wall or a steep roof, winds create positive static pressures that impede established flow or cause backdrafts in vents and chimneys. Locating a chimney termination near the surface of a low, flat roof can aid draft because the entire roof surface is under negative static pressure. Velocity is low, however, because of the cavity formed as wind sweeps up over the building. With greater chimney height, termination above the low-velocity cavity or negative-pressure zone subjects the chimney exit to greater wind velocity, thereby increasing draft from two causes: (1) height and (2) wind aspiration over an open top. As the termination is moved from the center of the building to the sides, its exposure to winds and pressure also varies. Terminations on pitched roofs may be exposed to either negative or positive static pressure, as well as to variation in wind velocity and direction. On the windward side, pitched roofs vary from complete to partial negative pressure as pitch increases from approximately flat to 30° (Chien et al. 1951). At a 45° pitch, the windward pitched roof surface is strongly positive; beyond this slope, pressures approach those observed on a vertical wall facing the wind. Wind pressure varies with its horizontal direction on a pitched roof, and on the lee (sheltered) side, wind velocity is very low, and static pressures are usually negative. Wind velocities and pressures vary not only with pitch, but with position between ridge and eaves. Reduction of these observed external wind effects to simple rules of termination for a wide variety of chimney and venting systems requires many compromises. In the wake zone or any higher location exposed to full wind velocity, an open top can create strong venting updrafts. The updraft effect relative to wind dynamic pressure is related to the Reynolds number. Open tops, however, are sensitive to the wind angle as well as to rain (Clarke 1967), and many proprietary tops have been designed to stabilize wind effects and improve the performance. Because of the many compromises made in vent termination design, this stability is usually achieved by sacrificing some of the updraft

Fig. 28 Eddy and Wake Zones for Low, Wide Buildings

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2020 ASHRAE Handbook—HVAC Systems and Equipment (SI)

created by the wind. Further, locating a vent cap in a cavity region frequently removes it from the zone where wind velocity could have a significant effect. Performance optimization studies of residential vent cap design indicate that the following performance features are important: (1) still-air resistance, (2) updraft ability with no flow, and (3) discharge resistance when vent gases are carried at low velocity in a typical wind (3 m/s vent velocity in a 9 m/s wind). Tests in UL Standard 441 for proprietary gas vent caps consider these three aspects of performance to ensure adequate vent capacity. Frequently, air supply to an appliance room is difficult to orient to eliminate wind effects. Therefore, the vent outlet must have a certain updraft capability, which can help balance a possible adverse wind. When wind flows across an inoperative vent termination, a strong updraft develops. Appliance start-up reduces this updraft, and in typical winds, the vent cap may develop greater resistance than it would have in still air. Certain vent caps can be made with very low still-air resistance, yet exhibit excessive wind resistance, which reduces capacity. Finally, because the appliance operates whether there is a wind or not, still-air resistance must be low. Some proprietary air ventilators have excessive still-air resistance and should be avoided on vent and chimney systems unless a considerably oversized vent is specified. Vertical-slot ventilators, for example, have still-air resistance coefficients of about 4.5. To achieve low still-air resistance on vents and chimneys, the verticalslot ventilator must be 50% larger than the diameter of the chimney or vent unless it has been specifically listed for such use. Freestanding chimneys high enough to project above the cavity zone require structurally adequate materials or guying and bracing for prefabricated products. The prefabricated metal building heating appliance chimney places little load on the roof structure, but guying is required at 2.4 to 3.6 m intervals to resist both overturning and oscillating wind forces. Various other expedients, such as spiral baffles on heavy-gage freestanding chimneys, have been used to reduce oscillation. The chimney height needed to carry the effluent into the undisturbed flow stream above the wake zone can be reduced by increasing the effluent discharge velocity. A 15 m/s discharge velocity avoids downward eddying along the chimney and expels the effluent free of the wake zone. Velocity this high can be achieved only with forced or induced draft. Rain entry is a problem for open, low-velocity, or inoperative systems. Good results have been obtained with drains that divert the water onto a roof or into a collection system leading to a sump. Figure 29 shows several configurations (Clarke 1967; Hama and Downing 1963). Runoff from stack drains contains acids, soot, and

metallic corrosion products, which can cause roof staining. Therefore, these methods are not recommended for residential use. An alternative procedure is to allow all water to drain to the base of the chimney, where it is piped from a capped tee to a sump. Rain caps prevent vertical discharge of high-velocity flue gases. However, caps are preferred for residential gas-burning equipment because it is easier to exclude rain than to risk rainwater leakage at horizontal joints or to drain it. Also, caps keep out debris and bird nests, which can block the chimney. Satisfactory vent cap performance can be achieved in the wind by using one of a variety of standard configurations, including the A cap and the wind band ventilator, or one of the proprietary designs shown in Figure 29. Where partial rain protection without excessive flow resistance is desired, and either wind characteristics are unimportant or windflow is horizontal, a flat disk or cone cap 1.7 to 2.0 diameters across located 0.5 diameter above the end of the pipe has a still-air resistance loss coefficient of about 0.5.

Fig. 29 Vent and Chimney Rain Protection

Table 11 List of U.S. National Standards Relating to Installationa NFPAb

ANSIc

CSAf

Subject

Materials Covered

Oil-burning equipment

Type L listed chimneys, single-wall, masonry

31

—

—

Gas appliances and gas piping

Type B, L listed chimneys, single-wall, masonry

54

Z223.1d

B149.1

Chimneys, fireplaces, vents, and solid-fuel-burning appliances

All types

211

—

CAN/ULC S605-M91

Recreational vehicles

Roof jacks and vents

1192

—

—

Gas piping and gas equipment on industrial premises

All types

54

Z223.1d

B149.1

Gas conversion burners

Chimneys

Z21.8e

—

Safe design

Z21.17e

—

Draft hoods

Part dimensions

Z21.12e

CAN1-6.2

Automatic vent dampers for use with gas-fired appliances

Construction and performance

Z21.66e

—

aThese

bNational

standards are subject to periodic review and revision to reflect advances in industry, as well as for consistency with legal requirements and other codes.

Fire Protection Association, Quincy, MA. cAmerican National Standards Institute, New York. dAvailable from American Gas Association, Washington, D.C.

eAvailable

from CSA America, Cleve-

land, OH. fCanadian Standards Association, Missis-

sauga, ON.

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Chimney, Vent, and Fireplace Systems See Chapter 46 of the 2019 ASHRAE Handbook—HVAC Applications for additional information on vent and chimney termination and wind effects.

15.

CODES AND STANDARDS

Building and installation codes and standards prescribe the installation and safety requirements of heat-producing appliances and their vents and chimneys. Chapter 52 lists the major national building codes, one of which may be in effect in a given area. Some jurisdictions either adopt a national building code with varying degrees of revisions to suit local custom or, as in many major metropolitan areas, develop a local code that agrees in principle but shares little common text with the national codes. Familiarity with applicable building codes is essential because of the great variation in local codes and adoption of modern chimney design practice. The national standards listed in Table 11 give greater detail on the mechanical aspects of fuel systems and chimney or vent construction. Although these standards emphasize safety aspects, especially clearances to combustibles for various venting materials, they also recognize the importance of proper flow, draft, and capacity.

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16. A Ac AF B Bo Cu cp df di Da Db Dp Dt F f

= = = = = = = = = = = = = = =

H I k kf kL L

= = = = = =

Lm = M Q q qm T Tm To Ts U

= = = = = = = = =

U V VF w x

= = = = =

y= p = t = te =

SYMBOLS

area of passage cross section, m2 chimney flue cross-sectional area, m2 fireplace frontal opening area, m2 existing or local barometric pressure, Pa standard pressure, Pa (101 325 Pa) temperature multiplier for heat loss, dimensionless specific heat of gas at constant pressure, kJ/(kg·K) inside diameter, m inside diameter, mm available draft, Pa boost (increase in static pressure by fan), Pa depressurization, Pa theoretical draft, Pa friction factor for L/di Darcy friction factor from Moody diagram (Figure 13, Chapter 3, 2017 ASHRAE Handbook—Fundamentals) height of vent or chimney system above grade or system inlet, m operating heat input, MJ/s (MW) system resistance loss coefficient, dimensionless fitting friction loss coefficient, dimensionless piping friction loss coefficient, dimensionless length of all piping in chimney system from inlet to exit, linear metres length of system from inlet to midpoint of vertical or to location of mean gas temperature, m mass flow to input ratio, kg/MJ volumetric flow rate, m3/s sensible heat at particular point in vent, MJ/s sensible heat at average temperature in vent, MJ/s absolute temperature, K mean flue gas temperature at average conditions in system, K ambient temperature, K standard temperature, K (288.15 K) overall heat transfer coefficient of vent or chimney wall material, referred to inside surface area, W/(m2·K) heat transfer coefficient for Equation (6), W/(m2·K) velocity of gas flow in passage, m/s frontal velocity of ambient air, m/s mass flow of gas, g/s length of one internal side of rectangular chimney cross section, m length of other internal side of rectangular chimney cross section, m system flow losses or pressure drop, Pa temperature difference, K temperature difference entering system, K

35.33 tm = temperature difference at average temperature location in system, K (tm = Tm – To) o = density of air at 288.15 K and 101 325 Pa, kg/m3 (1.226 kg/m3) m = density of chimney gas at average temperature and local barometric pressure, kg/m3

