ASHRAE Fundamentals of Water System Design IP 2015 9781936504664, 9781939200044
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EN Pages [350] Year 2015
Table of Contents
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Preface
Acknowledgments
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Water System Design Concepts
Water System Design Concepts
Study Objectives
Instructions
Introductory Concepts
Figure 1-1 Source/load.
Figure 1-2 Source–distribution–load.
Figure 1-3 Source–distribution–part-load.
Figure 1-4 Hydronic system fundamentals (closed system).
Figure 1-5 Cooling tower (open system).
Figure 1-6 Low-temperature water system—direct or reverse return.
Figure 1-7 Medium- or high-temperature water system.
Figure 1-8 Chilled-water system—direct return piping.
Figure 1-9 Condenser open water system (once through).
Figure 1-10 Condenser cooling tower system inside reservoir or heated sump.
Figure 1-11 Distribution orientation.
Basic System Components
Figure 1-12 Closed hydronic system fundamental components.
Source
Figure 1-13 Steam-to-water or water-to-water heat exchanger.
Figure 1-14 Multiple chiller variable-flow chilled-water system.
(1-1)
Load
Pump System
Distribution System
Expansion Chamber
Design Trade-Offs
System Temperatures
Figure 1-15 Psychrometric chart chilled-water example.
Heat Transfer in Hydronic Systems
Sensible Heating or Cooling of Air
(1-2)
(1-3)
Figure 1-16 Example system with heating coil.
(1-4)
Figure 1-17 Example system with heating coil.
Figure 1-18
Latent Cooling and Dehumidification of Air
Figure 1-18 Coil LMTD example.
(1-5)
(1-6)
(1-7)
Figure 1-19 Example system with cooling coil.
Heat Transferred to or from Water
(1-8)
(1-9)
(1-10)
Load Systems
Figure 1-20 Fan-coil unit.
Figure 1-21 Typical coil.
(1-11)
Figure 1-22 Example process diagrammed on ASHRAE Psychrometric Chart.
(1-12)
Figure 1-23 Example of manufacturer’s coil selection program.
Table 1-1 Coil Depth and Velocity to Achieve Dehumidification
Figure 1-24 Heat transfer versus water flow for a coil.
The Next Step
Summary
References
Skill Development Exercises for Chapter 1
Complete these questions by writing your answers on the worksheets at the back of this book.
1-1
1-2
1-3
1-4
1-5
1-6
1-7
1-8
1-9
1-10
1-11
Piping System Design
Piping System Design
Study Objectives
Instructions
Basic Considerations
Pressure Drop
Figure 2-1 Bernoulli’s theorem.
(2-1)
Figure 2-2 Bernoulli’s piping example.
Design Philosophy
Figure 2-3 Direct-return piping.
Figure 2-4 Reverse-return piping.
Figure 2-5 Combination of direct and reverse systems.
Figure 2-6 Piping expansion, offset piping.
Figure 2-7 Piping expansion, mechanical joint.
Sizing Piping
(2-2)
(2-3)
Figure 2-8 Experimental arrangement for determining head loss in a pipe.
(2-4)
(2-5)
Figure 2-9 Reynolds’s laminar versus turbulent flow demonstration.
(2-6)
Figure 2-10 Relation of Reynolds number, friction flow, and relative roughness for similar pipes.
(2-7)
Figure 2-11 Moody chart showing relationship between friction factors and Reynolds number for water flow.
Figure 2-12 Friction factors and relative roughness for various pipes.
Figure 2-13
Figure 2-13 Kinematic viscosity and Reynolds number determination nomogram.
Figure 2-14
(2-8)
(2-9)
Figure 2-14 Friction loss for water in commercial steel pipe (Schedule 40) and friction loss for water in copper tubing (Types K, L, M).
(2-10)
Flow-Rate Measurement
Direct and Indirect Flow Measurement Methods
Table 2-1 Friction Loss for Water in Feet for 100 ft, 2 in. Nominal Pipe Schedule 40
Table 2-2 Volumetric or Mass Flow-Rate Measurement
Venturi, Nozzle, and Orifice Flowmeters
Figure 2-15 Typical Herschel-type Venturi meter.
Figure 2-16 Dimensions of ASME long-radius flow nozzles.
Figure 2-17 Sharp-edge orifice with pressure tap locations.
(2-11)
(2-12)
(2-13)
Variable-Area Flowmeters (Rotameters)
Figure 2-18 Variable-area flowmeter.
Turbine Flowmeters
The Next Step
Summary
References
Skill Development Exercises for Chapter 2
Complete these questions by writing your answers on the worksheets at the back of this book.
2-1
2-2
2-3
2-4
2-5
2-6
2-7
2-8
2-9
Pipe Materials and Fittings
Pipe Materials and Fittings
Study Objectives
Instructions
Pipe Materials
Metal Pipe
Steel Pipe
Table 3-1 Allowable Stressesa for Pipe and Tube
Copper Tube
Table 3-2 Steel Pipe Data
Table 3-3 Copper Tube Data
Table 3-4 Internal Working Pressure for Copper Tube Joints
Ductile Iron and Cast Iron Pipe
Joining Methods for Metal Pipe
Threading
Soldering and Brazing
Flared and Compression Joints
Flanges
Welding
Reinforced Outlet Fittings
Other Joints
Threaded Unions
Special Systems
Plastic Pipe
Allowable Stresses
Table 3-5 Properties of Plastic Pipe Materialsa
(3-1)
Plastic Material Selection
Table 3-6 Manufacturers’ Recommendations for Plastic Materialsa,b
Corrosion
Valves and Fittings
(3-2)
Table 3-7 K Factors—Threaded Pipe Fittings
Table 3-8 K Factors—Flanged Welded Pipe Fittings
Table 3-9 Approximate Range of Variation for K Factors
Figure 3-1 Resistance coefficients like Figure 3-2 for valves and fittings.
Figure 3-2 Resistance coefficients for valves and fittings.
Figure 3-3 Resistance coefficients for increasers and diffusers.
Figure 3-4 Resistance coefficients for reducers.
Backflow Prevention Devices
Figure 3-5 Backflow prevention device.
Selection of Devices
Installation of Devices
Pipe Selection
1. Recalling that pipe selection guidelines suggest keeping friction loss less than 4 ft/100 ft and design velocity greater than 2 fps, Table 2-1 allows the designer to plot out a preliminary pipe selection. For 40 gpm, limited by a head loss of 4 ft...
Figure 3-6 Pipe loop sizing example.
2. Determine the pressure loss due to friction for the 300 ft of piping:
Dhf = 300 ft × 2.88 ft/100 ft = 8.4 ft
Dhf = 1.15 × 8.4 ft = 9.66 ft
3. Calculate the PD for the gate valve. Looking at Tables 3-7 and 3-8, K for a 2 in. gate valve is 0.17.
Figure 3-7 Pipe sizing using friction loss chart.
4. Determine the PD due to the 90° regular pattern elbows PD we know:
5. Similarly, the PD for the tee joints is as follows:
6. The total PD is the sum of all component contributions.
Total Dhf = coil + pipe + balance valve + control valve + gate valve + tee branches + elbows
= (3.0 ft) + (9.66 ft) + (2.31 ft) + (8.34 ft) + (0.04 ft) + (0.64 ft) + (0.91 ft)
= 24.89 ft
System Characteristic Curve
Figure 3-8 Exponential curve of flow versus pressure drop.
Figure 3-9 Curve rotates counterclockwise as the valve closes.
Figure 3-10 Open-system diagram.
Figure 3-11 Change in curve because of elevation difference.
Figure 3-12 Data for plotting Figure 3-11.
The Next Step
Summary
Bibliography
Skill Development Exercises for Chapter 3
Complete these questions by writing your answers on the worksheets at the back of this book.
3-1
3-2
3-3
3-4
3-5
3-6
3-7
3-8
3-9
3-10
Centrifugal Pumps
Study Objectives
Instructions
Types of Pumps
Common Applications
Figure 4-1 Diagram of a CHW system.
Figure 4-2 Condenser water circuits to cooling tower.
Figure 4-3 Boiler feed and condensate return pumping.
Figure 4-4 Driver for HVAC pumps.
Operation
Figure 4-5 Vanes directing fluid.
Figure 4-6 Vanes are inclined backward.
Figure 4-7 Energy captured and directed through the impeller enclosure, called a volute.
Figure 4-8 Centrifugal pump, impeller, and volute.
Figure 4-9 Impeller action on fluid.
(4-1)
Equipment
Figure 4-10 Variety of impeller shroud shape characteristics.
Source: HI 1994.
Figure 4-11 Single- and double-suction impellers.
Typical Pump Configurations
Figure 4-12 Circulator pump.
Figure 4-13 Close-coupled end-suction pump.
Figure 4-14 Frame-mounted end-suction pump.
Figure 4-15 Vertical inline pump.
Figure 4-16 Base-mounted horizontal split-case pump.
Figure 4-17 Vertical turbine pump, wet sump arrangement.
Pump Nomenclature
Pump Selection
Centrifugal Pump Characteristics
Figure 4-18 Manufacturer’s pump curve.
Figure 4-19 Head–capacity curve.
Figure 4-20 Flat versus steep pump curve.
Figure 4-21 Characteristic curves for pump models at given speed.
Figure 4-22 Selected pump head–capacity curve.
Figure 4-23 System curve and pump head–capacity curve.
Pump Horsepower
Water Horsepower
(4-2)
(4-3)
Figure 4-24 Comparing the increase of WHP as the pump increases flow.
(4-4)
Brake Horsepower
Figure 4-25 BHP on the pump curve above the pump head curve.
(4-5)
Pump Efficiency
(4-6)
Figure 4-26 Increase of pumping power required with pump flow.
Figure 4-27 Pump efficiency curves.
Figure 4-28 Recommended selection regions.
Figure 4-29 Pump performance data.
Figure 4-30 Pressures on impeller causing radial thrust.
Figure 4-31 Change in radial thrust versus pumping rate.
Radial Thrust
NPSH
Figure 4-32 Selected pump curve showing NPSHR.
(4-7)
Figure 4-33 NPSHA in a proposed installation.
(4-8)
Figure 4-34 NPSHA in an existing installation.
Figure 4-35 Factory test setup to determine pump’s NPSHR.
Pump Selection Process
1. Determine the load to be pumped (heating or cooling) in Btu/h.
2. Determine the design Dt across the water side of the load heat transfer coil or device. For example, 20°F drop for heating or 12°F rise for cooling. Calculate the required flow in gpm for each load.
3. Total the zone load flows to determine the total flow in gpm.
4. For the secondary pump, select the most resistant path within the secondary distribution and terminal piping.
5. Determine the method of mechanically mounting and supporting the pump on a pad in the equipment room, inline in the piping, or within a well or wet sump below floor level.
6. With the total head (in feet drop) and capacity (in gpm) determined, select a pump from the manufacturer’s family curves and the mounting required.
Figure 4-36 Pump selection process.
7. Refer to the manufacturer’s individual pump performance curve sheet showing pump efficiency, impeller diameter size, brake horsepower, and NPSHR. Select a flat curve pump for closed systems with control valves (to minimize variation in head for ...
System Design Considerations
Similarity Relationships: The Affinity Laws
Table 4-1 Equations for Speed Change and Impeller Diameter Change
Table 4-1 Equations for Speed Change and Impeller Diameter Change
Parallel Pumping
Figure 4-37 The old way: sketching the composite parallel pump curve by hand.
Figure 4-38 The new way: sketching the composite parallel pump curve by spreadsheet.
Pumps in Series
Figure 4-39 An example with two full-sized pumps.
Figure 4-40 Illustration of pump and system curves applied in parallel or in series and relative system flow effects.
Figure 4-41 Schematic modifications example to place pumps in series.
Variable-Speed Pumping
Figure 4-42 Example of pump curves at varying pump speeds.
Figure 4-43 Schematic of simplified variable-speed-pump-controlled system.
Figure 4-44 Typical constant-speed pump curve interaction with variable-speed pump.
Figure 4-45 Flow coefficient method of calculating balanced and unbalanced conditions.
Figure 4-46 System curve for variable-speed pumping.
Figure 4-47 Control area curve for example.
Figure 4-48 Calculation of total dynamic head for area curve due to valve change of states.
Figure 4-49 Control area curve with system curve overlaid.
The Next Step
Summary
References and Bibliography
Skill Development Exercises for Chapter 4
Complete these questions by writing your answers on the worksheets at the back of this book.
4-1
4-2
4-3
4-4
4-5
4-6
4-7
4-8
4-9
4-10
4-11
4-12
4-13
Terminal Unit Performance and Control
Study Objectives for Chapter 5
Instructions
Types of Terminals
Water Supply Systems to Terminals
Two-Pipe System
Figure 5-1 Two-pipe source/load concept.
Three-Pipe System
Four-Pipe System
Figure 5-2 Four-pipe dual-temperature water system.
Performance and Control
Figure 5-3 Space heat transfer in proportion to outdoor temperature.
Figure 5-4 As outdoor temperature increases, hot-water flow decreases.
Controlling Water Flow
Figure 5-5 Flow control using two-way valve.
Types of HVAC Control Valves
Figure 5-6 Flow control using three-way valve.
Figure 5-7 Two-pipe control valve body–single and double seat.
Figure 5-8 Three-way control valves—mixing and diverting types.
Globe Valves
Ball Valves
Figure 5-9 Actuated ball valve.
Butterfly Valves
Other Valves
Hydronic Accessories
Determining Valve Flow Rate
(5-1)
(5-1)
System Control Characteristics
Figure 5-10 Nonlinear heat transfer versus water flow in hydronic coil.
Figure 5-11 Core flow characteristics for a valve.
Equal Percentage Characteristic
(5-2)
Figure 5-12 Valve flow characteristic.
Valve Rangeability
Valve Authority
(5-3)
Figure 5-13 Distortion effect of low control valve authority caused by low pressure drop.
(5-4)
(5-5)
(5-6)
Modulating Control
Figure 5-14 Chilled-water coil heat transfer versus water flow.
Figure 5-15 Ideal combination of equal percentage valve curve with water coil emission curve.
Figure 5-16 Control effect of valve authority.
Two-Position Control
Figure 5-17 Two-position control.
Figure 5-18 Modulating proportional control.
Proportional and Proportional Integral Control
Figure 5-19 Proportional integral (PI) control.
System Control Configurations
Figure 5-20 Primary/secondary pumping with two-way valve.
Figure 5-21 Primary/secondary with check valve in common.
Figure 5-22 Primary/secondary with valve in common and differential pressure transmitter monitoring flow.
Figure 5-23 Primary/secondary with temperature sensors in bridge.
Figure 5-24 Terminal with face bypass control.
Figure 5-25 Terminal with face bypass control and conditioned bypass air.
Figure 5-26 Variable supply water temperature for part-load conditions.
Pressure-Independent Control Valves
Figure 5-27 Linear control characteristic.
Figure 5-28 Fluid-system-powered regulator.
Variable-Speed Circulator on Coil
Figure 5-29 Differential pressure regulator in same body as temperature control valve.
Figure 5-30 Variable-speed circulator on coil.
The Next Step
Summary
References
Skill Development Exercises for Chapter 5
5-1
5-2
5-3
5-4
5-5
5-6
5-7
5-8
5-9
5-10
5-11
5-12
5-13
5-14
Expansion Tanks and Air Elimination
Expansion Tanks and Air Elimination
Study Objectives
Instructions
Open and Closed Water Systems
Typical Open System
Figure 6-1 Typical open water system.
Typical Closed System
Figure 6-2 Typical hydronic system.
Hydronic Accessories
Pressure Relief Valve
Pressure-Reducing Valve
Figure 6-3 Pressure-reducing valve.
Expansion Tank
Figure 6-4 Expansion tanks.
Figure 6-5 Closed tank.
Figure 6-6 Expansion of water above 40°F.
Figure 6-7 Tank pressure related to system pressure.
Figure 6-8 Point of no pressure change.
Air Elimination
Air Separation
Figure 6-9 Internal operating mechanism for automatic air vent.
Figure 6-10 Automatic or manual air vents in system zones or coils for small pipe sizes.
Figure 6-11 Dynamic air separator.
Figure 6-12 Air separator.
Figure 6-13 Diaphragm expansion tank.
Figure 6-14 Compression tank.
Figure 6-15 Tank location for primary/secondary or compound pumping systems.
Sizing Expansion Tanks
Figure 6-16 Pressure effects of alternative tank locations: (a) pump suction side and (b) pump discharge side.
(6-1)
Figure 6-17 Henry’s constant versus temperature for air and water.
(6-2a)
Figure 6-18 Solubility versus temperature and pressure for air-water solutions.
Table 6-1 Volume of Water in Standard Pipe and Tube
(6-2b)
Pressure and Temperature Considerations
Figure 6-19 Flowchart for sizing expansion tanks.
The Next Step
Summary
References
Skill Development Exercises for Chapter 6
6-1
6-2
6-3
6-4
6-5
6-6
6-7
6-8
6-9
6-10
6-11
6-12
6-13
Piping System Development
Study Objectives
Instructions
Piping System Design
1. Know the building heat transfer load. In the initial approach, all aspects of the load must be known. What is the total load? How is the load distributed by time and location? How are controlled occupancy zones determined or laid out? As the syste...
Figure 7-1 Typical building layout.
2. Determine the heating and cooling loads based upon occupancy, comfort requirements, codes, and standards (see ANSI/ASHRAE/IES Standard 90.1 and 90.1 User’s Manual), and determine any special requirements for facilities like computer rooms, labor...
3. Develop a concept for part-load control:
Figure 7-2 Determine the loads and consult references.
4. Develop the piping and pumping system concept (see Figure 7-3), such as the following:
Consider modeling the system to determine the full- and part-load flows, the pressure distribution required, and this effect on components.
5. Develop a first-cost analysis versus energy operating costs over the projected life of the system.
6. Determine the maintenance and operating requirements and if they match with the personnel capabilities.
Figure 7-3 Develop piping/pumping system concept.
Direct-Return Analysis
Figure 7-4 Piping system design flowchart.
Figure 7-5 System requiring four AHUs, each with 100 gpm load.
Figure 7-6 Coil connections.
Figure 7-7 Friction loss, schedule 40 steel pipe.
Table 7-1 Schedule 40 Steel Pipe (3 in. Nominal Discharge)
Table 7-2 Schedule 40 Steel Pipe (4 in. Nominal Discharge)
Table 7-3 Schedule 40 Steel Pipe (5 in. Nominal Discharge)
Table 7-4 Unit 1 Supply and Return Sides
Figure 7-8 Direct-return piping layout.
Table 7-5 Direct Return for Units 1, 2, 3, and 4
Table 7-6 Lower Coil Pressure Drop Selected for Units 1, 2, 3, and 4
Reverse-Return Analysis
Figure 7-9 Reverse-return piping layout.
Table 7-7 Reverse-Return Piping Arrangement for Units 1, 2, 3, and 4
Summary
Summary of Pumping Horsepower— Direct Return Versus Reverse Return
Table 7-8 Lower Pressure Drop Coil for Units 1, 2, 3, and 4
Table 7-9 Direct and Reverse Return with Balanced and Unbalanced Flow
Table 7-10 Summary of Direct and Reverse Return
Primary/Secondary Analysis
Figure 7-10 Two-way valve in connecting bridge return with secondary pump.
Figure 7-11 Primary/secondary pumping.
Table 7-11 Pressure Drop from Main to Secondary Bridge for Units 1, 2, 3, and 4
Table 7-12 Pressure Drop in Typical Secondary Loop
Types of Pumps and Valves
Figure 7-12 Flat versus steep pump characteristics.
Effects of Control Valves
Primary/Secondary Application Study
Figure 7-13 Four-zone heating system.
Table 7-13 Four-Zone Heating System
Figure 7-14 Primary/secondary pumping, four-zone heating system.
Table 7-14 3 hp Motor Selected
Figure 7-15 Primary/secondary bridge energy.
Table 7-15 Primary/Secondary Bridge Energy
Figure 7-16 Primary/secondary pumping, four-zone heating system.
Antifreeze Solutions for Low-Temperature Applications
Figure 7-17 Coil with glycol heat exchanger and pump for low temperatures.
Figure 7-18 Pumped coil with face bypass dampers for low-temperature primary/secondary pumping.
(7-1)
Figure 7-19 Specific heats of aqueous solutions of industrially inhibited ethylene glycol (percent by volume).
Figure 7-20 Specific heats of aqueous solutions of industrially inhibited propylene glycol (percent by volume).
Pumping Design Factors
Figure 7-21 One-shot chemical feeder.
The Next Step
Summary
References and Bibliography
Skill Development Exercises for Chapter 7
7-1
7-2
7-3
7-4
7-5
7-6
7-7
7-8
7-9
7-10
7-11
7-12
7-13
7-14
7-15
7-16
7-17
Matching Pumps to Systems
Study Objectives
Instructions
Matching the Pump to the System
Figure 8-1 Typical open-system system curve.
Source: ASHRAE Handbook—HVAC Systems and Equipment (2012).
Figure 8-2 Pump curve and system curve intersection.
Figure 8-3 System with cooling tower.
Figure 8-4 Cooling tower curves.
System Curves
(8-1)
Series Flow Coefficients
(8-2)
(8-3)
(8-4)
(8-5)
(8-6)
(8-6)
Table 8-1 Theoretical Equal Percentage Control Valve Data for Example 8-1
Figure 8-5 Valve flow versus lift for Example 8-1.
Parallel Flow Coefficients
Figure 8-6 Head versus flow for pump serving three circuits.
(8-7)
Table 8-2 Balanced Path Example
Table 8-3 Unbalanced Path Example
Parallel Pumping
Figure 8-7 Unbalanced system curve intersection point with pump curve.
Figure 8-8 Effects of larger distribution pipes and resized coils and control valves.
Figure 8-9 Pump curve developed on paralleled pump curve line A-B-C.
Figure 8-10 Pump motor sized to prevent overloading during single pump operation.
Figure 8-11 Piping schematic of parallel pumps.
Series Pumping
Figure 8-12 Pump curve for series operation.
Figure 8-13 Piping schematic of series pumps.
Figure 8-14 Operating conditions for series pump installation.
Standby Pumps
Figure 8-15 Piping schematic of standby pump.
Trimming Pump Impellers and Adjusting Pump Speed
Figure 8-16 Pump operating points.
Two-Speed Pumping
Figure 8-17 Two-speed pumping.
Figure 8-18 Typical performance curve—6 in. suction × 8 in. discharge × 9.5 in. impeller.
Pumps with Two-Speed Motors (Stethem 1988)
Figure 8-19 Two-speed pumping example.
Table 8-4 Four Flow Selection Steps to Reduce Pumping Power
Variable-Speed Pumping
Figure 8-20 Typical direct-return system.
Figure 8-21 Head loss with two-way valves at full load and 50% part load.
Figure 8-22 Proportional controller and adjustable-frequency drive controlling pump.
Figure 8-23 Direct-return system with sensor/transmitter located at end of last riser.
Figure 8-24 Differential pressure control curve above piping friction loss.
Figure 8-25 Pump curve showing head reduction with change in pump speed.
Figure 8-26 Pump curve showing pumping power reduction with change in pump speed.
Source Distribution Pumping
Figure 8-27 Primary/secondary pumping concept.
Figure 8-28 Main-source primary/secondary variable-speed pumping.
Figure 8-29 Distributed variable-speed pumping.
The Next Step
Summary
References and Bibliography
Skill Development Exercises for Chapter 8
8-1
8-2
8-3
8-4
8-5
8-6
8-7
8-8
8-9
Water Chillers and Load Control
Study Objectives
Instructions
Basic Water Chiller Components
Refrigeration Cycle
Figure 9-1 Basic components of liquid chilling system.
Figure 9-2 Schematic of simple liquid chilling system.
Figure 9-3 Pressure-enthalpy diagram for a refrigerant.
Figure 9-4 Simplified pressure-enthalpy diagram for a refrigerant.
Figure 9-5 Refrigeration cycle shown on simplified pressure-enthalpy diagram.
Heat Transfer Chiller
(9-1)
(9-2)
(9-3)
(9-4)
(9-5)
(9-6)
(9-7)
Solution
Solution
Refrigeration Power
Chiller Types and Control
Figure 9-6 General guideline of the types of chillers available for air conditioning.
Source: ASHRAE Handbook—Refrigeration (2014).
Figure 9-7 Centrifugal compressor cross section.
Figure 9-8 Section of single-screw refrigeration compressor.
Figure 9-9 Sequence of compression process in single-screw compressor.
Figure 9-10 Vertical, discharge-cooled, hermetic twin-screw compressor.
Figure 9-11 Scroll compression process.
Source: Purvis (1987).
Figure 9-12 Bearings and Other Components of Scroll Compressor
Source: Elson et al. (1990).
Figure 9-13 Reciprocating compressor refrigeration system.
Figure 9-14 Two-shell lithium bromide cycle water chiller.
Figure 9-15 Schematic of double-effect, direct-fired absorption chiller with reverse parallel flow cycle.
Figure 9-16 Comparison of single-stage centrifugal, reciprocating, and screw compressor performance.
Figure 9-17 Reciprocating liquid chiller performance with three equal steps of unloading.
Figure 9-18 Parallel arrangement: water to be chilled is divided among liquid chillers and combined again in common header after chilling.
Figure 9-19 Series arrangement.
Figure 9-20 Chillers piped in parallel in primary production loop.
Figure 9-21 Schematic showing constant- and variable-flow arrangements.
Chiller Piping Arrangements
Figure 9-22 Three chillers in parallel, each with a dedicated constant-speed pump.
Figure 9-23 Parallel arrangement with common bridge between load and production sections.
Figure 9-24 Parallel arrangement with common bridge at opposite end of production section.
Figure 9-25 Primary/secondary allows for simple addition of extra chillers and loads.
Figure 9-26 Pumping system manifolded to sources gives flexibility for maintenance and failure operation modes as well as capability to provide more system flow.
Chiller Energy Performance
Figure 9-27 Chiller performance improvement versus percent load.
Figure 9-28 Chiller, control, and pumping alternatives versus design temperature rises.
Figure 9-29 Relationship between chiller power consumption and variable-speed pump power consumption.
Thermal Storage
Figure 9-30 Schematic of thermal storage concept using ice builder with a chilled-water system.
Summary
References
Skill Development Exercises for Chapter 9
9-1
9-2
9-3
9-4
9-5
9-6
9-7
9-8
9-9
Skill Development Exercises for Chapter 1
1-1
1-2
1-3
1-4
1-5
1-6
1-7
1-8
1-9
1-10
1-11
Skill Development Exercises
Skill Development Exercises
Skill Development Exercises for Chapter 2
2-1
2-2
2-3
2-4
2-5
2-6
2-7
2-8
2-9
Skill Development Exercises for Chapter 3
3-1
3-2
3-3
3-4
3-5
3-6
3-07
3-8
3-9
3-10
3-10
Diagram for exercise 3-10.
Skill Development Exercises for Chapter 4
4-1
4-2
Centrifugal Pump.
4-3
4-4
4-5
4-6
4-7
4-8
4-9
4-10
4-11
4-12
4-13
Skill Development Exercises for Chapter 5
5-1
5-2
5-3
5-4
5-5
5-6
5-7
5-8
5-9
5-10
5-11
5-12
5-13
5-14
Skill Development Exercises for Chapter 6
6-1
6-2
6-3
6-4
6-5
6-6
6-7
6-8
6-9
6-10
6-11
6-12
6-13
Skill Development Exercises for Chapter 7
7-1
7-2
7-3
7-4
7-5
7-6
7-7
7-8
7-9
7-10
7-11
7-12
7-13
7-14
7-15
7-16
7-17
Skill Development Exercises for Chapter 8
8-1
8-2
8-3
8-4
8-5
8-6
8-7
8-8
8-9
Skill Development Exercises for Chapter 9
9-1
9-2
9-3
Figure 9-1 Diagram for Exercise 9-3.
9-4
9-5
9-6
9-7
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9-9
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Fundamentals of Water System Design Second Edition Mark Hegberg Richard A. Hegberg
I-P Inch-Pound
A Course Book for Self-Directed or Group Learning
Includes Skill Development Exercises for PDH, CEU, or LU Credits
SDL_cover_I-P.indd 1
10/22/2015 1:52:53 PM
Fundamentals of Water System Design Second Edition
Mark Hegberg Richard A. Hegberg
Atlanta
Fundamentals of Water System Design (I-P), Second Edition A Course Book for Self-Directed or Group Learning ISBN 978-1-936504-66-4 (paperback) ISBN 978-1-939200-04-4 (PDF) SDL Number: 98020 © 1996, 2000, 2015 ASHRAE All rights reserved. ASHRAE is a registered trademark in the U.S. Patent and Trademark Office, owned by the American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. 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 publication 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. 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.
ASHRAE STAFF ASHRAE Learning Institute
Special Publications
Karen Murray Manager of Professional Development Martin Kraft Managing Editor of Professional Development
Mark Owen Editor/Group Manager of Handbook and Special Publications Cindy Sheffield Michaels Managing Editor Matt Walker Associate Editor Sarah Boyle Assistant Editor Lauren Ramsdell Assistant Editor Michshell Phillips Editorial Coordinator
For course information or to order additional materials, please contact: ASHRAE Learning Institute 1791 Tullie Circle, NE Atlanta, GA 30329
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Errors or omissions in the data should be brought to the attention of Special Publications via [email protected]. Any updates/errata to this publication will be posted on the ASHRAE Web site at www.ashrae.org/publicationupdates.
Your Source for HVAC&R Professional Development
1791 Tullie Circle, NE • Atlanta, GA 30329-2305 USA • Tel 404.636.8400 • Fax 404.321.5478 • www.ashrae.org
Karen M. Murray
Email: [email protected]
Manager of Professional Development
Dear Student, Welcome to the ASHRAE Learning Institute (ALI) Fundamentals of HVAC&R Series of self-directed or group learning courses. We look forward to working with you to help you achieve maximum results from this course. You may take this course on a self-testing basis (no continuing education credits awarded) or on an ALI-monitored basis with credits (PDHs, CEUs or LUs) awarded. ALI staff will provide support and you will have access to technical experts who can answer inquiries about the course material. For questions or technical assistance, contact us at 404-636-8400 or [email protected]. Skill Development Exercises at the end of each chapter will gauge your comprehension of the course material. If you take this course for credit, please complete the exercises and send copies from each chapter to [email protected] (preferred method) or ASHRAE Learning Institute, 1791 Tullie Circle, Atlanta, GA 303292305. Be sure to include your student ID number with each set of exercises. Your student ID can be the last five digits of your Social Security number or another unique 5-digit number you create. We will return answer sheets to the Skill Development Exercises and maintain records of your progress. Please keep copies of your completed exercises for your own records. When you finish all exercises, please submit the course evaluation, which is located at the back of your course book. Once we receive all chapter exercises and the evaluation, we will send you a Certificate of Completion indicating 35 PDHs/LUs or 3.5 CEUs of continuing education credit. The ALI does not award partial credit for SDLs. All exercises must be completed to receive full continuing education credit. You will have two years from the date of purchase to complete each Self-Directed Learning Course. We hope your educational experience is satisfying and successful. Sincerely,
Karen M. Murray Manager of Professional Development American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. AN INTERNATIONAL ORGANIZATION
Continuing Education Opportunities from the ASHRAE Learning Institute Self-Directed or Group Learning ASHRAE offers texts for self-study or group training with instructor materials. Texts cover the basics of what a practicing engineer needs, and skill development exercises are included to evaluate progress. ASHRAE offers the following course books: Fundamentals of Air System Design Fundamentals of Psychrometrics Fundamentals of Building Operation, MainteFundamentals of Refrigeration nance, and Management Fundamentals of Steam System Design Fundamentals of Heating and Cooling Loads Fundamentals of Thermodynamics Fundamentals of Heating Systems Fundamentals of Water System Design Each course book includes the following:
• Clear and concise discussion of the technical topic covered • Examples that show how to apply the lesson’s principles • Skill development exercises that test students’ ability to apply the newly acquired knowledge and answer sheets to assess progress in learning the material Those who complete a course receive a certificate designating continuing education (CE) credits. Note that individuals are responsible for contacting their relevant governing body to determine whether an activity qualifies for that body’s continuing education credits.
ASHRAE eLearning ASHRAE is a continuing education provider of the American Institute of Architects (AIA) and the Green Building Certification Institute (GBCI), a third-party certification administrator of the United States Green Building Council (USGBC). Continuing Education (CE) hours earned from ASHRAE courses may be applied toward maintenance of state-licensed professionals (Architects and Professional Engineers) and Leadership in Energy and Environmental Design® (LEED®) Green Building Rating System credentials. ASHRAE offers over 130 eLearning courses focusing on specific topics and 18 course packages that cover a topical area and include several related courses. ASHRAE offers the following course packages: AC Design Electrical System Design HVAC Systems Standard 62.1 HVAC Control Systems Standard 90.1 DDC Controls Standard 90.1 for Architects Small Office Buildings Standard 189.1 for High-Performance Small Retail Buildings Green Buildings Sustainable Buildings Data Center Equipment Load Trends AC and Refrigeration Principles and Planning AC and Refrigeration Equipment Data Center Thermal Guidelines Fundamentals: Electricity Data Center Liquid Cooling
Mark C. Hegberg: Mark Hegberg has spent major portions of his career involved in direct digital control system application for Tour & Andersson, consulting engineering for A. Epstein and Sons and others, and hydronic training and product development for ITT Bell & Gossett. In recent years Mark has been significantly involved in hydronic system testing and balancing and the issues associated with flow control in the field. Mark has chaired and been involved with several ASHRAE committees and is a recipient of the ASHRAE Exceptional Service Award. He too has been associated with the Illinois Chapter HVAC training class for many years. Richard (Dick) A. Hegberg, PE: Dick Hegberg spent the majority of his career in the area of temperature control application for Powers Regulator company (now Siemens Building Technology), where he lead their early application engineering efforts and development of their building automation systems. In particular, he devoted much time to valves and dampers. Dick was actively involved in several ASHRAE TCs regarding controls and hydronics, as well as other groups that used both as part of their processes. Dick is an ASHRAE Life Member, Fellow, and recipient of the Distinguished Service Award. Locally, he inaugurated and directed for over thirty years the Illinois Chapter HVAC training class which is taught to this day.