REFERENCES ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal.ashrae .org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org/bookstore. ANSI. 2012. Installation of domestic gas conversion burners. Standard Z21.8-1994 (R2012). American National Standards Institute, Washington, D.C. ANSI. 2000. Draft hoods. Standard Z21.12b-1994 (R2000). American National Standards Institute, Washington, D.C. ANSI. 2014. Domestic gas conversion burners. ANSI Standard Z21.171998/CSA 2.7-M1998 (R2014). American National Standards Institute, Washington, D.C., and Canadian Standards Association, Toronto, ON. ANSI. 2001. Automatic vent damper devices for use with gas-fired appliances. ANSI Standard Z21.66-1996 (R2001)/CGA 6.14-M-96. American National Standards Institute, Washington, D.C., and Compressed Gas Association, Chantilly, VA. Beaumont, M., D. Fitzgerald, and D. Sewell. 1970. Comparative observations on the performance of three steel chimneys. Institution of Heating and Ventilating Engineers (July):85. Butcher, T., and Y. Celebi. 1993. Fouling of oil-fired boilers and furnaces. Proceedings of the 1993 Oil Heat Technology Conference and Workshop (March): 140. CAN/ULC. 2008. Standard for gas vents. Standard S605-M91. Underwriters Laboratories Canada, Toronto, ON. Chien, N., Y. Feng, H. Wong, and T. Sino. 1951. Wind-tunnel studies of pressure on elementary building forms. Iowa Institute of Hydraulic Research. Clarke, J.H. 1967. Air flow around buildings. Heating, Piping and Air Conditioning (May):145. CSA. 2005. Natural gas and propane installation code. Standard B149.12005. Canadian Standards Association International, Mississauga, ON. CSA. 2006. Testing method for measuring annual fireplace efficiency. Standard P.4.1-02 (R2006). Canadian Standards Association International, Mississauga, ON. CSA. 2006. Draft hoods. Standard CAN1-6.2-M81 (R2006). Canadian Standards Association International, Mississauga, ON. Evans, B.H. 1957. Natural air flow around buildings. Texas Engineering Experiment Station Research Report 59. Fontana, M.G., and N.D. Greene. 1967. Corrosion engineering, p. 223. McGraw-Hill, New York. GTI. 2009. VENT-II, version 5.3 user’s guide and software. Gas Technology Institute, Des Plaines (formerly Gas Research Institute, Chicago), IL. Hama, G.M., and D.A. Downing. 1963. The characteristics of weather caps. Air Engineering (December):34. HYDI. 1989. Testing and rating standard for heating boilers. Hydronics Institute (now part of Air Conditioning, Heating and Refrigeration Institute [AHRI]), Berkeley Heights, NJ. Kam, V.P., R.A. Borgeson, and D.W. DeWerth. 1993. Masonry chimney for category I gas appliances: Inspection and relining. ASHRAE Transactions 99(1):1196-1201. Paper CH-93-13-4. Kinkead, A. 1962. Gravity flow capacity equations for designing vent and chimney systems. Proceedings of the Pacific Coast Gas Association 53. Koebel, M., and M. Elsener. 1989. Corrosion of oil-fired central heating boilers. Werkstoffe und Korrosion 40:285-94. Krajewski, R.F. 1996. Oil heat vent analysis program (OHVAP) users manual and engineering report. BNL Informal Report BNL-63668. www.osti.gov. Land, T. 1977. The theory of acid deposition and its application to the dewpoint meter. Journal of the Institute of Fuel (June):68. Lapple, C.E. 1949. Velocity head simplifies flow computation. Chemical Engineering (May):96-104. Mueller, G.R. 1968. Charts determine gas temperature drops in metal flue stacks. Heating, Piping and Air Conditioning (January):138. NFPA. 2006. Installation of oil-burning equipment. ANSI/NFPA Standard 31-2006. National Fire Protection Association, Quincy, MA.

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35.34

2020 ASHRAE Handbook—HVAC Systems and Equipment (SI)

NFPA. 2013. Chimneys, fireplaces, vents, and solid fuel-burning appliances. ANSI/NFPA Standard 211-2013. National Fire Protection Association, Quincy, MA. NFPA. 2015. Recreational vehicles. Standard 1192-2015. National Fire Protection Association, Quincy, MA. NFPA/AGA. 2015. National fuel gas code. ANSI/NFPA 54-2015/ANSI/ AGA Standard Z223.1-2015. National Fire Protection Association, Quincy, MA, and American Gas Association, Washington, D.C. Pray, H.R. et al. 1942-1953. The corrosion of metals and materials by the products of combustion of gaseous fuels. Battelle Memorial Institute Reports 1, 2, 3, and 4 to the American Gas Association. Ramsey, C.G., and H.R. Sleeper. 1956. Architectural graphic standards. John Wiley & Sons, New York. methodology. Rutz, A.L. 1992. VENT-II Solution sales.gastechnology.org/. Segeler, C.G., ed. 1965. Gas engineers’ handbook, Section 12. Industrial Press, New York. Stone, R.L. 1957. Design of multiple gas vents. Air Conditioning, Heating and Ventilating (July). Stone, R.L. 1969. Fireplace operation depends upon good chimney design. ASHRAE Journal (February):63. Stone, R.L. 1971. A practical general chimney design method. ASHRAE Transactions 77(1):91-100. Paper PH-2175. Stone, R.L. 2005. Letter to ASHRAE editor, revised Figure 19 (July 27, 2005). UL. 2001. Factory-built chimneys for residential type and building heating appliances. ANSI/UL Standard 103-2001. Underwriters Laboratories, Northbrook, IL. UL. 1999. Gas vents. Standard 441-99. Underwriters Laboratories, Northbrook, IL.

Verhoff, F.H., and J.T. Banchero. 1974. Predicting dew points of flue gases. Chemical Engineering Progress (August):71. Yeaw, J.S., and L. Schnidman. 1943. Dew point of flue gases containing sulfur. Power Plant Engineering 47(I and II).

BIBLIOGRAPHY ANSI. 2006. Gas-fired central furnaces. Standard Z21.47b-1994/CSA 2.3b2006. Canadian Standards Association, Cleveland, OH. Briner, C.F. 1984. Heat transfer and fluid flow analysis of sheet metal chimney systems. ASME Paper 84-WA/SOL-36. American Society of Mechanical Engineers, New York. Briner, C.F. 1986. Heat transfer and fluid flow analysis of three-walled metal factory-built chimneys. ASHRAE Transactions 92(1B):727-38. Paper SF-86-16-1. Hampel, T.E. 1956. Venting system priming time. Research Bulletin 74, Appendix E, 146. American Gas Association Laboratories, Cleveland, OH. Jakob, M. 1955. Heat transfer, vol. I. John Wiley & Sons, New York. Paul, D.D. 1992. Venting guidelines for Category I gas appliances. GRI-89/ 0016. Gas Technology Institute, Des Plaines (formerly Gas Research Institute, Chicago), IL. Ramsey, C.G., and H.R. Sleeper. 1970. Architectural graphic standards, 6th ed. John Wiley & Sons, New York. Reynolds, H.A. 1960. Selection of induced draft fans for heating boilers. Air Conditioning, Heating and Ventilating (December):51. Sepsy, C.F., and D.B. Pies. 1972. An experimental study of the pressure losses in converging flow fittings used in exhaust systems. Ohio State University College of Engineering, Columbus (December).

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HYDRONIC HEAT-DISTRIBUTING UNITS AND RADIATORS Description............................................................................... 36.1 Ratings of Heat-Distributing Units.......................................... 36.2

Design ...................................................................................... 36.3 Applications.............................................................................. 36.5

R

Sectional radiators are fabricated from welded sheet metal sections (generally two, three, or four tubes wide), and resemble freestanding cast-iron radiators. Panel radiators consist of fabricated flat panels (generally one, two, or three deep), with or without an exposed extended fin surface attached to the rear for increased output. These radiators are most common in Europe. Tubular steel radiators consist of supply and return headers with interconnecting parallel steel tubes in a wide variety of lengths and heights. They may be specially shaped to coincide with the building structure. Some are used to heat bathroom towel racks. Specialty radiators are fabricated of welded steel or extruded aluminum and are designed for installation in ceiling grids or floormounting. Various unconventional shapes are available.

ADIATORS, convectors, and baseboard and finned-tube units are heat-distributing devices used in hot-water and steam heating systems. They supply heat by a combination of radiation and convection and maintain the desired air temperature and/or mean radiant temperature in a space without fans. Figures 1 and 2 show sections of typical heat-distributing units. In heating systems, radiant panels are also used. Units are inherently self-adjusting in the sense that heat output is based on temperature differentials; cold spaces receive more heat and warmer spaces receive less heat.

1.

DESCRIPTION

Radiators The term radiator, though generally confined to sectional castiron column, large-tube, or small-tube units, also includes flat-panel types and fabricated steel sectional types. Small-tube radiators, with a length of only 45 mm per section, occupy less space than column and large-tube units and are particularly suited to installation in recesses (see Table 1). Column, wall-type, and large-tube radiators are no longer manufactured, although many of these units are still in use. See Tables 2, 3, and 4 in Chapter 28 of the 1988 ASHRAE Handbook—Equipment, Byrley (1978), or Hydronics Institute (1989) for principal dimensions and average ratings of these units. The following are the most common types of radiators:

Fig. 1

Pipe Coils Pipe coils have largely been replaced by finned tubes. See Table 5 in Chapter 28 of the 1988 ASHRAE Handbook—Equipment for the heat emission of such pipe coils.

Convectors A convector is a heat-distributing unit that operates with gravitycirculated air (natural convection). It has a heating element with a large amount of secondary surface and contains two or more tubes with headers at both ends. The heating element is surrounded by an

Terminal Units

The preparation of this chapter is assigned to TC 6.1, Hydronic and Steam Equipment and Systems.

Copyright © 2020, ASHRAE

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Fig. 2 Typical Radiators enclosure with an air inlet below and an air outlet above the heating element. Convectors are made in a variety of depths, sizes, and lengths and in enclosure or cabinet types. The heating elements are available in fabricated ferrous and nonferrous metals. Air enters the enclosure below the heating element, is heated in passing through the element, and leaves the enclosure through the outlet grille located above the heating element. Factory-assembled units comprising a heating element and an enclosure are common, and may be freestanding, wallhung, or recessed and may have outlet grilles or louvers and arched inlets or inlet grilles or louvers, as desired.

Baseboard Units Baseboard (or baseboard radiation) units are designed for installation along the bottom of walls in place of conventional baseboard. They may be made of cast iron, with a substantial portion of the front face directly exposed to the room, or with a finned-tube element in a sheet metal enclosure. They use gravity-circulated room air. Baseboard heat-distributing units are divided into three types: radiant, radiant convector, and finned tube. The radiant unit, which is made of aluminum, has no openings for air to pass over the wall side of the unit. Most of this unit’s heat output is by radiation. The radiant-convector baseboard is made of cast iron or steel. The units have air openings at the top and bottom to allow circulation of room air over the wall side of the unit, which has extended surface to provide increased heat output. A large portion of the heat emitted is transferred by convection. The finned-tube baseboard has a finned-tube heating element concealed by a long, low sheet metal enclosure or cover. A major portion of the heat is transferred to the room by convection. The output varies over a wide range, depending on the physical dimensions and the materials used. Avoid using a unit with a high relative output per unit length compared to overall heat loss, which would result in a concentration of the heating element over a relatively small area. Optimum comfort for room occupants is obtained when units are installed along as much of the exposed wall as possible.

Finned-Tube Units Finned-tube (or fin-tube) units are fabricated from metallic tubing, with metallic fins bonded to the tube. They operate with gravity-circulated room air. Finned-tube elements are available in several tube sizes, in either steel or copper (25 to 50 mm nominal steel or 22 to 35 mm nominal copper) with various fin sizes, spacings, and materials. Resistance to steam or water flow is the same as that through standard distribution piping of equal size and type.

Finned-tube elements installed in occupied spaces generally have covers or enclosures in a variety of designs. When human contact is unlikely, they are sometimes installed bare or provided with an expanded metal grille for minimum protection. A cover has a portion of the front skirt made of solid material. The cover can be mounted with clearance between the wall and the cover, and without completely enclosing the rear of the finned-tube element. A cover may have a top, front, or inclined outlet. An enclosure is a shield of solid material that completely encloses both the front and rear of the finned-tube element. The enclosure may have an integral back or may be installed tightly against the wall so that the wall forms the back, and it may have a top, front, or inclined outlet.

Heat Emission These heat-distributing units emit heat by a combination of radiation to the surfaces and occupants in the space and convection to the air in the space. Chapter 4 of the 2017 ASHRAE Handbook—Fundamentals covers the heat transfer processes and the factors that influence them. Those units with a large portion of their heated surface exposed to the space (i.e., radiator and cast-iron baseboard) emit more heat by radiation than do units with completely or partially concealed heating surfaces (i.e., convector, finned-tube, and finned-tube baseboard). Also, finned-tube elements constructed of steel emit a larger portion of heat by radiation than do finned-tube elements constructed of nonferrous materials. The heat output ratings of heat-distributing units are expressed in {watts or in square metres equivalent direct radiation (EDR).

2.

RATINGS OF HEAT-DISTRIBUTING UNITS

For convectors, baseboard units, and finned-tube units, an allowance for heating effect may be added to the test capacity (the heat extracted from the steam or water under standard test conditions). This heating effect reflects the ability of the unit to direct its heat output to the occupied zone of a room. The application of a heating effect factor implies that some units use less steam or hot water than others to produce an equal comfort effect in a room.

Radiators Current methods for rating radiators were established by the U.S. National Bureau of Standards publication, Simplified Practices Recommendation R174-65, Cast-Iron Radiators, which has been withdrawn (Table 1).

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Hydronic Heat-Distributing Units and Radiators Table 1

36.3

Small-Tube Cast-Iron Radiators Section Dimensions

Number of Tubes per Section 3 4

5

6

Catalog Rating per Section,a W

A Height, mmb

B Width, {mm Min. Max.

D Leg Height, mmb

113

635

83

89

45

65

113

483

113

122

45

65

127

559

113

122

45

65

141

635

113

122

45

65

148

559

143

160

45

65

169

635

143

160

45

65

162

483

173

203

45

65

211

635

173

203

45

65

260

813

173

203

45

65

a Ratings based on steam at {101.7°C and air at {21.1°C. They apply only to installed radi-

b Overall height and leg height, as produced by some manufacturers, are {25 mm greater

ators exposed in a normal manner, not to radiators installed behind enclosures, behind grilles, or under shelves. For ratings at other temperatures, multiply table values by factors found in Table 2.

than shown in columns A and D. Radiators may be furnished without legs. Where greater than standard leg heights are required, leg height should be {115 mm. c Length equals number of sections multiplied by {45 mm.

Convectors Licensed for single user. © 2020 ASHRAE, Inc.

C Spacing, mmc

The generally accepted method of testing and rating ferrous and nonferrous convectors in the United States was given in Commercial Standard CS 140-47, Testing and Rating Convectors (Dept. of Commerce 1947), but it has been withdrawn. This standard contained details covering construction and instrumentation of the test booth or room and procedures for determining steam and water ratings. Under the provisions of Commercial Standard CS 140-47, the rating of a top outlet convector was established at a value not in excess of the test capacity. For convectors with other types of enclosures or cabinets, a percentage that varies up to a maximum of 15% (depending on the height and type of enclosure or cabinet) was added for heating effect (Brabbee 1926; Willard et al. 1929). The addition made for heating effect must be shown in the manufacturer’s literature. The testing and rating procedure set forth by Commercial Standard CS 140-47 does not apply to finned-tube or baseboard radiation.

Baseboard Units The generally accepted method of testing and rating baseboards in the United States is covered in the Testing and Rating Standard for Baseboard Radiation (AHRI 2005). This standard contains details covering construction and instrumentation of the test booth or room, procedures for determining steam and hotwater ratings, and licensing provisions for obtaining approval of these ratings. Baseboard ratings include an allowance for heating effect of 15% in addition to the test capacity. The addition made for heating effect must be shown in the manufacturer’s literature.

Finned-Tube Units The generally accepted method of testing and rating finned-tube units in the United States is covered in AHRI Standard 1410. This standard contains details covering construction and instrumentation of the test booth or room, procedures for determining steam and water ratings, and licensing provisions for obtaining approval of these ratings. The rating of a finned-tube unit in an enclosure that has a top outlet is no greater than the test capacity. For finned-tube units with other types of enclosures or covers, a percentage is added for heating effect that varies up to a maximum of 15%, depending on the very short runs designed for conventional temperature drops (i.e.,

height and type of enclosure or cover. The addition made for heating effect must be shown in the manufacturer’s literature (Pierce 1963).

Other Heat-Distributing Units Unique radiators and radiators from other countries generally are tested and rated for heat emission in accordance with prevailing standards. These other testing and rating methods have basically the same procedures as the Hydronics Institute and AHRI standards, which are used in the United States. See Chapter 6 for information on the design and sizing of radiant panels.

Corrections for Nonstandard Conditions The heating capacity of a radiator, convector, baseboard, finnedtube heat-distributing unit, or radiant panel is a power function of the temperature difference between the air in the room and the heating medium in the unit, shown as q = c(ts – ta ) n

(1)

where q = heating capacity, {W c = constant determined by test ts = average temperature of heating medium, {°C. For hot water, use arithmetic average of entering and leaving water temperatures. ta = room air temperature, {°C. Air temperature {1.5 m above floor is generally used for radiators, whereas entering air temperature is used for convectors, baseboard units, and finned-tube units. n = exponent that equals {1.2 for cast-iron radiators, {1.31 for baseboard radiation, {1.42 for convectors, 1.0 for ceiling heating and floor cooling panels, and 1.1 for floor heating and ceiling cooling panels. For finned-tube units, n varies with air and heating medium temperatures. Correction factors to convert heating capacities at standard rating conditions to heating capacities at other conditions are given in Table 2.

Equation (1) may also be used to calculate heating capacity at nonstandard conditions.

3.

DESIGN

Effect of Water Velocity Designing for high temperature drops through the system (as much as {35 to 45 K in low-temperature water (LTW) systems and as much as {110 K in high-temperature systems) can result in low water velocities in the finned-tube or baseboard element. Applying {10 K) can also result in low velocities.

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2020 ASHRAE Handbook—HVAC Systems and Equipment (SI) Table 2 Correction Factors c for Various Types of Heating Units

Steam Pressure (Approx.), kPa (absolute) 9.5 15.8 25.0 38.6 57.9 84.6 120.9 169.2 232.3 313.4 415.8

Steam or Water Temp., °C 45 55 65 75 85 95 105 115 125 135 145

Radiator

Convector

Finned-Tube

Baseboard

Room Temp., °C 25 20 15

Air Temp., °C 25 20 15

Air Temp., °C 25 20 15

Air Temp., °C 25 20 15

— — 0.33 0.47 0.61 0.77 0.93 1.11 1.30 1.50 1.70

0.15 0.26 0.37 0.50 0.64 0.78 0.94 1.09 1.26 1.42 1.60

— — 0.40 0.54 0.68 0.83 0.99 1.15 1.32 1.50 1.68

— — 0.47 0.61 0.76 0.91 1.07 1.24 1.41 1.59 1.77

— 0.40 0.54 0.68 0.83 0.99 1.15 1.32 1.50 1.68 1.86

— — 0.40 0.54 0.69 0.85 1.02 1.20 1.40 1.60 1.81

— 0.33 0.47 0.61 0.77 0.93 1.11 1.30 1.50 1.70 1.92

0.21 0.32 0.44 0.57 0.71 0.86 1.01 1.18 1.34 1.51 1.69

0.26 0.37 0.50 0.64 0.78 0.94 1.09 1.26 1.42 1.60 1.78

0.14 0.24 0.36 0.49 0.63 0.78 0.94 1.11 1.29 1.47 1.66

0.19 0.30 0.43 0.56 0.70 0.86 1.02 1.20 1.38 1.57 1.76

0.24 0.36 0.49 0.63 0.78 0.94 1.11 1.29 1.47 1.66 1.86

vectors and finned-tube or baseboard units represents the same room comfort conditions as 21°C room air temperature for a radiator. Standard conditions for radiant panels are 50°C heating medium temperature and 20°C for room air temperature; c depends on panel construction. To determine output of a heating unit under nonstandard conditions, multiply standard heating capacity by appropriate factor for actual operating heating medium and room or inlet air temperatures.