Table of Contents Preface . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . viii Acknowledgments . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . ix Chapter 1: Water System Design Concepts . . . . . . . . . . . . . . . . . . . . 1 Introductory Concepts. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1 Basic System Components. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 10 Heat Transfer in Hydronic Systems . . . . . . . . . . . . . . . . . . . . . . . . 15 Load Systems . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 21 Chapter 2: Piping System Design. . . . . . . . . . . . . . . . . . . . . . . . . . . . 31 Basic Considerations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 31 Design Philosophy . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 34 Sizing Piping . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 38 Flow-Rate Measurement . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 46 Chapter 3: Pipe Materials and Fittings . . . . . . . . . . . . . . . . . . . . . . . . 59 Pipe Materials . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 59 Corrosion . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 73 Valves and Fittings . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 75 Backflow Prevention Devices . . . . . . . . . . . . . . . . . . . . . . . . . . . . 80 Pipe Selection . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 81 Chapter 4: Centrifugal Pumps . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 93 Types of Pumps. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 93 Operation . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 95 Pump Selection . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 104 System Design Considerations . . . . . . . . . . . . . . . . . . . . . . . . . . . 120 Variable-Speed Pumping . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 127 Chapter 5: Terminal Unit Performance and Control . . . . . . . . . . . 139 Types of Terminals . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 139 Performance and Control . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 142 System Control Characteristics . . . . . . . . . . . . . . . . . . . . . . . . . . 150 System Control Configurations . . . . . . . . . . . . . . . . . . . . . . . . . . 159 Chapter 6: Expansion Tanks and Air Elimination . . . . . . . . . . . . . . 171 Open and Closed Water Systems . . . . . . . . . . . . . . . . . . . . . . . . 171 Hydronic Accessories . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 173 Sizing Expansion Tanks . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 184 Chapter 7: Piping System Development . . . . . . . . . . . . . . . . . . . . . 193 Piping System Design. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 193 Direct-Return Analysis . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 197
Reverse-Return Analysis. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Primary/Secondary Analysis . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Types of Pumps and Valves . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Primary/Secondary Application Study . . . . . . . . . . . . . . . . . . . . . . Antifreeze Solutions for Low-Temperature Applications . . . . . . . . Pumping Design Factors. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Chapter 8: Matching Pumps to Systems . . . . . . . . . . . . . . . . . . . . . Matching the Pump to the System . . . . . . . . . . . . . . . . . . . . . . . . Parallel Pumping . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Series Pumping . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Standby Pumps . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Trimming Pump Impellers and Adjusting Pump Speed . . . . . . . . . Two-Speed Pumping. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Variable-Speed Pumping . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Source Distribution Pumping . . . . . . . . . . . . . . . . . . . . . . . . . . . . Chapter 9: Water Chillers and Load Control . . . . . . . . . . . . . . . . . Basic Water Chiller Components . . . . . . . . . . . . . . . . . . . . . . . . . Refrigeration Cycle . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Heat Transfer Chiller . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Refrigeration Power. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Chiller Types and Control . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Chiller Piping Arrangements. . . . . . . . . . . . . . . . . . . . . . . . . . . . . Chiller Energy Performance . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Thermal Storage . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Skill Development Exercises
210 215 218 221 226 229 235 235 247 251 255 255 258 261 266 273 273 274 277 279 279 288 295 298
Preface Fundamentals of Water System Design, Second Edition is an entry level text for the principles and fundamentals of water, or hydronic, HVAC system design and calculations. It uses the ASHRAE Handbook, and its related chapters on hydronics, as well as various research projects, ASHRAE documents, and articles on hydronics. The text represents the fundamentals of design and is an equally sound footing for work in commissioning, operation, and troubleshooting. The manual has been evolving over thirty years under the sponsorship of ASHRAE and is supported by ASHRAE Technical Committee (TC) 6.1, Hydronic and Steam System Design. For the majority of its history, this writing effort was lead by Richard (Dick) A. Hegberg. Hydronic systems are a key foundation technology of HVAC system design. Hydronics provides an important tool in the designer’s toolbox of providing effective and efficient HVAC systems to buildings around the world. One can argue exactly when hydronics started to be used for heating buildings; however, the modern application of hot-water systems may have started in the 1870s in the United States, and the technology has been refined ever since. Of the many benefits that water plays as a heat transfer medium for heating or cooling, fundamentally important is the ability to transport a lot of energy in the small area of a pipe (especially when compared to the equivalent size of an air duct). That benefit allowed for the development of central systems, eliminating the necessity of a local heating production system for each heated zone. The advancement allowed us to exercise individuality in our own building zone comfort control and lead to widespread application of hydronic systems in HVAC. From the early days of gravity circulation hot-water heating systems to the most complex of district energy chilled-water systems that are used today, the fundamentals presented here can be employed to successfully design the modern hydronic system.
Acknowledgments ASHRAE TC 6.1 and its members have played a key role in collecting, researching, and communicating the details and experiences for the successful application of hydronic systems over the years. TC 6.1 is an ad-hoc collection of individuals interested in hydronic system technology who take the time to volunteer and attend meetings and then communicate what they have learned to the other 53,000 members of ASHRAE. Key in the educational organizing efforts of TC 6.1 in hydronic systems and ASHRAE, a few individuals stand out when this text and the general concept of the ASHRAE’s professional development series (be they Self Directed Learning manuals or the face-to-face version of the Professional Development Seminar) are considered. At the apex of this is William (Bill) J. Coad, former ASHRAE President and longtime member of several ASHRAE committees. When the professional development series was first developed, Bill Coad, J. Barry Graham, and Gerry Williams developed the first two classes on Air System Design and Energy Efficient HVAC. Bill’s significant involvement in leading many ASHRAE activities left little time to develop the Water Systems manual, which lead to Dick Hegberg taking up the charge. All of these volunteer member contributors are owed a deep debt of gratitude for their devotion to helping ASHRAE develop the initial active training efforts that have evolved into the many practical training programs that are available today. ASHRAE has, from its very inception as ASHVE, represented a unique cooperative collaboration between all aspects of the HVAC industry. Designers, owners, operators, academicians, and manufacturers have all met and debated the merits of all things HVAC&R for over 100 years, seeking the answers to the problems associated with the HVAC system challenges of the day and providing the fundamental knowledge required by all professionals in a noncommercial manner. In that regard, while there are many professionals from the manufacturing or “product” side of the business that have contributed to advancing the area of hydronics, we would be remiss to not mention the significant contributions of Gil Carlson of Bell & Gossett. As hydronics shifted from steam heating system application to hot- and chilled-water systems in the 1950s and 1960s and then rolled into the early days of the first energy crisis of the 1970s, Gil’s application engineering, research expertise, and ability to communicate was an invaluable resource to the industry and ASHRAE TC 6.1. His significant ASHRAE research and journal articles on hydronic application, as well as the simple design tools he developed, were the educational precursor to what later became the professional development series. The industry is indebted to Gil’s contributions and to Bell & Gossett, now a division of Xylem Corporation, for their support to this day of the fundamentals of application engineering education in the area of hydronics. On a personal note, thanks to all those within ASHRAE who, for no other reason than their commitment to the spirit of mentoring that ASHRAE represents, helped provide me the practical education that comes from beyond col-
lege corridors. I thank you all and in particular Dick Hegberg, Bill Coad, and Roy Ahlgren. I have had the great fortune to come to know many of the ASHRAE members that contributed to this manual over the years, successfully designing and operating systems even before the rise of computers and spreadsheets. Amazingly, we can still employ those fundamentals to the same end today, keeping in mind the guidance of Carlson, “A difference to be a difference must be a big enough difference to make a difference,” and Coad paraphrasing Einstein, “The secret is to keep it simple… but not too simple.” Mark Hegberg
Water System Design Concepts
Study Objectives After studying the material in this chapter, you should be able to do the following: K K K K K K K K K
know what determines the load; understand the difference between closed and open systems; know the components of a hydronic system; understand heating versus cooling source devices; understand how systems meet part-load conditions; identify temperature and pressure ranges for low-, medium-, and hightemperature water systems; know what sensible, latent, and total heat loads are and how they affect design water flow; identify examples of heating and cooling load devices; and know how load diversity suggests a reduction in the total cooling capacity required.
Instructions Read Chapter 1, and answer all of the questions at the end.
Introductory Concepts Water system design depends on the designer’s ability to evaluate the space loads, occupancy patterns, and indoor environment requirements. This chapter examines the actual process of water system design and provides information on how to evaluate space loads. It also provides strategies and formulas for mastering the key requirements for water systems. Water systems that convey heat to or from a conditioned space or process with hot or chilled water are frequently called hydronic systems. In general, these systems employ centrifugal pumps to force water flow from a heating or a cooling source to the conditioned space or load by means of various piping, pumping, control, and terminal arrangements (ASHRAE 2012).
2
Chapter 1 Water System Design Concepts Given the design requirements, it is the designer’s task to evaluate the space loads resulting from building construction, weather distribution, occupancy patterns, indoor environment requirements, and other internal loads to determine the total load (Figure 1-1) subject to the local building codes. The loads include transmission, solar radiation, infiltration, ventilation air, people, lights, power, appliances, and materials in and out (Sauer and Howell 2013). The historical weather distribution for the project location is important, and a means to control the systems at part-load conditions to maintain comfort conditions for the occupants must be studied for proper design. The designer must weigh the cost of the source utilities available and also the efficiencies of boilers and chillers or other hydronic sources to determine the most efficient system design (Figure 1-2). The system must be able to operate between part-load and full-load conditions. In many cases, the hydronic system is a support system providing the heating or cooling medium for heat transfer equipment in an air distribution system. This course is intended to acquaint the student with the various hydronic principles and practices available for consideration in a project’s design concept. Figure 1-3 shows the basic components of a hydronic system that the designer must define for an HVAC system, namely, a source of heating or cooling, a distribution system, and the load components. There are different classifications of hydronic systems; the most common types are summarized below. As you familiarize yourself with the various systems, remember that different factors may come into play for each. Water systems may be closed or open types. The fundamental difference between them is the interface of the water with a compressible gas (such as air) or an elastic surface (such as a diaphragm). A closed water system is
Figure 1-1
Source/load.
Fundamentals of Water System Design I-P
Figure 1-2
Source–distribution–load.
Figure 1-3
Source–distribution–part-load.
3
defined as one with no more than one point of interface with a compressible gas (air) or surface (Figure 1-4). This definition is fundamental to understanding the hydraulic dynamics of these systems (to be discussed in the section Expansion Chamber).
4
Chapter 1 Water System Design Concepts An open system has more than one such interface. For example, a cooling tower has at least two points of interface: the tower basin and the discharge pipe or nozzles entering the tower. Comparing Figure 1-4 and Figure 1-5, the differences between the hydraulics of the systems become evident as one analyzes the two systems. However, one major difference is that certain hydraulic characteristics of open systems do not occur in closed systems. For example, in a closed system, • • •
flow cannot be motivated by static head difference, pumps do not provide static lift, and the entire piping system is always filled with water.
These factors affect the installation and operating costs of the system over its service life. Closed water systems are classified by operating temperature: •
Figure 1-4
Low-temperature water (LTW) system. This hydronic heating system operates within the pressure and temperature limits of the ASME Boiler and Pressure Vessel Code for low-pressure boilers (Figure 1-6). The maximum allowable working pressure for low-pressure boilers is 160 psig, with a maximum temperature of 250°F. The usual maximum working pressure for LTW boiler systems is 30 psig; however, boilers specifically designed, tested, and stamped for higher pressures are frequently used. Steam-towater and water-to-water heat exchangers are also used for heating LTW.
Hydronic system fundamentals (closed system).
Fundamentals of Water System Design I-P
Figure 1-5
5
•
Medium-temperature water (MTW) system. This hydronic heating system operates at temperatures between 250°F and 350°F, with pressures not exceeding 160 psig (Figure 1-7). The design supply water temperature is approximately 250°F to 325°F, with a pressure rating of 150 psig for boilers and equipment.
•
High-temperature water (HTW) system. This hydronic heating system operates at temperatures over 350°F, with pressures not exceeding 300 psig. The maximum design supply water temperature is about 400°F, with a pressure rating for boilers and equipment of 300 psig (Figure 1-7). The pressure/ temperature rating of each component should be checked for compliance with the system’s design versus the manufacturer’s rating.
Cooling tower (open system).
6
Chapter 1 Water System Design Concepts
Figure 1-6
Low-temperature water system—direct or reverse return.
Figure 1-7
Medium- or high-temperature water system.
Fundamentals of Water System Design I-P
7
•
Chilled-water (CHW) system. A hydronic cooling system normally operates with a design supply water temperature of 40°F to 55°F (usually 44°F or 45°F) with a pressure rating of 120 psig. Figure 1-8 shows a small- to medium-sized system with constant-speed pumping using three-way valves to ensure constant flow in the chiller source and balancing valves on each load for flow measurement and adjustment. Larger systems may employ two-way control valves and different chiller piping and pumping arrangements to reduce pumping power.
•
Condenser water (CW) system. Open water systems are typically used in refrigeration CW systems as once-through or cooling tower systems. Figure 1-9 shows a water-cooled condenser using city, well, or river water. The return is run higher than the condenser so that the condenser is always full of water. Water flow through the condenser is modulated by a control valve in the supply line. This is usually actuated by a condenser head pressure to maintain a constant condensing temperature with load variations.
Figure 1-10 shows two cooling tower applications to protect against low outdoor temperature conditions. Water flows to the pump from the tower basin, and the level should be above the top of the pump casing for positive prime, and piping pressure drop should be minimized.
Figure 1-8
Chilled-water system—direct return piping.
8
Chapter 1 Water System Design Concepts
Figure 1-9
Figure 1-10
Condenser open water system (once through).
Condenser cooling tower system inside reservoir or heated sump.
Fundamentals of Water System Design I-P
9
Antifreeze or brine solutions may be used for applications (process applications) that require temperatures below 40°F or for coil freeze protection. Well water systems can use supply temperatures of 60°F or higher. In addition, there are other forms of system classification that tend to blur the lines between system operating temperatures, pipe layout styles, and/or a reference to the operated system style. Four-pipe dual temperature water systems are one such example of this blur. This older example of system design combines the heating and cooling system, circulating hot and/or chilled water through separate supply and return pipes to common terminal coils. Figure 1-11 shows a very simplified piping diagram for this type of system, which would not normally be piped this way. This system operates within the pressure and temperature limits of LTW systems, with usual winter design supply water temperatures of about 100°F to 150°F and summer supply water temperatures of 40°F to 45°F. A reading of ANSI/ASHRAE/IES Standard 90.1 (ASHRAE 2013) can suggest that this type of system (especially as shown) is not allowed due to energy performance issues associated with potential mixing of hot or cold conditioned water. Of course, given enough thought and control logic, this issue may be overcome. A more common application is to have a two-pipe dual-temperature system where a common piping system serves common loads with both hot and chilled water supplied seasonally to the system. In either case, some of the issues a designer must deal with start to show themselves so that the system operates properly. Think through the components and issues like the designer!
Figure 1-11
Distribution orientation.
10
Chapter 1 Water System Design Concepts The design should consider source protection to prevent temperature shocks to the chiller or boiler on cycle changeover. The designer might consider timing mechanisms, temperature mechanisms, or combinations of both so that the two sources never operate together and that neither source gets water that is too hot or too cold. Think about heat transfer. In general, a hot-water heating coil with its higher operating-water temperatures and differential temperatures as compared to CHW requires a much smaller heat transfer coil. The water, on the other hand, does not know or care about this, it just acts as it does because there is heat transfer surface area. Therefore, it performs to whatever capabilities are inherently a part of that seasonal operation. The designer then has to carefully consider the design temperature differentials and flow rates associated with the coil selection and whether the sources are capable of handling the developed differentials. Consider how the coil will be controlled. For example, will flow be proportionally modulated or be positioned for either full or no flow? Consider whether seasonal changes in operating flow will affect the water flow velocity, possibly impacting air control and pressure management in the system. There are other issues to be considered, too. This manual teaches a framework of issues to be dealt with in design that can be adapted to deal with any type of system.
Basic System Components Figure 1-12 shows the fundamental components of a closed hydronic system. Actual systems generally have additional components (such as valves, vents, etc.), but these are not essential to the basic principles underlying the concept of the system. These fundamental components are as follows:
Figure 1-12
Closed hydronic system fundamental components.
Fundamentals of Water System Design I-P • • • • •
11
Source system Load system Pump system Distribution system Expansion chamber
Source The source is the point where heat is added in a heating system or removed from a cooling system. Ideally, the amount of energy entering or leaving the source equals the amount entering or leaving through the load system. Under steady-state conditions, the load energy and source energy are equal and opposite. In reality, energy conversion and/or transfer is not perfect and the source has an efficiency of less than 100%. Each type of source has its own efficiency characteristics as a function of load. You must consider this source efficiency in the system design process. Any device that can be used to heat or cool water under controlled conditions can be used as a source device. Sources typically function in one of two ways: • •
By converting chemical, electrical, or solar energy to heat, which is then transferred to water in the system By transferring heat from one system to another
The most common source devices for heating and cooling systems are the following: •
•
Heating-source devices: Hot-water generator or boiler, steam-to-water heat exchanger (Figure 1-13), water-to-water heat exchanger (Figure 1-13), solar collector or panels, heat recovery or salvage heat device, exhaust gas heat exchanger, incinerator heat exchanger, heat pump condenser, and airto-water heat exchanger Cooling-source devices: Electric compression chiller (Figure 1-14), thermal absorption chiller (Figure 1-14), heat pump evaporator, air-to-water heat exchanger, and water-to-water heat exchanger
A typical large CHW system with multiple chillers, various load controls, and compound pumping is shown in Figure 1-14. This system provides variable flow, constant supply temperature CHW, multiple chillers, two-way valve control, and the advantage of adding CHW storage. One design issue shown is the placement of the common pipe for the chillers. With the common pipe located at the opposite end of the chiller production section, the chillers will unload from right to left. With the common pipe in the alternative location (between the CHW production and the loads), the chillers will load and unload equally in proportion to their capacity (see Chapter 8 for further discussion).
12
Chapter 1 Water System Design Concepts
Figure 1-13
Steam-to-water or water-to-water heat exchanger.
Figure 1-14
Multiple chiller variable-flow chilled-water system.
The two primary considerations in selecting a source device are the design capacity and the part-load capability, which, when combined, define the turndown ratio. The turndown ratio, expressed in percent of design capacity, is Minimum Capacity Turndown Ratio (%) = 100 ---------------------------------------------Design Capacity
(1-1)
Fundamentals of Water System Design I-P
13
The reciprocal of the turndown ratio is sometimes used. For example, a turndown ratio of 25% may also be expressed as a turndown ratio of four. The turndown ratio has a significant effect on the successful performance of a system, and lack of consideration for this capability of the source system has been responsible for many systems that do not function properly or that excessively consume energy. The turndown ratio has a significant impact on the ultimate system design selection because operating efficiencies tend to decrease as the turndown ratio decreases. Generally the larger the boiler or chiller on a single unit basis, the more difficult it is to achieve acceptable operation efficiencies and therefore acceptable costs for the low-load portion of its seasonal operation. Seasonal operation at less than 50% load can, in the case of heating, be over 50% of its total duty hours; in northern climates, a similar ratio may occur during cooling seasons. This presents the designer with a dilemma. You must specify a system that meets the maximum requirements, but in doing that the system may operate ineffectively for most of its duty season. One approach to solving this dilemma is by using multiple sources of lower capacity (Figure 1-14). Another design consideration is the diversity of the cooling load. Carrier (1965) states diversity of cooling load results from the probable non-occurrence of part of the cooling load on a design day. Diversity factors are applied to the refrigeration capacity in large air-conditioning systems. These factors vary with location, type, and size of the application, and are based entirely on the judgment of the engineer.
The diversity factor, as a ratio of actual load/design load, can be applied to people and lighting loads in large multistory office, hotel, or apartment buildings. For example, in an office building the diversity factor for the refrigeration capacity due to people may be from 0.75 to 0.90 and due to the lighting from 0.70 to 0.85. In addition, the design engineer must consider the storage load factors for heat gain due to lighting and solar gains by glass. Specific methods of load control for CHW plants are discussed in Chapter 9.
Load The load is the point where heat flows out of or into the system from the space or process; it is the independent variable to which the remainder of the system must respond. Outward heat flow characterizes a heating system, and inward heat flow characterizes a cooling system. The quantity of heating or cooling is calculated by one of the means discussed in the following section. Typically, loads are a form of heat exchanger, either air to water or water to water.
14
Chapter 1 Water System Design Concepts
Pump System If we consider the modern era of hydronic system application to be the last 100 or so years, historically the first hydronic systems were heating systems, and they did not apply a pump to the system to circulate water. These systems were gravity fed, relying on the difference in head caused in the system by water being heated in a boiler and cooled in a radiator and pipes. These were exceptionally simple systems, circulating water with differential pressures (less than a few inches of pressure). They tended to be small in size and used a large pipe (2 to 3 in.) to circulate flows less than 1 gpm. These systems were slow to react and operation-sensitive to how well the pipe was sized. There was no real control short of turning the flow on or off. The first recorded use of a pump in a heating system appeared in 1926 in the Transactions of ASHVE, forerunner to ASHRAE (Giesecke 1926). Giesecke, a researcher and professor at Texas A&M, delivered a paper on testing pipe fittings, and an engineer from Toronto, H.H. Angus, commented that they had solved flow problems in gravity heating systems by applying a small pump to boost the water flow. A new era of hydronic system design dawned and a term that is still used today, booster pump, came into use. Pumps create flow by converting captured centrifugal force into a higher pressure of water. In essence, the pump is a differential pressure machine with the flow of water being created as the pressure moves to a point of lower pressure.
Distribution System The distribution system in a hydronic system is a series of pipes, transporting water to the location of the loads where it is required.
Expansion Chamber Recognizing that water is transported in rigid pipes and that water expands and contracts as it is heated and cooled, water must have a place to expand into to prevent damage to the distribution system. The expansion chamber is this place, and there are two types of chambers that are commonly applied. By using this device, the pressure of the water in the system can be controlled and managed so that useful work is done with the system without damage.
Design Trade-Offs Remember, improved efficiency comes with an initial installation cost penalty. As the designer, you must work out the acceptable trade-off in the initial installation cost that a customer will accept for reduced operational expenses over the life of the system. This will be easier to accomplish when backup system requirements are considered in the decision.
Fundamentals of Water System Design I-P
15
System Temperatures As the designer, you must design temperatures and temperature ranges by considering the performance and economics of the components. For example, for a cooling system that must maintain 50% rh at 75°F (Figure 1-15), the dew-point temperature is 55°F. This sets the maximum return water temperature near 55°F (60°F maximum); the lowest practical temperature for refrigeration, considering the freezing point and the economics, is about 40°F. This temperature spread then sets constraints for a CHW system. For a heating system, the maximum hot-water temperature is normally established by the ASME Boiler and Pressure Vessel Code (2015) as 250°F, and, with space temperature requirements of little over 75°F, the actual operating supply temperatures and the temperature ranges are set by the design of the load devices. Most economic considerations relating to the distribution and pumping systems favor the use of the maximum possible temperature range t.
Heat Transfer in Hydronic Systems Sensible Heating or Cooling of Air The quantity of heat entering or leaving the airstream is expressed by q = 60 a c p Q a t where cp = q =
Figure 1-15
specific heat of air, Btu/lb·°F heat transfer rate, Btu/h
Psychrometric chart chilled-water example.
(1-2)
16
Chapter 1 Water System Design Concepts Qa t a
= = =
airflow rate, cfm air temperature change, °F density of air, lb/ft3
For standard air with a density of 0.075 lb/ft3 and a specific heat of 0.24 Btu/lb·°F, this equation becomes 60 min
0.24 Btu 0.075 lb
q Btu/h = 1h
lb·°F
Qa ft3
ft3
q Btu/h = 1.08 Q a cfm t
t°F
min
(1-3)
Example 1-1 In the system shown in Figure 1-16, the air upstream of the heating coil is 60°F and the air temperature leaving the heating coil is 130°F. Given that cp = 0.24 Btu/lb·°F, Qa = 5000 cfm, and the density is 0.075 lb/ft3, calculate the heat transfer rate of the system. q = 1.08 Qa t = 1.08 Btu/h·cfm·°F q = (1.08)(5000 cfm)(130°F – 60°F) q = 378,000 Btu/h
Solution
Figure 1-16
The heat exchanger or coil must then transfer this heat to the water. The quantity of sensible heat transferred to the heated or cooled medium in a specific heat exchanger is a function of the surface area, the mean temperature difference between the water and the medium, and the overall heat transfer coefficient, which is a function of the fluid velocities, properties of the medium, geometry of the heat transfer surfaces, and other factors (Figure 1-16). It may be expressed by
Example system with heating coil.
Fundamentals of Water System Design I-P q = UA LMTD
17 (1-4)
where A = surface area, ft2 LMTD = logarithmic mean temperature difference, heated medium to water, °F Q = heat transfer rate, Btu/h U = overall coefficient of heat transfer, Btu/h·ft2·°F
Example 1-2 In Figure 1-17, assume that the coil has a U of 155 Btu/h·ft2·°F per row. The upstream temperature is 60°F, and the air temperature leaving the coil is 130°F. Water enters the heating coil at 160°F and leaves at 140°F. The duct size is 48 × 36 in. and the coil has four rows. Find the heat transfer rate (see also Figure 1-18). q = UA LMTD
Solution
t max – t min LMTD = -------------------------------t max Ln ------------ t
t max = 140 – 60 = 80 t min = 160 – 30 = 30
min
80 – 30- = 51F LMTD = ----------------80 Ln ------ 30 Next, using LMTD, find q.
Figure 1-17
Example system with heating coil.
18
Chapter 1 Water System Design Concepts
Figure 1-18
Coil LMTD example.
q = UA LMTD 2
Btu/h·ft ·F = 155 ----------------------------- 3 ft 4 ft 51F 4 Rows Row q = 379,440 Btu/h
Latent Cooling and Dehumidification of Air The quantity of heat removed from the cooled medium (Figure 1-18) when both sensible cooling and dehumidification are present is expressed by q t = Wh where qt = W = h =
(1-5)
total heat transfer rate, Btu/h mass flow rate of cooled medium, lb/h enthalpy difference of entering and leaving conditions of cooled medium, Btu/lb
Expressed for a cooling coil, this equation becomes q t = 60Q a a h where
(1-6)
Fundamentals of Water System Design I-P
Figure 1-19
19
Example system with cooling coil.
Qa a
= airflow rate, cfm = density of air, lb/ft3 For standard air, the density is 0.075 lb/ft3 and the formula reduces to q t = 4.5Q a h
(1-7)
Example 1-3
Solution
For the system shown in Figure 1-19, determine the heat transfer rate for the sensible cooling/dehumidification process, assuming entering air is 82°F, enthalpy is 38.5 Btu/lb, and discharge air is 52°F and saturated at enthalpy of 21.4 Btu/lb. qt = 4.5Qah = (4.5)(5000 cfm)(38.5 – 21.4) Btu/h = 384,750 Btu/h
Heat Transferred to or from Water The quantity of heat transferred to or from the water is a function of the flow rate, the specific heat, and the temperature drop or rise of the water as it passes through the heat exchanger. The heat transferred to or from the water is expressed by q t = mc p t where cp m qt t
= = = =
specific heat of water, Btu/lb·°F mass flow rate of water, lb/h total heat transfer rate, Btu/h water temperature increase or decrease across unit, °F
(1-8)
20
Chapter 1 Water System Design Concepts These equations are also used to express the heat-carrying capacity of the piping or distribution system of any portion of that system. In this regard, the temperature differential, sometimes called the temperature range, is established or identified. For any flow rate through the piping, qw is called the heat-carrying capacity. With water systems, it is common to express the flow rate in gallons per minute (gpm), in which case the equation becomes q t = 8.02 w c p Q w t
(1-9)
where Qw = water flow rate, gpm w = density of water, lb/ft3 For standard conditions in which density is 62.4 lb/ft3 and specific heat is 1 Btu/lb·°F, the equation becomes q t = 500Q w t
(1-10)
Equations 1-9 and 1-10 can be used to express the heat transfer across a single load or source device or any quantity of such devices connected across a piping system. In the design or diagnosis of a system, the load side may be balanced with the source side by these equations.
Example 1-4
Solution
For a single system similar to that shown in Figure 1-16, assume the heat transfer rate across the coil is 378,000 Btu/h and the t of the water supplying the coil is 20°F (water and air at standard conditions). Find the water flow rate required for the system. q t = 500Q w t qt Q w = -------------------500 t 378,000 Btu/h Q w = ----------------------------------------------------------------------- 500 Btu·min/°F·h·gal 20F Q w = 37.8 gpm
Example 1-5 Assume qt (cooling coil) = 384,750 Btu/h and t = 12°F.
Solution
384,750 Btu/h Q w = ------------------------------------------------------------------------- 500 Btu·min/F·h·gal 12F Q w = 64.1 gpm
Fundamentals of Water System Design I-P
21
Load Systems Load systems are the devices (terminal units) that convey heat from the water for heating or to the water for cooling of the space or process. Most load systems are basically water-to-air finned-coil heat exchangers or water-to-water heat exchangers. The specific configuration is usually used to describe the device. Common configurations include the following: •
•
Heating load devices: Preheat coils in central air-handling units, heating coils in central air-handling units, zone or central reheat coils, finned-tube radiation, baseboard radiation, convectors, unit heaters, fan-coil units, waterto-water heat exchangers, radiant heating panels, and snow-melting panels Cooling load devices: Coils in central units, fan-coil units (Figure 1-20), induction unit coils, radiant cooling panels, and water-to-water heat exchangers
It is at this point that further consideration of the coil selection becomes important. Figure 1-21 shows a representative coil. Typically, as air enters the coil it passes over tubes carrying the water back to the source to be reconditioned. Those tubes are arranged in a geometrical pattern selected to achieve the required heat transfer and air and water pressure drop considerations. Tubes are connected to a supply or return header, which either supplies water to the coil or directs it back to the source. The designer has a number of criteria that can be adjusted in achieving the selection, albeit they are bounded by the heat transfer characteristic and the psychrometric process the HVAC process needs to achieve. While the application of the psychrometric chart has been previ-
Figure 1-20
Fan-coil unit.
22
Chapter 1 Water System Design Concepts
Figure 1-21
Typical coil.
ously noted, it becomes useful in determining what to select for a heat transfer device. Load calculation programs are often applied by designers for calculation of the required heating or cooling, which generates a gross value for the required heating or cooling. While space and outdoor design conditions are specified by the designer, these programs do not select the required heat transfer components, because it is just too early in the deign process to do so. The load calculation program does not have the capability to evaluate how many air-handling units are required and what their respective sizes should be. The designer’s link then is to take the load calculation data and apply it to the correct heat transfer device based on the design criteria of applied equipment. This implies selecting physical equipment to perform the air-conditioning function. Often getting the equipment to perform like the theoretical equation can be daunting. For example, consider the determination of a cooling coil and its characteristics for a CHW coil per the following criteria: • • • • • • •
Office space with a sensible cooling load of 42,000 Btu/h and a total latent load of 18,000 Btu/h Space design conditions are 75°F/50% rh Outdoor design conditions are 95°F db/75°F wb Ventilation air is 30% of the total Supply air is delivered at 55°F 65% efficient draw-through blower at 5.0 in. total pressure 12°F design t CHW rise through the coil
Fundamentals of Water System Design I-P
23
In this example, we have to incorporate our basic knowledge of psychrometrics and load calculation but also account for the energy use as the unit operates, represented by the fan energy causing a temperature rise. Similarly, we could also take into account temperature rise for the airflow in the ductwork. We know the following: Sensible heating Q Q = 1.08 cfm t cfm = ---------------------- 1.08 t Fan heat energy balance Q = 1.08 cfm t 0.1175 cfm P IN P WATTS = -------------------------------------------------F P BTU = P WATTS 3.412 Energy balance 0.1175 3.412 cfm P IN 1.08 cfm t = ---------------------------------------------------------------------F 0.371 P IN 0.1175 3.412 cfm P IN t = ---------------------------------------------------------------------- = ------------------------------1.08 cfm F F
(1-11)
The process could be sketched psychrometrically per Figure 1-22. From the psychrometric chart, the desired space conditions are fixed, as is the design temperature for the supply air temperature. In order for the process to work psychrometrically, the supply air condition must be along the sensible heat ratio (SHR) line. Calculating SHR at 0.7, using either ASHRAE Psychrometric Chart No. 1 (1992) compass or other means, a line can be drawn at the SHR slope and through the space condition and the design supply air temperature, showing the entering point at about 55°F/80% rh (51.5°F wb). Blower energy must be accounted for, which is a sensible rise calculated to be about 3°F. This is also plotted and shows the coil leaving conditions to be 52°F/90% rh (about 50°F wb). Sensible heating Q = 1.08 cfm t cfm 42,000 Q = ---------------------- = --------------------------------------- 1950 cfm 1.08 t 1.08 75 – 55
Figure 1-22 Example process diagrammed on ASHRAE Psychrometric Chart.