Fig. 3 Water Velocity Correction Factor for Baseboard and Finned-Tube Radiators

Fig. 4 Effect of Air Density on Radiator Output

Figure 3 shows the effect of water velocity on the heat output of typical sizes of finned-tube elements. The figure is based on work done by Harris (1957) and Pierce (1963) and tests at the Hydronics Institute. The velocity correction factor Fv is

The designer should check water velocity throughout the system and select finned-tube or baseboard elements on the basis of velocity as well as average temperature. Manufacturers of finnedtube and baseboard elements offer a variety of tube sizes, ranging from {15 mm copper tubes for small baseboard elements to {50 mm for large finned-tube units, to aid in maintenance of turbulent flow conditions over a wide range of flow.

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Note: Use these correction factors to determine output ratings for radiators, convectors, and finned-tube and baseboard units at operating conditions other than standard. Standard conditions in the United States for a radiator are 102°C heating medium temperature and 21°C room temperature (at center of space and at 1.5 m level). Standard conditions for convectors and finned-tube and baseboard units are 102°C heating medium temperature and 18°C inlet air temperature at 101.3 kPa atmospheric pressure. Water flow is 0.9 m/s for finned-tube units. Inlet air at 18°C for con-

Fv = (V/0.9)0.0

(2)

where V = water velocity, {m/s. Heat output varies little over the range from {0.15 to 0.9 m/s, where Fv ranges from 0.93 to 1.00. The factor drops rapidly below {0.15 m/s because flow changes from turbulent to laminar at around {0.03 m/s. Avoid such a low velocity because the output is difficult to predict accurately when designing a system. In addition, the curve is so steep in this region that small changes in actual flow have a significant effect on output. Not only does the heat transfer rate change, but the temperature drop and, therefore, the average water temperature change (assuming a constant inlet temperature).

Effect of Altitude The effect of altitude on heat output varies depending on the material used and the portion of the unit’s output that is radiant rather than convective. The reduced air density affects the convective portion. Figure 4 shows the reduction in heat output with air density (Sward and Decker 1965). The approximate correction factor FA for determining the reduced output of typical units is

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Hydronic Heat-Distributing Units and Radiators FA = ( ppo) n

36.5 (3)

where p = local station atmospheric pressure po = standard atmospheric pressure n = 0.9 for copper baseboard or finned tube = 0.5 for steel finned-tube or cast-iron baseboard = 0.2 for radiant baseboard and radiant panels

The value of p/po at various altitudes may be calculated as follows: –4

p/po = e

– 1.2210 h

(4)

where h = altitude, {m. The following are typical values of p/po: Altitude h, {m

p/po

600

0.93

1200

0.86

1500

0.83

1800

0.80

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Sward and Harris (1970) showed that some components of heat loss are affected in the same manner.

Effect of Mass Mass of the terminal unit (typically cast-iron versus copperaluminum finned element) affects the heat-up and cooldown rates of equipment. It is important that high- and low-mass radiation not be mixed in the same zone. High-mass systems have historically been favored for best comfort and economy, but tests by Harris (1970) showed no measurable difference. The thermostat or control can compensate by changing the burner’s cycle rate. The effect of mass is further reduced by constantly circulating modulated temperature water to the unit. The only time that mass can have a significant effect is in response to a giant shift in load. In that situation, the lowmass unit will respond faster.

Performance at Low Water Temperatures Table 2 summarizes the performance of baseboard and finnedtube units with an average water temperature down to {38°C. Solarheated water, industrial waste heat in a low- to medium-temperature district heating system, condensing boilers, heat pump system cooling water, and ground-source heat pumps are typical applications in this range. To compensate for heating capacity loss, either heatdistributing equipment should be oversized, or additional heatdistributing units should be installed. Minimize capital and operating costs (Kilkis 1998).

Effect of Enclosure and Paint An enclosure placed around a direct radiator restricts airflow and diminishes the proportion of output resulting from radiation. However, enclosures of proper design may improve heat distribution in the room compared to that obtained with an unenclosed radiator (Allcut 1933; Willard et al. 1929). For a radiator or cast-iron baseboard, the finish coat of paint affects the heat output. Oil paints of any color give about the same results as unpainted black or rusty surfaces, but an aluminum or a bronze paint reduces the heat emitted by radiation. The net effect may reduce the total heat output of the radiator by 10% or more (Allen 1920; Rubert 1937; Severns 1927).

4.

APPLICATIONS

Radiators Radiators can be used with steam or hot water. They are installed in areas of greatest heat loss: under windows, along cold walls, or at

doorways. They can be freestanding, semirecessed, or with decorative enclosures or shields (although these affect output) (Willard et al. 1929). Unique and imported radiators are generally not suitable for steam applications, although they have been used extensively in low-temperature water systems with valves and connecting piping left exposed. Various combinations of supply and return locations are possible, and may alter heat output. Although long lengths may be ordered for linear applications, lengths may not be reduced or increased by field modification. The small cross-sectional areas often inherent in unique radiators require careful evaluation of flow requirements, water temperature drop, and pressure drops.

Convectors Convectors can be used with steam or hot water. Like radiators, they should be installed in areas of greatest heat loss. They are particularly applicable where wall space is limited, such as in entryways and kitchens.

Baseboard Radiation Baseboard units are used almost exclusively with hot water. When used with one-pipe steam systems, tube sizes of {32 mm nominal diameter must be used to allow drainage of condensate counterflow to the steam flow. The basic advantage of the baseboard unit is that its normal placement is along the cold walls and under areas where the greatest heat loss occurs. Other advantages are that it (1) is inconspicuous, (2) offers minimal interference with furniture placement, and (3) distributes the heat near the floor. This last characteristic reduces the floor-to-ceiling temperature gradient to about {1 to 2 K and tends to produce uniform temperature throughout the room. It also makes baseboard heat-distributing units adaptable to homes without basements, where cold floors are common (Kratz and Harris 1945). Heat loss calculations for baseboard heating are the same as those used for other types of heat-distributing units. Hydronics Institute (1989) describes a procedure for designing baseboard heating systems.

Finned-Tube Radiation The finned-tube unit can be used with either steam or hot water. It is advantageous for heat distribution along the entire outside wall, thereby preventing downdrafts along the walls in buildings such as schools, churches, hospitals, offices, airports, and factories. It may be the principal source of heat in a building or a supplementary heater to combat downdrafts along the exposed walls in conjunction with a central conditioned air system. Its placement under or next to windows or glass panels helps to prevent fogging or condensation on the glass. Normal placement of a finned tube is along the walls where the heat loss is greatest. If necessary, the units can be installed in two or three tiers along the wall. Hot-water installations requiring two or three tiers may run a serpentine water flow if the energy loss is not excessive. A header connection with parallel flow may be used, but the design must not (1) allow water to short circuit along the path of least resistance, (2) reduce capacity because of low water velocity in each tier, or (3) cause one or more tiers to become air-bound. Many enclosures have been developed to meet building design requirements. The wide variety of finned-tube elements (tube size and material, fin size, spacing, fin material, and multiple-tier installation), along with the various heights and designs of enclosures, give great flexibility of selection for finned-tube units that meet the needs of load, space, and appearance. In areas where zone control rather than individual room control can be applied, all finned-tube units in the zone should be in series. In such a series loop installation, however, temperature drop must be considered in selecting the element for each separate room in the loop.

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Radiant Panels Hydronic radiant heating panels are controlled-temperature surfaces on the floor, walls, or ceiling; a heated fluid circulates through a circuit embedded in or attached to the panel. More than 50% of the total heating capacity is transmitted by radiant heat transfer. Usually, {50°C mean fluid temperature delivers enough heat to indoor surfaces. With such low temperature ratings, hydronic radiant panels are suitable for low-temperature heating. See Chapter 6 for more information.

REFERENCES

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ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal.ashrae .org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org/bookstore. AHRI. 2005. Testing and rating standard for baseboard radiation, 8th ed. Air-Conditioning, Heating, and Refrigeration Institute, Arlington, VA. AHRI. 2017. Performance rating of commercial finned tube radiation. Standard 1410-2017. Air-Conditioning, Heating, and Refrigeration Institute, Arlington, VA. Allcut, E.A. 1933. Heat output of concealed radiators. School of Engineering Research Bulletin 140. University of Toronto, Canada. Allen, J.R. 1920. Heat losses from direct radiation. ASHVE Transactions 26:11. Brabbee, C.W. 1926. The heating effect of radiators. ASHVE Transactions 32:11. Byrley, R.R. 1978. Hydronic rating handbook. Color Art Inc., St. Louis. Department of Commerce. 1947. Commercial standard for testing and rating convectors. Standard CS 140-47. Withdrawn. Washington, D.C. Harris, W.H. 1957. Factor affecting baseboard rating test results. Engineering Experiment Station Bulletin 444. University of Illinois, UrbanaChampaign. Harris, W.H. 1970. Operating characteristics of ferrous and non-ferrous baseboard. IBR 8. University of Illinois, Urbana-Champaign.

Hydronics Institute. 1989. Installation guide for residential hydronic heating systems. IBR 200, 1st ed. Hydronics Institute. Available from Air Conditioning Contractors of America, Arlington, VA. Kilkis, B.I. 1998. Equipment oversizing issues with hydronic heating systems. ASHRAE Journal 40(1):25-31. Kratz, A.P., and W.S. Harris. 1945. A study of radiant baseboard heating in the IBR research home. Engineering Experiment Station Bulletin 358. University of Illinois, Urbana-Champaign. National Bureau of Standards (currently NIST). 1965. Cast-iron radiators. Simplified Practices Recommendation R174-65. Withdrawn. Pierce, J.S. 1963. Application of fin tube radiation to modern hot water systems. ASHRAE Journal 5(2):72. Rubert, E.A. 1937. Heat emission from radiators. Engineering Experiment Station Bulletin 24. Cornell University, Ithaca, NY. Severns, W.H. 1927. Comparative tests of radiator finishes. ASHVE Transactions 33:41. Sward, G.R., and A.S. Decker. 1965. Symposium on high altitude effects on performance of equipment. ASHRAE. Sward, G.R., and W.S. Harris. 1970. Effect of air density on the heat transmission coefficients of air films and building materials. ASHRAE Transactions 76:227-239. Paper KA-2158. Willard, A.C., A.P. Kratz, M.K. Fahnestock, and S. Konzo. 1929. Investigation of heating rooms with direct steam radiators equipped with enclosures and shields. ASHVE Transactions 35:77 or Engineering Experiment Station Bulletin 192. University of Illinois, Urbana-Champaign.

BIBLIOGRAPHY Kratz, A.P. 1931. Humidification for residences. Engineering Experiment Station Bulletin 230:20, University of Illinois, Urbana-Champaign. Laschober, R.R., and G.R. Sward. 1967. Correlation of the heat output of unenclosed single- and multiple-tier finned-tube units. ASHRAE Transactions 73(I):V.3.1-15. Willard, A.C, A.P. Kratz, M.K. Fahnestock, and S. Konzo. 1931. Investigation of various factors affecting the heating of rooms with direct steam radiators. Engineering Experiment Station Bulletin 223. University of Illinois, Urbana-Champaign.

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Related Commercial Resources CHAPTER 37

SOLAR ENERGY EQUIPMENT AND SYSTEMS SOLAR HEATING SYSTEMS .................................................. 37.2 Air-Heating Systems ................................................................ 37.2 Liquid-Heating Systems ........................................................... 37.2 Solar Thermal Energy Collectors ............................................ 37.3 Row Design .............................................................................. 37.6 Array Design............................................................................ 37.7

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S

OLAR energy use is becoming more economical, primarily due to technological maturity, economies of scale, and governmental policies with immediate interest in low-carbon and sustainable construction (Kavlak et al. 2018). In addition, many countries consider solar and renewable energy a security measure, to ensure the availability of power under adverse conditions. While the United States continues to grow its solar industry, China, Europe, Asia, and the Mediterranean basin are leading development of advanced manufacturing techniques and applications. However, equipment and systems are still very similar in all markets; therefore, this chapter primarily discusses the basic equipment used, with particular attention to collectors. More detailed descriptions of systems and designs can be found in Chapter 36 of the 2019 ASHRAE Handbook— HVAC Applications. Commercial and industrial solar thermal energy systems are generally classified according to the heat transfer medium used in the collector loop (i.e., air or liquid). Although both systems share basic fundamentals of conversion for solar radiant energy, the equipment used in each is entirely different. Air systems are primarily limited to forced-air space heating and industrial and agricultural drying processes. Liquid systems are suitable for a broader range of applications, such as hydronic space heating, service water heating, industrial process water heating, energizing heat-driven air conditioning, and pool heating, and as a heat source for series-coupled heat pumps. Because of this wide range in capability, liquid systems are more common than air systems in commercial and industrial applications. As shown in Table 1, global installed thermal capacity of solar collectors at the end of 2017 reached 473.8 gigawatts (GWth), which corresponds to an annual solar thermal energy yield of at least 390 TWh (Weiss and Spork-Dur 2018). Unglazed water collectors are used mainly for low-temperature (e.g., swimming pool) heating; their main markets are in North America (i.e., the United States and Canada) with an installed capacity of 16.1 GWth, followed by Australia and Brazil (3.7 GWth each). Glazed flat-plate and evacuated-tube collectors, which are mainly used to generate domestic hot water and space heating, dominate the market in China (334.5 GWth), Turkey (16.3 GWth), and Germany (13.3 GWth). By the end of 2017, the total thermal capacity in operation in the European Union exceeded 36.5 GWth, corresponding to 52.1 million m2 of collector area. Germany is the leader in terms of market volume, with about 38% of the European market, whereas Austria, Greece, Italy, Spain, France, and Poland together account for about 45%. Solar hot-water systems are now mandatory in new buildings according to solar ordinances in Cyprus, Greece, Italy, Portugal, Slovenia, Spain, and elsewhere in Europe. China is the leader in the global solar heating industry, producing an estimated 334.5 GWth or 477.8 million m2 of collectors in 2017. Most installations in China use evacuated-tube collectors for domesThe preparation of this chapter is assigned to TC 6.7, Solar Energy Utilization.

Copyright © 2020, ASHRAE

Thermal Collector Performance .............................................. 37.9 Thermal Energy Storage ........................................................ 37.11 Heat Exchangers .................................................................... 37.16 Controls .................................................................................. 37.17 PHOTOVOLTAIC SYSTEMS ................................................. 37.19

Table 1 Worldwide Solar Capacity in Operation by Type Capacity, Equiv. Glazed Collector Area, 1 000 000 m2 GWth Glazed flat-plate collectors Evacuated-tube collectors Unglazed collectors Glazed and unglazed air collectors Total

107.2 336.6 28.6 1.4

153.1 480.9 41.0 1.8

473.8

676.8

tic hot water only. The trend in Europe is towards larger solar-combi systems that provide both domestic hot water and space heating, accounting for half of the annual market in Austria and Germany (Balaras et al. 2010), as well as multipurpose heat-pump-assisted solar systems (Todorovic et al. 2010). The U.S. market for solar hotwater collectors (excluding unglazed swimming pool heating) is still relatively small but is gaining ground (especially in California): some states have set solar thermal carve-outs in their renewable portfolio standards (RPS), or allow electric utilities to meet RPS requirements with solar water heating systems (REN21 2018). Other markets that are growing quickly include Denmark, India, Mexico, Australia, and South Africa, although some of these markets are challenged by a lack of standards, leading to use of inferior products and poor installations, which have undermined the systems’ reputation. Photovoltaic (PV) systems convert light from the sun directly into electricity for a wide variety of applications. The cost of PV systems has dropped significantly worldwide. At the end of 2017, the average price in the United States for installed systems varied from $2.88/W for grid-connected commercial systems to $0.98/W for gridconnected, ground-mounted ones (IEA 2018). This cost reduction has made PV systems considerably more economical, particularly when combined with federal tax credits and local incentives, and the amount of installed capacity has thus increased dramatically. In 2018, the global cumulative installed photovoltaic capacity crossed the 500 GWp mark and produced at least 660 TWh annually (IEA 2019a). China is the leader in terms of cumulative installed capacity (176.1GWp), followed by the European Union (115 GWp), United States (62.2 GWp), and Japan (56 GWp). Germany is also a leader, with about 39% of the cumulative European installed capacity, whereas Italy, the United Kingdom, France, Spain, Belgium, the Netherlands, and Austria together account for about 50%. Other markets that are also growing quickly include India, Australia, Mexico, Korea, and Turkey. PV systems with thermal heat recovery (PVT) are a hybrid solar technology that integrates photovoltaic and thermal solar energy conversion within a single system. In PVT systems, the fraction of incident solar radiation that is absorbed is converted partially into electricity and partially into thermal energy that is recovered by a heat recovery fluid (air, water, or a refrigerant), reaching combined (electrical and thermal) efficiencies of up to 70%. By the end of 2018, a total of 524.2 MWth thermal capacity of PVT modules and

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2020 ASHRAE Handbook—HVAC Systems and Equipment (SI)

178.2 MWp of PV power was installed globally, corresponding to 1 million m² of PVT modules. Early adopters of this hybrid solar technology are France (215.6 MWth and 70.1 MWp), Korea (137.6 MWth and 47.8 MWp), China (65.6 MWth and 22.8 MWp), and Germany (53.6 MWth and 18.6 MWp). More information on PVT systems and designs can be found in Chapter 36 of the 2019 ASHRAE Handbook—HVAC Applications.

1. SOLAR HEATING SYSTEMS

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Solar energy system design requires careful attention to detail because solar radiation is a low-intensity form of energy, and the equipment to collect and use it can be expensive. A brief overview of air and liquid systems is presented here to show how the equipment fits into each type of system. Chapter 36 of the 2019 ASHRAE Handbook—HVAC Applications covers solar energy use, and books on design, installation, operation, and maintenance are also available (ASHRAE 1988, 1990, 1991). Solar energy and HVAC systems often use the same components and equipment. This chapter covers only the following elements, which are either exclusive to or have specific uses in solar energy applications: • • • •

Collectors and collector arrays Thermal energy storage Heat exchangers Controls

Thermal energy storage is covered in Chapter 50, heat exchangers in Chapter 47, pumps in Chapter 44, and fans in Chapter 21.

1.1

AIR-HEATING SYSTEMS

Air-heating systems circulate air through ducts to and from an air heating collector (Figure 1). Air systems are effective for space heating because a heat exchanger is not required, and the collector inlet temperature is low throughout the day (approximately room temperature). Air systems do not need protection from freezing, overheat, or corrosion. Furthermore, air costs nothing and does not cause disposal problems or structural damage. However, air ducts and airhandling equipment require more space than pipes and pumps, ductwork is hard to seal, and leaks are difficult to detect. Fans consume more power than the pumps of a liquid system, but if the unit is installed in a facility that uses air distribution, only a slight power cost is chargeable against the solar space-heating system. Thermal storage for hot-air systems has been problematic as well because of the difficulty in controlling humidity and mold growth in pebble beds and other such devices, particularly in humid climates.