24 Chapter 1 Water System Design Concepts
Fundamentals of Water System Design I-P
25
This allows the linear process line, sometimes called the grand sensible heat ratio (GSHR), from the coil entering condition at the mixed-air point for 30% outdoor air at 81°F/48% rh (66.5°F wb) through the leaving-air point. Extend the graphic plot of the line so that it intersects with the saturation line of the psychrometric chart; it must intersect for the process to work psychrometrically. If it does not, the process must be modified to one that does. The intersection point is known as the apparatus dew point (ADP) and has a specific function as it relates to coil selection—it helps set the entering CHW temperature for the coil. This is also a relative indicator of how well the process works. If the entering water temperature is too low, again reevaluate the process. Blower temperature rise 0.371 P IN 0.371 5 in. t = ------------------------------- = ------------------------------ = 2.9F F 0.65 Blower sensible heat Q FAN = 1.08 1480 2.9F = 4600 Btu/h The ADP is an average temperature representing the coil surface and is what allows condensation for dehumidification to occur. The ADP in this example is about 47°F (depending on either how good one’s eyesight is or how large the psychrometric chart has been printed!) The coil entering water temperature (EWT) can be estimated using the following equation: Coil EWT = 2 ADP – T WBLVG
(1-12) Coil EWT = 2 47F – 50F = 44F It appears plotted on the ASHRAE Psychrometric Chart in Figure 1-22. Coil selection does not end with the establishment of the selection data; this merely gives the required data to search out the heat transfer device from a manufacturer. In that regard, most manufacturers will provide the selection service or provide complimentary software that will find either a single coil or a series of coils that provide varying levels of performance. An example of a manufacturer’s coil selection program is shown in Figure 1-23. Coil selection is an iteration process, and rarely does a real product provide exactly the desired psychrometric process as specified by the designer. The only datum not established for selection is that of the required flow rate, and to do that the designer must set design criteria for the water-side temperature difference for the coil. The ADP calculation established an entering temperature of 44°F. For the example, we will use a moderate t for design of 12°F, implying a leaving water temperature of 56°F, just as, previously, design flow rate is calculated as follows:
Q COIL = 4.5 cfm h = 4.5 1950 31.2 – 21 = 89,500 Btu/h Q COIL 89,500 - = --------------------- = 15 gpm q gpm = -----------------------500 t W 500 12
26
Chapter 1 Water System Design Concepts A variety of options are up to the designer in coil selection. Typical tube sizes are 0.375, 0.5, and 0.625 in. in water coils. The quantity of fins per inch (FPI) of tube has a large effect on overall heat transfer as well as dehumidification. In that regard, the depth of the coil generally referred to as the rows with the quantity of tubes is equally important. All air that flows through the coil does not come into physical contact with the metal surfaces of the coil. The air that does not is called bypass air. Getting a coil to provide the required heat transfer requires playing with all of the dimensions that affect the surface area exposed to the air as it passes through the coil as well as giving the air enough time to come into contact with that metal. For CHW coils, air velocities are normally limited to a maximum of 500 fpm to reduce the potential for condensate carryover in the airstream. Velocity can be lower and often is to achieve the required latent energy removal. Table 1-1 provides guidance on coil depth and velocity to achieve dehumidification. In addition, carefully consider the quantity of FPI of the coil. While packing in the fins increases heat transfer and reduces coil
Figure 1-23
Example of manufacturer’s coil selection program.
Table 1-1
Coil Depth and Velocity to Achieve Dehumidification
Coil Load, cfm/ton
Face Velocity, fpm
Coil Rows
600
500
3
500
500
4
400
400
4
300
300
6
200
200
8
Fundamentals of Water System Design I-P
27
size, it may degrade latent energy transfer and become a maintenance problem, collecting dirt on the fin surface. (Remember, they get wet!) Generally, try to limit fin spacing to a maximum of 8 FPI in cooling and dehumidification applications. It may not always be possible, but it is a good starting point for evaluation. Note that in this selection, we are very close to the psychrometric process that we plotted on the chart. We are not exactly at 89,500 Btu/h, but we are very close. Similarly, we are at a little less than the required 52°F discharge temperature and similarly at slightly less than the ideal water-side T. All things considered though, this is not a bad first pass at selection. Air-side pressure drop is a reasonable 0.4 in., and water-side pressure drop is a very respectable 1 ft. Note though that it is 8 tubes deep by 18 tubes high, or 144 copper tubes. Coil cost is all about the quantity of metal and the labor to connect everything together, so this might be a pricey coil selection (only your coil provider knows for sure), and you may only know by selecting a few comparable coils and asking for estimates. Less easy to get will be a coil characteristic (as portrayed in Figure 1-24), which either will require data from the manufacturer for your calculation or may be given to you by them. This is a helpful graph in determining control strategy
Figure 1-24
Heat transfer versus water flow for a coil.
28
Chapter 1 Water System Design Concepts and part-load performance. It is a normalized graph (percentage heat transfer versus percentage water flow) so that it is not influenced by the magnitude of heat transfer (Btu/h versus flow), and generally it is shown in this form holding the entering air conditions to the coil as fixed. Generally, the characteristic plots retain this type of shape. It is the sensible heat curve that is important to the control valve selection and performance, with the shape typically being the influence in selection of equal percentage valve characteristics, as is discussed in Chapter 5. Recall that this is generated with a fixed entering condition, and if the same type of evaluation is performed where specific entering conditions at specific loads are examined, one might find a wide range of shapes and required flow rates. Often, it makes the shape more linear, but that is not always the case. Similarly, note how dehumidification starts at about 30% flow to the coil. That take-off point too can vary based on the geometry of the selected coil. Typically, with less depth or higher airflow velocities, that starting point happens to the right, at larger flow rates. In this particular coil selection (which is slightly different from the previous example), a water-side t of 20°F was used, and note that as flow decreases to the coil, t rises, as would be expected for any throttled condition. This needs to be checked against the chiller to make sure that entering water to the chiller does not have a maximum criteria. Another observation of this coil characteristic is that it is for throttled flow. Were the coil pumped, the entering water temperature would be controlled by throttling a percentage of the chilled water being circulated to the coil. That makes the coil characteristic more linear in performance with respect to the percentage of chilled water provided to the coil. Why do such a thing? Specifically, when there may be concern about freezing a coil in situations where that can arise or in situations where as a designer you always want to provide the full latent energy (dehumidification) heat transfer effect of the coil. Options such as pumping the coil and changing the heat transfer characteristic are some of the great tools that chilled-water systems provide in air-conditioning system design.
The Next Step In Chapter 2, you will be introduced to piping system design.
Summary Chapter 1 covered the following topics: • • • • •
What determines the load The difference between closed and open systems Components of a hydronic system Heating versus cooling source devices How systems need to meet part-load conditions
Fundamentals of Water System Design I-P • • • •
29
Temperature and pressure ranges for low-, medium-, and high-temperature water systems Sensible, latent, and total heat loads and how they affect design water flow Examples of heating and cooling load devices How load diversity suggests a reduction in total cooling capacity required
References ASHRAE. 1992. ASHRAE Psychrometric Chart No. 1. Atlanta: ASHRAE ASHRAE. 2012. ASHRAE Handbook—HVAC Systems and Equipment. Atlanta: ASHRAE. ASHRAE. 2013. ANSI/ASHRAE/IES Standard 90.1-2013, Energy Standard for Buildings Except Low-Rise Residential Buildings. Atlanta: ASHRAE. ASME. 2015. Boiler and Pressure Vessel Code. New York: ASME. Carrier. 1965. Handbook of Air Conditioning System Design, Chapter 3. New York, NY: McGraw-Hill. Giesecke, F.E. 1926. Friction of water in elbows. ASHRAE Transactions 32:303–20. New York: ASHRAE. Sauer, H., and R. Howell. 2013. Principles of Heating, Ventilating and AirConditioning, Chapter 5. Atlanta: ASHRAE.
30
Chapter 1 Water System Design Concepts
Skill Development Exercises for Chapter 1 Complete these questions by writing your answers on the worksheets at the back of this book. 1-1
Water systems that convey heat to or from a conditioned space or process with hot or chilled water are frequently called _____________________________.
1-2
What is the fundamental difference between closed and open types of water systems?
1-3
A cooling tower has at least two points of interface with air. Where are they?
1-4
What is the maximum working pressure for LTW boiler systems?
1-5
What is a CHW system? How is it different from a CW system?
1-6
What are the fundamental components of a closed hydronic system?
1-7
Explain the most common source devices for heating and cooling systems.
1-8
Explain what load means.
1-9
What factors influence the heating and cooling load requirements?
1-10
Define sensible heat transfer.
1-11
Name five heating load devices and describe how each is used in system applications.
Piping System Design
Study Objectives After completing this chapter, you should be able to do the following: K K K K K K K K
understand Bernoulli’s principle, know the three steps in design of a fluid distribution system, understand the difference between direct-return and reverse-return piping and if they can be combined, know methods to allow thermal expansion, determine pressure drop in piping, know the difference between laminar and turbulent flow and what index quantifies it, understand piping roughness factors, and know what governs pressure drop in a piping system.
Instructions Read Chapter 2 and answer all of the questions at the end.
Basic Considerations The piping system is a key component of the distribution system, and good design practice can significantly affect the performance and energy efficiency of an HVAC system. This chapter discusses the key aspects of piping system design, including pipe sizing and system design philosophy. In the design of any fluid distribution system, you must consider the following steps (Coad 1985): • • •
Establish the piping design philosophy and objectives. Size the pipes. Calculate or determine the pressure drop in the system as a whole or in various subparts or branches.
To achieve the best energy efficiency, you may need to repeat these steps several times to optimize the design.
32
Chapter 2 Piping System Design
Pressure Drop From an instructional standpoint, it is important to understand the concept of pressure drop before discussing design philosophy and sizing. In general, to direct a flow of water through a piping system, a pressure difference must be created to overcome the friction head due to the piping length, type of fittings, elevation changes, and pressure requirements at the receiving end. In Hydrodynamica, Bernoulli (1738) analyzed the flow of water through a piping system and theorized that it must obey the law of the conservation of energy, where the energy can never be created or destroyed but only transformed or directed in its flow (Figure 2-1). Many engineering texts on thermodynamics and fluid mechanics have explained this concept and developed a general energy equation for analyzing the fluid flow in a process. An energy balance is made by equating all of the energy entering the process to that leaving plus the heat added or subtracted and the work done by or on the fluid per unit of time. Bernoulli and Euler developed this concept into the well-known Bernoulli equation for the flow of an incompressible liquid with the addition of a term for head loss hL due to flow in the pipe (Euler 1750): 2
2
V P V P Z 1 + ------1 + -----1- = Z 2 + ------2 + -----2- + h L 2g 1 2g 2
Figure 2-1
where g =
acceleration due to gravity, 32.2 ft/s2
hL
head loss in feet of fluid flowing
=
Bernoulli’s theorem.
(2-1)
Fundamentals of Water System Design I-P P1, P2 =
pressure, lb/ft2
V1, V2 =
velocity, ft/s
Z1, Z2 =
elevation above or below datum, ft
1, 2 =
density, lb/ft3
33
Figure 2-2 shows an example of a piping system where two gage readings are taken, the elevation is measured, and the pipe size is the same for the entering and leaving conditions. According to the Bernoulli theorem (Equation 2-1): 2
2
V P V P Z 1 + ------1 + -----1- = Z 2 + ------2 + -----2- + h L 2g 1 2g 2 Substituting values of pressure (in feet of fluid head) into the equation and making sure units are consistent yields the following: 2
2 2 2 V 100 lb/in. 144 in. /ft 0 + ------1 + ---------------------------------------------------------------3 2g 62.4 lb/ft 2
2 2 2 V 30 lb/in. 144 in. /ft = 100 + ------2 + ------------------------------------------------------------- + h L 3 2g 62.4 lb/ft
h L = 230.8 – 169.2 = 61.6 ft
Figure 2-2
Bernoulli’s piping example.
34
Chapter 2 Piping System Design So, we see a total head loss of 61.6 ft due to the piping and fitting friction and the elevation head loss (assuming the same size pipe internal diameter at 1 and 2, V1 = V2).
Design Philosophy In a direct-return system, the length of supply and return piping through the subcircuits is unequal. This may cause unbalanced flow rates and requires careful balancing to provide each subcircuit with design flow. Ideally, a reverse-return system provides nearly equal total lengths for all terminal circuits. Will the design provide the most economical geometrical solution through vertical and/or horizontal distribution? Figure 2-3 shows a direct-return piping system, as compared to a reversereturn piping system shown in Figure 2-4. As a designer, you must make several important decisions based on design philosophy before starting the design process. Will this be a basic plan or a combination of direct and reverse systems (as shown in Figure 2-5)? Design philosophy and objectives are most often overlooked by designers. It is in this step that the why and how of the system are addressed, including other considerations such as the following:
Figure 2-3
Direct-return piping.
Fundamentals of Water System Design I-P
Figure 2-4
Reverse-return piping.
Figure 2-5
Combination of direct and reverse systems.
35
36
Chapter 2 Piping System Design •
•
• •
•
•
•
•
Is the system to be constant flow? For example, are three-way valves inserted to handle part-load conditions by reducing flow through the load while maintaining nearly constant flow through the source? Or is the flow through the load heat transfer coils to be constant while allowing the flow through the source to vary? Will the system have intermittent flow, such as on/off control to start/stop a pump for a zone or a load coil or, if this is a small system, to start/stop the distribution pumping? Is variable flow being considered? For example, two-way valves vary the flow in the load coils that result in variable flow in the source. Will the pump speeds be varied with the load? How will the variable system flow affect the flow through the source? (Variable-speed concepts and valve arrangements are discussed in Chapters 5 and 8.) In the pressure distribution of a direct-return system (Figure 2-3), the available pressure drop for a load circuit is greatest near the system pump and decreases the farther away the load is from the source pump. Care must be taken by the designer to size and select the control valves to ensure adequate flow distribution and proper close-off. In the pressure distribution of a reverse-return system (Figure 2-4), the pressure drop for a load circuit is uniform (if the load pressure drops are similar) even as the distance is increased from the source pump. A key reason for the reverse-return design is to assist the two-way control valve with a more uniform pressure drop. Selection of control valves must ensure adequate flow and proper close-off, but this is not as critical as in the direct-return design. The designer should consider balancing valves because the control valve may provide a larger flow than the design flow, and the balancing device will permit field measurement and readjustment. Is thermal expansion to be handled through geometrical offset configurations (Figure 2-6) or by using mechanical joints (Figure 2-7)? For example, steel pipe may increase its length by 0.53 in. per 100 ft if its temperature increases from 0°F to 70°F, or by 1.52 in. if from 0°F to 200°F. The system must be designed to handle thermal expansion and contraction of the piping. Failure to properly design for thermal expansion can result in piping distortion, noise, and possible system failures.
Arrangement of the piping and its suspension from the building structure must also be given specific attention by the designer if quiet operation is to be attained. Rigid attachment of the pipe to the structural members of a building, especially at midspan, provides a direct link that will transmit objectionable vibration and sound. The exception would be at preselected points of the piping that must serve as anchors to control the amount and direction of movement due to expansion and contraction. Chapter 46 of the 2012 ASHRAE Handbook—HVAC Systems and Equipment contains recommendations regarding anchor specification and spacing (ASHRAE 2012). These are examples of concerns and questions that should be answered in the early steps of the design process.
Fundamentals of Water System Design I-P
Figure 2-6
Piping expansion, offset piping.
Figure 2-7
Piping expansion, mechanical joint.
37
38
Chapter 2 Piping System Design
Sizing Piping Sizing the pipe is not to be confused with pressure drop calculations, although the pressure drop is generally used as a primary consideration in the sizing. In most fluid systems, the size of the piping is established on the basis of the friction loss per running foot of pipe. The fluid velocity is then used as a limiting selection parameter. The equation most often used that relates pressure drop, flow rate, and pipe size is the Darcy-Weisbach equation: 2
L V h f = f ---- ------ D 2g
(2-2)
where D =
pipe diameter, ft
f
=
friction factor (0.10 to 0.010)
g
=
gravitational constant, 32.2 ft/s2
hf
=
energy lost through friction, expressed as fluid head, feet of fluid flowing
L
=
pipe length, ft
V
=
average fluid velocity, ft/s
Figure 2-8 shows an experimental arrangement for determining head loss in a pipe. Fluid velocity is calculated from the continuity equation (Streeter and Wylie 1985): V = Q ---A where A =
cross-sectional area of the pipe, ft2
Q
=
flow rate, ft3/s
V
=
fluid velocity, ft/s
(2-3)
Pipe sizing is covered in Chapter 22 of the 2013 ASHRAE Handbook— Fundamentals (ASHRAE 2013). The general range of pipe friction loss used for designing hydronic systems is between ~1 and 4 ft/100 ft. For controlling the velocity noise, ASHRAE Handbook—Fundamentals suggests a velocity limit of 4 ft/s for a 2 in. pipe and smaller. A pressure drop limit of 4 ft/100 ft for pipe sizes above a 2 in. size is suggested, but this is subject to the designer’s selection. Maximum water velocity versus operation hours to minimize erosion may also be considered in the design. After a pipe size has been selected for a known pipe material, flow rate, and friction factor, the Darcy-Weisbach equation can be used directly to calculate the head loss in feet of fluid flowing.
Fundamentals of Water System Design I-P
Figure 2-8
39
Experimental arrangement for determining head loss in a pipe.
One approach to simplify the determination of pipe diameter (assuming turbulent flow, combining some of the formulas, and assuming an approximation of f = 0.03) reduces the Darcy-Weisbach equation to (Coad 1985): D = SQ where D = S = = = Q =
0.4
(2-4)
pipe diameter, in. pipe sizing constant 0.44 assuming 4 ft pressure drop per 100 ft length 0.50 assuming 2 ft pressure drop per 100 ft length flow rate, ft3/s
This simplified method is useful for quick approximations. It is important to know some further classic approaches for accuracy and if it is required to evaluate a pipe sizing program. In an experiment in 1883, Osborne Reynolds showed that fluids can flow through a pipe under two different conditions: laminar flow and turbulent flow. He demonstrated that when dye was injected in a glass pipe with low water velocities (Figure 2-9), the stream of dye stayed in layers (laminar flow) up to Re = 2000. However, as the velocity was increased, the layer of dye wavered and then broke up, diffusing with the water because of intermingling of the particles or of the water in what was turbulent flow (Re > 2000). Reynolds defined the Reynolds number (Re) as Re = DV -- where D V
= = = =
inside pipe diameter, ft average fluid velocity, ft/s dynamic viscosity, lbm/ft·s density, lb/ft3
(2-5)
40
Chapter 2 Piping System Design
Figure 2-9
Reynolds’s laminar versus turbulent flow demonstration.
The Reynolds number is a nondimensional parameter relating pipe diameter, fluid velocity, and the fluid viscous properties. Relating these fluid flow parameters enables the development of charts relating flow conditions and pipe characteristics. These charts present experimental data that can be used in pipe sizing and pipe system design. Reynolds’s study showed that the friction factor in the laminar flow range is equal to the following: 64 f = ----Re
(2-6)
Johann Nikuradse (1933) demonstrated the effect of pipe surface roughness on friction for both the laminar and turbulent regions (Figure 2-10) and defined a roughness factor, /D. Figure 2-10 can be used to determine the friction factor when the Reynolds number and the pipe roughness factor /D are known. Lewis Moody (1944) demonstrated that a transition region appears between Re of 2000 and 10,000, as shown in his diagram (Figure 2-11). The Moody diagram portrays friction factor f from 0.01 to 0.08 as a function of the relative roughness /D of the pipe or tubing and the Reynolds number DV/. The relative roughness /D can be determined from another Moody graph (Figure 2-12) portraying pipe diameter, pipe material, and relative roughness of pipe. The kinematic viscosity may also be used in the Reynolds formula since = / (ft2/s). Substituting the Reynolds number can be written as follows: V Re = D ---
(2-7)
Fundamentals of Water System Design I-P
Figure 2-10
41
Relation of Reynolds number, friction flow, and relative roughness for similar pipes. Source: Moody (1944).
Figure 2-13 is a nomogram that helps simplify determination of kinematic viscosity () and the Reynolds number (Re) for different fluids, knowing the fluid temperature in °F, pipe diameter in in., and velocity in ft/s. The 2013 ASHRAE Handbook—Fundamentals also refers to the Colebrook equation for determining the friction factor f in the turbulent flow range (ASHRAE 2013): 1 18.7 -------- = 1.74 – 2log 2 ---- + --------------- D Re f f
(2-8)
The Hazen-Williams equation is also mentioned as an alternative to the Darcy-Weisbach equation: 1.852 1 - ----------h f = 3.0221 V ---- C 1.67 D
(2-9)
where C is the pipe roughness factor. Recommended values of the C factor are as follows: 150 for plastic pipe and copper tubing, 140 for new steel pipe, 100 for steel pipe after 20 years of use, and down to 80 after 30 years or for badly corroded or very rough pipe (Streeter and Wylie 1985; Karassik et al. 1986).
Figure 2-11
Moody chart showing relationship between friction factors and Reynolds number for water flow.
Source: Moody (1944).
42 Chapter 2 Piping Systems Design
Figure 2-12
Source: Engineering Data Book, Figure IIIA-4. (HI 1990).
Friction factors and relative roughness for various pipes.
Fundamentals of Water Systems Design I-P 43
Figure 2-13
Kinematic viscosity and Reynolds number determination nomogram.
Source: 2013 ASHRAE Handbook—Fundamentals, Chapter 3, Figure 13.
44 Chapter 2 Piping Systems Design
Fundamentals of Water System Design I-P
Figure 2-14
45
Friction loss for water in commercial steel pipe (Schedule 40) and friction loss for water in copper tubing (Types K, L, M). Source: 2013 ASHRAE Handbook—Fundamentals, Chapter 22, Figures 4 and 5.
The 2013 ASHRAE Handbook—Fundamentals notes that the Darcy-Weisbach equation with friction factors from the Moody chart, the Colebrook equation, or the Hazen-Williams equation are fundamental to calculating pressure drop in hot and chilled-water piping. Charts calculated from these equations (such as Figure 2-14) show flow rates and head loss for schedule 40 steel pipe (ASHRAE 2013). The 2013 ASHRAE Handbook—Fundamentals has similar charts for K, L, and M copper tubing and schedule 80 plastic pipe. Also, tables and charts can be found in the Crane Company’s Technical Paper #410 (Crane 1988) and the Hydraulic Institute’s Engineering Data Book (1990). Reviewing the Engineering Data Book, we find another version of the Colebrook equation:
46
Chapter 2 Piping System Design 1 2.51 -------- = 2log ------------ + --------------- 3.7D Re f f
(2-10)
Be aware that this equation is used for calculation of the friction factor (f) and, consequently, also the values in the tables of friction loss (hf) in feet per 100 feet for new pipes from 1/8 to 5 in. nominal schedule 40 steel pipe and from 6 to 84 in. for schedule 40 steel and cast-iron pipe. These values will differ slightly from those found in the 2013 ASHRAE Handbook— Fundamentals. A typical table for 2 in. nominal pipe is shown as Table 2-1. In commercial installations, the tables suggest adding 15% to the friction loss to allow for aging.
Flow-Rate Measurement Taken from the 2013 ASHRAE Handbook—Fundamentals, Table 2-2 lists various means of measuring fluid flow rate. The values for volume or mass flow-rate measurement are often determined by measuring pressure difference across an orifice, nozzle, or venturi tube (ASME 1972; Benedict 1984). These types of meters have different advantages and disadvantages. For example, the orifice plate is more easily changed than the complete nozzle or venturi tube assembly. However, the nozzle is often preferred to the orifice because its discharge coefficient is more precise. The venturi tube is a nozzle followed by an expanding recovery section to reduce net pressure loss. Fluid meters use a wide variety of physical techniques to make flow measurements; those more prevalently used are described in the following section (ASME 1972; Miller 1983; DeCarlo 1984). The search for high-accuracy flow measurement includes the arrangement of appropriate calibration procedures. While these used to be available only in calibration laboratories, they are now frequently purchased along with flowmeters so that flow measurements can be efficiently and effectively ensured and validated at high levels of performance. To ensure and validate calibration facilities and procedures, realistic traceability should be established and maintained for the calibration facilities and procedures.
Direct and Indirect Flow Measurement Methods Both gas and liquid flow can be measured quite accurately by timing a collected amount of fluid that is determined gravimetrically or volumetrically. While this method is commonly used for calibrating other metering devices, it is particularly useful where the flow rate is low or intermittent and where a high degree of accuracy is required. These systems are generally large and slow, but in their simplicity they can be considered primary devices.
Fundamentals of Water System Design I-P Table 2-1
47
Friction Loss for Water in Feet for 100 ft, 2 in. Nominal Pipe Schedule 40 Schedule 40 Steel Pipe Inside Diameter = 2.067 in. /D = 0.00087
2 in. Nominal Discharge cfs
gpm
V, ft/s
V2/2g, ft
ft per 100 ft of Pipe
0.00446
2
0.191
0.00057
0.0151
0.00669
3
0.287
0.00128
0.0302
0.00892
4
0.383
0.00228
0.0497
0.0112
5
0.479
0.00356
0.0733
0.0134
6
0.574
0.00512
0.1008
0.0156
7
0.670
0.00698
0.132
0.0178
8
0.766
0.00911
0.167
0.0201
9
0.861
0.0115
0.206
0.0223
10
0.957
0.0142
0.249
0.0268
12
1.15
0.0205
0.344
0.0312
14
1.34
0.0279
0.454
0.0357
16
1.53
0.0364
0.577
0.0401
18
1.72
0.0461
0.714
0.0446
20
1.91
0.0569
0.865
0.0491
22
2.11
0.0689
1.03
0.0535
24
2.30
0.0820
1.21
0.0580
26
2.49
0.0962
1.40
0.0624
28
2.68
0.112
1.60
0.0669
30
2.87
0.128
1.82
0.0781
35
3.35
0.174
2.42
0.0892
40
3.83
0.228
3.09
0.100
45
4.31
0.288
3.85
0.112
50
4.79
0.356
4.69
0.123
55
5.26
0.431
5.61
0.134
60
5.74
0.512
6.61
0.145
65
6.22
0.601
7.69
0.156
70
6.70
0.698
8.85
0.167
75
7.18
0.801
10.1
48
Chapter 2 Piping System Design
Table 2-1
Friction Loss for Water in Feet for 100 ft, 2 in. Nominal Pipe Schedule 40 (Continued) Schedule 40 Steel Pipe Inside Diameter = 2.067 in. /D = 0.00087
2 in. Nominal Discharge cfs
gpm
V, ft/s
V2/2g, ft
ft per 100 ft of Pipe
0.178
80
7.66
0.911
11.4
0.190
85
8.14
1.03
12.8
0.201
90
8.61
1.15
14.3
0.212
95
9.09
1.28
15.8
0.223
100
9.57
1.42
17.5
0.245
110
10.53
1.72
21.0
0.268
120
11.5
2.05
24.8
0.290
130
12.4
2.41
28.9
0.312
140
13.4
2.79
33.4
0.335
150
14.4
3.20
38.1
0.357
160
15.3
3.64
43.2
0.379
170
16.3
4.11
48.6
0.401
180
17.2
4.61
54.3
0.424
190
18.2
5.14
60.3
0.446
200
19.1
5.69
66.6
0.491
220
21.1
6.89
80.2
0.535
240
23.0
8.20
95.1
0.580
260
24.9
9.62
111
0.624
280
26.8
11.16
129
0.669
300
28.7
12.81
147
0.714
320
30.6
14.6
167
0.758
340
32.5
16.5
188
0.803
360
34.5
18.4
211
0.847
380
36.4
20.6
234
0.892
400
38.3
22.8
259
Fundamentals of Water System Design I-P Table 2-2
49
Volumetric or Mass Flow-Rate Measurement
Measurement Means Application
Range
Precision
Limitations
1%–5%
Discharge coefficient and accuracy influenced by installation conditions
0.5%–2.0%
Discharge coefficient and accuracy influenced by installation conditions
Orifice and differential pressure measurement system
Flow through pipes, ducts, and plenums for all fluids
Above Reynolds number of 5000
Nozzle and differential pressure measurement system
Flow through pipes, ducts, and plenums for all fluids
Above Reynolds number of 5000
Venturi tube and differential pressure measurement system
Flow through pipes, ducts, and plenums for all fluids
Above Reynolds number of 5000
0.5%–2.0%
Discharge coefficient and accuracy influenced by installation conditions
Liquids or gases; Timing given mass or used to calibrate volumetric flow other flowmeters
Any
0.1%–0.5%
System is bulky and slow
Rotameters
Liquids or gases
Any
0.5%–5.0%
Should be calibrated for fluid being metered
Displacement meter
Relatively small volumetric flow with high pressure loss
As high as 1000 cfm, depending on type
0.1%–2.0% Most types require depending on calibration with fluid type being metered
Gasometer or volume displacement
Short-duration tests; used to calibrate other flowmeters
Total flow limited by available 0.5%–1.0% volume of containers
—
Thomas meter (temperature rise of stream caused by electrical heating)
Elaborate setup justified by need for good accuracy
Any
1%
Uniform velocity; usually used with gases
Element of resistance to flow and differential pressure measurement system
Used for check Lower limit set by where system has readable pressure calibrated drop resistance element
1%–5%
Secondary reading depends on accuracy of calibration
Turbine flowmeters
Liquids or gases
0.25%–2.0%
Uses electronic readout
Single or multipoint instrument for measuring velocity at specific point in flow
Primarily for installed airhandling systems with no special provision for flow measurement
Any
Lower limit set by accuracy of velocity 2%–10% measurement instrumentation
Accuracy depends on uniformity of flow and completeness of traverse; may be affected by disturbances near point of measurement
50
Chapter 2 Piping System Design Table 2-2
Volumetric or Mass Flow-Rate Measurement (Continued)
Measurement Means Application
Range
Precision
Limitations
Heat input and temperature changes with steam and water coil
Check value in heater or cooler tests
Any
1%–3%
—
Laminar flow element and differential pressure measurement system
Measure liquid or gas volumetric flow rate; nearly linear relationship 0.0001–2000 cfm with pressure drop; simple and easy to use
1%
Fluid must be free of dirt, oil, and other impurities that could plug meter or affect its calibration
Magnetohydrodynamic flowmeter (electromagnetic)
Measures electrically conductive fluids, slurries; meter does not obstruct flow; no moving parts
0.1–10,000 gpm
1%
At present state of the art, conductivity of fluid must be greater than 5 mho/cm
Measure liquid or Swirl flowmeter and gas flow in pipe; vortex shedding meter no moving parts
Above Reynolds number of 104
1%
—
Source: 2013 ASHRAE Handbook—Fundamentals, Table 5, Chapter 36.
The variable-area meter or rotameter is a convenient direct-reading flowmeter for liquids and gases. This is a vertical, tapered tube in which the flow rate is indicated by the position of a float suspended in the upward flow. The position of the float is determined by its buoyancy and the upwardly directed fluid drag. A velocity traverse (made using a pitot tube or other velocity-measuring instrument) measures airflow rates in the field or calibrates large nozzles. This method can be imprecise at low velocities and impractical where many test runs are in progress.
Venturi, Nozzle, and Orifice Flowmeters Flow in a pipeline can be measured by a venturi meter (Figure 2-15), flow nozzle (Figure 2-16), or orifice plate (Figure 2-17). The ASME publication MFC3M (ASME 2004) describes measurement of fluid flow in pipes using the orifice, nozzle, and venturi; ASME Performance Test Code 19.5-72 specifies their construction (ASME 1972). Assuming an incompressible fluid (liquid or slow-moving gas), uniform velocity profile, frictionless flow, and no gravitational effects, the principle of conservation of mass and energy can be applied to the venturi and nozzle geometries to give the following:
Fundamentals of Water System Design I-P
Figure 2-15
Typical Herschel-type Venturi meter.
Figure 2-16
Dimensions of ASME long-radius flow nozzles.
51
52
Chapter 2 Piping System Design
Figure 2-17
Sharp-edge orifice with pressure tap locations.
2g p 1 – p 2 w = V 1 A 1 = V 2 A 2 = A 2 -------------------------------4 1–
(2-11)
where A1 = area 1 A2
= area 2
g
=
acceleration due to gravity, 32.2 ft/s2
p1, p2 =
absolute pressure, lb/ft2
V1, V2 =
velocity, ft/s
w
=
flow rate, lb/s
=
D2/D1 for venturi and sharp-edged orifice and d/D for flow nozzle
=
density, lb/ft3
Because the flow through the meter is not frictionless, a correction factor C is deemed to account for friction losses. If the fluid is at a high temperature, an additional correction factor Fa should be included to account for thermal
Fundamentals of Water System Design I-P
53
expansion of the primary element. Because this amounts to less than 1% at 500°F, it can usually be omitted. Equation 2-11 then becomes 2g p 1 – p 2 w = C A 2 -------------------------------4 1–
(2-12)
The factor C is a function of geometry and Reynolds number. Values of C are given in ASME Performance Test Code 19.5-72 (ASME 1972). The approach factor can be combined with the discharge coefficient, as described later. The jet passing through an orifice plate contracts to a minimum area at the vena contracta, which is located a short distance downstream from the orifice plate. The contraction coefficient, energy loss coefficient, and approach factor can be combined into a single constant K, which is a function of geometry and Reynolds number. The orifice flow rate equations then become Q = K A 2 2g p 1 – p 2
(2-13)
where A2 = orifice area, ft2 g = acceleration due to gravity, 32.2 ft/s2 K = values are shown in ASME Performance Test Code 19.5-72 (ASME 1972) p1, p2 = pressure drop, lb/ft2, as obtained by pressure taps Q = flow rate, ft3/s = density, lb/ft3 Valves, bends, and fittings upstream from the flowmeter can cause errors. Long, straight pipes should be installed upstream and downstream of the flow devices to ensure fully developed flow for proper measurement. ASHRAE Standard 41.8 specifies upstream and downstream pipe lengths for measuring flow of liquids with an orifice plate (ASHRAE 1989). ASME Performance Test Code 19.5-72 gives the piping requirements between various fittings and valves and the venturi, nozzle, and orifice (ASME 1972). The 2015 ASHRAE Handbook—HVAC Applications recommends straight pipe a minimum of 15 pipe diameters upstream and 5 diameters downstream for any flow-measuring device (ASHRAE 2015).
Variable-Area Flowmeters (Rotameters) In permanent installations where high precision, ruggedness, and operational ease are important, the variable-area flowmeter is satisfactory. It is frequently used to measure liquids or gases in small-diameter pipes. However, for ducts or pipes over 6 in. diameter, the expense of this meter may not be warranted. In larger systems, the meter can be placed in a bypass line and used with an orifice. The variable-area meter (Figure 2-18) commonly consists of a float that is free to move vertically in a transparent tapered tube. The fluid to be metered enters at
54
Chapter 2 Piping System Design the narrow bottom end of the tube and moves upward, passing at some point through the annulus formed between the float and the inside wall of the tube. At any particular flow rate, the float assumes a definite position in the tube; a calibrated scale on the tube shows the float’s location and the fluid flow rate.
Turbine Flowmeters Turbine flowmeters are volumetric flow-rate-sensing meters with a magnetic stainless steel turbine rotor suspended in the flow stream of a nonmagnetic meter body. The fluid stream exerts a force on the blades of the turbine rotor, setting it in motion and converting the fluid’s linear velocity to an angular velocity. Design motivation for turbine meters is to have the rotational speed of the turbine proportional to the average fluid velocity and thus to the volume rate of fluid flow (Miller 1983; DeCarlo 1984; Mattingly 1992). The rotational speed of the rotor is monitored by an externally mounted pickoff assembly. Magnetic and radio frequency are the most commonly used pickoffs. The magnetic pickoff contains a permanent magnet and coil. As the turbine rotor
Figure 2-18
Variable-area flowmeter.
Fundamentals of Water System Design I-P
55
blades pass through the field produced by the permanent magnet, a shunting action induces alternating-current voltage in the winding of the coil wrapped around the magnet. A sine wave with a frequency proportional to the flow rate develops. With the radio frequency pickoff, an oscillator applies a high-frequency carrier signal to a coil in the pickoff assembly. The rotor blades pass through the field generated by the coil and modulate the carrier signal by shunting action on the field shape. The carrier signal is modulated at a rate corresponding to the rotor speed, which is proportional to the flow rate. With both pickoffs, the frequency of the pulses generated becomes a measure of flow rate, and the total number of pulses measures total volume (Mattingly 1992; Woodring 1969; Shafer 1961). The lubricity of the process fluid and the type and quality of rotor bearings determine whether the meter is satisfactory for the particular application. When choosing turbine flowmeters for use with fluorocarbon refrigerants, attention must be paid to the type of bearings used in the meter and to the oil content of the refrigerant. For these applications, sleeve-type rather than standard ball bearings are recommended. The amount of oil in the refrigerant can severely affect calibration and bearing life. In metering liquid fluorocarbon refrigerants, the liquid must not flash to a vapor (cavitate). This would cause a tremendous increase in flow volume. Flashing results in erroneous measurements and rotor speeds that can damage the bearings or cause a failure. Flashing can be avoided by maintaining an adequate backpressure on the downstream side of the meter (Liptak 1972).
The Next Step In Chapter 3, you will learn about pipe materials and fittings.
Summary Chapter 2 covered the following topics: • • • • • • • • • •
Bernoulli’s principle The three steps in designing a fluid distribution system The difference between direct and return piping and if they can be combined Methods to allow for thermal expansion How to determine pressure drop and sizing of piping The difference between laminar and turbulent flow, and what index quantifies it Piping roughness factors The factors that govern pressure drop in a piping system Flow-rate measurement Flow-measuring device location
56
Chapter 2 Piping System Design
References ASHRAE. 1989. Standard 41.8-1989, Standard Methods of Measurement of Flow of Liquids in Pipes Using Orifice Flowmeters. Atlanta: ASHRAE. ASHRAE. 2012. ASHRAE Handbook—HVAC Systems and Equipment. Atlanta: ASHRAE. ASHRAE. 2013. ASHRAE Handbook—Fundamentals. Atlanta: ASHRAE. ASHRAE. 2015. ASHRAE Handbook—HVAC Applications. Atlanta: ASHRAE. ASME. 1972. Application of fluid meters. ASME Performance Test Code PTC 19.5-72. New York, NY: ASME. ASME. 2004. MFC-3M-2004, Measurement of Fluid Flow in Pipes Using Orifice, Nozzle and Venturi. New York: ASME. Benedict, R. 1984. Fundamentals of Temperature, Pressure and Flow Measurements. New York: John Wiley and Sons. Bernoulli, D. 1738. Hydrodynamica. New York: Dover Publications (1968). Coad, W. 1985. Sizing of pipes & ducts. Heating, Piping, Air-Conditioning: July. Crane. 1988. Flow of fluids through valves, fittings and pipe. Technical Paper #410. Joliet, IL: Crane Co. DeCarlo, J. 1984. Fundamentals of Flow Measurement. Research Triangle Park, NC: Instrumentation Society of America. Euler, L. 1750. HI. 1990. Engineering Data Book, 2nd ed. Cleveland, OH: Hydraulic Institute. Karassik, I., J. Messina, P. Cooper, and C. Heald. 1986. Pump Handbook. New York, NY: McGraw-Hill. Liptak, B., ed. 1972. Instrument Engineers Handbook. Philadelphia: Chiton Book Co. Mattingly, G. 1992. The characterization of a piston displacement-type flowmeter calibration facility and the calibration and use of pulsed output type flowmeters. Journal of Research of the National Institute of Standards and Technology. Gaithersburg, MD: NIST. Miller, R. 1983. Measurement Engineering Handbook. New York: McGrawHill. Moody, L. 1944. Friction factors for pipe flow. ASME Transactions 66(8). Nikuradse, J. 1933. Strommung Gesetze in rouhen Rohren. Forsch. Arb. Ing.Wes. No.361. Reynolds, O. 1883. An experimental investigation of the circumstances which determine whether the motion of water shall be direct or sinuous, and of the law of resistance in parallel channels. Philosophical Transactions of the Royal Society. London. Shafer, M. 1961. Performance characteristics of turbine flowmeters. Proceedings of the ASME Winter Annual Meeting, New York, NY. Streeter, V., and E. Wylie. 1985. Fluid Mechanics. New York, NY: McGraw-Hill. Woodring, E. 1969. Magnetic turbine flowmeters. Instruments and Control Systems 6:133.
Fundamentals of Water System Design I-P
57
Skill Development Exercises for Chapter 2 Complete these questions by writing your answers on the worksheets at the back of this book. 2-1
What causes unbalanced flow rates in direct-return piping arrangements?
2-2
What is the most often used equation that relates to pressure drop?
2-3
Fluids can flow through a pipe under two different conditions. Name them. Explain the differences between these flow conditions.
2-4
Explain friction factor.
2-5
In commercial installations, it is suggested that _____% should be added to the friction loss to allow for aging.
2-6
What is the Bernoulli principle?
2-7
What factors determine pressure drop in piping?
2-8
What methods allow thermal expansion?
2-9
What is the minimum distance upstream and downstream for a water flowmeasuring device (in pipe diameters)?
Pipe Materials and Fittings
Study Objectives After completing this chapter, you should be able to K K K K
K K
name the U.S. organizations that issue codes and standards for piping systems and components; list the different types of pipe used in HVAC water system applications and describe the characteristics of each; list the pipe-joining methods commonly encountered in HVAC water systems and describe the characteristics of each; list common factors that support or promote corrosion, the five methods of corrosion control, and the two corrosion environments of particular concern to the HVAC piping system designer; describe the function, selection, and installation of backflow prevention devices; and solve a basic pipe selection problem.
Instructions Read Chapter 3 and answer all of the questions at the end.
Pipe Materials This section covers pipe materials commonly used for heating, air-conditioning, and refrigerating systems. When selecting and applying pipe, applicable local codes, state or provincial codes, and voluntary industry standards (some of which have been adopted by code jurisdictions) must be followed. The following U.S. organizations issue codes and standards for piping systems and components: • • • • •
ASME ASTM International National Fire Protection Association (NFPA) International Code Council (ICC) Manufacturers Standardization Society, Valve and Fitting Industry (MSS)
60
Chapter 3 Pipe Materials and Fittings • •
American National Standards Institute (ANSI) American Water Works Association (AWWA)
Parallel federal specifications also have been developed by government agencies for many public works projects. ANSI/ASME Standard B31.9, Building Services Piping, lists applicable American codes and standards for HVAC piping (ASME 1988). In addition, it gives the requirements for safe design and construction of piping systems for building heating and air conditioning.
Metal Pipe Steel Pipe Steel pipe is manufactured by several processes. Seamless pipe made by piercing or extruding has no longitudinal seam. Other manufacturing methods roll skelp into a cylinder and weld a longitudinal seam. A continuous-weld (CW) furnacebutt-welding process forces and joins the edges together at a high temperature. An electric current welds the seam of electric resistance welded (ERW) pipe. ASTM Standards A53 and A106 specify steel pipe (2012, 2014a). Both specify A and B grades. The A grade has a lower tensile strength and is not widely used. The ASME (2004, 2014) pressure piping codes require that a longitudinal joint efficiency factor E be applied to each type of seam when calculating the allowable stress, as listed in Table 3-1. ASME Standard B36.10 (2004) specifies the dimensional standard for steel pipe. Up to 12 in. diameter, nominal pipe sizes (NPSs) are used, which do not match the internal or external diameters. For 14 in. and larger pipe, the size corresponds to the outside diameter. Steel pipe is manufactured with wall thicknesses identified by schedule and weight. Although schedule numbers and weight designations are related, they are not constant for all pipe sizes. Standard-weight (ST) and schedule 40 pipe have the same wall thickness through 10 in. NPS. For 12 in. and larger standard-weight pipe, the wall thickness remains constant at Table 3-1
Allowable Stressesa for Pipe and Tube
Minimum Basic Allowable Joint Allowable Available ASTM Manufacturing Tensile Allowable Stress Efficiency Stressb SE, Grade Type Sizes, Specification Process Strength, Stress Rangec Factor E in. psi S, psi psi SA, psi A53 Steel
—
F
CW
1/2 to 4
45,000
11,250
0.6
6,800
16,900
A53 Steel
B
S
Seamless
1/2 to 26
60,000
15,000
1.0
15,000
22,500
A53 Steel
B
E
ERW
2 to 20
60,000
15,000
0.85
12,800
22,500
A106 Steel
B
S
Seamless
1/2 to 26
60,000
15,000
1.0
15,000
22,500
B88 Copper
—
—
Hard Drawn
1/4 to 12
36,000
9,000
1.0
9,000
13,500
a Listed stresses are for temperatures to 650°F for steel pipe (to 400°F for Type F) and to 250°F for copper tubing. b To be used for internal pressure stress calculations in Equations 1 and 2 of ASHRAE 2012, Chapter 46. c To be used only for piping flexibility calculations; see Equations 3 and 4 of ASHRAE 2012, Chapter 46.
Source: 2012 ASHRAE Handbook—HVAC Systems and Equipment, Table 1, Chapter 46.
Fundamentals of Water System Design I-P
61
0.375 in., while schedule 40 wall thicknesses increase with each size. A similar equality exists between extra strong (XS) and schedule 80 pipe up through 8 in.; afterward, XS pipe has a 0.500 in. wall, while schedule 80 increases in wall thickness. Table 3-2 lists properties of representative steel pipe. Joints in steel pipe are made by welding or using threaded, flanged, grooved, or welded-outlet fittings. Unreinforced welded-in-branch connections weaken a main pipeline, and added reinforcement is necessary unless the excess wall thicknesses of both mains and branches are sufficient to sustain the pressure. ANSI/ ASME Standard B31.1, Power Piping, gives formulas for determining when reinforcement is required (ASME 2014). Such calculations are seldom needed in HVAC applications because standard-weight pipe through 20 in. NPS at 300 psig requires no reinforcement, full-size branch connections are not recommended, and fittings such as tees and reinforced outlet fittings provide inherent reinforcement.
Copper Tube Because of their inherent corrosion resistance and ease of installation, copper and copper alloys are often used in heating, air conditioning, refrigeration, and water supply installations. There are two principal classes of copper tube. ASTM Standard B88 (ASTM 1995a) includes types K, L, M, and drain, waste, and vent (DWV) for water and drain service. ASTM Standard B280 (ASTM 1995b) specifies air conditioning and refrigeration (ACR) tube for refrigeration service. Types K, L, M, and DWV designate descending wall thicknesses for copper tube. All types have the same outside diameter for corresponding sizes. Table 3-3 lists properties of ASTM Standard B88 copper tube (ASTM 1995a). In the plumbing industry, a tube of nominal size approximates the inside diameter (ID). The heating and refrigeration trades specify copper tube by the outside diameter (OD). ACR tubing has a different set of wall thicknesses. Types K, L, and M tube may be hard-drawn or annealed (soft) temper. Copper tubing is joined with soldered or brazed wrought or cast copper capillary socket-end fittings. Table 3-4 lists the pressure temperature ratings of soldered and brazed joints. Small copper tubes are also joined by flare or compression fittings. Hard-drawn tubing has a higher allowable stress value than annealed, but if hard tubing is joined by soldering or brazing, the annealed allowable stress value should be used. Brass pipe and copper pipe are also made in steel pipe thicknesses for threading. High cost has eliminated these materials from the market, except for special applications. The heating and air-conditioning industry generally uses types L and M tubing, which have higher internal working pressure ratings than the solder joints used at fittings. Type K may be used with brazed joints for higher pressure and temperature requirements or for direct burial. Type M should be used with care where exposed to potential external damage. Copper and brass should not be used in ammonia refrigerating systems. The Special Systems section in this chapter covers other limitations on refrigerant piping.
62
Chapter 3 Pipe Materials and Fittings Table 3-2
Steel Pipe Data Working Pressurec ASTM A53 to 400°F
Cross Section
Surface Area Weight Schedule Wall Inside Nominal Pipe Number Thickness Diameter Size, OD, or t, d, Metal Flow in. in. Outside, Inside, Pipe, Water, Mfr. Joint Weighta in. in. Area, Area, psig 2 2 ft /ft ft /ft lb/ft lb/ft Process Typeb in2 in2 1/4 3/8 1/2 3/4 1 1-1/4 1-1/2 2 2-1/2 3 4 6 8
0.54 0.675 0.84 1.05 1.315 1.66 1.9 2.375 2.875 3.5 4.5 6.625 8.625
40 ST
0.088
0.364
0.141
0.095 0.125 0.104 0.424 0.045
CW
T
188
80 XS
0.119
0.302
0.141
0.079 0.157 0.072 0.535 0.031
CW
T
871
40 ST
0.091
0.493
0.177
0.129 0.167 0.191 0.567 0.083
CW
T
203
80 XS
0.126
0.423
0.177
0.111 0.217 0.141 0.738 0.061
CW
T
820
40 ST
0.109
0.622
0.22
0.163
0.25 0.304 0.85
0.131
CW
T
214
80 XS
0.147
0.546
0.22
0.143
0.32 0.234 1.087 0.101
CW
T
753
40 ST
0.113
0.824
0.275
0.216 0.333 0.533 1.13
0.231
CW
T
217
80 XS
0.154
0.742
0.275
0.194 0.433 0.432 1.47
0.187
CW
T
681
40 ST
0.133
1.049
0.344
0.275 0.494 0.864 1.68
0.374
CW
T
226
80 XS
0.179
0.957
0.344
0.251 0.639 0.719 2.17
0.311
CW
T
642
40 ST
0.14
1.38
0.435
0.361 0.669
1.5
2.27
0.647
CW
T
229
80 XS
0.191
1.278
0.435
0.335 0.881 1.28
2.99
0.555
CW
T
594
40 ST
0.145
1.61
0.497
0.421 0.799 2.04
2.72
0.881
CW
T
231
80 XS
0.2
1.5
0.497
0.393 1.068 1.77
3.63
0.765
CW
T
576
40 ST
0.154
2.067
0.622
0.541
1.07
3.36
3.65
1.45
CW
T
230
80 XS
0.218
1.939
0.622
0.508
1.48
2.95
5.02
1.28
CW
T
551
40 ST
0.203
2.469
0.753
0.646
1.7
4.79
5.79
2.07
CW
W
533
80 XS
0.276
2.323
0.753
0.608
2.25
4.24
7.66
1.83
CW
W
835
40 ST
0.216
3.068
0.916
0.803
2.23
7.39
7.57
3.2
CW
W
482
80 XS
0.3
2.9
0.916
0.759
3.02
6.6
10.25
2.86
CW
W
767
40 ST
0.237
4.026
1.178
1.054
3.17 12.73 10.78
5.51
CW
W
430
80 XS
0.337
3.826
1.178
1.002
4.41
11.5 14.97
4.98
CW
W
695
40 ST
0.28
6.065
1.734
1.588
5.58 28.89 18.96
12.5
ERW
W
696
80 XS
0.432
5.761
1.734
1.508
8.4
26.07 28.55 11.28
ERW
W
1209
30
0.277
8.071
2.258
2.113
7.26 51.16 24.68 22.14
ERW
W
526
40 ST
0.322
7.981
2.258
2.089
8.4
50.03 28.53 21.65
ERW
W
643
80 XS
0.5
7.625
2.258
1.996 12.76 45.66 43.35 19.76
ERW
W
1106
Fundamentals of Water System Design I-P Table 3-2
63
Steel Pipe Data (Continued)
Working Pressurec Cross ScheSurface Area Weight ASTM A53 Inside Section Pipe dule Wall to 400°F Nominal DiaOD, Number ThickSize, in. meter d, Metal Flow or in. ness t, in. Outside, Inside, Pipe, Water, Mfr. Joint in. a Area, Area, psig 2 2 Weight ft /ft ft /ft lb/ft lb/ft Process Typeb in2 in2 10
12
14
16 18
20
10.75
12.75
14
16 18
20
30
0.307
10.136
2.814
2.654 10.07 80.69 34.21 34.92
ERW
W
485
40 ST
0.365
10.02
2.814
2.623 11.91 78.85 40.45 34.12
ERW
W
606
XS
0.5
9.75
2.814
2.552
16.1 74.66 54.69 32.31
ERW
W
887
80
0.593
9.564
2.814
2.504 18.92 71.84 64.28 31.09
ERW
W
1081
30
0.33
12.09
3.338
3.165 12.88 114.8 43.74 49.68
ERW
W
449
ST
0.375
12
3.338
3.141 14.58 113.1 49.52 48.94
ERW
W
528
40
0.406
11.938
3.338
3.125 15.74 111.9 53.48 48.44
ERW
W
583
XS
0.5
11.75
3.338
3.076 19.24 108.4 65.37 46.92
ERW
W
748
80
0.687
11.376
3.338
2.978 26.03 101.6 88.44 43.98
ERW
W
1076
30 ST
0.375
13.25
3.665
3.469 16.05 137.9 54.53 59.67
ERW
W
481
40
0.437
13.126
3.665
3.436 18.62 135.3 63.25 58.56
ERW
W
580
XS
0.5
13
3.665
3.403 21.21 132.7 72.04 57.44
ERW
W
681
80
0.75
12.5
3.665
3.272 31.22 122.7 106.05 53.11
ERW
W
1081
30 ST
0.375
15.25
4.189
3.992 18.41 182.6 62.53 79.04
ERW
W
421
40 XS
0.5
15
4.189
3.927 24.35 176.7 82.71 76.47
ERW
W
596
ST
0.375
17.25
4.712
4.516 20.76 233.7 70.54 101.13 ERW
W
374
30
0.437
17.126
4.712
4.483 24.11 230.3 81.91 99.68
ERW
W
451
XS
0.5
17
4.712
4.45
93.38 98.22
ERW
W
530
40
0.562
16.876
4.712
4.418 30.79 223.7 104.59 96.8
ERW
W
607
20 ST
0.375
19.25
5.236
5.039 23.12 291
78.54 125.94 ERW
W
337
30 XS
0.5
19
5.236
4.974 30.63 283.5 104.05 122.69 ERW
W
477
40
0.593
18.814
5.236
4.925 36.15 278 122.82 120.3
W
581
27.49 227
ERW
a Numbers are schedule numbers per ASME Standard B36.10M; ST = standard weight, XS = extra strong. b T = thread, W = weld. c Working
pressures were calculated per ANSI/ASME B31.9 using furnace butt-weld (CW) pipe through 4 in. and electric resistance welded (ERW) thereafter (ASME 1988). The allowance A has been taken as (1) 12.5% of t for mill tolerance on pipe wall thickness, plus (2) an arbitrary corrosion allowance of 0.025 in. for pipe sizes through NPS 2 and 0.065 in. for NPS 2.5 in. through 20, plus (3) a thread cutting allowance for sizes through NPS 2. Because the pipe wall thickness of threaded standard pipe is so small after deducting the allowance A, the mechanical strength of the pipe is impaired. It is good practice to limit standard-weight threaded pipe pressure to 90 psig for steam and 125 psig for water. Source: 2012 ASHRAE Handbook—HVAC Systems and Equipment, Table 2, Chapter 40.
64
Chapter 3 Pipe Materials and Fittings Table 3-3
Copper Tube Data
Diameter Surface Area Wall Nominal Thickness Diameter, Type t, in. Outside Inside in. Outside, Inside, D, d, ft2/ft ft2/ft in. in. 1/4 K 0.035 0.375 0.305 0.098 0.08 L 0.03 0.375 0.315 0.098 0.082 3/8 K 0.049 0.5 0.402 0.131 0.105 L 0.035 0.5 0.43 0.131 0.113 M 0.025 0.5 0.45 0.131 0.118 1/2 K 0.049 0.625 0.527 0.164 0.138 L 0.04 0.625 0.545 0.164 0.143 M 0.028 0.625 0.569 0.164 0.149 5/8 K 0.049 0.75 0.652 0.196 0.171 L 0.042 0.75 0.666 0.196 0.174 3/4 K 0.065 0.875 0.745 0.229 0.195 L 0.045 0.875 0.785 0.229 0.206 M 0.032 0.875 0.811 0.229 0.212 1 K 0.065 1.125 0.995 0.295 0.26 L 0.05 1.125 1.025 0.295 0.268 M 0.035 1.125 1.055 0.295 0.276 1-1/4 K 0.065 1.375 1.245 0.36 0.326 L 0.055 1.375 1.265 0.36 0.331 M 0.042 1.375 1.291 0.36 0.338 DWV 0.04 1.375 1.295 0.36 0.339 1-1/2 K 0.072 1.625 1.481 0.425 0.388 L 0.06 1.625 1.505 0.425 0.394 M 0.049 1.625 1.527 0.425 0.4 DWV 0.042 1.625 1.541 0.425 0.403 2 K 0.083 2.125 1.959 0.556 0.513 L 0.07 2.125 1.985 0.556 0.52 M 0.058 2.125 2.009 0.556 0.526 DWV 0.042 2.125 2.041 0.556 0.534 2-1/2 K 0.095 2.625 2.435 0.687 0.637 L 0.08 2.625 2.465 0.687 0.645 M 0.065 2.625 2.495 0.687 0.653
Cross Section Metal Area, in2 0.037 0.033 0.069 0.051 0.037 0.089 0.074 0.053 0.108 0.093 0.165 0.117 0.085 0.216 0.169 0.12 0.268 0.228 0.176 0.168 0.351 0.295 0.243 0.209 0.532 0.452 0.377 0.275 0.755 0.64 0.523
Flow Area, in2 0.073 0.078 0.127 0.145 0.159 0.218 0.233 0.254 0.334 0.348 0.436 0.484 0.517 0.778 0.825 0.874 1.217 1.257 1.309 1.317 1.723 1.779 1.831 1.865 3.014 3.095 3.17 3.272 4.657 4.772 4.889
Weight
Working Pressurea,b,c ASTM B88 (1995a) to 250°F
Tube, Water, Annealed, Drawn, lb/ft lb/ft psig psig 0.145 0.126 0.269 0.198 0.145 0.344 0.285 0.203 0.418 0.362 0.641 0.455 0.328 0.839 0.654 0.464 1.037 0.884 0.682 0.65 1.361 1.143 0.94 0.809 2.063 1.751 1.459 1.065 2.926 2.479 2.026
0.032 0.034 0.055 0.063 0.069 0.094 0.101 0.11 0.144 0.151 0.189 0.209 0.224 0.336 0.357 0.378 0.527 0.544 0.566 0.57 0.745 0.77 0.792 0.807 1.304 1.339 1.372 1.416 2.015 2.065 2.116
851 730 894 638 456 715 584 409 596 511 677 469 334 527 405 284 431 365 279 265 404 337 275 236 356 300 249 180 330 278 226
1596 1368 1676 1197 855 1341 1094 766 1117 958 1270 879 625 988 760 532 808 684 522 497 758 631 516 442 668 573 467 338 619 521 423
Fundamentals of Water System Design I-P Table 3-3
Nominal Wall Diameter, Type Thickness in. t, in. 3
3-1/2
4
5
6
8
10
12
K L M DWV K L M K L M DWV K L M DWV K L M DWV K L M DWV K L M K L M
0.109 0.09 0.072 0.045 0.12 0.1 0.083 0.134 0.11 0.095 0.058 0.16 0.125 0.109 0.072 0.192 0.14 0.122 0.083 0.271 0.2 0.17 0.109 0.338 0.25 0.212 0.405 0.28 0.254
Diameter Outside D, in. 3.125 3.125 3.125 3.125 3.625 3.625 3.625 4.125 4.125 4.125 4.125 5.125 5.125 5.125 5.125 6.125 6.125 6.125 6.125 8.125 8.125 8.125 8.125 10.125 10.125 10.125 12.125 12.125 12.125
65
Copper Tube Data (Continued)
Surface Area
Inside Outside, Inside, d, in. ft2/ft ft2/ft 2.907 0.818 0.761 2.945 0.818 0.771 2.981 0.818 0.78 3.035 0.818 0.795 3.385 0.949 0.886 3.425 0.949 0.897 3.459 0.949 0.906 3.857 1.08 1.01 3.905 1.08 1.022 3.935 1.08 1.03 4.009 1.08 1.05 4.805 1.342 1.258 4.875 1.342 1.276 4.907 1.342 1.285 4.981 1.342 1.304 5.741 1.603 1.503 5.845 1.603 1.53 5.881 1.603 1.54 5.959 1.603 1.56 7.583 2.127 1.985 7.725 2.127 2.022 7.785 2.127 2.038 7.907 2.127 2.07 9.449 2.651 2.474 9.625 2.651 2.52 9.701 2.651 2.54 11.315 3.174 2.962 11.565 3.174 3.028 11.617 3.174 3.041
Cross Section Metal Area, in2 1.033 0.858 0.691 0.435 1.321 1.107 0.924 1.68 1.387 1.203 0.741 2.496 1.963 1.718 1.143 3.579 2.632 2.301 1.575 6.687 4.979 4.249 2.745 10.392 7.756 6.602 14.912 10.419 9.473
Flow Area, in2 6.637 6.812 6.979 7.234 8.999 9.213 9.397 11.684 11.977 12.161 12.623 18.133 18.665 18.911 19.486 25.886 26.832 27.164 27.889 45.162 46.869 47.6 49.104 70.123 72.76 73.913 100.554 105.046 105.993
Weight Tube, lb/ft 4.002 3.325 2.676 1.687 5.12 4.291 3.579 6.51 5.377 4.661 2.872 9.671 7.609 6.656 4.429 13.867 10.2 8.916 6.105 25.911 19.295 16.463 10.637 40.271 30.054 25.584 57.784 40.375 36.706
Water, lb/ft 2.872 2.947 3.02 3.13 3.894 3.987 4.066 5.056 5.182 5.262 5.462 7.846 8.077 8.183 8.432 11.201 11.61 11.754 12.068 19.542 20.28 20.597 21.247 30.342 31.483 31.982 43.51 45.454 45.863
Working Pressurea,b,c ASTM B88 (1995a) to 250°F Annealed, Drawn, psig psig 318 596 263 492 210 394 131 246 302 566 252 472 209 392 296 555 243 456 210 394 128 240 285 534 222 417 194 364 128 240 286 536 208 391 182 341 124 232 304 570 224 421 191 358 122 229 304 571 225 422 191 358 305 571 211 395 191 358
a When using soldered or brazed fittings, the joint determines the limiting pressure. b Working pressures were calculated using ANSI/ASME Standard B31.9 allowable stresses (ASME 1988). A 5% mill tolerance has been used on the wall
thickness. Higher tube ratings can be calculated using the allowable stress for lower temperatures. c If soldered or brazed fittings are used on hard-drawn tubing, use the annealed ratings. Full-tube allowable pressures can be used with suitably rated flare or compression-type fittings. Source: ASHRAE Handbook—HVAC Systems and Equipment, Table 3, Chapter 46, 2012.
66
Chapter 3 Pipe Materials and Fittings Table 3-4
Internal Working Pressure for Copper Tube Joints Internal Working Pressure, psi
Alloy Used for Joints
Service Temperature, °F
Water and Noncorrosive Liquids and Gasesa
Nominal Tube Size (Types K, L, M), in. 1/4 to 1 1 1/4 to 2 2 1/2 to 4 5 to 8a
Tin/leadb
50-50 solder (ASTM B32 Gr 50A [2014b])
95-5 Tin/antimonyc solder (ASTM B32 Gr 50TA [2014b])
Brazing alloys melting at or above 1000°F
Sat. Steam and Condensate
10 to 12a
1/4 to 8
100
200
175
150
130
100
—
150
150
125
100
90
70
—
200
100
90
75
70
50
—
250
85
75
50
45
40
15
100
500
400
300
270
150
—
150
400
350
275
250
150
—
200
300
250
200
180
140
—
250
200
175
150
135
110
15
100 to 200
d
d
d
d
d
—
250
300
210
170
150
150
—
350
270
190
150
150
150
120
a Solder joints are not to be used for (1) flammable or toxic gases or liquids or (2) gas, vapor, or compressed air in tubing over 4
in., unless maximum pressure is limited to 20 psig.
b Lead solders must not be used in potable-water systems. c Tin/antimony solder is allowed for potable-water supplies in some jurisdictions. d Rated pressure for up to 200°F applies to the tube being joined.
Source: ASHRAE Handbook—HVAC Systems and Equipment, Table 4, Chapter 46, 2012 and based on ANSI/ASME Standard B31.9, Building Services Piping (ASME 1988).
Ductile Iron and Cast Iron Pipe Ductile iron pipe is used for city water mains and waste drainage piping per ANSI/AWWA C150/A21.50 (AWWA 1996). These pipes use bell and spigot joints, or mechanical or flanged joints. Cast iron is not used for pressure piping and has been replaced by ductile iron pipe.
Joining Methods for Metal Pipe Threading Threading is the most commonly used method for joining small-diameter steel or brass pipe, as shown in ANSI/ASME Standard Bl.20.1 (ASME 1983). Pipe with a wall thickness less than standard weight should not be threaded. ANSI/ ASME Standard B31.5, Refrigeration Piping, limits the threading for various refrigerants and pipe sizes (ASME 1992).
Fundamentals of Water System Design I-P
67
Soldering and Brazing Copper tube is usually joined by soldering or brazing socket end fittings. Brazing materials melt at temperatures over 1000°F and produce a stronger joint than solder. Health concerns have caused many jurisdictions to ban solders containing lead or antimony for joining pipe in potable-water systems. In particular, lead-based solders must not be used for potable-water systems.
Flared and Compression Joints Flared and compression fittings can be used to join copper, steel, stainless steel, and aluminum tubing. Properly rated fittings can keep the joints as strong as the tube.
Flanges Flanges can be used for large pipes and all piping materials. They are commonly used to connect to equipment, valves, and wherever it may be necessary to open the joint to permit service or replacement of components. For steel pipe, flanges are available in pressure ratings to 2500 psig. For welded pipe, weld neck, slip-on, or socket weld connections are available. Thread-on flanges are available for threaded pipe. Flanges are generally flat faced or raised face. Flat-faced flanges with fullfaced gaskets are most often used with cast iron and materials that cannot take high bending loads. Raised-face flanges with ring gaskets are preferred with steel pipe because they facilitate increasing the sealing pressure on the gasket to help prevent leaks. Other facings (such as O-rings and ring joints) are available for special applications. All flat-faced, raised-face, and lap-joint flanges require a gasket between the mating flange surfaces. Gaskets are made from rubber, synthetic elastomers, cork, fiber, plastic, polytetrafluorethylene, metal, or a combination of these materials. The gasket must be compatible with the flowing media and the temperatures at which the system is operating.
Welding Welding steel pipe joints over 2 in. in diameter offers the following advantages: • • • •
Welded joints do not age, dry out, or deteriorate as do gasketed joints. Welded joints can accommodate greater vibration and water hammer and higher temperatures and pressures than other joints. For critical service, pipe joints can be tested by any of several nondestructive examination (NDE) methods (such as by radiography or ultrasound). Welded joints provide maximum long-term reliability.