Most air space-heating systems also preheat domestic hot water through an air-to-liquid heat exchanger. In this case, tightly fitting dampers are required to prevent reverse thermosiphoning at night, which could freeze water in the heat exchanger coil. If this system heats only water in the summer, the parasitic power consumption must be charged against the solar energy system because no space heating is involved and there are no comparable energy costs associated with conventional water heating. In some situations, solar water-heating systems could be more expensive than conventional water heaters, particularly if electrical energy costs are high. To reduce parasitic power consumption, some systems use the low speed of a two-speed fan.

1.2

Liquid-heating systems circulate a liquid, often a water-based fluid, through a solar collector (Figures 2 and 3). The liquid in solar collectors must be protected against freezing, which could damage the system. Freezing is the principal cause of liquid system failure. For this reason, freeze tolerance is an important factor in selecting the heat transfer fluid and equipment in the collector loop. A solar collector radiates heat to the cold sky and freezes at air temperatures well above 0 °C. Where freezing conditions are rare, small solar heating systems are often equipped with low-cost protection devices that depend on simple manual, electrical, and/or mechanical components (e.g., electronic controllers and automatic valves) for freeze protection. Because of the large investment associated with most commercial and industrial installations, solar designers and installers must consider designs providing reliable freeze protection, even in the warmest climates. Thermosiphon solar domestic hot-water heating systems operate at low, self-regulated collector flow rates. Their performance is generally better in temperate climates than active (using a pump to circulate the fluid) systems operating at conventional flow rates and is comparable to pump systems operating at low collector flow rates. Thermosiphon systems dominate the market in residential

Fig. 2

Simplified Schematic of Indirect Nonfreezing System

Fig. 3 Fig. 1 Air-Heating Space and Domestic Water Heater System

LIQUID-HEATING SYSTEMS

Simplified Schematic of Indirect Drainback Freeze Protection System

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Solar Energy Equipment and Systems and low-rise building applications in southern Europe and China, and Australia. The collector fluid is circulated by natural convection, eliminating the need for the pump and controller of an active system. The hot-water storage tank is placed above the collector area. The flow rate varies depending on the absorbed solar radiation, fluid temperatures, system geometry, etc.

Direct and Indirect Systems In a direct liquid system, city water circulates through the collector. In an indirect system, the collector loop is separated from the high-pressure city water supply by a heat exchanger. In areas of poor water quality, isolation protects the collectors from fouling by minerals in the water. Indirect systems also offer greater freeze protection, so they are used almost exclusively in commercial and industrial applications.

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Freeze Protection Direct systems, used where freezing is infrequent and not severe, can avoid freeze damage by (1) recirculating warm storage water through the collectors, (2) continually flushing the collectors with cold water, or (3) isolating collectors from the water and draining them. Systems that can be drained to avoid freeze damage are called draindown systems. Although all of these methods can be effective, none of them are generally approved by sanctioning bodies or recommended by manufacturers. Use caution when designing direct systems in freezing climates with any of these freeze protection schemes. Indirect systems use two methods of freeze protection: (1) nonfreezing fluids and (2) drainback. Nonfreezing Fluid Freeze Protection. The most popular solar energy system for commercial use is the indirect system with a nonfreezing heat transfer fluid to transmit heat from the solar collectors to storage (Figure 2). The most common heat transfer fluid is water/ propylene glycol, although other heat transfer fluids such as silicone oils, hydrocarbon oils, or refrigerants can be used. Because the collector loop is closed and sealed, the only contribution to pump pressure is friction loss; therefore, the location of solar collectors relative to the heat exchanger and storage tank is not critical. Traditional hydronic sizing methods can be used for selecting pumps, expansion tanks, heat exchangers, and air removal devices, as long as the heat transfer liquid’s thermal properties are considered. When the control system senses an increase in solar panel temperature, the pump circulates the heat transfer liquid, and energy is collected. The same control also activates a pump on the domestic water side that circulates water through the heat exchanger, where it is heated by the heat transfer fluid. This mode continues until the temperature differential between the collector and the tank is too slight for meaningful energy to be collected. At this point, the control system shuts the pumps off. At low temperatures, the nonfreezing fluid protects the solar collectors and related piping from bursting. Because the heat transfer fluid can affect system performance, reliability, and maintenance requirements, fluid selection should be carefully considered. Because the collector loop of the nonfreezing system remains filled with fluid, it allows flexibility in routing pipes and locating components. However, a double-separation (double-wall) heat exchanger is generally required (by local building codes) to prevent contamination of domestic water in the event of a leak. The doublewall heat exchanger also protects the collectors from freeze damage if water leaks into the collector loop. However, the double-wall heat exchanger reduces efficiency by forcing the collector to operate at a higher temperature. The heat exchanger can be placed inside the tank, or an external heat exchanger can be used, as shown in Figure 2. The collector loop is closed and, therefore, requires an expansion tank and pressure-relief valve. Air purge is also necessary to expel air during filling and to remove air that has been absorbed into the heat transfer fluid.

37.3 Overtemperature protection is necessary to ensure that the system operates within safe limits and to prevent collector fluid from corroding the absorber or heat exchanger. For maximum reliability, glycol should be replaced every few years. In some cases, systems have failed because the collector fluid in the loop thermosiphoned and froze the water in the heat exchanger. If the water side is exposed to the city water system, this disastrous situation must be avoided by design. Drainback Freeze Protection. A drainback solar water-heating system (see Figure 3) uses ordinary water as the heat transport medium between the collectors and thermal energy storage. Reverse-draining (or back-siphoning) the water into a drainback tank located in a nonfreezing environment protects the system from freezing whenever the controls turn off the circulator pump or a power outage occurs. For drainback systems with a large amount of working fluid in the collector loop, heat loss can be significantly decreased and overall efficiency increased by including a tank for storing the heat transfer fluid at night. Using a night storage tank in large systems is an appropriate strategy even for regions with favorable meteorological conditions. The drainback tank can be a sump with a volume slightly greater than the collector loop, or it can be the thermal energy storage tank. The collector loop may or may not be vented to the atmosphere. Many designers prefer the nonvented drainback loop because makeup water is not required and the corrosive effects of air that would otherwise be ingested into the collector loop are eliminated. The drainback system is virtually fail-safe because it automatically reverts to a safe condition whenever the circulator pump stops. Furthermore, a 20 to 30% glycol solution can be added to drainback loops for added freeze protection in case of controller or sensor failure. Because the glycol is not exposed to stagnation temperatures, it does not decompose. A drainback system requires space for the necessary pitching of collectors and pipes for proper drainage. Also, a nearby heated area must have a room for the pumps and the drainback tank. Plumbing exposed to freezing conditions drains to the drainback tank, making the drainback design unsuitable for sites where the collector cannot be elevated above the storage tank. Both dynamic and static pressure losses must be considered in drainback system design. Dynamic pressure loss is due to friction in the pipes, and static pressure loss is associated with the distance the water must be lifted above the level of the drainback tank to the top of the collector. There are two distinct designs of drainback systems: the oversized downcomer (or open-drop) and the siphon return. The static pressure requirement remains constant in the open-drop system and decreases in the siphon return. Drainback performs better than other systems in areas with low temperatures or high irradiance. Drainback has the advantage that time and energy are not lost in reheating a fluid mass left in the collector and associated piping (as in the case of antifreeze systems). Also, water has a higher heat transfer capacity and is less viscous than other heat transfer fluids, resulting in smaller parasitic energy use and higher overall system efficiency. In closed-return (indirect) designs, there is also less parasitic energy consumption for pumping because water is the heat transfer fluid. Drainback systems can be worked on safely under stagnation conditions, but should not be restarted during peak solar conditions to avoid unnecessary thermal stress on the collector.

1.3

SOLAR THERMAL ENERGY COLLECTORS

Collector Types Solar collectors depend on air heating, liquid heating, or liquidvapor phase change to transfer heat. The most common type for

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2020 ASHRAE Handbook—HVAC Systems and Equipment (SI)

commercial, residential, and low-temperature (100 MW) and can facilitate snow removal. A dual-axis system tracks the sun from east to west and adjusts the PV array’s north/south tilt angle to account for the sun’s solar elevation change. An azimuth tracker has a fixed slope and rotates around a vertical axis. Proper use of these systems can increase annual solar energy harvest by approximately 30% or more if the site has adequate solar resources. Due to the higher capital and

Fig. 35 PV System Configurations Using (A) String Inverter, (B) Microinverters, and (C) String Inverter with Power Optimizers Source: Natural Resources Canada 2019

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maintenance cost associated with tracking systems and the declining cost per watt of solar modules, tracking systems are typically used in off-grid and/or mission-critical solar power systems.

REFERENCES

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ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal.ashrae .org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org/bookstore. Argiriou, A.A. 1997. CSHPSS systems in Greece: Test of simulation software and analysis of typical systems. Solar Energy 60(3-4):159-170. ASHRAE. 1988. Active solar heating systems design manual. In cooperation with Solar Energy Industries Association and American Consulting Engineers Council Research & Management Foundation. ASHRAE. 1990. Guide for preparing active solar heating systems operation and maintenance manuals. ASHRAE. 1991. Active solar heating systems installation manual. ASHRAE. 2016. Energy standard for buildings except low-rise residential buildings. ANSI/ASHRAE/IES Standard 90.1-2016. ASHRAE. 2018. Energy efficient design of low-rise residential buildings. ANSI/ASHRAE Standard 90.2-2018. ASME. 2019. Boiler and pressure vessel code. American Society of Mechanical Engineers, New York. Balaras, C.A., H-M. Henning, G. Grossman, E. Podesser, and C.A. Infante Ferreira. 2006. Solar cooling: An overview of European applications and design guidelines. ASHRAE Journal 48(6):14-22. Balaras, C.A., E. Dascalaki, P. Tsekouras, and A. Aidonis. 2010. High solar combi systems in Europe. ASHRAE Transactions 116(1):408-415. Paper OR-10-044. California Public Utilities Commission. 2017. California solar initiative program handbook. www.gosolarcalifornia.ca.gov/documents/CSI _HANDBOOK.PDF. Cerón, I., E. Caamaño-Martín, F.J. Neila. 2013. “State-of-the-art” of building integrated photovoltaic products. Renewable Energy 58:127-133. Cole, R.L., K.J. Nield, R.R. Rohde, and R.M. Wolosewicz. 1980. Design and installation manual for thermal energy storage. ANL-79-15. Argonne National Laboratory, Argonne, IL. Duffie, J.A., and W.A. Beckman. 2013. Solar engineering of thermal processes. John Wiley & Sons. Franta, G. 1981. Solar design workbook. Solar Energy Research Institute, Golden, CO. IEA. 2018. Trends 2018 in photovoltaic applications: Survey report of selected IEA countries between 1992 and 2017. Photovoltaic Power Systems Programme (PVPS). Report IEA PVPS T1-34:2018. International Energy Agency, Paris, France. IEA. 2019a. Snapshot of global PV markets. Photovoltaic Power Systems Programme (PVPS). International Energy Agency, Paris, France. IEA. 2019b. Building Integrated PV. Photovoltaic Power Systems Programme (PVPS), Task 15. www.iea-pvps.org/index.php?id=task15. International Energy Agency, Paris, France. ISO. 2017. Solar energy–solar thermal collectors–test methods. ANSI/ISO Standard 9806:2017. Jelle, B.P., Breivik, C., and H.D. Røkenes. 2012. Building integrated photovoltaic products: A state-of-the-art review and future research opportunities. Solar Energy Materials and Solar Cells 100:69-96. Kalogirou, S., Y. Tripanagnostopoulos, and M. Souliotis. 2005. Performance of solar systems employing collectors with colored absorber. Energy and Buildings 37(8):824-835. Kavlak, G., McNerney, J., and J.E. Trancik. 2018. Evaluating the causes of cost reduction in photovoltaic modules. Energy Policy 123:700-710. Knowles, D.S. 1981. A simple balancing technique for liquid cooled flat plate solar collector arrays. International Solar Energy Society, Phoenix, AZ. Kreider, J. 1982. The solar heating design process—Active and passive systems. McGraw-Hill, New York. Kutscher, C.F., R.L. Davenport, D.A. Dougherty, R.C. Gee, P.M. Masterson, and E.K. May. 1982. Design approaches for solar industrial process heat systems. SERI/TR-253-1356. Solar Energy Research Institute, Golden, CO. Newton, A.B., and S.F. Gilman. 1983. Solar collector performance manual. ASHRAE SP32.

NFPA. 2017. National electrical code®. ANSI/NFPA Standard 70-17. National Fire Protection Association, Quincy, MA. Norton, B., P.C. Eames, T.K. Mallick, M.J. Huang, S.J. McCormack, J.D. Mondol, and Y.G. Yohanis. 2011. Enhancing the performance of building integrated photovoltaics. Solar Energy 85(8):1629-1664. NREL. 2019. System advisor model. National Renewable Energy Laboratory, Lakewood, CO. sam.nrel.gov. Radue, C., and E.E. van Dyk. 2010. A comparison of degradation in three amorphous silicon PV module technologies. Solar Energy Materials and Solar Cells 94(3):617-622. REN21. 2018. Renewables 2018 global status report. Renewable Energy Policy Network for the 21st Century, Paris, September. www.ren21.net/ ren21activities/globalstatusreport.aspx. SUPSI. 2017. Building integrated photovoltaics: Product overview for solar building skins. Status Report 2017. La Scuola universitaria professionale della Svizzera italiana (SUPSI) and Solar Energy Application Centre (SEAC). www.bipv.ch/index.php/en/component/content/article?id=227: pubblicazioni-posters&catid=58. Todorovic, M.S., Pejkovic, and V. Zenovic, 2010. 3.5 MW seawater heat pump assisted multipurpose solar system’s 25 years of operation. ASHRAE Transactions 116(1): 27-241. Weiss W., and M. Spork-Dur. 2018. Solar heat worldwide: The Solar Heating and Cooling Programme. International Energy Agency, Paris, France. www.iea-shc.org.

BIBLIOGRAPHY Architectural Energy Corporation. 1991. Maintenance and operation of stand-alone photovoltaic systems, vol. 5. Naval Facilities Engineering Command, Charleston, SC. ASHRAE. 1994. Design guide for cool thermal storage. ASHRAE. 2018. ASHRAE greenguide: Design, construction, and operation of sustainable buildings, 5th ed. ASHRAE. 2010. Methods of testing to determine the thermal performance of solar collectors. ANSI/ASHRAE Standard 93-2010. ASHRAE. 1989. Methods of testing to determine the thermal performance of unglazed flat-plate liquid-type solar collectors. ANSI/ASHRAE Standard 96-1980 (R1989). ASHRAE/USGBC. 2017. Standard for the design of high-performance green buildings. ANSI/ASHRAE/USGBC/IES Standard 189.1-2017. ASHRAE and U.S. Green Building Council, Washington, D.C. ASTM. 2015. Standard test method for photovoltaic modules in cyclic temperature and humidity environments. ASTM Standard E1171-15 (2019). ASTM International. ASTM. 2015. Standard test method for saltwater immersion and corrosion testing of photovoltaic modules for marine environments. ASTM Standard E1597-10 (2019). American Society for Testing and Materials, West Conshohocken, PA. Davidson, J., and F. Orner. 2008. New solar electric home: The complete guide to photovoltaics for your home, 3rd ed. Chelsea Green Publishing, VT. Eiffert, P., and G.J. Kiss. 2000. Building-integrated photovoltaics for commercial and institutional structures: A sourcebook for architects and engineers. NREL/BK-520-25272. National Renewable Energy Laboratory, Golden, CO. German Energy Society. 2008. Planning and installing photovoltaic systems: A guide for installers, architects and engineers, 2nd ed. Earthscan, London. Henning, H-M., M. Motta, and D. Mugnier, eds. 2013. Solar cooling handbook—A guide to solar assisted cooling and dehumidification processes. AMBRA|V part of Walter de Gruyter GmbH, Vienna. ICBO. 1997. Uniform building code. International Conference of Building Officials, Whittier, CA. IEA. Photovoltaic Power Systems Programme. International Energy Agency. www.iea-pvps.org. IEA. Solar Heating and Cooling Programme. International Energy Agency. www.iea-shc.org. Klein, S.A., and W.A. Beckman. 2001. Solar systems analysis, F-chart user’s manual. F-Chart Software, Madison, WI. Klein, S.A., and W.A. Beckman. 2001. Photovoltaic systems analysis, PV Fchart user’s manual. F-Chart Software, Madison, WI. Klise, G.T., and J.S. Stein. 2009. Models used to assess the performance of photovoltaic systems. SAND2009-8258. Sandia National Laboratories, Albuquerque, NM.

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Solar Energy Equipment and Systems Lawrence, T., A.K. Darwich, and J.K. Means, eds. 2013. ASHRAE greenguide: Design, construction, and operation of sustainable buildings, 4th ed. Marion, W., and S. Wilcox. 1994. Solar radiation data manual for flat-plate and concentrating collectors. NREL/TP-463-5607. National Renewable Energy Laboratory, Golden, CO.

Risser, V., and H. Post, eds. 1995. Stand-alone photovoltaic systems: A handbook of recommended design practices. SAND87-7023. Photovoltaic Design Assistance Center, Sandia National Laboratories, Albuquerque, NM. Strong, S., and W.G. Scheller. 1993. The solar electric house: Energy for the environmentally-responsive, energy-independent home, 2nd ed. Chelsea Green, White River Junction, VT. Wiles, J. 2003. Photovoltaic systems and the National Electrical Code—Suggested practices. Photovoltaic Design Assistance Center, Sandia National Laboratories, Albuquerque, NM.

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Marion, W., and S. Wilcox. 1995. Solar radiation data manual for buildings. NREL/TP-463-7904. National Renewable Energy Laboratory, Golden, CO.

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Related Commercial Resources CHAPTER 38

COMPRESSORS POSITIVE-DISPLACEMENT COMPRESSORS .................................................................. 38.1 Performance............................................................................. 38.2 Abnormal Operating Conditions, Hazards, and Protective Devices................................................................ 38.4 Motors ...................................................................................... 38.6 RECIPROCATING COMPRESSORS ...................................... 38.7 ROTARY COMPRESSORS .................................................... 38.13 Rolling-Piston Compressors .................................................. 38.13 Rotary-Vane Compressors ..................................................... 38.15

Screw Compressors ................................................................ Scroll Compressors ................................................................ Trochoidal Compressors ........................................................ CENTRIFUGAL COMPRESSORS......................................... Application ............................................................................. Mechanical Design................................................................. Isentropic Analysis ................................................................. Polytropic Analysis ................................................................ Operation and Maintenance .................................................. Symbols ..................................................................................