The applicable section of the ASME Boiler and Pressure Vessel Code (2015) provides rules for welding. The code requires that all welders and welding procedure specifications (WPSs) be qualified. Separate WPSs are needed
68
Chapter 3 Pipe Materials and Fittings for different welding methods and materials. The qualifying tests and the variables requiring separate procedure specifications are set forth in Section IX of the code. The manufacturer, fabricator, or contractor is responsible for the welding procedure and welders. ASME Standard B31.9, Building Services Piping (1988), requires visual examination of welds and outlines limits of acceptability. The following welding processes are often used in the HVAC industry: • •
•
Shielded-metal arc welding (SMAW), also called stick welding. The molten weld metal is shielded by the vaporization of the electrode coating. Gas metal arc welding (GMAW), also called metal inert gas (MIG). The electrode is a continuously fed wire, which is shielded by argon or carbon dioxide gas from the welding gun nozzle. Gas tungsten arc welding (GTAW), also called tungsten inert gas or heliarc. This process uses a nonconsumable tungsten electrode surrounded by a shielding gas. The weld material may be provided from a separate noncoated rod.
Reinforced Outlet Fittings Reinforced outlet fittings are used to make branch and takeoff connections and are designed to permit welding directly to pipe without supplemental reinforcing. Fittings are available with threaded, socket, or butt-weld outlets.
Other Joints Grooved joint systems require that a shallow groove be cut or rolled into the pipe end. These joints can be used with steel, cast iron, ductile iron, and plastic pipes. A segmented clamp engages the grooves, and the seal is provided by a special gasket designed so that internal pressure tightens the seal. Some clamps are designed with clearance between tongue and groove to accommodate misalignment and thermal movements, while others are designed to limit movement and provide a rigid system. Manufacturers’ data give temperature and pressure limitations. Another form of mechanical joint consists of a sleeve slightly larger than the OD of the pipe. The pipe ends are inserted into the sleeve, and gaskets are packed into the annular space between the pipe and coupling and held in place by retainer rings. This type of joint can accept some axial misalignment, but it must be anchored or otherwise restrained to prevent axial pullout or lateral movement. Manufacturers provide pressure and temperature data. Ductile iron pipe may be furnished with a bell-spigot end adapted for caulk, gasket, and retainer ring, and mechanical or flanged joints. This joint is also not restrained.
Threaded Unions Unions allow disassembly of threaded pipe systems. Unions are three-part fittings with a mating machined seat on the two parts that thread onto the pipe ends. A threaded locking ring holds the two ends tightly together. A union also allows threaded pipe to be turned at the last joint connecting two pieces of equipment. Companion flanges (a pair) for small pipe serve the same purpose.
Fundamentals of Water System Design I-P
69
Special Systems Certain piping systems are governed by separate codes or standards, which are summarized below. Generally, any failure of the piping in these systems is dangerous to the public, so local areas have adopted laws enforcing the codes. •
•
• • •
Boiler piping: ANSI/ASME Standard B31.1 (ASME 2014) and the ASME Boiler and Pressure Vessel Code (2015) specify the piping inside the coderequired stop valves on boilers that operate up to 15 psig with steam or 160 psig with water and are limited to 250°F. Above these conditions, codes require fabricators to be certified for such work. The field or shop work must also be inspected while it is in progress by inspectors commissioned by the National Board of Boiler and Pressure Vessel Inspectors. Refrigeration piping: ASME Standard B31.5, Refrigeration Piping and Heat Transfer Components (1992), and ANSI/ASHRAE Standard 15, Safety Standard for Refrigeration Systems, cover the requirements for refrigerant piping (ASHRAE 2013). Plumbing systems: Local codes cover piping for plumbing systems. Sprinkler systems: NFPA Standard 13, Standard for the Installation of Sprinkler Systems, covers this field (2016). Fuel gas: ANSI Z223.1/NFPA 54, National Fuel Gas Code, prescribes fuel gas piping in buildings (ANSI 2015).
Plastic Pipe Plastic pipe is gaining wider usage in HVAC and plumbing systems where local building codes permit. Plastic is usually lighter in weight than metal, generally inexpensive, and corrosion-resistant. It also has a higher C factor (see Chapter 2, Equation 2-9), requiring lower pumping power and allowing smaller pipe sizes. The disadvantages of plastic pipe include the rapid loss of strength at temperatures above ambient and the high coefficient of linear expansion. The modulus of elasticity of plastics is low, resulting in short support span distances. Some jurisdictions do not allow certain plastics in buildings because of toxic products emitted under fire conditions. Plastic piping materials fall into two main categories: thermoplastic and thermoset. Thermoplastics melt and are formed by extruding or molding. They are usually used without reinforcing filaments. Thermosets are cured and cannot be reformed. They are normally used with glass fiber reinforcing filaments. Plastic piping materials include the following: • • • • • • • •
Polyvinyl chloride (PVC) Chlorinated polyvinyl chloride (CPVC) Polybutylene (PB) Polyethylene (PE) Cross-linked polyethylene (PEX) Polypropylene (PP) Acrylonitrile butadiene styrene (ABS) Polyvinylidene fluoride (PVDF)
70
Chapter 3 Pipe Materials and Fittings Thermosetting piping systems used in the HVAC industry are referred to as reinforced-thermosetting resin (RTR) and fiberglass-reinforced plastic (FRP). RTR and FRP are interchangeable and refer to pipe and fittings commonly made of fiberglass-reinforced epoxy resin, fiberglass-reinforced vinyl ester, and fiberglass-reinforced polyester. Because pipe and fittings made from epoxy resin are generally stronger and operate at higher temperatures than those made from polyester or vinylester resins, they are often used in HVAC applications. PEX is generally made from high-density polyethylene (HDPE). Crosslinked polymer bonds are formed in the manufacturing process, changing the characteristics to a thermoset. Bonds may be formed by irradiating the tube with an electron beam or using other chemical means. PEX is often associated with radiant heating or cooling systems; however, it is also applied in domestic water distribution. Upper-end HVAC design criteria limit application to systems under 180°F operating temperature. Combined with application of energy design standards that might require high system design differential temperatures, implying the use of condensing hot-water heating systems, PEX then may also be considered as a viable piping alternative where allowed by code. What makes this consideration interesting, aside from beneficial HVAC system operating characteristics of the material, is consideration of material economics. While steel piping is often applied for larger piping runs, it has not been uncommon to see piping predominantly installed with copper piping. However, copper prices have been quite volatile. During the period from 2004 to 2011, copper went through two significant increases in price at the metals market exchange level, rising from a low of about $1.50 per pound to just under $5.00 per pound. Prices for finished product, such as pipe, have risen accordingly. Compared to rigid metal piping systems, alternative methods of pipe support are required; the significant difference in material costs may make this attractive.
Allowable Stresses Both thermoplastics and thermosets have allowable stresses derived from a hydrostatic design basis stress (HDBS). The HDBS is determined by a statistical analysis of both static and cyclic stress-rupture test data as set forth in ASTM Standard D2837 (1992) for thermoplastics and ASTM Standard D2992 (1991) for glass fiber RTRs. The allowable stress, which is called the hydrostatic design stress (HDS), is obtained by multiplying the HDBS by a service factor. The HDS values recommended by some manufacturers and those allowed by ANSI/ASME B31.9 (1988) are listed in Table 3-5. The pressure design thickness for plastic pipe can be calculated using the code stress values and the following formula: pD t = --------------------2S + p
(3-1)
Type L
Drawn
ERW
5-2-2 3-5-5 4-4-5
6-3-3
355434-C
12454-B 12454-C 14333-D 23447-B
Cell No.
9,000
44,000
36,000
60,000
8,000
1,275
44,000
7,000
1,600 705
2,000 1,000
8,000 4,800
5,000 5,000 5,500
2,000
Mfr.
7,500
Tensile Strength, psi (at 73°F)
9,000
12,800
1,000 1,600 1,250
2,000 2,000 2,000 2,000 1,000 630 630 630 800
ASME B31
200
280
140 212 176
210 180
140
Mfr.
400
800
200
300
180 180 180 275
150 150 150 210 210 140 160 180 180 210
ASME B31
Upper Temperature Limit, °F
8,200
9,200
5,000
7,000
640 1,000 800 306
800
320 T-3 = 1°F. The monitoring of the zone room sensor and the zone water temperatures can be accomplished by a DDC cabinet or a dedicated microprocessor controller. Using a face bypass damper (Figure 5-24). The airflow can be throttled to meet load conditions by mixing terminal coil discharge airflow at the terminal coil and bypass airflow to maintain constant airflow. The control valve is sequenced to open before the face damper is opened. On an outdoor air application, a low-temperature detection thermostat is located on the leaving air side of the coil in the return pipe from the coil. Another concept is to open the coil valve when the outdoor air drops below 40°F. The face bypass principle can be used on the preheat coils of air-handling units to reduce the possibility of coil freeze up.
Primary/secondary with temperature sensors in bridge.
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Chapter 5 Terminal Unit Performance and Control
Figure 5-24
Terminal with face bypass control.
•
Mixing the airflow through the coil face with bypass air from the return air duct in an HVAC supply unit cooling coil (Figure 5-25). The bypass air is preconditioned return air to reduce additional moisture load from outdoor air/mixed air at partial loads.
•
Varying the supply water temperature to a heating coil for part-load conditions (Figure 5-26). Note that for a given entering air temperature, a flow of 100% at a baseline of 160°F delivers 100% heat transfer. At 180°F, 120% of the heat transfer is provided, and at 200°F, 140% heat transfer is established. In all cases, the coil maintains a relatively curved shape similar to the base coil characteristic. This can be contrasted with a similar coil operation that controls through changing entering water temperature. Flow is kept at 100% to the coil through a circulating pump, and capacity is varied as a function of entering water temperature. Note the linear control characteristic that is established in Figure 5-27.
Pressure-Independent Control Valves An alternative form of hybrid control valve that is discussed in Chapter 8 is the pressure-independent control valve (PICV). Throughout this section, the working premise of proper control valve operation and selection comes from the implied relationship of a stable pump selection. Traditionally, this means that in larger systems, a relatively flat performance curve for the pump is selected and as
Fundamentals of Water System Design I-P
Figure 5-25
Terminal with face bypass control and conditioned bypass air.
Figure 5-26
Variable supply water temperature for part-load conditions.
165
166
Chapter 5 Terminal Unit Performance and Control
Figure 5-27
Linear control characteristic.
control valves throttle, reducing flow, the relatively minor increases in system head attributed to the pump increase yield a very minor increase in flow across the throttled control valves. In variable-speed pumping systems, as flow decreases yielding an increase in system differential pressure, pump speed is reduced and greatly reduces system differential pressure. In theory, energy savings are harvested through the pump speed reduction. However, these same differential head savings may also cause zone control loop instability due to the hydraulic interactions of the control valve and pump. One method of limiting or eliminating the pump speed change as a source of system instability is to add a pressure control valve to the circuit of the temperature control valve or of a number of temperature control valves. Figure 5-28 schematically shows a fluid-system-powered regulator (a self-contained control device) that opens and closes in response to changes in system pressure. It acts to maintain a constant differential pressure available for the temperature control valve orifice. Doing this stabilizes available differential pressure for the control valve (as long as the pump differential pressure setpoint is properly controlled), which makes heat transfer control potential capability constant. In this way, if the controller were to react to a
Fundamentals of Water System Design I-P
Figure 5-28
167
Fluid-system-powered regulator.
zone disturbance, the capability to get the required flow for control is there. As a by-product of the regulator, this also stabilizes the inherent characteristic of the control valve, reducing or eliminating hydraulic interaction effects associated with control valve authority. What is making these valves more noteworthy is that manufacturing technologies and materials have been developed that allow the differential pressure regulator to be implemented in the same body as the temperature control valve, down to a small level (Figure 5-29). The valve concept has been available for a long time; however, it was implementation size and cost that generally made the concept less applicable. All things considered, these actions are good things from a control standpoint. On the other hand, the regulating device adds an extra pressure drop in the circuit, so it does use some energy (head). Most of these types of valves offer the capability to limit the maximum flow that the valve controls. In this regard, they eliminate one of the functions of a balancing valve. However, they do not allow for proportional balancing unless special control considerations are taken in the programming of the control system. This is discussed with variable-speed pumping in Chapter 8.
Variable-Speed Circulator on Coil Figure 5-30 shows another method of zone automatic temperature control and is a variation of the primary/secondary/tertiary pumping concept, applying a small circulating pump to directly control the coil (Green 1994). This can be done with either a variable-speed circulating pump or an on/off circulating pump.
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Chapter 5 Terminal Unit Performance and Control
Figure 5-29
Differential pressure regulator in same body as temperature control valve.
Figure 5-30
Variable-speed circulator on coil.
Research into the control method has shown that the temperature control can be maintained quite close to the setpoint. On the other hand, and especially when applied to hot-water heating systems, care must be exercised to prevent gravity circulation from the supply or return main through the coil. This is typically accommodated through application of a flow control check valve.
Fundamentals of Water System Design I-P
169
The Next Step In the next chapter, we will discuss expansion tanks and air elimination.
Summary In this chapter, we covered the following: • • • • • • • • • •
The variables involved with terminal control What mechanical components are suggested at a terminal How a terminal control valve is selected The four types of terminal control action (off/on position, proportional [P], proportional integral [PI], and proportional-integral-derivative [PID]) The emission characteristics of heating versus cooling terminals The types of control valve characteristics available and what works best with a hydronic coil Valve authority, rangeability, and selection The advantages and disadvantages of two-way and three-way valves The principles of primary/secondary pumping systems and different methods of control What types of control method vary the flow of air through a terminal
References ASHRAE. 2012. ASHRAE Handbook—HVAC Systems and Equipment. Atlanta: ASHRAE. ASHRAE. 2013. ANSI/ASHRAE/IES Standard 90.1-2013, Energy Standard for Buildings Except Low-Rise Residential Buildings. Atlanta: ASHRAE. ASHRAE. 2015. ASHRAE Handbook—HVAC Applications. Atlanta: ASHRAE. Bell & Gossett. 2013. Primary secondary pumping application manual. Technical Report 775A. Morton Grove, IL: Xylem Inc. Green, R.H. 1994. An air-conditioning control system using variable-speed water pumps. ASHRAE Transactions 94(1):463–70. ISA. 2008. ANSI/ISA Standard 75.02.01-2008 (IEC 60534-2-3 Mod), Control Valve Capacity Test Procedures. Research Triangle Park, NC: International Society of Automation. Landis and Gyr. n.d. Fundamentals of Control. L&G Appl. Bul. Mannion, G. 1988. High temperature rise piping design for variable volume systems: Key to chiller energy management. ASHRAE Transactions 94(2):1427–43.
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Chapter 5 Terminal Unit Performance and Control
Skill Development Exercises for Chapter 5 Complete these questions by writing your answers on the worksheets at the back of this book. 5-1
A typical fan-coil terminal requires 4 gpm. What valve Cv should be selected if a control valve is specified for a 9 ft drop?
5-2
What piping components should be specified at the terminal?
5-3
What type of control action should be considered to reduce discharge temperature cycle?
5-4
A control valve is to be selected for a 44 gpm terminal coil; coil drop is 18 ft. Select the correct size of control valve if the valve is specified for 50% of the coil drop, and the typical valve sizes and Cv available (Cv is in parentheses) are •
0.5 in. (2.5), 0.75 in. (6), 1 in. (10)
•
1.25 in. (16), 1.5 in. (20), 2 in. (36)
5-5
What control valve flow characteristic should be specified for proportional control of a hot water heating control?
5-6
An on/off thermostat controls a cabinet heater in a hallway. What valve flow characteristic should be specified?
5-7
A three-way valve is to be applied to a refrigeration condenser and cooling tower to maintain a 95°F condensing temperature. What type of three-way valve arrangement should be applied?
5-8
It is desirable to control flow in a chilled-water coil down to a minimum of 5% of design flow before close off. In addition to proper valve sizing for design flow capacity and proportional control, what else should be specified?
5-9
What should be specified in the bypass circuit of a three-way valve?
5-10
What type of control method varies airflow through a terminal coil?
5-11
Explain the difference between primary and secondary pumping systems.
5-12
Define valve authority, rangeability, and selection.
5-13
How is terminal control valve size selected?
5-14
What are the three types of terminal control action?
Expansion Tanks and Air Elimination
Study Objectives After completing this chapter, you should be able to K K K K K K K
understand the differences between open and closed systems, understand where air comes from in a hydronic system, know what maintains minimum and maximum pressures in a closed hydronic system and how they can be set, know the types of compression tanks and where they are located in a hydronic system, understand what solubility of air is in a hydronic system and what factors determine its increase, know what factors are needed to size and select a compression tank, and know where the point of no pressure change is.
Instructions Read Chapter 6 and answer all of the questions at the end.
Open and Closed Water Systems The fundamental difference between an open and a closed water system is the interface of the water with a compressible gas (such as air) or an elastic surface (such as a diaphragm). An open system (such as a cooling tower system) has at least two points of interface with the air. In a cooling tower, these points of interface are the surface of the tower basin water and the discharge pipe or nozzles entering the tower. A closed water system has only one point of interface with a compressible gas or surface, and flow may not be caused by elevation differences. Both open and closed systems demand different design considerations.
Typical Open System An open system is shown in Figure 6-1, illustrating a water-cooled refrigeration condenser employing a cooling tower to provide atmospheric cooling of the condenser water. A float-type valve is used in the tower sump to provide
172
Chapter 6 Expansion Tanks and Air Elimination
Figure 6-1
Typical open water system.
makeup water to maintain a predetermined water level. In addition, water treating and screening equipment are provided to maintain proper water flow conditions. Local building codes should be checked to ensure compliance of the system design with local requirements.
Typical Closed System A typical hydronic heating or cooling system, as shown in Figure 6-2, is fundamentally a closed system. In addition to the source of heating or cooling, the distribution pumping, and the piping arrangement, it must include a means of system pressure control. In closed systems, the objectives of system pressure control are to • • •
limit the pressure of all system equipment to its allowable working pressure; maintain minimum pressure for all normal operating temperatures, vent air, and prevent cavitation of the pump suction and boiling of system water; and accomplish these objectives with a minimum addition of new water. (Air is often admitted into closed hydronic heating or cooling systems as makeup water is introduced. The makeup water may be at a different temperature than the system water and subject to expansion or contraction as it is heated or cooled.)
Fundamentals of Water System Design I-P
Figure 6-2
173
Typical hydronic system.
Hydronic Accessories System pressure control is affected using the following hydronic accessories: • • • •
ASME-rated pressure relief valve Automatic pressure-reducing fill valve Expansion tank Manual or automatic air-venting equipment
Pressure Relief Valve A pressure relief valve is selected to limit the maximum operating pressure of the piping system and the source. The main function is to prevent danger to occupants, operating personnel, and equipment. In boilers, the relief valve size and capacity are usually recommended by the boiler manufacturer. However, the valve is sized for the pressure and thermal capacity of the boiler and in keeping with the ASME Boiler and Pressure Vessel Code (2015). Boilers are not the only source that require a relief valve; chillers also require a relief to prevent damage when the chiller is turned off and the contained water temperature rises to ambient conditions. Reliefs are meant to be used only as safety devices and should not normally operate. As such, the pressure required will account for system static pressure, operating
174
Chapter 6 Expansion Tanks and Air Elimination increase of the source thermally, and the increase in head related to the location with respect to the pump. If a relief valve is improperly sized, it may begin to relieve as a normal part of system operation, opening to relieve internal pressure. It is possible that the fluid flowing through the relief could carry particulate and suspended mineral solids that could precipitate across the valve seat and prevent its proper operation, possibly clogging the port and keeping it closed when it should open. This obviously causes a potentially dangerous problem where the weakest part of the system could rupture, possibly causing injury.
Pressure-Reducing Valve A pressure-reducing valve (PRV) is used for system fill and where required for reducing the pressure of the city main to the level required in the system. An example is shown in Figure 6-3. As can be seen, there is an adjusting screw compressing a spring, which serves as a setpoint for a simple system-powered regulator. The spring exerts force on the upper section of a diaphragm. Underneath the diaphragm is a complementary plate connected to the plug of a globe valve. Porting connects the downstream portion of the valve under the diaphragm, transferring the pressure to counteract the force of the spring and thus controlling the pressure.
Figure 6-3
Pressure-reducing valve.
Fundamentals of Water System Design I-P
175
Expansion Tank An expansion tank is a partially filled tank operating at or above atmospheric pressure and located in a water system to accommodate the volume expansion and contraction of water (Figure 6-4). Expansion tanks are of three basic configurations: •
•
Figure 6-4
Open tank. An open tank is a tank open to atmosphere. These were typically found in older residential heating systems (pre-1950s) and may have been in larger systems of a similar time period. It is not good practice to base designs on this type of system today. Compression tank. A compression tank (ASME rated) (see Figures 6-4 and 6-5) contains a captured volume of system air compressed by the expansion of water in the tank with an air/water interface (often called a
Expansion tanks.
176
Chapter 6 Expansion Tanks and Air Elimination compression tank) and is just a plain steel tank. Use of the compression tank makes the system an air management system, as air that is in the system needs to be captured at one common point and directed to the tank for the system pressure control. Automatic air vents may not be used in this system. In addition to the tank, an interface fitting (Figure 6-5) is required that directs the air into the tank while providing a restriction to prevent gravity thermal circulation of the tank fluid with the system fluid. •
Diaphragm or bladder tanks. A diaphragm or bladder expansion tank (ASME rated) is one in which a flexible membrane is located between the air and the water. Diaphragm tanks represent air elimination systems as no air needs to be captured and directed to the tank. As a result, air management devices with automatic vents may be used in as many places as required for the system. That said, it should be noted that if automatic air vents are used, they must always be placed in locations that have pressure greater than atmospheric pressure so that air is not drawn into the piping system.
In the open tank and the compression tank, air can enter the water through the interface and can affect system performance over long periods of time.
Figure 6-5
Closed tank.
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177
As defined, a closed system should have only one expansion tank. The presence of more than one tank can cause the closed system to behave in unexpected ways and can cause damage from water hammer or shock waves. An expansion tank is required to serve thermal and hydraulic functions: •
Figure 6-6
In its thermal function, the expansion tank provides a space into which the noncompressible liquid can expand or contract as the liquid undergoes volumetric changes due to temperature. For example, at 40°F, 1 lb of water occupies 0.01602 ft3, and at 220°F it occupies 0.01677 ft3. Figure 6-6 shows a graph of the expansion of water above 40°F.
Expansion of water above 40°F. Source: ASHRAE Handbook—HVAC Systems and Equipment (1987).
178
Chapter 6 Expansion Tanks and Air Elimination •
As a hydraulic device, the expansion tank serves as the reference pressure point in the system, which is analogous to a ground in an electrical system. Where the tank connects to the piping, the pressure equals the pressure of the air in the tank plus or minus any fluid pressure due to the elevation difference between the tank liquid surface and the pipe (see Figure 6-7). The point of tank interface often is referred to as a point of no pressure change in a system (see Figure 6-8), making the pump suction the preferred point of installation for the tank. In this way essentially all of the pump energy is added to the system pressure, which is set via the tank. While this tank location is preferred, it is not a requirement. Tank placement can be
Figure 6-7
Tank pressure related to system pressure.
Figure 6-8
Point of no pressure change.
Fundamentals of Water System Design I-P
179
managed as part of the design to manage the size of the required tank. For example, in a chilled-water (CHW) system, it is not uncommon to install the tank at the top of the piping system. This can reduce the size and expense of the device.
Air Elimination Excessive amounts of undissolved air in a piping system due to improper venting or removal can make the system operate poorly. It has been shown that air will separate from water at low velocities, and it is recommended that a minimum velocity of 1.5 to 2 ft/s in 2 in. pipe be used to reduce this risk (ASHRAE 2013a). This leaves the designer with the choice of implementing either the air management style of pressure control system, using a steel compression tank and appropriate fitting, or the air elimination style of system, using a diaphragm expansion tank. When air management is selected, a common point of collection is selected for the placement of an air separator. The air separator directs the air-entrained fluid to the compression tank and tank fitting so that air collects in the tank. While terminals are vented manually on start-up to remove large amounts of excess air, after flow and fill are established, these are kept closed so that air in solution is directed towards the tank. Automatic air vents should not be used in the system. When air elimination is selected, air separators may be located anywhere and should be installed with automatic air vents Figure 6-9 shows the internal operating mechanism for an example automatic air vent. There are many varieties of construction. Figure 6-10 shows the installation of automatic or manual air vents in the system zones or coils for small pipe sizes. It is suggested to place the vent on the leaving side of a coil or other heat transfer device and ahead of the control valve (on the return) to enable constant pressure on the vent.
Air Separation On the larger supply or return mains, a dynamic air separator (see Figure 6-11) is used to allow the air to separate centrifugally from the water. The lighterweight air mixture is collected by a perforated collector tube in the center and flows up to the closed compression tank, displacing water downward (Figure 6-12). Some versions combine a strainer and blowdown connection in the lower flange. In chiller applications, the air separator, expansion tank, and distribution pump may be located on the return main pumping into the chiller. The warmer return water releases air more readily than the CHW. An air separator is also used with a diaphragm tank and large-capacity automatic air vent, as shown in Figure 6-13. The air vent should be chosen with adequate capacity because the diaphragm tank stores the expanding water volume but not the released air.
180
Chapter 6 Expansion Tanks and Air Elimination
Figure 6-9
Figure 6-10
Internal operating mechanism for automatic air vent.
Automatic or manual air vents in system zones or coils for small pipe sizes.
Fundamentals of Water System Design I-P
Figure 6-11
Dynamic air separator.
Figure 6-12
Air separator.
181
182
Chapter 6 Expansion Tanks and Air Elimination Both Figure 6-13 for diaphragm expansion tanks and Figure 6-14 for compression tanks show the accessories that are required and are important to system operation: •
Relief valve protects for maximum system pressure.
•
PRV is set to provide minimum system pressure.
•
Service valve permits maintenance.
•
When using a compression tank, a manual three-way hand valve permits draining excess water from the expansion tank to provide proper air cushion in the tank (if it is not part of the tank fitting). When the cushion is lost, the system does not permit expansion and the relief valve will open, dumping water down the drain. The PRV fill valve may add water and system pressure may cycle, causing fill-dump-fill, etc.
Figure 6-13
Diaphragm expansion tank.
Figure 6-14
Compression tank.
Fundamentals of Water System Design I-P
183
° A glass sight gage might be added to the expansion tank to indicate water level and air cushion and may be required by local code. In some cases, this is required by code. This should be carefully considered, though, as the compression seals used to join the glass tube tank fittings must be maintained so that the joint does not become a point of air or water leakage, which would then defeat the purpose of having the tank. •
Air separator should have a blowdown valve to drain collected sediment to the sewer.
Other considerations related to the tank connection include the following (refer to Figure 6-2):
Figure 6-15
•
A tank with an air/water interface is generally used with an air control system that continually reverts air into the tank and should be located in a place where air can best be released from the circulating pump.
•
Within reason, the lower the overall pressure in a tank, the smaller the tank (as will be seen in the tank size calculations); thus, in a vertical system, the higher the tank is placed, the smaller it can be.
•
In primary/secondary or compound pumping systems, the tank should only be at one location, preferably on the suction side of the source pump (Figure 6-15).
Tank location for primary/secondary or compound pumping systems.
184
Chapter 6 Expansion Tanks and Air Elimination
Sizing Expansion Tanks With a single tank on a system, and assuming isothermal conditions for the air, the air pressure changes as a result of displacement by the water. Assuming no water is being added or removed from the system, the only thing to cause water to move in or out of the tank is the expansion or shrinkage of the water on the system. Thus, in sizing the tank, thermal expansion is related to the pressure extremes of the air in the tank (as will be seen in the tank calculations from the examples). The connection point of the tank to the system is very important and should be based on the pressure requirements of the system, remembering that the pressure at the tank connection will not change as the pump is turned on or off. This is also commonly called the point of no pressure change. For example, consider a system containing an expansion tank at 30 psig pressure and a pump with a head of 23.1 ft or 10 psig, as shown in Figures 6-16a and 6-16b. Alternative locations for connecting the expansion tank are shown. In either case, with the pump off the pressure will be 30 psig on both the pump suction and the discharge. •
•
With the tank connected to the pump suction side (Figure 6-16a), the pressure increases on the pump discharge by an amount equal to the pump pressure (e.g., 30 + 10 = 40 psig). With the tank connected to the pump discharge side (Figure 6-16b), the pressure decreases on the suction side by the same amount (e.g., 30 – 10 = 20 psig), and pressure at the pump discharge remains at 30 psi (point of no pressure change).
Sizing the tank is the primary thermal consideration in incorporating a tank into a system. However, prior to sizing the tank, the control or elimination of air must be considered. The amount of air that will be absorbed and can be held in solution with the water is expressed by Henry’s equation (ASHRAE 2012): Px = ---H where H =
Henry’s constant
P
=
absolute pressure, psia
x
=
solubility of air in water, percentage by volume
(6-1)
Henry’s constant varies with temperature and pressure. As shown in Figure 6-17, it is only constant at a given temperature. While difficult to determine molecularly, this concept of air absorption has been proven in test. The solubility in percentage air by volume can be seen for system temperatures and pressures in Figure 6-18. For example, if a system is at
Fundamentals of Water System Design I-P
185
(a)
(b)
Figure 6-16
Pressure effects of alternative tank locations: (a) pump suction side and (b) pump discharge side.
186
Chapter 6 Expansion Tanks and Air Elimination
Figure 6-17
Henry’s constant versus temperature for air and water.
50 psig and 140°F, the water can contain about 6% air by volume; if the pressure is dropped to 40 psig at the same temperature, the water contains about 5% air by volume, or 1% by volume will be released and must be vented or work its way into the compression tank. The equation for determining the size of a closed compression tank in a closed hydronic system is v----2- – 1 – 3t v 1 V t = V s ------------------------------------------P P a a ------ – ------ P P 1 2 where Pa =
atmospheric pressure, psia
P1
lower-temperature pressure, psia
=
(6-2a)
Fundamentals of Water System Design I-P
Figure 6-18
Solubility versus temperature and pressure for air-water solutions.
P2
=
higher-temperature pressure, psia
t1
=
lower temperature, °F
t2
=
higher temperature, °F
v1
=
specific volume of water (low temperature), ft3/1b
v2
=
specific volume of water (high temperature), ft3/1b
Vs
=
volume of water in system, gal
Vt
=
volume of expansion tank, gal
=
linear coefficient of thermal expansion, in/in·°F ( = 6.5 × 10–6 in/in·°F [steel]) ( = 9.5 × 10–6 in/in·°F [copper])
t
=
(t2 – t1), °F
187
188
Chapter 6 Expansion Tanks and Air Elimination Vs is typically taken from tables such as Table 6-1 (for steel pipe and copper tube); v1 and v2 are found in Table 3 of Chapter 1 of the 2013 ASHRAE Handbook—Fundamentals (ASHRAE 2013a). Similarly, the equation for determining the size of a diaphragm compression tank in a closed hydronic system is v----2- – 1 – 3t v 1 V t = V s ------------------------------------------P 2 1 – ------ P 1 Table 6-1
Volume of Water in Standard Pipe and Tube Standard Steel Pipe
Nominal Pipe Size
(6-2b)
Inside Diameter
Type L Copper Tube
Volume
Inside Diameter
Volume
in.
(mm)
Schedule No.
in.
(mm)
gal/ft
(L/m)
in.
(mm)
gal/ft
(L/m)
3/8
(10)
—
—
—
—
—
0.430
(10.9)
0.0075
(0.09)
1/2
(15)
40
0.622
(15.8)
0.0157
(0.19)
0.545
(13.8)
0.0121
(0.15)
5/8
(16)
—
—
—
—
—
0.666
(16.9)
0.0181
(0.22)
3/4
(20)
40
0.824
(20.9)
0.0277
(0.34)
0.785
(19.9)
0.0251
(0.31)
1
(25)
40
1.049
(26.6)
0.0449
(0.56)
1.025
(26.0)
0.0429
(0.53)
1 1/4
(32)
40
1.380
(35.0)
0.0779
(0.97)
1.265
(32.1)
0.0653
(0.81)
1 1/2
(40)
40
1.610
(40.9)
0.106
(1.32)
1.505
(38.2)
0.0924
(1.15)
2
(50)
40
2.067
(52.5)
0.174
(2.16)
1.985
(50.4)
0.161
(2.00)
2 1/2
(65)
40
2.469
(62.7)
0.249
(3.09)
2.465
(62.6)
0.248
(3.08)
3
(80)
40
3.068
(77.9)
0.384
(4.77)
2.945
(74.8)
0.354
(4.40)
3 1/2
(90)
40
3.548
(90.1)
0.514
(6.38)
3.425
(87.0)
0.479
(5.95)
4
(100)
40
4.026
(102.3)
0.661
(8.21)
3.905
(99.2)
0.622
(7.73)
5
(125)
40
5.047
(128.2)
1.04
(12.92)
4.875
(123.8)
0.970
(12.05)
6
(150)
40
6.065
(154.1)
1.50
(18.63)
5.845
(148.5)
1.39
(17.26)
8
(200)
30
8.071
(205.0)
2.66
(33.03)
7.725
(196.2)
2.43
(30.18)
10
(250)
30
10.136
(257.5)
4.19
(52.04)
9.625
(244.5)
3.78
(46.95)
12
(300)
30
12.090
(307.1)
5.96
(74.02) 11.565 (293.8)
5.46
(67.81)
Source: ASHRAE Pocket Guide for Air Conditioning, Heating, Ventilation, and Refrigeration (2013), Chapter 3, Table 3.7.
Fundamentals of Water System Design I-P
189
Example 6-1 Size an expansion tank for a heating system that will operate at 180°F to 220°F. For the design, it is given that the minimum pressure is 10 psig (24.7 psia) and the maximum pressure is 25 psig (39.7 psia). Atmospheric pressure is assumed to be 14.7 psia. The steel piping in the system has an estimated volume of 3000 gal and will have minimum water fill temperature of 40°F.
Solution
From the problem stated, we know v1 = 0.01602 ft3/lb at 40°F (assumed lower temperature) v2 = 0.01677 ft3/lb at 220°F (highest temperature per design) Vs = 3000 gal Therefore, applying the equation, v----2- – 1 – 3t v 1 V t = V s ------------------------------------------P P a a ------ – ------ P P 1 2 –6 0.01677 ------------------- – 1 – 3 6.5 10 220 – 40 0.01603 V t = 3000 -------------------------------------------------------------------------------------------------------14.7 14.7 ---------- – ---------24.7 39.7
V t = 578 gal
Example 6-2 Size a diaphragm tank for a heating system that will operate at 180°F to 220°F. The system will operate with a minimum pressure of 10 psig (24.7 psia) and a maximum pressure of 25 psig (39.7 psia). Atmospheric pressure is 14.7 psia. The steel piping in the system has an estimated volume of 3000 gallons and will have a minimum water-fill temperature of 40°F.
Solution
We have been given Vs = 3000 gal. Applying Equation 6.2 for a diaphragm tank: v----2- – 1 – 3t v 1 V t = V s ------------------------------------------P 2 1 – -----P 1
190
Chapter 6 Expansion Tanks and Air Elimination –6 0.01677 ------------------- – 1 – 3 6.5 10 220 – 40 0.01602 V t = 3000 -------------------------------------------------------------------------------------------------------24.7 1 – ---------- 39.7
V t = 344 gal
Pressure and Temperature Considerations Selection of the lower and higher pressures P1 and P2 is critical in these determinations. Pipe, tubing, boilers, chillers, and coils must be evaluated to determine the total system volume (see Figure 6-19). The lower temperature for a heating system is normal ambient temperature at fill conditions (for example, 40°F to 50°F) and the higher temperature is the operating supply water temperature for the system. For a CHW system, the lower temperature is the design CHW supply and the higher temperature is the ambient (86°F to 95°F). For a hot/CHW system, the lower is the CHW design temperature and the higher is the heating water design supply temperature. The specific volume data (v1 and v2) are found in the 2013 ASHRAE Handbook—Fundamentals, Chapter 6, Table 1 (ASHRAE 2013a).
Figure 6-19
Flowchart for sizing expansion tanks.
Fundamentals of Water System Design I-P
191
At the tank connection point, the pressure in closed tank systems increases as the water temperature increases. Pressures at the expansion tank are generally set by the following parameters: • •
The lower pressure is usually selected to hold a positive pressure at the highest point in the system (set on the water PRV fill valve). The higher pressure is normally set by the safety relief valve selected for the maximum pressure allowable at the location of the valve (without opening the valve). Other considerations are to ensure the following:
•
•
The pressure at any point will not drop below the saturation pressure at the operating temperature (again refer to the 2013 ASHRAE Handbook— Fundamentals, Chapter 6, Table 1 [ASHRAE 2013a]). All pumps have adequate net positive suction head (NPSH) available to prevent cavitation.
The Next Step In the next chapter, you will cover piping system development.
Summary In this chapter, we covered the following: • • • • • • •
The differences between open and closed systems How air enters a hydronic system How minimum and maximum pressures are maintained and set in a closed hydronic system The types of compression tanks and where they are located in a hydronic system What solubility of air in a hydronic system is and what factors determine its increase The factors that are needed to size and select a compression tank Where the point of no pressure change is
References ASHRAE. 1987. ASHRAE Handbook—HVAC Systems and Applications. Atlanta: ASHRAE. ASHRAE. 2012. ASHRAE Handbook—HVAC Systems and Equipment. Atlanta: ASHRAE. ASHRAE. 2013a. ASHRAE Handbook—Fundamentals. Atlanta: ASHRAE. ASHRAE. 2013b. ASHRAE Pocket Guide for Air Conditioning, Heating, Ventilation, and Refrigeration. Atlanta: ASHRAE. ASME. 2015. Boiler and Pressure Vessel Code. New York: ASME.
192
Chapter 6 Expansion Tanks and Air Elimination
Skill Development Exercises for Chapter 6 Complete these questions by writing your answers on the worksheets at the back of this book. 6-1
What maintains the maximum and minimum pressure limits of a hydronic system?
6-2
What must a closed water system have to permit the expansion and contraction of the water volume?
6-3
What should be specified for terminal coil returns and high points in the piping system to provide service for air in the hydronic system?
6-4
Where does air come from in a hydronic system?
6-5
What routine tasks should be performed by the building maintenance staff to the hydronic systems prior to the heating and the cooling seasons?
6-6
How much air can be present in water at 120°F and 30 psig?
6-7
A steel pipe system has 1000 gal total volume and will operate periodically in the cooling season with 40°F water, but when it is off it will reach 100°F ambient; minimum pressure is 10 psig (24.7 psia) and maximum is 25 psig (39.7 psia). What size diaphragm expansion tank is required?
6-8
A pump is selected for 100 gpm at 50 ft head, the system has 20 psig static pressure when off, and the expansion tank is improperly located on the pump discharge. What will the gages read on the pump suction and discharge when the pump is started? What will happen at the float-type air vents in the boiler room?
6-9
Explain what factors are needed to size and select an expansion tank.
6-10
What are the differences between open and closed systems?
6-11
What are the three types of expansion tanks?
6-12
Give three examples of where expansion tanks are used.
6-13
Why is it important to have the correct size expansion tank?
Piping System Development
Study Objectives After studying the material in this chapter you should be able to K K K K K K K
know the factors to consider before starting a piping design, understand what piping equipment should be considered at a load coil, know which piping system design gives more uniform pressure drop and why, understand how flow can be determined in a given loop, know what type of pump curve should be selected if two-way valves are to be used, know the two advantages of primary/secondary hydronic systems, and know how the possibility of freezing coils and piping in an HVAC system can be reduced.
Instructions Read Chapter 7 and answer all of the questions at the end.
Piping System Design To develop the piping system approach, the designer must consider several factors to arrive at a project solution. In short, there is no one way to satisfy all the conditions; there are many ways. The purpose of this course is to help identify possible design methods and to help select those options that best meet design and energy efficiency requirements. Looking at a project, a designer must consider some logical steps and questions to develop a design philosophy: 1. Know the building heat transfer load. In the initial approach, all aspects of the load must be known. What is the total load? How is the load distributed by time and location? How are controlled occupancy zones determined or laid out? As the system designer, some of these calculations may be up to you or may be determined by how the space to be conditioned is pro-
194
Chapter 7 Piping System Development grammed. For example, in a simple office system, there may be a ring of private offices surrounded by open floor space or vice versa. Each office probably represents a separate point of temperature control for the occupant, pointing the designer towards individual heating or cooling coils, while the open areas present a gross overall area to condition. Often the space occupancy program logically determines where the water system loads (e.g., the terminal heat transfer coils) and the source (e.g., boilers and chillers) are to be located. There are several decisions to be made to determine the best design approach depending on the building layout and type, whether it is a single-story or multistory building and whether it is a single building or a complex of buildings (such as a campus). Should there be a mechanical room, or will the equipment be distributed throughout the structure? Where should the mechanical rooms be located? How will the piping and ductwork be routed? All of these factors influence the design decisions that must be made to specify equipment location. These design observations may also lead to modification of the space planning program, indicating the importance of being involved in the initial project conceptual phases, and also indicating the potential iterative nature of a design. Be prepared for changes. Construct some of your initial design concepts with flexibility in mind, allowing for balance between first costs, design modification for the inevitable changes that will occur up to and during construction, and changes to occupancy that will require building modification.
Figure 7-1
Typical building layout.
Fundamentals of Water System Design I-P
195
2. Determine the heating and cooling loads based upon occupancy, comfort requirements, codes, and standards (see ANSI/ASHRAE/IES Standard 90.1 and 90.1 User’s Manual), and determine any special requirements for facilities like computer rooms, laboratories, and clean rooms (see Figure 7-2). 3. Develop a concept for part-load control: • • • • • • • •
Are large zones or individual rooms to be controlled? Are part-load heat transfer conditions to be obtained by varying water flow, water temperature, or airflow and temperature? Will fan-coil terminals or large air-handling units (AHUs) be considered? Will the system satisfy the full-load conditions for all building zones or must the designer consider some diversity factors? How is the source of cooling or heating to be operated at part-load conditions? Are control valves to be selected with adequate characteristics and pressure drop to provide good coil control at various loads? Will the piping system design consider methods to measure, balance, and adjust flows in each zone to ensure performance at various loads? Will backup be required? If so, which zones will be included?
4. Develop the piping and pumping system concept (see Figure 7-3), such as the following: • • •
Figure 7-2
Direct-return piping Reverse-return piping Primary/secondary piping
Determine the loads and consult references.
196
Chapter 7 Piping System Development
Figure 7-3
Develop piping/pumping system concept.
•
Combinations of the above
•
Constant-speed pumping
•
Two-speed pumping
•
Multiple pumps
•
Primary/secondary pumps
•
Variable-speed pumping
•
Distributed pumping
Consider modeling the system to determine the full- and part-load flows, the pressure distribution required, and this effect on components. 5. Develop a first-cost analysis versus energy operating costs over the projected life of the system. 6. Determine the maintenance and operating requirements and if they match with the personnel capabilities. The piping system design may not address all of these issues, but the issues need to be identified by the designer in the system layout and the project specifications.
Fundamentals of Water System Design I-P
197
The flowchart in Figure 7-4 summarizes the key steps in the piping system design process. In the section Introductory Concepts in Chapter 1, we introduced the basic direct- and reverse-return piping concepts. In the section Basic Considerations in Chapter 2, pressure drops and pipe sizing were discussed.
Direct-Return Analysis Now, examine a model of a system requiring four AHUs with loads of 100 gpm each at locations 100 ft apart in a square floor area configuration (see Figure 7-5). At each load, the design is planned to control the load flow with two-way control valves and to determine the effect that 10.7 and 5.3 ft drop coils have on the piping design for direct-return and reverse-return layouts.
Figure 7-4
Piping system design flowchart.
198
Chapter 7 Piping System Development
Figure 7-5
System requiring four AHUs, each with 100 gpm load.
Each load coil should have a manual shutoff valve for servicing, a flow strainer on its supply, an automatic air vent at the coil return, a two-way control valve (one size smaller than the branch pipe) on the leaving side of the coil, and a manual balancing valve with test ports and shutoff (line size) after the control valve (see Figure 7-6). In Figure 7-7, for a 100 gpm flow condition, 3 in. pipe has a head loss over 2 ft/100 ft at 4+ fps velocity; a 2.5 in. pipe has 7 ft/100 ft head loss at 7 fps, so a 3 in. pipe is a good starting point. The general guideline for velocity is between 1 to 4 fps. Air separates from the water below 1.5 fps, and above 4 to 5 fps noise might be noticed. Similarly, 200 gpm in 4 in. pipe is over 2 ft/100 ft (5 fps), 300 gpm in 4 in. pipe is about 5 ft/100 ft (7.5 fps), and 400 gpm in 5 in. pipe is nearly 3 ft/100 ft (6.5 fps). These drops will be checked against the values in the Hydraulic Institute’s Engineering Data Book (see Tables 7-1 to 7-3) (HI 1990). It is important to note that the selection of pipe size and the resulting velocity is based on the designer’s judgment and experience. For the example, assume 4 in. pipe for main supply and return connections and 3 in. pipe for the rest of the system. To assess each unit’s piping for pressure drop from the main to the coil see Table 7-4. Similarly, we check units 2, 3, and 4 in the model (Figure 7-5) and find they look close, so 2 ft for supply and 9 ft for return will be assumed for this example. Now look at the system pressure drops for a direct-return piping layout (in Figure 7-8) using 10.7 ft drop coils and the control valve selection of 5.27 ft drop (assuming one size smaller than branch size). For direct return, see Table 7-5. The highest pressure drop path is unit 3, with a differential of 29.66 ft required at A-F to provide 100 gpm flow.
Fundamentals of Water System Design I-P
Figure 7-6
Coil connections.
Figure 7-7
Friction loss, schedule 40 steel pipe.
199
Because the other units in the model have lower pressure drops, this means they will have a greater flow than design unless balanced. The unbalanced branches can be estimated by the use of the Darcy-Weisbach relationship, where Q2 ~ h. Q2 = Unit 1 =
2 h Q 1 --------2h 1
2 29.66 ft 100 ------------------- = 112.1 23.06 ft
Unit 2 =
2 29.66 ft 100 ------------------- = 103.2 27.84 ft
Unit 4 =
2 29.66 ft 100 ------------------- = 109.2 24.88 ft
200
Chapter 7 Piping System Development Table 7-1
Schedule 40 Steel Pipe (3 in. Nominal Discharge) Schedule 40 Steel Pipe ID = 3.066 in. D = 0.000587
3 in. Nominal Discharge ft3/s
gpm
V, ft/s
V2/2g, ft
Pressure Drop, ft per 100 ft of pipe
0.0111
5
0.217
0.000732
0.0112
0.0223
10
0.434
0.00293
0.0372
0.0334
15
0.651
0.0659
0.0762
0.0446
20
0.868
0.0117
0.126
0.0557
25
1.085
0.0183
0.189
0.0668
30
1.3
0.0263
0.262
0.078
35
1.52
0.0359
0.347
0.0891
40
1.74
0.0468
0.443
0.1
45
1.95
0.0593
0.547
0.111
50
2.17
0.0732
0.662
0.123
55
2.39
0.0885
0.789
0.134
60
2.6
0.105
0.924
0.145
65
2.82
0.124
1.07
0.156
70
3.04
0.143
1.22
0.167
75
3.25
0.165
1.39
0.178
80
3.47
0.187
1.57
0.189
85
3.69
0.211
1.76
0.201
90
3.91
0.237
1.96
0.212
95
4.12
0.264
2.17
0.223
100
4.34
0.2927
2.39
0.245
110
4.77
0.354
2.86
0.267
120
5.21
0.421
3.37
0.29
130
5.64
0.495
3.92
0.312
140
6.08
0.574
4.51
0.334
150
6.51
0.659
5.14
0.356
160
6.94
0.749
5.81
0.379
170
7.38
0.846
6.53
0.401
180
7.81
0.948
7.28
Fundamentals of Water System Design I-P Table 7-1
201
Schedule 40 Steel Pipe (3 in. Nominal Discharge) (Continued) Schedule 40 Steel Pipe ID = 3.066 in. /D = 0.000587
3 in. Nominal Discharge ft3/s
gpm
V, ft/s
V2/2g, ft
Pressure Drop, ft per 100 ft of pipe
0.423
190
8.25
1.06
8.07
0.446
200
8.68
1.17
8.9
0.49
220
9.55
1.42
10.7
0.535
240
10.4
1.69
12.6
0.579
260
11.3
1.98
14.7
0.624
280
12.2
2.29
16.9
0.668
300
13
2.63
19.2
0.713
320
13.9
3
22
0.758
340
14.8
3.38
24.8
0.802
360
15.6
3.79
27.7
0.847
380
16.5
4.23
30.7
0.891
400
17.4
4.68
33.9
0.936
420
18.2
5.16
37.3
0.98
440
19.1
5.67
40.9
1.025
460
20
6.19
44.6
1.069
480
20.8
6.74
48.5
1.114
500
21.7
7.32
52.5
1.225
550
23.9
8.85
63.2
1.337
600
26
10.5
74.8
1.448
650
26.2
12.4
87.5
1.56
700
30.4
14.3
101
1.671
750
32.5
16.5
116
1.782
800
34.7
18.7
131
1.894
850
36.9
21.1
148
2.005
900
39.1
23.7
165
2.117
950
41.2
26.4
184
2.228
1000
43.4
29.27
204
202
Chapter 7 Piping System Development Table 7-2
Schedule 40 Steel Pipe (4 in. Nominal Discharge) Schedule 40 Steel Pipe ID = 4.0286 in. /D = 0.0004
4 in. Nominal Discharge ft3/s
gpm
V, ft/s
V2/2g, ft
Pressure Drop, ft per 100 ft of pipe
0.0111
5
0.126
0.000247
0.0031
0.0223
10
0.252
0.000987
0.01017
0.0446
20
0.504
0.00395
0.0344
0.0668
30
0.756
0.00888
0.0702
0.0891
40
1.01
0.0158
0.118
0.111
50
1.26
0.0247
0.176
0.134
60
1.51
0.0355
0.245
0.156
70
1.76
0.0484
0.325
0.178
80
2.02
0.0632
0.415
0.201
90
2.27
0.08
0.515
0.223
100
2.52
0.0987
0.624
0.245
110
2.77
0.199
0.744
0.267
120
3.02
0.142
0.877
0.29
130
3.28
0.167
1.017
0.312
140
3.53
0.193
1.165
0.334
150
3.78
0.222
1.32
0.356
160
4.03
0.253
1.49
0.379
170
4.28
0.285
1.67
0.401
180
4.54
0.32
1.86
0.423
190
4.79
0.356
2.06
0.446
200
5.04
0.395
2.27
0.49
220
5.54
0.478
2.72
0.535
240
6.05
0.569
3.21
0.579
260
6.55
0.667
3.74
0.624
280
7.06
0.774
4.3
0.668
300
7.56
0.888
4.89
0.713
320
8.06
1.01
5.51
0.758
340
8.57
1.14
6.19
Fundamentals of Water System Design I-P Table 7-2
203
Schedule 40 Steel Pipe (4 in. Nominal Discharge) (Continued) Schedule 40 Steel Pipe ID = 4.0286 in. /D = 0.0004
4 in. Nominal Discharge ft3/s
gpm
V, ft/s
V2/2g, ft
Pressure Drop, ft per 100 ft of pipe
0.802
360
9.07
1.28
6.92
0.847
380
9.58
1.43
7.68
0.891
400
10.1
1.58
8.47
0.936
420
10.6
1.74
9.3
0.98
440
11.1
1.91
10.2
1.025
460
11.6
2.09
11.1
1.069
480
12.1
2.27
12
1.114
500
12.6
2.47
13
1.225
550
13.9
2.99
15.7
1.337
600
15.1
3.55
18.6
1.448
650
16.4
4.17
21.7
1.56
700
17.6
4.84
25
1.671
750
18.9
5.55
28.6
1.782
800
20.2
6.32
32.4
1.894
850
21.4
7.13
36.5
2.005
900
22.7
8
40.8
2.117
950
23.9
8.91
45.3
2.228
1000
25.2
9.87
50.2
2.451
1100
27.7
11.9
60.5
2.674
1200
30.2
14.2
72
2.896
1300
32.8
16.7
84.3
3.119
1400
35.3
19.3
97.6
3.342
1500
37.8
22.2
112
3.565
1600
40.3
25.3
127
3.788
1700
42.8
28.5
143
4.01
1800
45.4
32
160
4.233
1900
47.9
35.6
178
4.456
2000
50.4
39.5
196
204
Chapter 7 Piping System Development Table 7-3
Schedule 40 Steel Pipe (5 in. Nominal Discharge) Schedule 40 Steel Pipe ID = 5.047 in. /D = 0.000357
5 in. Nominal Discharge ft3/s
gpm
V, ft/s
V2/2g, ft
Pressure Drop, ft per 100 ft of pipe
0.0111
5
0.0802
0.0000999
0.00107
0.0223
10
0.16
0.0004
0.00348
0.0446
20
0.321
0.0016
0.0116
0.0668
30
0.481
0.0036
0.0237
0.0891
40
0.641
0.00639
0.0395
0.111
50
0.802
0.00999
0.0587
0.134
60
0.962
0.0144
0.0814
0.156
70
1.12
0.0196
0.1076
0.178
80
1.28
0.0256
0.137
0.201
90
1.44
0.0324
0.169
0.223
100
1.6
0.04
0.204
0.267
120
1.92
0.0576
0.286
0.312
140
2.25
0.0783
0.38
0.356
160
2.57
0.102
0.487
0.401
180
2.89
0.129
0.606
0.446
200
3.21
0.16
0.736
0.49
220
3.53
0.193
0.879
0.535
240
3.85
0.23
1.035
0.579
260
4.17
0.27
1.2
0.624
280
4.49
0.313
1.38
0.668
300
4.81
0.36
1.58
0.713
320
5.13
0.409
1.78
0.758
340
5.45
0.462
2
0.802
360
5.77
0.518
2.22
0.847
380
6.09
0.577
2.46
0.891
400
6.41
0.639
2.72
0.936
420
6.74
0.705
2.98
0.98
440
7.06
0.774
3.26
Fundamentals of Water System Design I-P Table 7-3
205
Schedule 40 Steel Pipe (5 in. Nominal Discharge) (Continued) Schedule 40 Steel Pipe ID = 5.047 in. /D = 0.000357
5 in. Nominal Discharge ft3/s
gpm
V, ft/s
V2/2g, ft
Pressure Drop, ft per 100 ft of pipe
1.02
460
7.38
0.846
3.55
1.07
480
7.7
0.921
3.85
1.11
500
8.02
0.999
4.16
1.23
550
8.82
1.21
4.98
1.34
600
9.62
1.44
5.88
1.45
650
10.4
1.69
6.87
1.56
700
11.2
1.96
7.93
1.67
750
12
2.25
9.05
1.78
800
12.8
2.56
10.22
1.89
850
13.6
2.89
11.5
2.01
900
14.4
3.24
12.9
2.12
950
15.2
3.61
14.3
2.23
1000
16
4
15.8
2.45
1100
17.6
4.84
19
2.67
1200
19.2
5.76
22.5
2.9
1300
20.8
6.75
26.3
3.12
1400
22.5
7.83
30.4
3.34
1500
24.1
8.99
34.8
3.56
1600
25.7
10.2
39.5
3.79
1700
27.3
11.6
44.5
4.01
1800
38.8
12.9
49.7
4.23
1900
30.5
14.4
55.2
4.46
2000
32.1
16
61
4.68
2100
33.7
17.6
67.1
4.9
2200
35.3
19.3
73.5
5.12
2300
36.9
21.1
80.1
5.35
2400
38.5
23
87
5.57
2500
40.1
25
94.2
206
Chapter 7 Piping System Development
Table 7-4
Unit 1 Supply and Return Sides
Length, ft
Unit 1: Supply Side (100 gpm)
Pressure Drop, ft
1 in.
4 in. tee branch
0.4
1 in.
4 × 3 in. bushing
0.1
1 in.
3 in. butterfly valve
0.44
3 in.
3 in. ells @ 0.1
0.3
1 in.
3 in. strainer
0.35
25 in.
3 in. pipe (2.39 ft/100 ft)
0.5975
Supply branch
2.1875
(report as 2.0 ft) Length, ft
Unit 1: Return Side (100 gpm)
Pressure Drop, ft
1 in.
4 × 3 in. tee branch
0.4
1 in.
4 × 3 in. bushing
0.02
3 in.
3 in. ells @ 0.1
0.3
1 in.
2½ in. control valve
5.27
2 in.
3 × 2½ in. bushing at 0.02
0.04
1 in.
3 in. balancing and service valve (open)
1.9
25 in.
3 in. pipe (2.39 ft/100 ft)
0.5975
Return branch
8.5275
(report as 9 ft)
In summary, for direct-return piping and 10.7 ft coils: Unit 1 = 112 gpm Unit 2 =
103 gpm
Unit 3 =
100 gpm
Unit 4 =
109 gpm
Total =
424 gpm
This example shows the effect of the uneven flow in a direct-return piping example without balancing. If a lower coil pressure drop is selected in the example (by changing the coil design), say 5.3 ft, the unbalance gets worse. The calculations are performed as seen in Table 7-6. As you can see, unit 3 is still the highest pressure drop path, with 24.26 ft required at A-F to provide it with a design flow of 100 gpm. Again applying the Darcy-Weisbach relationship Q2, the overflows can be approximated:
Fundamentals of Water System Design I-P
Figure 7-8
207
Direct-return piping layout.
Q2 =
2 h Q 1 --------2h 1
Unit 1 =
2 24.26 ft 100 ------------------- = 117.2 17.66 ft
Unit 2 =
2 24.26 ft 100 ------------------- = 104 22.44 ft
Unit 4 =
2 24.26 ft 100 ------------------- = 111.6 19.48 ft
In summary, for direct-return piping and 5.3 ft coils: Unit 1 = 117.2 gpm Unit 2 = 104.0 gpm Unit 3 = 100.0 gpm Unit 4 = 111.6 gpm Total = 432.8 gpm The summary shows a greater flow unbalance in the direct-return piping due to the lower coil drop.
208
Chapter 7 Piping System Development Table 7-5
Length, ft 30
30
Direct Return for Units 1, 2, 3, and 4
Unit 1 (Path A-B-B'-F): Supply Main Supply main: 4 in. pipe, 200 gpm = 2.27 ft/100 ft
Pressure Drop, ft 0.681
Supply branch
2
Coil (100 gpm)
10.7
Return branch
9
Return main: 4 in. pipe, 200 gpm = 2.27 ft/100 ft
0.681 Total 23.06
Length, ft
Unit 2 (Path A-B-C-C'-B'-F)
Pressure Drop, ft
30
Supply main: 4 in. pipe, 200 gpm = 2.27 ft/100 ft
0.681
100
Supply main: 3 in. pipe, 100 gpm = 2.39 ft/100 ft
2.39
Supply branch
2
Coil (100 gpm)
10.7
Return branch
9
100
Return main: 3 in. pipe, 100 gpm = 2.39 ft/100 ft
2.39
30
Return main: 4 in. pipe, 200 gpm = 2.27 ft/100 ft
0.681 Total 27.84
Length, ft
Unit 3 (Path A-E-D-D'-E'-F)
Pressure Drop, ft
70
Supply main: 4 in. pipe, 200 gpm = 2.27 ft/100 ft
1.589
100
Supply main: 3 in. pipe, 100 gpm = 2.39 ft/100 ft
2.39
Supply branch
2
Coil (100 gpm)
10.7
Return branch
9
100
Return main: 3 in. pipe, 100 gpm = 2.39 ft/100 ft
2.39
70
Return main: 4 in. pipe, 200 gpm = 2.27 ft/100 ft
1.589 Total 29.66
Length, ft 70
70
Unit 4 (Path A-E-E'-F) Supply main: 4 in. pipe, 200 gpm = 2.27 ft/100 ft
Pressure Drop, ft 1.589
Supply branch
2
Coil (100 gpm)
10.7
Return branch
9
Return main: 4 in. pipe, 200 gpm = 2.27 ft/100 ft
1.589 Total 24.88
Fundamentals of Water System Design I-P Table 7-6 Length, ft 30
30
Lower Coil Pressure Drop Selected for Units 1, 2, 3, and 4
Unit 1 (Path A-B-B'-F): Supply Main Supply main: 4 in. pipe, 200 gpm = 2.27 ft/100 ft
Pressure Drop, ft 0.681
Supply branch
2
Coil (100 gpm)
5.3
Return branch
9
Return main: 4 in. pipe, 200 gpm = 2.27 ft/100 ft
0.681 Total 17.66
Length, ft
Unit 2 (Path A-B-C-C'-B'-F)
Pressure Drop, ft
30
Supply main: 4 in. pipe, 200 gpm = 2.27 ft/100 ft
0.681
100
Supply main: 3 in. pipe, 100 gpm = 2.39 ft/100 ft
2.39
Supply branch
2
Coil (100 gpm)
5.3
Return branch
9
100
Return main: 3 in. pipe, 100 gpm = 2.39 ft/100 ft
2.39
30
Return main: 4 in. pipe, 200 gpm = 2.27 ft/100 ft
0.681 Total 22.44
Length, ft
Unit 3 (Path A-E-D-D'-E'-F)
Pressure Drop, ft
70
Supply main: 4 in. pipe, 200 gpm = 2.27 ft/100 ft
1.589
100
Supply main: 3 in. pipe, 100 gpm = 2.39 ft/100 ft
2.39
Supply branch
2
Coil (100 gpm)
5.3
Return branch
9
100
Return main: 3 in. pipe, 100 gpm = 2.39 ft/100 ft
2.39
70
Return main: 4 in. pipe, 200 gpm = 2.27 ft/100 ft
1.589 Total 24.26
Length, ft 70
70
Unit 4 (Path A-E-E'-F) Supply main: 4 in. pipe, 200 gpm = 2.27 ft/100 ft
Pressure Drop, ft 1.589
Supply branch
2
Coil (100 gpm)
5.3
Return branch
9
Return main: 4 in. pipe, 200 gpm = 2.27 ft/100 ft
1.589 Total 19.48
209
210
Chapter 7 Piping System Development
Reverse-Return Analysis Let’s look at a reverse-return piping arrangement for the same example, per Figure 7-9. We will assume the supply and return branches are also 2.0 ft and 9.0 ft. As you can see in Table 7-7, units 2 and 3 are the highest pressure-drop paths, with 36.47 ft required at A-F to provide them with a design flow of 100 gpm each. Again applying the Darcy-Weisbach relationship Q2 ~ h , we can estimate the overflow in the other coils: 2 h Q 1 --------2h 1 In summary, for reverse-return piping and 10.7 ft coils:
Q2 =
Unit 1 = Unit 2 =
100
Unit 3 =
100
Unit 4 =
Figure 7-9
2 36.47 ft 100 -------------------- = 103.6 34.0 ft
2 36.47 ft 100 -------------------- = 103.6 34.0 ft
Reverse-return piping layout.
Fundamentals of Water System Design I-P Table 7-7 Length, ft 30
100 100 100 70 Length, ft 30 100
100 100 70 Length, ft 30 100 100
100 70 Length, ft
Reverse-Return Piping Arrangement for Units 1, 2, 3, and 4
Unit 1 (Path A-B-B'-F): Supply Main Supply main: 5 in. pipe, 400 gpm at 2.72 ft/100 ft Supply branch Coil (100 gpm) Return branch Return main: 3 in. pipe, 100 gpm at 2.39 ft/100 ft Return main: 4 in. pipe, 200 gpm at 2.27 ft/100 ft Return main: 4 in. pipe, 300 gpm at 4.89 ft/100 ft Return main: 5 in. pipe, 400 gpm at 2.72 ft/100 ft Unit 2 (Path A-B-C-C'-D'-E'-F) Supply main: 5 in. pipe, 400 gpm at 2.72 ft/100 ft Supply main: 4 in. pipe, 300 gpm at 4.89 ft/100 ft Supply branch Coil (100 gpm) Return branch Return main: 4 in. pipe, 200 gpm at 2.27 ft/100 ft Return main: 4 in. pipe, 300 gpm at 4.89 ft/100 ft Return main: 5 in. pipe, 400 gpm at 2.72 ft/100 ft Unit 3 (Path A-B-C-D-D'-E'-F) Supply main: 5 in. pipe, 400 gpm at 2.72 ft/100 ft Supply main: 4 in. pipe, 300 gpm at 4.89 ft/100 ft Supply main: 4 in. pipe, 200 gpm at 2.27 ft/100 ft Supply branch Coil (100 gpm) Return branch Return main: 4 in. pipe, 300 gpm at 4.89 ft/100 ft Return main: 5 in. pipe, 400 gpm at 2.72 ft/100 ft Unit 4 (Path A-B-C-D-E-E'-F)
30
Supply main: 5 in. pipe, 400 gpm at 2.72 ft/100 ft
100
Supply main: 4 in. pipe, 300 gpm at 4.89 ft/100 ft
100
Supply main: 4 in. pipe, 200 gpm at 2.27 ft/100 ft
100
Supply main: 3 in. pipe, 100 gpm at 2.39 ft/100 ft Supply branch Coil (100 gpm) Return branch
70
Return main: 5 in. pipe, 400 gpm at 2.72 ft/100 ft
Pressure Drop, ft 0.82 2 10.7 9 2.39 2.27 4.89 1.90 Total 34.0 Pressure Drop, ft 0.82 4.89 2 10.7 9 2.27 4.89 1.90 Total 36.5 Pressure Drop, ft 0.82 4.89 2.27 2 10.7 9 4.89 1.90 Total 36.5 Pressure Drop, ft 0.82 4.89 2.27 2.39 2 10.7 9 Total 34.0
211
212
Chapter 7 Piping System Development Total =
407.2 gpm
Note that the reverse-return lowered the total flow from 424 gpm to 407 gpm, and each branch’s pressure drop is more balanced. If the lower pressure drop coil of 5.3 ft is substituted, the results are as shown in Table 7-8. Estimating the coil flows for the reverse-return piping and 5.3 ft coils, we see a reduced flow as compared to direct-return piping:
Unit 2 =
2 31.1 ft 100 ---------------- = 104.3 28.6 ft 100
Unit 3 =
100
Unit 1 =
Unit 4 = Total =
2 31.1 ft 100 ---------------- = 104.3 28.6 ft 408.6 gpm
Summary By going through these examples for direct-return and reverse-return piping system designs, we can see how the use of the reverse-return design lowers the overall flow requirement and the magnitude of imbalance in the circuits. Furthermore, we can see that coil flow selection (lower flow rating) increases this flow imbalance for the direct-return system but has only a small effect for the reverse-return system.