38.16 38.26 38.30 38.32 38.37 38.39 38.41 38.42 38.43 38.44

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A

COMPRESSOR is one of the four essential components of the basic vapor compression refrigeration system; the others are the condenser, evaporator, and expansion device. The compressor circulates refrigerant through the system and increases refrigerant vapor pressure to create the pressure differential between the condenser and evaporator. This chapter describes the design features of several categories of commercially available refrigerant compressors. There are two broad categories of compressors: positive displacement and dynamic. Positive-displacement compressors increase refrigerant vapor pressure by reducing the volume of the compression chamber through work applied to the compressor’s mechanism. Positive-displacement compressors include many styles of compressors currently in use, such as reciprocating, rotary (rolling piston, rotary vane, single screw, twin screw), and orbital (scroll, trochoidal). Dynamic compressors increase refrigerant vapor pressure by continuous transfer of kinetic energy from the rotating member to the vapor, followed by conversion of this energy into a pressure rise. Centrifugal compressors function based on these principles. There are many reasons to consider each compressor style. Some compressors have physical size limitations that may limit their application to smaller equipment; some have associated noise concerns; and some have efficiency levels that make them more or less attractive. Each piece of equipment using a compressor has a certain set of design parameters (refrigerant, cost, performance, sound, capacity, etc.) that requires the designer to evaluate various compressor characteristics and choose the best compressor type for the application. Figure 1 addresses volumetric flow rate of the compressor as a function of the differential pressure (discharge pressure minus suction pressure) against which the compressor is required to work. Three common compressor styles are represented on the chart. Positive-displacement compressors tend to maintain a relatively constant volumetric flow rate over a wide range of differential pressures, because this compressor draws a predetermined volume of vapor into its chamber and compresses it to a reduced volume mechanically, thereby increasing the pressure. This helps to keep the equipment operating near its design capacity regardless of the conditions. Centrifugal compressors dynamically compress the suction gas by converting velocity energy to pressure energy. Therefore, they do not have a fixed volumetric flow rate, and the capacity can vary over a range of pressure ratios. This tends to make centrifugal-based equipment much more application specific. The preparation of this chapter is assigned to TC 8.1, Positive Displacement Compressors, and TC 8.2, Centrifugal Machines.

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Fig. 1 Comparison of Single-Stage Centrifugal, Reciprocating, and Screw Compressor Performance

1.

POSITIVE-DISPLACEMENT COMPRESSORS

Major types of positive-displacement compressors classified by compression mechanism design are shown in Figure 2. Compressors also can be further classified as single-stage or multistage, and by type of motor drive (electrical or mechanical), capacity control (single speed, variable speed, single speed with adjustable compression chamber volume), and drive enclosure (hermetic, semihermetic, open). Open-drive compressors are those in which the shaft or other moving part extends through a seal in the crankcase for an external drive. Ammonia compressors are predominantly manufactured only in the open design because of the incompatibility of the refrigerant and common hermetic motor materials. Most automotive compressors are also open-drive type. Hermetic compressors contain the motor and compressor in the same gastight housing, which is permanently sealed with no access for servicing internal parts in the field, with the motor shaft integral with the compressor crankshaft and the motor in contact with the

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Fig. 2 Types of Positive-Displacement Compressors (Classified by Compression Mechanism Design)

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refrigerant. Hermetic compressors normally have the motorcompressor pump assembly mounted inside a steel shell, which is sealed by welding. A semihermetic compressor (also called bolted, accessible, or serviceable) is a compressor of bolted construction that is amenable to field repair. The seal in the bolted joints is provided by O rings or gaskets.

1.1

PERFORMANCE

Compressor performance depends on an array of design compromises involving characteristics of the refrigerant, compression mechanism, and motor. The goal is to provide the following: • • • • •

Greatest trouble-free life expectancy Most refrigeration effect for least power input Lowest applied cost Wide range of operating conditions Acceptable vibration and sound level

Fig. 3

Ideal Compressor Cycle

The useful measure of compressor performance is the coefficient of performance (COP). The COP is the ratio of the compressor’s refrigerating capacity to the input power. For a hermetic or semihermetic compressor, the COP includes the combined operating efficiencies of the motor and the compressor: Capacity, W COP (hermetic or semihermetic) = ----------------------------------------------------------Input power to motor, W The COP for an open compressor does not include motor efficiency: Capacity, W COP (open) = --------------------------------------------------------Input power to shaft, W Because capacity and motor/shaft power vary with operating conditions, COP also varies with operating conditions. Power input per unit of refrigerating capacity (W/W) is used to compare different compressors at the same operating conditions, primarily with open-drive industrial equipment. W in Power input to shaft, W = --------------------------------------------------------------------W out Compressor capacity, W

Ideal Compressor During operation, pressure and volume in the compression chamber vary as shown in Figure 3. There are four sequential processes: first, gas is drawn into the compression chamber during the suction process (1-2); next is compression (2-3); and then higher-pressure gas is pushed out during the discharge process (34), followed by the next cycle. The capacity of a compressor at a given operating condition is a function of the mass of gas compressed per unit time. Ideally, mass

Fig. 4 Pressure-Enthalpy Diagram for Ideal Refrigeration Cycle flow is equal to the product of the compressor displacement per unit time and the gas density, as shown in Equation (1): m = sVd

(1)

where m = ideal mass flow of compressed gas, kg/s s = density of gas entering compressor (at suction port), kg/m3 Vd = geometric displacement of compressor, m3/s

The ideal refrigeration cycle, discussed in detail in Chapter 2 of the 2017 ASHRAE Handbook—Fundamentals, consists of four processes, as shown in Figure 4: 1–2: isentropic (reversible and adiabatic) compression 2–3: desuperheating, condensing, and subcooling at constant pressure

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Compressors

38.3

3–4: adiabatic expansion 4–1: boiling and superheating at constant pressure The following quantities can be determined from the pressureenthalpy diagram in Figure 4 using m, the mass flow of gas from Equation (1), Qo = mQrefrigeration effect = m(h1 – h4)

(2)

Poi = mQwork of compression = m(h2 – h1) = mwoi

(3)

Volumetric efficiency v is the ratio of actual volumetric flow to ideal volumetric flow (i.e., the geometric compressor displacement). Compression isentropic efficiency oi considers only what occurs within the compression volume and is a measure of the deviation of actual compression from isentropic compression. It is defined as the ratio of work required for isentropic compression of the gas wio to work delivered to the gas within the compression volume wa. oi = woi /wa

where woi = specific work of isentropic compression, J/kg Qo = ideal capacity, W Poi = ideal power input, W

Actual Compressor Ideal conditions never occur, so actual compressor performance differs from ideal performance. Various factors contribute to decreased capacity and increased power input. Depending on compressor type, some or all of the following factors can have a major effect on compressor performance.

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• Pressure drops in compressor • • • • • • • • •

Through shutoff valves Through suction accumulator Across suction strainer/filter Across motor (hermetic compressor) In manifolds (suction and discharge) Through valves and valve ports (suction and discharge) In internal muffler Through internal lubricant separator Across check valves

(4)

For a multicylinder or multistage compressor, this equation applies only for each individual cylinder or stage. Mechanical efficiency m is the ratio of work delivered to the gas (measured) to work input to the compressor shaft wm. m = wa /wm

(5)

Isentropic efficiency i is the ratio of work required for isentropic compression of the gas woi to work input to the compressor shaft wm. i = woi /wm = oim

(6)

Motor efficiency e is the ratio of work input to the compressor shaft wm to work input to the motor we. e = wm/we

(7)

Total compressor efficiency com is the ratio of work required for isentropic compression woi to work input to the motor we. com = woi /we = oime

(8)

• Heat gain by refrigerant from • Cooling the hermetic motor • Internal heat exchange between compressor and suction gas • Power losses because of • Friction • Lubricant pump power consumption • Motor losses • • • •

Valve inefficiencies caused by imperfect mechanical action Internal gas leakage Oil circulation Reexpansion (clearance losses). The gas remaining in the compression chamber after discharge reexpands into the compression chamber during the suction cycle and limits the mass of fresh gas that can be brought into the compression chamber. • Over- and undercompression. Overcompression occurs when pressure in the compression chamber reaches discharge pressure before finishing the compression process. Undercompression occurs when the compression chamber reaches the discharge pressure after finishing the compression process. • Deviation from isentropic compression. In the actual compressor, the compression process deviates from isentropic compression primarily because of fluid and mechanical friction and heat transfer in the compression chamber. The actual compression process and work of compression must be determined from measurements.

Compressor Efficiency, Subcooling, and Superheating Deviations from ideal performance are difficult to evaluate individually. They can, however, be grouped together and considered by category. Their effect on ideal compressor performance is characterized by the following efficiencies:

Actual shaft compressor power is a function of the power input to the ideal compressor and the compression, mechanical, and motor efficiencies of the compressor, as shown in the following equation: Pe = Pm /e = Pa /(me) = Poi /(oime)

(9)

where Pe Pm Pa Poi

= = = =

power input to motor power input to shaft power input to the compression mechanism (pump) power required for isentropic compression

Actual capacity is a function of the ideal capacity and volumetric efficiency v of the compressor: Q = Qov

(10)

Total heat rejection is the sum of refrigeration effect and heat equivalent of power input to the compressor. Heat radiation or using means for additional cooling may reduce this value. The quantity of heat rejection must be known in order to size condensers. Note that compressor capacity with a given refrigerant depends on saturation suction temperature (SST), saturation discharge temperature (SDT), superheating (SH), and subcooling (SC). Saturation suction temperature (SST) is the temperature of twophase liquid/gas refrigerant at suction pressure. SST is often called evaporator temperature; however, in real systems, there is a difference because of pressure drop between evaporator and compressor. Saturated discharge temperature (SDT) is the temperature of two-phase liquid/gas refrigerant at discharge pressure. SDT is often called condensing temperature; however, in

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real systems, there is a difference because of the pressure drop between compressor and condenser. Liquid subcooling is not accomplished by the compressor. However, the effect of liquid subcooling is included in compressor ratings by some manufacturers. Note: Air-Conditioning, Heating, and Refrigeration Institute (AHRI) Standard 540 and European Committee for Standardization (CEN) European Norm (EN) 12900 do not include subcooling. Suction Superheat. In general, no liquid refrigerant should be present in suction gas entering the low-pressure side of the compressor (especially reciprocating and scroll), because it causes oil dilution and gas formation in the lubrication system. If liquid carryover is severe enough to reach the cylinders, excessive wear of valves, valve stops, pistons, and rings can occur; liquid slugging can break valves, pistons, and connecting rods. Measuring suction superheat can be difficult, and the indication of a small superheat (