Summary of Pumping Horsepower— Direct Return Versus Reverse Return Using the pump horsepower equation discussed in Chapter 4 (Equation 4-3), we can make a judgment on the relative power used in this example and Table 7-9: Flow (gpm) Head (ft) WHP = --------------------------------------------------------, hp 3960 Table 7-9 shows the theoretical power at 100% flow and does not take into account the pump or motor efficiencies, which would increase the values. Direct return uses less horsepower than reverse return but is more unbalanced at 100% flow. A further comparison might be made at part-load conditions to determine operating hours versus percent of flow based on local weather patterns to get a closer look at pumping horsepower. Further, decreasing pipe diameter selection to increase flow velocities could be examined, as well as pressure drop selected for the control valves. A basic comparison is to look at the piping lengths of the direct-return versus reverse-return piping in this example:
Fundamentals of Water System Design I-P Table 7-8 Length, ft 30 ft
100 ft 100 ft 100 ft 70 ft Length, ft 30 ft 100 ft
100 ft 100 ft 70 ft Length, ft 30 ft 100 ft 100 ft
100 ft 70 ft Length, ft 30 ft 100 ft 100 ft 100 ft
70 ft
Lower Pressure Drop Coil for Units 1, 2, 3, and 4
Unit 1 (Path A-B-B'-F): Supply Main Supply main: 5 in. pipe 400 gpm at 2.72 ft/100 Supply branch Coil (100 gpm) Return branch Return main: 3 in. pipe 100 gpm at 2.39 ft/100 Return main: 4 in. pipe 200 gpm at 2.27 ft/101 Return main: 4 in. pipe 300 gpm at 4.89 ft/102 Return main: 5 in. pipe 400 gpm at 2.72 ft/103 Unit 2 (Path A-B-C-C'-D'-E'-F) Supply main: 5 in. pipe 400 gpm at 2.72 ft/100 Supply main: 4 in. pipe 300 gpm at 4.89 ft/100 Supply branch Coil (100 gpm) Return branch Return main: 4 in. pipe 200 gpm at 2.27 ft/101 Return main: 4 in. pipe 300 gpm at 4.89 ft/102 Return main: 5 in. pipe 400 gpm at 2.72 ft/103 Unit 3 (Path A-B-C-D-D'-E'-F): Supply main: 5 in. pipe 400 gpm at 2.72 ft/100 Supply main: 4 in. pipe 300 gpm at 4.89 ft/100 Supply main: 4 in. pipe 200 gpm at 2.27 ft/101 Supply branch Coil (100 gpm) Return branch Return main: 4 in. pipe 300 gpm at 4.89 ft/102 Return main: 5 in. pipe 400 gpm at 2.72 ft/103
Unit 4 (Path A-B-C-D-E-E'-F): Supply main: 5 in. pipe 400 gpm at 2.72 ft/100 Supply main: 4 in. pipe 300 gpm at 4.89 ft/100 Supply main: 4 in. pipe 200 gpm at 2.27 ft/100 Supply main: 3 in. pipe 100 gpm at 2.39 ft/100 Supply branch Coil (100 gpm) Return branch Return main: 5 in. pipe 400 gpm at 2.72 ft/100
Pressure Drop, ft 0.82 ft 2 ft 5.3 ft 9 ft 2.39 ft 2.27 ft 4.89 ft 1.90 ft Total 28.6 ft Pressure Drop, ft 0.82 ft 4.89 ft 2 ft 5.3 ft 9 ft 2.27 ft 4.89 ft 1.90 ft Total 31.1 ft Pressure Drop, ft 0.82 ft 4.89 ft 2.27 ft 2 ft 5.3 ft 9 ft 4.89 ft 1.90 ft Total 31.1 ft Pressure Drop, ft 0.82 ft 4.89 ft 2.27 ft 2.39 ft 2 ft 5.3 ft 9 ft 1.90 ft Total 28.6 ft
213
214
Chapter 7 Piping System Development Direct Return: Supply main (4 in.)
= 30 + 70 = 100 ft
Return main (4 in.)
= 30 + 70 = 100 ft
Supply main (3 in.)
= 100 + 100 = 200 ft
Return main (3 in.)
= 100 + 100 = 200 ft
Unit branches (3 in.)
= 8 × 25 = 200 ft
Total
= 800 ft
Reverse Return: Return mains Main (5 in.)
= 30 + 70 = 100 ft
Main (4 in.)
= 100 + 100 + 100 + 100 = 400 ft
Main (3 in.)
= 100 + 100 = 200 ft
Unit branches (3 in.)
= 8 × 25 = 200 ft
Total
= 900 ft
See Table 7-10 for a summary. Table 7-9
Direct and Reverse Return with Balanced and Unbalanced Flow
Direct Return (Unbalanced Flow) 424 gpm and 30 ft = 3.2 hp
10.7 ft coils
433 gpm and 24 ft = 2.6 hp
Direct Return (Balanced Flow) 400 gpm and 30 ft = 3.0 hp
5.3 ft coils
400 gpm and 24 ft = 2.4 hp
Reverse Return (Unbalanced Flow) 407 gpm and 36 ft = 3.7 hp
10.7 ft coils
409 gpm and 31 ft = 3.2 hp
Reverse Return (Balanced Flow) 400 gpm and 36 ft = 3.6 hp
5.3 ft coils
400 gpm and 31 ft = 3.1 hp
Table 7-10
Summary of Direct and Reverse Return
5 in.
4 in.
3 in.
Total
Direct
—
200 ft
600 ft
800 ft
Reverse
100 ft
400 ft
400 ft
900 ft
Fundamentals of Water System Design I-P
215
As you can see, the analysis of the piping layout takes time, but it is valuable to determine the design options available, amount of piping needed, and how to reduce pumping horsepower. There are many design trade-offs to be considered. The direct-return system requires less piping and theoretically less horsepower than the reverse-return system. However, the reverse-return system balances the system flow better, which translates into better efficiency and performance in load distribution. If the imbalance in the system is left unchecked, uneven distribution of flow to the loads may result, producing poor performance. In the next section, we explore alternative designs that eliminate some of these performance shortcomings.
Primary/Secondary Analysis Another consideration in the piping design strategy is to select a secondary pump for each unit coil and move the two-way valve to the connecting bridge return (see Figure 7-10). One advantage of this concept is to reduce the burden of the pressure drop of the unit coil from the distribution pump and allow constant flow in the unit coil to improve coil heat transfer and response to load. Adjusting the coil flow permits flexibility of higher coil ts as compared to the primary system. Reviewing the previous example, modified for primary/secondary design but limiting the study to the 10.7 ft drop coil and a direct-return piping system (see Figure 7-11), we again assess each unit’s piping for pressure drop from the main to the secondary bridge (see Table 7-11). Unit 3 is the highest pressure drop path, with 19.86 ft required at A-F to provide a design flow of 100 gpm. Again applying the Darcy-Weisbach relationship, Q2 ~ h:
Figure 7-10
Two-way valve in connecting bridge return with secondary pump.
216
Chapter 7 Piping System Development
Figure 7-11
Primary/secondary pumping.
Q2 =
2 h Q 1 --------2h 1
Unit 1 =
2 19.86 ft 100 ------------------- = 122.4 13.26 ft
Unit 2 =
2 19.86 ft 100 ------------------- = 104.9 18.04 ft
Unit 4 =
2 24.26 ft 100 ------------------- = 114.8 15.08 ft
In summary: Unit 1 = 122.4 gpm Unit 2 =
104.9 gpm
Unit 3 =
100.0 gpm
Unit 4 =
114.8 gpm
Total =
442.1 gpm
Fundamentals of Water System Design I-P
217
Again using the water horsepower equation, the primary loop power used at 442 gpm and 19.86 ft is 2.2 hp, which is less than the similar unbalanced direct-return case above at 3.2 hp. When the system is balanced, the primary loop uses 2 hp of energy, versus 3 hp for the direct return. Table 7-11
Pressure Drop from Main to Secondary Bridge for Units 1, 2, 3, and 4
Length, ft
Unit 1: Supply Side (100 gpm)
Pressure Drop, ft
1
4 in. tee branch
0.4
1
4 × 3 in. bushing
0.1
1
3 in. butterfly valve
0.44
3
3 in. ells at 0.1
0.3
1
3 in. strainer
0.35
1
3 in. tee-thru
0.22
25
3 in. pipe (2.39 ft/100 ft)
0.60 Total 2.4
Supply branch
Length, ft
Unit 1: Return Side (100 gpm)
Pressure Drop, ft
1
3 in. tee-thru
0.22
3
3 in. ells at 0.1
0.3
1
2 1/2 in. control valve
5.27
2
3 × 2 1/2 in. bushing at 0.41
0.82
1
3 in. balancing and service valve (open)
1.9
1
4 × 3 in. bushing
0.2
1
4 in. tee branch
0.4
25
3 in. pipe (2.39 ft/100 ft)
0.60 Total 9.49
Return branch
Length, ft 30
30
Unit 1 (Path A-B-B'-F): Supply Main Supply main: 4 in. pipe, 200 gpm = 2.27 ft/100 ft
Pressure Drop, ft 0.681
Supply branch
2.4
Common pipe
0
Return branch
9.5
Return main: 4 in. pipe, 200 gpm = 2.27 ft/100 ft
0.681 Total 13.26
Length, ft
Unit 2 (Path A-B-C-C'-B'-F)
Pressure Drop, ft
30
Supply main: 4 in. pipe, 200 gpm = 2.27 ft/100 ft
0.681
100
Supply main: 3 in. pipe, 100 gpm = 2.39 ft/100 ft
2.39
Supply branch
2.4
Common pipe
0
Return branch
9.5
100
Return main: 3 in. pipe, 100 gpm = 2.39 ft/100 ft
2.39
30
Return main: 4 in. pipe, 200 gpm = 2.27 ft/100 ft
0.681 Total 18.04
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Chapter 7 Piping System Development
Table 7-11
Pressure Drop from Main to Secondary Bridge for Units 1, 2, 3, and 4 (Continued)
Length, ft
Unit 3 (Path A-E-D-D'-E'-F)
Pressure Drop, ft
70
Supply main: 4 in. pipe, 200 gpm = 2.27 ft/100 ft
1.59
100
Supply main: 3 in. pipe, 100 gpm = 2.39 ft/100 ft
2.39
Supply branch
2.4
Common pipe
0
Return branch
9.5
100
Return main: 3 in. pipe, 100 gpm = 2.39 ft/100 ft
2.39
70
Return main: 4 in. pipe, 200 gpm = 2.27 ft/100 ft
1.59 Total 19.86
Length, ft 70
70
Unit 4 (Path A-E-E'-F)
Pressure Drop, ft
Supply main: 4 in. pipe, 200 gpm = 2.27 ft/100 ft
1.59
Supply branch
2.4
Common pipe
0
Return branch
9.5
Return main: 4 in. pipe, 200 gpm = 2.27 ft/100 ft
1.59 Total 15.08
Table 7-12
Pressure Drop in Typical Secondary Loop
Length, ft
Pressure Drop, ft
4
3 in. ells at 0.1
0.4
1
Coil
10.7
50
3 in. pipe
1.2 Total 12.3
The pressure drop in a typical secondary loop is shown in Table 7-12. At 100 gpm coil flow, 0.31 hp pumping power is required to handle the secondary loop and coil. Comparing total horsepower, primary/secondary pumping would use 2 + 4(0.31) = 3.24 hp, as compared to 3 hp for the direct return. Why use primary/secondary then in this type of application? Primarily for heat transfer. Keeping full flow on the coil at all times maximizes heat transfer and maintains design differential temperature on the coil. For CHW applications, the coil always has full latent energy capability for the maintained inlet temperature condition. In other examples, primary/secondary pumping would be used to manage system differential pressure, organize supply water distribution to systems, and provide for flexibility with respect to future system additions.
Types of Pumps and Valves Pump performance must be considered not only at the design point but across its entire characteristic curve. Centrifugal pumps are available with steep curves that drop from high head at low flow to low head at high flow versus
Fundamentals of Water System Design I-P
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those with flat curves that show a small change in head between shutoff to design flow (see Figure 7-12). Some designers like to limit this to a 15% to 25% rise-to-shutoff curve. These flat-curve pumps are always recommended where two-way valves are applied to unit terminals. At part loads, the valves will be operating at lower flows and this will move the system operating differential pressure up the pump curve. Another factor is that when the pressure drop through the terminals is low and the system balance is less than ideal, there may be a tendency to short-circuit the flow in the units closest to the pump. With a steep-curve pump, this results in a drop in head and less flow is available to the units farthest from the pump. A flat-curve pump will show a minimum drop in head and give better flow to the remote units.
Effects of Control Valves Factors affecting control valve performance in a typical load’s control loop are as follow: •
Figure 7-12
Valve size. It is most important that a valve be sized for the required load flow in gallons per minute and adequate pressure drop. Some size a valve for one size less than coil inlet/outlet size, but this is not accurate enough.
Flat versus steep pump characteristics.
220
Chapter 7 Piping System Development
•
•
•
•
•
•
Higher pressure drops yield higher valve authority. Ideally, the designer calculates the coil characteristic and the valve characteristic to understand the system control gain. The closer to linear the better for proportionally throttled modulating control. The designer can also consider alternative control strategies, such as using open/close control valves that are proportionally controlled through the open and close time. Control valve actuator. An adequate control valve actuator with sufficient power to hold the valve’s commanded position at maximum pressure drop should be selected and specified. Valve characteristic. The valve characteristic selected for throttled hydronic heating or cooling units should be an equal percentage characteristic to give a linear output of Btu emission in relation to valve stroke. Valve authority. Valve authority is determined by the valve pressure drop at full load compared to the load coil and piping pressure drop. The pressure drop selected should be at least 30% to 50% of the controlled loop pressure drop (which may be the system head or some other controlled pressure drop). In the direct-return example, a valve drop of 5.27 ft was selected as compared to the loop drop of 29.66 ft, which results in 5.27/ 29.7 = 17.7%. When the coil drop of 5.3 ft was selected, the authority increased to 5.27/24.26 = 21.7%; an authority of 25% would require a drop of about 9 ft for the 29.7 ft loop pressure drop. Valve body rangeability. The valve body rangeability should be 30:1 or more. That is the ratio of its maximum controlled flow; say, 100 gpm to a minimum flow of at least 3 gpm (100/3 = 30:1). The designer may not be given the valve body rangeability, with the manufacturer opting to formulate a combined value of valve body and actuator. In these cases, a minimum rangeability of 75:1 should be considered. Valve body style. In general, a two-way valve is recommended because it modulates the volume of flow in relation to the load. Use of three-way valves on terminals is not recommended because the flow is bypassed and does not reduce the pumping power at low loads. Traditionally, globe valves are used for proportionally throttled control systems, especially in hot-water systems where cavitation can become a concern. Segmented ball valves are also an option for throttled control, although with all ball valve installations, cavitation and pipe/valve size correction factors should be carefully examined when specifying the valve. Balancing valve. The balancing valve is a multifunction valve that is used to measure flow, stop flow for servicing, and proportionally balance the flow in circuits that have excessive flow. This valve should be selected for the design flow when wide open and a minimum but measurable pressure drop. When applying variable-speed pumping, automatic flow-limiting valves are generally a better choice instead of manual balancing valves. This is due to the changing system differential pressures as the pump differential pressure controller changes speed in relation to the controlled pressure of the system. Care should be taken in the selection of these devices for design flow at a minimum
Fundamentals of Water System Design I-P
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pressure drop. Also, a shutoff valve is required in the return for servicing, and a venturi or orifice may be required to determine the flow in the field.
Primary/Secondary Application Study The use of a secondary pump to reduce the pump head of the main distribution pump is always an option when designing a hydronic system. Each system should be evaluated on its own requirements and the resulting pumping energy reductions that may be obtained. Figure 7-13 shows a four-zone heating system using a single 7.5 hp pump delivering 230 gpm at 60 ft head with a 20°F t. The zones are as follows: • • • •
One-pipe heating zone: 30 gpm at 12 ft Two-pipe reverse-return heating zone: 60 gpm at 20 ft Two-pipe direct-return heating zone: 100 gpm at 40 ft Heat exchanger zone: 40 gpm at 15 ft The distribution main has a total friction loss of 30 ft from the distribution pump through the mains and returning through the boiler to the pump suction. The pump has to deliver the total flow of 230 gpm and must overcome the worst-case pressure drop of 60 ft. The zone 3 head loss calculations are shown in Table 7-13.
Figure 7-13
Four-zone heating system.
222
Chapter 7 Piping System Development A redesign of the system to a primary/secondary concept is shown in Figure 7-14. Each zone is connected to a primary/secondary bridge and a secondary pump is selected for each zone’s load. Because the primary/secondary bridge has a minimal pressure drop, the distribution pump is required to flow 230 gpm at 30 ft instead of 60 ft, so a 3 hp motor can be selected, as shown in Table 7-14. Note that the flow and head are large enough that a larger pump is selected and that there is a duty point operation of 2.3 hp. Each zone is pumped by a circulating pump that is significantly smaller, in both size and cost but also flow and head. Small circulating pumps do not come with a published duty point of operation, as the motors are generally small, fractional horsepower and designed for non-overloading operation for any potential operation of the pump. This means the original 7.5 hp pump is replaced by five pumps totaling about 6 hp, which gives a reduction of 12% in connected pumping energy. If the system operates properly, the primary/secondary system Table 7-13 Zone, ft
Four-Zone Heating System Mains, ft
Total Head Required
Zone 1
12
+
0
=
= 12 ft
Zone 2
20
+
5+5
=
= 30 ft
Zone 3
40
+
10 + 10
=
= 60 ft
Zone 4
15
+
15 + 15
=
= 45 ft
Figure 7-14
Primary/secondary pumping, four-zone heating system.
Fundamentals of Water System Design I-P
223
will probably operate with slightly less energy than the 4.75 hp of a primary-only pump, even though the duty point of the primary pump and the nominal horsepower of the secondary pumps total 5.25 hp. This example assumes the load flows are based on a 20°F drop; further study of the primary/secondary bridge flows shows an ability to have different ts as follows for a typical secondary bridge (see Figure 7-15):
Table 7-14
3 hp Motor Selected
Required Power, hp
Operating Power, hp
Motor Power, hp
Zone 1
30 gpm at 12 ft
0.09
0.3
0.3
Zone 2
60 gpm at 20 ft
0.30
0.75
0.75
Zone 3
100 gpm at 40 ft
1.01
1.5
1.5
Zone 4
40 gpm at 15 ft
0.15
0.4
0.4
Secondary
230 gpm
1.56
2.95
2.95
Primary
230 gpm at 30 ft
1.74
2.3
3
Total
3.30
5.25
5.95
3.48
4.73
7.5
Primary only 230 gpm at 60 ft
Figure 7-15
Primary/secondary bridge energy.
224
Chapter 7 Piping System Development Energy In = Energy Out Primary Energy = Secondary Energy 500 GPM t = 500 GPM t Here is the benefit of a primary/secondary system from an energy perspective: If we design for a greater drop of 40°F in the primary and maintain a 20°F drop for the design flow in the secondary loops (as shown in Figure 7-15), we see a 50% flow reduction in the primary supply required. The 50% reduction in the supply flow means one-fourth of the original main pressure drop, so the primary distribution pump size and horsepower are reduced to 115 gpm at 7.5 ft, which means the supply pump can be a 0.5 hp inline circulator. Now our summary becomes the layout and data shown in Figure 7-15 and Table 7-15. The original design requiring a single 7.5 hp has been reduced to a total of 3.5 hp using the primary/secondary design principles per Figure 7-16. A further review of pipe sizing may provide additional economies. Another review may yield further flexibilities in the zoning capabilities to provide diversity and energy economy by treating the secondary circuit as a controlled supply temperature and t when used with a two-way valve in the bridge return. Note: By increasing t in the primary piping circuits, a 50% reduction in supply flow was achieved. This allowed significant energy savings by reducing the total pump horsepower requirements. Primary/secondary pumping techniques not only provide this powerful design technique but also have other important design abilities. Implicit in the discussion is that because of the common pipe, which is shared between the two systems (primary and secondary), the two systems are hydraulically isolated from each other. This allows for system pressures to be managed vis-à-vis the design process. Managing pressures is ultimately what system balance is all about—getting the design flow where it is needed when it is needed. It also allows for changes to be made to parts of the system without necessarily affect-
Table 7-15
Primary/Secondary Bridge Energy Required Power, hp
Operating Power, hp
Motor Power, hp
Zone 1
30 gpm at 12 ft
0.09
0.3
0.3
Zone 2
60 gpm at 20 ft
0.30
0.75
0.75
Zone 3
100 gpm at 40 ft
1.01
1.5
1.5
Zone 4
40 gpm at 15 ft
0.15
0.4
0.4
Secondary
230 gpm
1.56
2.95
2.95
Primary
115 gpm at 8 ft
0.22
0.5
0.5
Total
1.77
3.45
3.45
230 gpm at 60 ft
3.48
4.73
7.5
Primary only
Fundamentals of Water System Design I-P
Figure 7-16
225
Primary/secondary pumping, four-zone heating system.
ing the other connected systems. For example, a chiller plant is designed with primary/secondary techniques and it is decided to add more loads to the system, requiring another chiller on the primary side. Primary/secondary design allows for the primary plant to be modified without having an impact on the secondary side. In a primary-only system, chances are there would have to be a complete modification of the pumping system. The counter to this, though, is that there is disparate control of zone loads/ flow and the affect it has on the pump. If one zone is closed or closes while others are open to flow, there can be a large flow interaction affecting all attached zone circuits. When zones are laid out so that all zone valves are on similar thermally reacting zones, acting in the same way opening and closing, flow control tends to stay proportional without large swings in flow. This specifically was one of the early problems of variable-speed pumping being applied to primary/secondary CHW plants. Running a variable-speed, variable-flow secondary pumping system with a constant-speed primary pump allowed for poor system t (and associated plant efficiency) to be established at some operating conditions. At others it created chiller sequencing problems. In large part, this was due to chillers that required constant flow even at part load due to the inherent operating characteristics of the chiller control. As chiller controls evolved, they too allowed for flow to vary through them; however, designers failed to follow along with variable-primary/variable-secondary systems to accommodate maximizing t and system efficiency. As a design technique, primary/secondary is still quite applicable, even in the world of variable-speed pumping, and brings the benefit of being able to break up a system into smaller manageable hydraulic zones.
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Chapter 7 Piping System Development
Antifreeze Solutions for Low-Temperature Applications Another topic that relates to piping system development is how to treat low-temperature applications. In most cases, protection is required in the form of circulating an antifreeze solution that is capable of preventing bursting of coils, fittings, and piping. Figure 7-17 shows a dedicated heat exchanger and pump for a glycol subsystem complete with an expansion tank and detachable water makeup (or other water makeup devices to meet local code requirements). Other approaches assume constant water flow in a coil by treating it as a secondary pumping circuit with constant flow to provide 4 to 6 ft/s tube velocities. On the air side, face bypass dampers are required to provide temperature control to blend warm air from the coil side with bypass air to maintain a controlled discharge temperature (see Figure 7-18). In addition, a separate low-temperature thermostat with a long sensing element (within which the coldest 1 ft length controls) set at 40°F is recommended on the coil air discharge surface. An alternative method is a bulb inserted in the leaving water to provide an alarm to the operating personnel before the discharge or water temperature reaches freezing conditions. Sometimes this is wired to the fan circuit to shut down the fan and close outside dampers to reduce the possibility of damage.
Figure 7-17
Coil with glycol heat exchanger and pump for low temperatures.
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When applying water/glycol solutions, it is important to recognize the reduced heat carrying capacity due to the lower specific heat and higher density of the mixture. Previously, in Chapters 1 and 2, we discussed that the heat transferred to or from water qw (Btu/h) is expressed as q w = mc p t where cp =
specific heat, 1.0 Btu/lb·°F
m
=
mass flow of water, lb/h
t
=
temperature change across unit, °F
(7-1)
If water flow Qw is expressed in gpm, q w = 8.02 w c p Q w t where w is density of water, 62.4 lb/ft3. For water/glycol mixtures, the specific heat cp and density w of the mixture have to be used. For example, a 35% propylene water/glycol solution at 180°F with 20°F drop will transfer as follows: qw = 8.02 wcpQw t
Figure 7-18
Pumped coil with face bypass dampers for low-temperature primary/secondary pumping.
228
Chapter 7 Piping System Development =
8.02(62.2)(0.95)(1)(20)
=
9478 Btu/h
In other words, at 20°F t, the mixture will transfer 9478 Btu/h per 1 gpm,, as compared to 10,000 Btu/h per gpm for water. Chapter 31 in the 2013 ASHRAE Handbook—Fundamentals shows complete reference tables of freezing points and graphs of specific heat versus percentage water mixture and temperature (see Figures 7-19 and 7-20) (ASHRAE 2013b). In addition to specific heat, the Handbook displays density, viscosity, and thermal conductivity for ethylene/glycol and propylene/glycol water mixtures. The chapter also states that for winterizing coils in HVAC systems, a 30% ethylene/glycol or 35% propylene/glycol mixture with water can be used. It states that as the fluid freezes, it forms a slush that expands and flows to any available space. Therefore, expansion volume must be included with this type of protection. It also recommends that if the application requires the fluid to remain entirely liquid, a concentration with a freezing point 5°F below the lowest expected temperature should be chosen. Further information regarding corrosion inhibition is also provided (ASHRAE 2013b). Without inhibitors, glycols oxidize into acidic end products.
Figure 7-19 Specific heats of aqueous solutions of industrially inhibited ethylene glycol (percent by volume).
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Provision should be made for testing and filling glycol and glycol inhibitor into the piping system. Figure 7-21 is one design of a one-shot chemical feeder that should be considered.
Pumping Design Factors By completing this chapter, you should have learned the following general design concepts:
Figure 7-20
•
The higher the head losses through the terminal units of a hydronic system, relative to the main piping losses, the closer the system comes to a natural balance.
•
Reverse-return systems are closer to a natural balance of flows than direct-return systems.
•
If automatic control valves are employed, the design pressure drop selected should be as high as practical. A pressure drop at least equal to the drop in the terminal unit coil is a desirable goal. The valve should be sized for the
Specific heats of aqueous solutions of industrially inhibited propylene glycol (percent by volume).
230
Chapter 7 Piping System Development
Figure 7-21
One-shot chemical feeder.
• •
•
•
design flow with the Cv flow formula (Equation 5-1), which may not be the same size as the coil inlet piping. Centrifugal pumps with flat characteristics should be selected for systems with control valves. Two-way valves should be considered over three-way valves because they vary the volume of water flow in direct relationship with the control signal. Three-way valves provide a continuous flow regardless of the load and are not suitable with variable-volume pumping systems. Manual balancing valves should be chosen for a minimal pressure drop and provide the means to measure flows in various loops in the field as well as provide a shutoff valve for coil servicing. Performance is best ensured by requiring proportional balancing after the system is operating. Variable-volume pumping systems should be checked and adjusted for balance at 50%, 75%, and 100% design flows.
The Next Step In Chapter 8, you will learn how to match pumps to systems.
Summary In this chapter, we covered the following:
Fundamentals of Water System Design I-P • • • • • • •
231
The factors that must be considered before starting a piping design What piping equipment should be considered at a load coil Which piping system design gives more uniform pressure drop and why How flow can be determined in a given loop What type of pump curve should be selected if two-way valves are to be used Two advantages of primary/secondary hydronic systems How the possibility of freezing coils and piping in an HVAC system can be reduced
References and Bibliography ASHRAE. 2012. ASHRAE Handbook—HVAC Systems and Equipment. Atlanta: ASHRAE. ASHRAE. 2013a. ANSI/ASHRAE/IES Standard 90.1-2013, Energy Standard for Buildings Except Low-Rise Residential Buildings. Atlanta: ASHRAE. ASHRAE. 2013b. ASHRAE Handbook—Fundamentals. Atlanta: ASHRAE. ASHRAE. 2014. 90.1-2013 User’s Manual. Atlanta: ASHRAE. Beaty, F. 1987. Sourcebook of HVAC Details. New York: McGraw-Hill Book Co. HI. 1990. Engineering Data Book, 2d ed. Parsippany, NJ: Hydraulic Institute.
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Chapter 7 Piping System Development
Skill Development Exercises for Chapter 7 Complete these questions by writing your answers on the worksheets at the back of this book. 7-1
In the direct-return example (see Figure 7-8), which unit’s piping path dictates the pump head from A to F? How much head is required?
7-2
What size pump capacity and head would be required to handle the four AHUs in Figure 7-8 for the conditions shown from A to F?
7-3
What size pump capacity and head would be required in Figure 7-8 (direct return for supplying four identical floors), assuming 10 ft between floors and allowing a 25% head drop in the main for fittings and a 20 ft drop across the chiller?
7-4
In the reverse-return example in Figure 7-9, which unit’s piping path dictates the pump head from A to F? How much head is required?
7-5
What size pump capacity and head would be required to handle the four AHUs in Figure 7-9 for the conditions shown from A to F?
7-6
What size pump capacity and head would be required in Figure 7-9, assuming 10 ft between floors and allowing a 25% head drop in the main for fittings and a 20 ft drop across the chiller?
7-7
In the primary/secondary examples in Figures 7-10 and 7-11, which unit’s piping path dictates the pump head from A to F? How much head is required?
7-8
What size distribution pump capacity and head would be required to handle the four AHUs in the Figure 7-11 primary/secondary example for the conditions shown from A to F?
7-9
What size distribution pump capacity and head would be required for supplying four identical floors (similar to the Figure 7-11 primary/secondary example), assuming 10 ft between floors and allowing a 25% head drop in the main for fittings and a 20 ft drop across the chiller?
7-10
What is the cooling capacity of 100 gpm of water flow with a 50% propyleneglycol mixture at an average temperature of 50°F and a 10°F rise?
7-11
What is the cooling capacity of 100 gpm of water flow at a 10°F rise? How many tons of cooling?
7-12
What is the pumping horsepower for the propylene-glycol mixture (specific gravity = 1.05) in Exercise 7-10, compared to the plain water in Exercise 7-11, if the coil pressure drop is 20 ft, assuming a pump efficiency of 75% and motor efficiency of 85%?
7-13
How can the possibility of frozen coils and piping in an HVAC system be reduced?
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7-14
What are two advantages of primary and secondary hydronic systems?
7-15
What type of pump curves should be selected if two-way valves are to be used?
7-16
How can flow be determined in a given loop?
7-17
Which piping system design gives more uniform pressure drop? Why?
Matching Pumps to Systems
Study Objectives After studying the material in this chapter, you should be able to do the following: • • • • • • • • •
determine the operating point of a pump and a system by plotting the system head curve and the pump head–capacity curve; determine static pressure in a system by turning the system pumps off and reading the gage pressure at the pump; plot system head and pump capacity curves, including the incremental effect of static pressure; know what methods are available to match pumping operation to a system for full- and part-load flow; develop operating curves for pumps connected in series and in parallel; plan for emergency flow in case of a pump failure; determine how many combinations of capacity and head two two-speed pumps can provide when operating in parallel; identify what the variable-volume controller should measure to control variable-speed pumps; and describe what methods are available to provide flow to buildings that are remote from a central chiller plant.
Instructions Read Chapter 8 and answer all of the questions at the end.
Matching the Pump to the System The section Pipe Selection in Chapter 3 discusses pressure drop in piping systems and the system curve that results from plotting pressure drop versus flow (see Figure 8-1). Similarly, in the section Pump Selection in Chapter 4, the concept of pump head curves was developed, noting that each type of pump has a unique curve governed by its size and design. The concept of the pump operating at the intersection point of the system curve was also introduced.
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Chapter 8 Matching Pumps to Systems
Figure 8-1
Typical open-system system curve. Source: ASHRAE Handbook—HVAC Systems and Equipment (2012).
Superimposing the system curve and pump head curve, the operating point is defined by the intersection of the curves (see Figure 8-2). Under actual operating conditions, control valves are varied to meet changing load conditions and the system curve changes, moving the operating point along the pump curve. In this manual, the premise was that the intersection point was based on a system curve for a closed-loop piping system. That makes sense considering the energy implementation of the Bernoulli equation (Equation 2-1), which shows that pump energy is equal to the friction loss of the piping in a closed-loop system. However, Bernoulli refers to three layers of pressure, one of which is due to changes in elevation. When the system is piped in such a manner that a change in elevation occurs within the piping circuit, this too must be accounted for on the system curve. One such example is with a cooling tower. Using Figure 8-3 as an example, we will define that the pump is supposed to move 100 gpm and that there is 30 ft of pipe friction loss. (We are not dealing with extra additions for pipe aging in this case, and those allowances could be significant). From
Fundamentals of Water System Design I-P
Figure 8-2
Pump curve and system curve intersection.
Figure 8-3
System with cooling tower.
237
238
Chapter 8 Matching Pumps to Systems
Figure 8-4
Cooling tower curves.
Figure 8-4, we will define d1 and d2 as 10 and 20 ft, respectively. These distances represent both the difference in elevation and the distance between the surface of the water in the tower sump serving the pump and either the discharge of the pipe (d1) or the maximum pipe elevation (d2). The pump selected is slightly oversized, being provided with a 7 in. diameter pump impeller. Under ideal circumstances and if the system ran continuously, the pump would always have to elevate the water 10 ft, the difference between pipe discharge and sump height. This is a constant head loss due to elevation, regardless of what the flow rate is and the varying head loss due to friction found in the pipe. Plotting the 30 ft variable system curve from a starting point of 10 ft, the pump will provide 110 gpm. To get to the design flow, either a balancing valve is throttled, adding an extra 7.5 ft of head loss, making the pump flow at 100 gpm and 47.5 ft head, or the pump impeller is trimmed to slightly larger than 6 1/2 in. and operating at 40 ft of head. Why does curve 2 appear in Figure 8-4? In cooling towers this helps illustrate a simple point: the pump has to first elevate the water to the top of the pipe, and it must have the energy to do so. What happens on start-up in this case? We will stipulate that there is 10 ft of elevation from the sump water level to the centerline of the pump suction. With the pump off, water is in the sump, floods the pump case, and goes up the supply pipe 10 ft. As the pump begins to add energy, water is removed from the sump and the elevation goes down, increasing the elevation difference as water starts to go up the pipe. An important question here is whether there is enough sump capacity to fill the pipe and allow the period in time when
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239
the system is off as the water backs up into the sump. As the pump works, and it will happen rather slowly, water will eventually make it to the top of the pipe and establish a flow with the discharge pipe (sometimes called a downcomer), eventually filling. Cooling tower start-up has always been a troublesome subject when it comes to start-up and having enough pump energy to initiate the process of design flow. Adding to the issues are sump volume and coordination with sump makeup water control so that water is not wasted at start-up or shutdown. One suggested option for reducing issues is leaving the piping system in a filled state (perhaps incorporating a control valve in the discharge that closes prior to the pump being turned off) and possibly a check valve so that flow moves from the sump to the pump but not vice versa. Regardless, remember that there are effects such as this that must be allowed for when dealing with open systems. Independent head is the system static pressure with the pump off. A good piping system design will match the system characteristics to the pump head curve to provide the best system performance with the best economics over the life of the system.
System Curves System curves can be a powerful tool for the designer in determining the effects that various components can have on system performance. Chapter 13 of ASHRAE Handbook—HVAC Systems and Equipment (2012) defines the use of a system coefficient as Q C s = ----------P
(8-1)
and is the system version of the control valve flow coefficient Cv. As such, it represents a defined flow in gallons per minute at a pressure drop of 1 psi or a head of 2.31. Combined then with the use of the Darcy flow and head relationship (see Equation 2-2), one can imply flow coefficients for any device or component that has a defined head loss at a specific flow and plot the component curve just as one might for a system curve. Note: because the flow coefficient is based on units of pressure in psi, we will stipulate that as we discuss a change in pressure we will treat it as an equal change in head by the factor of 2.31 ft/psi.
Series Flow Coefficients If one were to take this a few more steps, then we could form a logical path such as the following: To calculate system head loss, we add all of the pressure (head) losses for devices and components such that P s = P 1 + P 2 + + P n Pn PS
= =
pressure drop of components pressure drop of system or subsystem
(8-2)
240
Chapter 8 Matching Pumps to Systems Algebraically, if the system pressure loss is equal to the sum of the component losses, we could rearrange the equation such that Q P = -----Q- 2 C s = ----------C P s
(8-3)
and Q 2 Q 2 Q 2 Q 2 P = ------- = ---------- + ---------- + + ---------- C C C C s s1 s2 sn
(8-4)
Knowing that the flow in a circuit Q is the same for all components under evaluation, then the equation for a system flow coefficient made up of a group of components in series is 1 1 1 1 ------ = -------- + -------- + + -------2 2 2 2 Cs C s1 C s2 C sn
(8-5)
As shown in Chapter 5, this equation is used for calculation of the effects of control valve authority. Commonly behaving devices can be grouped into one composite flow coefficient. Considering how a system operates, pipe, fittings, coils, and other nonmoving pieces represent these types of components. They all have a fixed flow coefficient. The premise of a control valve, though, is that every incremental position has a unique flow coefficient so that flow may be decreased or increased according to the valve position. When tested in a flow laboratory, the control valve is tested at every position, maintaining a constant differential pressure across the valve. In so doing, the system effects of components are factored out of the test, so the true flow capability of the valve is accounted for. In application, it is rare that differential pressure is controlled across the control valve, so the variable nature of the valve flow coefficients must be weighed against the constant component flow coefficients. Equation 8-5 may be rearranged algebraically to yield the following: Cs – v Cs – C C s = -----------------------------------2 2 Cs – v + Cs – C where Cs =
(8-6)
composite system coefficient
Cs – C =
component flow coefficient calculated from the pressure drop of the components (not including the control valve)
Cs – v =
control valve flow coefficient at each incremental position
Fundamentals of Water System Design I-P
241
Example 8-1 For an equal percentage control valve, calculate and plot the composite flow characteristic for the valve operating in a system with 50%, 30%, and 10% control valve authority. For the example, use design parameters as follows: Control valve rangeability = 35:1 System head = 100 ft Control valve flow = 100 gpm
Solution
There are a couple of equations presented in previous chapters that will be repeated here for this example. To begin with, recall how a theoretical equal percentage control valve operates. In general it follows the equation C v = C vmax R where Cv
l – 1
=
partial operating height flow coefficient, gpm/psi Cvmax = flow coefficient at full operating height, gpm/psi l = partial lift of the valve R = rangeability of valve, dimensionless and tested In organizing a solution, it becomes obvious that we will need to do a lot of repetitive calculations that reference each other. In this regard, using a spreadsheet to calculate and organize the data makes a lot of sense. Therefore, step 1 has us calculating the theoretical flow coefficient required for a flow rate of 100 gpm using all available head, which would be 100 ft. To do that, we use the form of the flow equation incorporating head as follows: Q 100 C v = -------------- -------------- = 15.2 h 100 ------------------2.31 2.31 This sequence is repeated for each valve position, the results of which are shown for increments of 10% position (raw valve Cv). Note that at 0% valve position there is an incremental flow due to the math of the calculation. As mentioned previously, most manufactured valves are a modified equal percentage characteristic of some sort. For the purposes of this problem, we will simply modify the percentage by subtracting the closed value from the 0% through 90% valve positions (biased valve Cv). The biased valve flow coefficient is then normalized (% Cv) and is plotted (curve 1) as the control valve characteristic. This is the reference equal percentage characteristic, as if it were tested in a fluids laboratory. See Table 8-1 for the data.
3.23
2.13
1.37
0.83
0.45
0.19
0.00
60% 3.67
50% 2.57
40% 1.80
30% 1.26
20% 0.88
10% 0.62
0%
0.43
4.80
70% 5.23
0%
1%
3%
5%
9%
14%
21%
32%
0.61
0.88
1.25
2
3
4
5
7
11
7.03
80% 7.46
0.0
0.3
0.6
1.2
1.9
3.0
4.6
6.8
9.9
14.4
15
90% 10.65 10.22 67%
46%
21.5
100% 15.20 15.20 100% 21.5
21.5
21.5
21.5
21.5
21.5
21.5
21.5
21.5
21.5
0.00
0.26
0.64
1.17
1.92
2.99
4.47
6.47
9.02
21.5 11.99
0%
2%
4%
8%
13%
20%
29%
43%
59%
79%
21.5 15.20 100%
27.7
0.79
1.13
1.61
2.30
3.29
4.69
6.69
9.55
0.0
0.3
0.8
1.5
2.5
3.9
5.9
8.8
13.63 12.8
19.45 18.7
27.7
18.2
18.2
18.2
18.2
18.2
18.2
18.2
18.2
0.00
0.34
0.82
1.51
2.47
3.81
5.61
7.89
18.2 10.48
18.2 13.01
0%
2%
5%
10%
16%
25%
37%
52%
69%
86%
18.2 15.20 100%
1.37
1.96
2.80
3.99
5.69
8.12
11.59
16.54
23.60
33.68
48.1
0.0
0.6
1.4
2.6
4.3
6.8
10.2
15.2
22.2
32.3
48.1
16.0
16.0
16.0
16.0
16.0
16.0
16.0
16.0
16.0
16.0
16.0
System Cv
0.00
0.59
1.42
2.58
4.17
6.22
8.62
11.01
13.00
14.35
15.20
System Cv
Design P = 10 ft
Design Flow = 100
0%
4%
9%
17%
27%
41%
57%
72%
86%
94%
100%
Percent System Cv
Equal Percentage 10% Authority
Sys- Percent Raw Biased tem System Valve Valve Cv Cv Cv Cv
Design P = 30 ft
Design Flow = 100
Equal Percentage 30% Authority
Sys- Percent Raw Biased System System Valve Valve tem Cv Cv Cv Cv Cv
Design P = 50 ft
Design P = 100 ft Raw Biased Sys% Valve Valve tem Cv Cv Cv Cv
Design Flow = 100
Design Flow = 100
Raw Biased Posi- Valve Valve tion C Cv v
Equal Percentage 50% Authority
Theoretical Equal Percentage Control Valve Data for Example 8-1
Theoretical Equal Percentage
Table 8-1
Fundamentals of Water Systems Design I-P 242
Fundamentals of Water System Design I-P
243
The next step in the process is to calculate each control valve based on the specified authority. Recall that control valve authority is the ratio of the control valve pressure drop to the controlled system drop. In this case, the controlled system drop is the 100 ft of the system as specified. Therefore, a 50% authority would yield 50 ft head across the control valve, 30% 30 ft on the valve and 70 ft on the system components, and 10% 10 ft on the valve and 90 ft on the system components. 50% valve authority Q 100 C v = -------------- -------------- = 21.5 h 50 ------------------2.31 2.31 30% valve authority Q -------------100 = 27.7 C v = -------------h 30 ------------------2.31 2.31 10% valve authority Q 100 C v = -------------- -------------- = 48.1 h 10 ------------------2.31 2.31 Similar to the development of the theoretical equal percentage coefficient, each valve is biased and incremental positions are calculated. However, to see the effect on the system flow control, the pressure drop effect of the system components must also be calculated. To do this, Equation 8-5 is used and is shown here for the 50% authority control valve operating at 50% position Cs – v Cs – C C s = -----------------------------------2 2 Cs – v + Cs – C 3 21.5 64.5 64.5 C s = ----------------------------- = ------------------------------ = ---------- = 2.99 21.7 2 2 9 + 462.25 3 + 21.5 This value is about 20% of the total system composite flow coefficient and provides about 50% more water flow than implied by the theoretical control valve coefficient. The rest of the calculations are repeats of this same process. When the normalized values are plotted, the graph shown in Figure 8-5 is developed. In Figure 8-5, characteristic 1 is the theoretical characteristic, characteristic 2 is the 50% authority characteristic, characteristic 3 is the 30% authority characteristic, and characteristic 4 is the 10% authority characteristic.
244
Chapter 8 Matching Pumps to Systems
Figure 8-5
Valve flow versus lift for Example 8-1.
Parallel Flow Coefficients Knowing how to calculate the flow coefficient when we have components in series leads to dealing with parallel flow paths. For discussion purposes, Figure 8-6 shows head versus flow for a pump serving three circuits, which we will treat this as three balanced paths (e.g., that all paths have the same pressure drop). The system coefficient can be calculated just as it would be if we were developing a single value number akin to the system curve. If each circuit has flows as designated in the figure, with each pressure drop being equal, then it stands to reason that we could also calculate a single flow coefficient for the path. In doing this, it becomes apparent that the system flow coefficient is the sum of each single path coefficient as designated in Equation 8-7. If we wanted to see what happens in the system, we could use the approach of the components in series to relate what happens when adjustments are made to specific
Fundamentals of Water System Design I-P
Figure 8-6
245
Head versus flow for pump serving three circuits.
components. When this is done, it is important to remember that shared piping pressure drops are used with respect to each path’s individual flow rate to establish the flow coefficient for that path under evaluation. C s = C sP1 + C sP2 + + C sPn
(8-7)
For example, in circuit 2 shown in Figure 8-6, the flow rate is 80 gpm. The path shares pipe segments A-B, with a head loss of 4 (2 × 2), and B-C, with a head loss of 12 (2 × 6). If we were to consider just the flow coefficient for segment A-B, we could calculate it as about CsA – B = 140. However, while the pipe carries the full flow, from a circuit perspective the path head loss is calculated as the total head loss for the incremental flow path. So if we were calculating the path flow coefficient from the sum of its components in series, the coefficient would be close to 86 for CsA – B, as part of circuit 2. Figure 8-6 shows how each path curve interacts with a pump curve chosen for this system example and also shows how the system curve interacts. It is based on the calculations shown in Table 8-2, which were done in a spreadsheet. It is a balanced path example, meaning that all paths have the same head loss, 42.5 ft. What would have happened if the system had been left unbalanced? The effect may be easily calculated by leaving the balancing valve pressure drop effect out of the path calculation, as shown in the calculations in Table 8-2. Note that for each path, the balanced flow coefficient is shown above the unbalanced, and it is always less, indicating that unbalanced will have more flow as determined by the unbalanced system curve intersection point with the pump curve
0.0004
B-C
85.98
0.0001
A-B
Flow Coefficient
1/ C S
0.0029
0.001
1/ C S
2
18.61
32.24
Flow Coefficient
6
2
Head
30
30
Flow
2
49.64
2
Head 6
80
Flow 80
0.0024
20.39
5
30
C-D
Comps
C-D
A-B
B-C
0.0027
0.0022
1/ C S
2
19.22
21.49
Flow Coefficient
0.0005
45.83
0.99
30
Comps
7E-05
121.59
1
80
2.5
2
20
Comps
Head
C-D
20
B-C
Flow
A-B
0.0015
25.49
3.2
30
Coil
0.0002
70.20
3
80
Coil
0.013
8.77
12
20
Coil
0.0048
14.42
10
30
Valve
Path 3
0.0003
60.79
4
80
Valve
Path 2
0.013
8.77
12
20
Valve
Path 1
0.0011
30.00
2.31
30
Balance
0.0013
28.27
18.5
80
Balance
0.013
8.77
12
20
Balance
5
30
E-F
E-F
E-F
0.0024
20.39
Balanced Path Example
Balanced Path Flow
Table 8-2
0.0029
18.61
6
30
F-G
0.0004
49.64
6
80
F-G
F-G
0.001
32.24
2
30
G-H
0.0001
85.98
2
80
G-H
0.0022
21.49
2
20
G-H
0.0204
42.5
30
Total
0.0029
42.5
80
Total
0.046
42.5
20
Total
6.9941
Cv
6.9941
Cv
18.651
Cv
18.651
Cv
4.6627
Cv
4.6627
Cv
246 Chapter 8 Matching Pumps to Systems
Fundamentals of Water System Design I-P
247
(Table 8-3). This is then plotted so that it can be evaluated, as seen in Figure 8-7: the unbalanced system flows about 155 gpm, which is about 20% more than design flow. Many scenarios can be altered to show different effects. In Figure 8-8, distribution pipes were made larger, substantially reducing friction losses in part of the system, and the coils and control valves were resized to reflect less pressure loss. As can be seen, the system curve nearly goes off the pump curve to a point of unsatisfactory operation of the pump. Similarly, we can make the curve come closer to the actual system curve by redistributing where the head losses are located in the piping network. The idea being presented is not to prove a particular point, it is actually to show a technique that is not difficult to implement in a spreadsheet that allows the designer to evaluate and make system decisions to optimize the control of the system and minimize the energy use. With respect to balancing or not balancing a system, using normal design guidance on pipe selection with reasonable pressure drops for control valves and coils (preferably more than the pipe losses), systems that are unbalanced will typically overflow from 8% to 25%. Of course, these examples are based on the idea that the pump head has actually been calculated and selected for what is actually required. It is very common in the field to see a variety of what can only be described as gross errors, such as installing the wrong speed pump, unapproved modification of the pump selection through the addition of extra head or flow as a safety factor, guesstimating the system head, using diversity factors improperly, and installing components that are not what was called for in the design. This technique, because it can establish a flow coefficient based on components, allows for system evaluation in a simple and reasonably accurate way.
Parallel Pumping As mentioned in Chapter 4, designers should consider pumps in parallel, especially when the designer is installing two pumps, one as primary and the other as backup. In backup mode, two pumps of the same size are selected. They are connected to the system in parallel only from the perspective of pump failure (e.g., when one pump is inoperable, it is hydraulically disconnected from the system through valves and the other is connected so that the system may operate). Parallel pumping operates two pumps to provide design flow and head conditions. A controller turns one of the pumps off at conditions of load less than design and complementary to the flow required for heat transfer. The advantage to this design is that two smaller pumps may be purchased, saving money. Also, when it is operating in single pump mode, pump energy may be reduced. When two smaller pumps are purchased and operated in parallel, a composite pump curve is created by the designer and compared with the design system curve. Pumps are selected for one-half of the system design flow rate and the full design head of the system. The composite curve is sketched on a pump curve or by using a spreadsheet by adding flow rates on a line of constant
32.24
Flow Coefficient
1/ C s
0.001
2
Head
2
30
Flow
0.0029
18.61
6
30
B-C
A-B
49.64 0.0004
85.98
Flow Coefficient
6
80
0.0001
2
Head
2
80
1/ C s
C-D
0.0024
20.39
5
30
C-D
0.0005
45.83
0.99
30
Comps
7E-05
121.59
1
80
Comps
B-C
19.22
2.5
20
Comps
A-B
Flow
1/ C s
C-D
0.0027
21.49
Flow Coefficient
B-C
0.0022
2
Head
2
20
Flow
A-B
0.0015
25.49
3.2
30
Coil
0.0002
70.20
3
80
Coil
0.0013
8.77
12
20
Coil
0.0048
14.42
10
30
Valve
Path 3
0.0003
60.79
4
80
Valve
Path 2
0.0013
8.77
12
20
Valve
Path 1
2.31
30
Balance
18.5
80
Balance
12
20
Balance
0.0024
20.39
5
30
E-F
E-F
E-F
Unbalanced Path Example
Unbalanced Path Flow
Table 8-3
0.0029
18.61
6
30
F-G
0.0004
49.64
6
80
F-G
F-G
0.001
32.24
2
30
G-H
0.0001
85.98
2
80
G-H
0.0022
21.49
2
20
G-H
0.0193
42.5
30
Total
0.0016
42.5
80
Total
0.033
42.5
20
Total
7.1923
Cv
6.994115
Cv
24.819
Cv
18.65097
Cv
5.5041
Cv
4.662743
Cv
248 Chapter 8 Matching Pumps to Systems
Fundamentals of Water System Design I-P
Figure 8-7
Unbalanced system curve intersection point with pump curve.
Figure 8-8
Effects of larger distribution pipes and resized coils and control valves.
249
250
Chapter 8 Matching Pumps to Systems head for the various points of the single pump curve. The points are then connected to represent a pump curve when two pumps are operational. At each head, the horizontal vector of flow is added to the first as follows: Y 1 + Y 2 = Total flow at B and X 1 + X 2 = Total flow at A A new pump curve is developed (Figure 8-9) on the paralleled pump curve line A-B-C. Plotting a system curve across the parallel pump curve shows the operating points for both single and parallel pump operation (see Figure 8-9). It is interesting to note that parallel pumping tends to flatten the combined pump curve to make the system pressure more suitable for control valve operation. Pumps should be identical in rating of flow and head, design, impeller diameter, and speed. It may cause problems if one of the pump’s characteristics is greater than those of the other (such as closing the discharge check valve), thereby making start-up unpredictable. Pumps of unequal pressures may result in one pump creating a pressure across the other pump in excess of its cutoff pressure, causing the flow through the second pump to reduce or
Figure 8-9
Pump curve developed on paralleled pump curve line A-B-C.
Fundamentals of Water System Design I-P
251
cease. This can cause flow problems or pump damage. By examination of the system and composite pump curves, when the system is in a design condition (meaning that flow may go to all connected paths), running a single pump yields more than 50% flow. While dissimilar pumps may be pumped in parallel, care should be taken so that when both pumps are operated together, the system flow does not fall low enough that the operation of a pump in the shadow of a larger pump occurs. When this occurs, it is possible for uneven forces and excess radial thrusts to be forced on the smaller pump, with the potential for cavitation. Typically two (or more) same-sized pumps are used and are preferred. The system curve should intersect both pump curves as shown in Figure 8-10. Where the system curve intersects the single pump curve, it will be further out on the curve than the nominal one-half design point conditions, yielding a flow greater than one-half of the design flow. This intersection point is sometimes referred to as the standby flow, and generally when a flat-curve pump is selected, the percentage of standby flow yields a high degree of design heat transfer with respect to the system. The system curve crosses the single pump curve to the left of the combined pumps’ operating point. This factor leads to two important points: • •
The pump motor must be adequately sized to prevent overloading during single pump operation (see Figure 8-10). The flow rate associated with where the pump system curve and the single pump curve intersect divided by the design flow rate yields a percentage of design flow for the system in single pump operation. A single pump may provide standby service as high as 80% of design flow; the actual amount depends on the specific pump curve and the system curve. Recall that in a heating system designed for 20°F T, 75% of design flow yields 97.5% heat transfer and 60% yields about 90%.
The piping of parallel pumps (Figure 8-11) should be laid out with provision to run either pump with a bypass around the other. Hand valves (gate or butterfly) must be manually positioned, or two-position automatic control valves can be tied in with the pump selection controls. The check valve in the pump discharge closes when the pump is shut down. The alternate pump draws from the return bypass and discharges into the supply bypass.
Series Pumping When pumps are operated in series, each operates at the same flow rate and provides its share of the total pressure at that flow (see Figure 8-12). At each flow, the vertical vector of head is added to the first (shown as Xl + X2 = total head at A, and Yl + Y2 = total head at B, etc.) until a new pump curve is developed (curve C-A-B).
252
Chapter 8 Matching Pumps to Systems
Figure 8-10
Pump motor sized to prevent overloading during single pump operation.
Figure 8-11
Piping schematic of parallel pumps.
Fundamentals of Water System Design I-P
Figure 8-12
253
Pump curve for series operation.
The series pump curve is drawn with full flow at low head and then with double the head of the single pump curve at each flow value to construct a similar pump curve. A system curve plotted across the series pump curve shows the operating points for both single and series pump operation (see Figure 8-13). Note that the single pump provides up to about 80% flow as a standby and at a lower power requirement. As with parallel pumps, it is important that pumps in series be identical in rating of flow and head, design, impeller diameter, and speed. Pumps of different flow capacity connected in series can result in problems; the pump of greater capacity can overflow the pump of lesser capacity, causing damage in the smaller pump due to cavitation. This can also cause a pressure drop rather than a pressure rise across that pump. It is important to discuss the proposed parallel, series, or multiple pump selection with the pump manufacturer to prevent a potential problem and to get the manufacturer’s suggestions and horsepower requirements. The piping of series pumps (see Figure 8-14) should be laid out with provision to run either pump with a bypass (B-C in Figure 8-14) around the other. Hand valves (gate or butterfly) must be manually positioned, or two-position automatic control valves can be selected.
254
Chapter 8 Matching Pumps to Systems
Figure 8-13
Piping schematic of series pumps.
Figure 8-14
Operating conditions for series pump installation.
Fundamentals of Water System Design I-P
255
Standby Pumps It is always good practice to consider a backup pump of equal capacity and proper valves to permit operation when the normal pump is inoperable. Usually this is an application for a parallel pump (see Figure 8-15). Failure often occurs at the worst possible time, meaning that heating pumps fail in extremely cold weather or chilled-water or condenser water pumps fail in the middle of a hot spell during the cooling season. The original investment costs of bypass piping and pumps will be trivial compared to the inconvenience or damage for the building occupants or the operator. Standby pumps are similar to parallel pumps except that they are sized for full design flow operation. This is why sizing for parallel operation is considered a convenience: it cuts down the pump cost with almost no impact on system heat transfer or temperature control when properly considered.
Trimming Pump Impellers and Adjusting Pump Speed The pump affinity laws were introduced in Chapter 4. You will remember that • • •
Figure 8-15
flow is directly related to pump impeller diameter, head is related to the square of the pump impeller diameter, and power is related to the cube of the pump impeller diameter.
Piping schematic of standby pump.
256
Chapter 8 Matching Pumps to Systems As shown in Table 4-1, Chapter 4, BHP1 = original pump horsepower, hp BHP2 =
new pump horsepower, hp
D1
=
original pump diameter, in.
D2
=
new pump diameter, in.
h1
=
original head, ft
h2
=
new head, ft
N1
=
original pump speed, rpm
N2
=
new pump speed, rpm
Q1
=
original flow, gpm
Q2
=
new flow, gpm
These conditions in Table 4-1 hold true as long as the pump head curve is for a centrifugal pump and the system curve is followed. Based on these relationships, optimizing the installed pump construction so that the impeller matches the system (rather than matching the system to the pump by providing excessive throttling losses) is something that should be carefully considered. This can be done in two different ways, by varying either the speed of the pump or the diameter of the impeller. Depending on the connected horsepower, variable-speed drives can be a highly economical method of getting system adjustment with little commissioning headache. In 2011, variable-speed drives could be purchased for as little as $150 for a 1/4 hp drive or $700 for a 10 hp drive. This is a pittance, especially considering that a 10 hp drive was thousands of dollars just a few decades ago. Considering that energy codes tend to require the use of a drive on the pump at about 5 hp as a rule of thumb (about $400), the speed drive, even if left uncontrolled by a separate controller, allows for the maximum pump speed to be limited to an amount equivalent to a reduced-size impeller, thus saving horsepower after efficiencies are accounted for. The alternative adjustment is to trim the pump impeller. In this operation, the pump is operated, the system is commissioned and adjusted, and, when convenient, the system is shut down with the pump isolated. The pump is disassembled, removing the impeller from the pump volute, and machined, reducing the impeller diameter using the affinity laws. The pump is then reassembled, and any throttling adjustments that were made during system commissioning are removed from the system by opening the pump throttling device. If we were to examine a modified pump curve to show the difference in effects from speed adjustment versus impeller trimming, the curves would look exactly the same. Either method is an acceptable method for reducing pump horsepower to just that necessary to make the system operate properly. Prior to any of these adjustments, it is imperative that the operating point of a pump be considered when the system terminals operate at full load and part load with thermostat control of two-way valves. In Figure 8-16, point 1 shows the pump operating at the design flow and the calculated design pressure drop of
Fundamentals of Water System Design I-P
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the system. Typically, the actual system curve is slightly different than the design curve. As a result, the pump operates at point 2 and produces a flow rate higher than design. In the process of balancing the system, a terminal balancing valve in the most hydraulically significant circuit should be open, or throttled only enough to provide a measurable differential pressure to calculate terminal flow. All other circuits should be proportionally balanced with respect to this circuit. To reduce the actual flow to the design flow at point 1, a balancing valve downstream from the pump can be adjusted when all terminal valves are in a wide-open position. This pump discharge balancing valve imposes a pressure drop equal to the pressure difference between point 1 and point 3. This alternative simply adds extra head loss to the system by throttling a pump discharge valve to limit maximum pump flow after the system has been proportionally balanced. Lest we think that this is not energy efficient, a simple comparison of operating horsepower in the adjusted and nonadjusted states shows that whenever flow is reduced in a centrifugal pumping system, operating energy use will be reduced. However, that reduction is not optimal in the sense that getting the
Figure 8-16
Pump operating points.
258
Chapter 8 Matching Pumps to Systems pump to provide only the energy required to the system without extra throttling will save more pump energy than throttling alone. A note on this, though: depending on just how flawed a system and its connected pump are, it may be impossible to get the pump adjusted so that it provides only the minimum energy required to create the design flow in the system. In these cases, minor throttling losses and pump speed or impeller modification might be used to match the connected system and the pump. The manufacturer’s pump curve will dictate exactly how much the capacity may be reduced by substituting a new impeller with a smaller diameter, trimming the existing pump impeller, or adjusting the speed. The manufacturer’s pump curve shows the boundary limits to impeller sizes in a particular volute. Pump speed, on the other hand, can usually be taken down to a minimum of 30% speed, a larger range than normally is achieved by trimming the impeller. After adjustment, reopening the balancing valve in the pump discharge eliminates artificial drop and the pump operates at point 3 (see Figure 8-16). Points 3 and 4A demonstrate the effect that a trimmed impeller has on reducing flow (Figure 8-16).
Two-Speed Pumping Multiple-speed motors are also an option to reduce system overpressure at reduced flow. Standard two-speed motors are available in models with speeds of 1750/1150 rpm, 1750/850 rpm, 1150/850 rpm, and 3500/1750 rpm. These are common options on small fractional horsepower circulating pumps. In larger pumps, as previously mentioned, application of a variable-speed drive may be much more economical on a single-speed motor rather than purchasing the two-speed motor. However, as mentioned regarding the pump affinity laws (Chapter 4), •
flow is directly related to pump speed,
•
head is related to the square of the pump speed, and
•
power is related to the cube of the pump speed.
Figure 8-17 shows the performance of a system with a 1750/1150 rpm multiple-speed pump. In Figure 8-17, curve A shows the system’s response when the pump runs at 1750 rpm. When the pump runs at 1150 rpm, operation is at point 1 and not at point 2 (as pump affinity laws calculate), because the pump head curve (at 1150 rpm) and the system curve must be satisfied. Another concern is if the system were designed to operate as shown in curve B, the pump would operate at shutoff and be damaged if it was run at 1150 rpm. This example demonstrates that the designer must analyze the system carefully to determine the pump’s limitations and the effect of lower speed on performance.
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Pumps with Two-Speed Motors (Stethem 1988) If a pump has a two-speed motor, a simple approach to reducing pump capacity is selecting a relay to direct the power to the high or low windings of the pump motor. Figure 8-18 shows an example of a pump performance curve for a 6 × 8 × 9.5 in. pump. If this application calls for a capacity of 1670 gpm and 50 ft head (at point C in Figure 8-19), a two-speed motor of 1750/ 1150 rpm will permit this flow to be reduced to 985 gpm and 26.3 ft (point A). If the example further calls for a design increase of 500 gpm in the future, the two-speed pump can be combined with a duplicate two-speed pump in parallel to give the results shown in Table 8-4. If 2170 gpm represents the ultimate load, then you can see this arrangement permits four flow selection steps to reduce pumping power at a minimum of equipment investment:
Figure 8-17
Two-speed pumping.
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Figure 8-18
Typical performance curve—6 in. suction × 8 in. discharge × 9.5 in. impeller.
Figure 8-19
Two-speed pumping example.
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261
100% gpm at 100% power point D 77% gpm at 57% power point C 58% gpm at 27% power point B 45% gpm at 16% power point A
Variable-Speed Pumping To understand what happens in a constant-speed pumping system, look at a direct-return system with two-way control valves (see Figure 8-20). At full load, all two-way control valves are wide open and the pump operates at point A (see Figure 8-21) to deliver 1100 gpm at 100 ft total system head loss. The design head loss is 80 ft for piping, coil, and fittings, and 20 ft for the control valve. As
Table 8-4
Four Flow Selection Steps to Reduce Pumping Power
Pumps, rpm
Capacity, gpm
Head, ft
bhp
Both at 1750
2170
75.0
48.0
One at 1750
1670
50.5
27.5
Both at 1150
1250
33.0
13.0
One at 1150
985
26.3
7.5
Figure 8-20
Typical direct-return system.
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Figure 8-21
Head loss with two-way valves at full load and 50% part load.
the load reduces, the two-way valves assume new positions (under command of their zone or room thermostats), and the system flow is reduced to 50% (to 550 gpm). The system curve moves along the pump head curve from point A to point B. This means that the total system head loss is 110 ft, of which 90 ft must be absorbed by the control valve and 20 ft is the head loss for piping, coil, and fittings at 50% flow. The control valve must be chosen to hold its position against this high differential. If one wants the total system to operate down to extremely low flows (10% to 25%), then the valve actuators have to be chosen for an even higher h (100 to 110 ft). This is a serious concern because unit valves may be mounted in limited space conditions of unit cabinets where larger-diameter actuators may not fit. An increase in valve cost is another factor, and so the problem requires other system design features. Staging fewer parallel pumps off or selecting a lower pump speed should be considered. Staging fewer pumps may result in an operating cost penalty.
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Because variable-speed drives have become more economical, their use for controlling pump speed is now a typical application. Generally, the economics are so favorable that building energy codes now make them a requirement starting at a very low horsepower, often around 5 hp. A pump controller is installed, typically using a proportional-integral-derivative (PID) controller where differential pressure becomes the controlled variable and pump speed is the manipulated variable through the variable-speed drive. Differential pressure is the most common control application; however, other variables could be controlled to implement the designer’s control concept. Differential pressure is most common because it closely aligns with the primary functional response of the building system (space temperature control for comfort), as discussed in Chapter 5. Differential pressure is sensed at the most hydraulically significant point in the system, which is the flow path with the greatest head loss used to select the pump. The controller acts to maintain a constant differential pressure across a portion of the piping system that must include a modulating control valve. As the valve opens and closes, the P controller senses the difference between the measured value and the design head loss for the controlled segment of pipe and transmits a proportional control signal (usually 0 to 10 vdc or 4 to 20 ma) to the variable-speed drive to vary the pump speed proportionally to maintain the desired P setpoint. The device may be calibrated in differential pressure for feet head or psi. The sensor/transmitter continually sends a signal to the controller, making this a feedback control application. In some applications, multiple sensors/transmitters are located in different zones and are used to maintain each zone’s setpoint. That said, with only one signal output to increase or decrease a pump speed, multiple zone setpoints would suggest that only one could be satisfied at any given moment. Typically, controllers that incorporate multiple sensors select the zone with the greatest control error, using that as the PID controller input for variable-speed drive adjustment. Note: To ensure continuous signal monitoring within direct digital control (DDC) systems, it is recommended that the differential pressure sensor be hardwired to the PID controller rather than transmitted across a digital network (see Figure 8-22). In a basic direct-return system, a sensor/transmitter is usually located at the end of the last riser (see Figure 8-23) or in a zone with the highest estimated drop to maintain a set differential, nominally 20 or 30 ft. As the controller changes the pump speed to maintain the P (H) setpoint, it creates a control curve (see Figure 8-24) maintained by the control system as the load on the system is varied by the zone or local loop temperature controls. Reviewing the pump affinity laws, we see in Figure 8-25 how the pump speed directly varies the pump head curve; if flow is reduced to 1/2, then the head is reduced to (1/2)2 = 1/4. Figure 8-26 shows the horsepower reduction with pump speed; as flow is reduced to 1/2, the power is reduced to (1/2)3 = 1/8. We can see how the variable-speed pumping system then reduces high h at part loads but also takes advantage of the pump laws to reduce the operating cost. The electric motor should be selected to be equivalent to a National Electrical Manufacturers Association (NEMA) class B design, and the motor manufacturer should be consulted regarding minimum practical operating speeds. In general,
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Figure 8-22
Proportional controller and adjustable-frequency drive controlling pump.
Figure 8-23
Direct-return system with sensor/transmitter located at end of last riser.
Fundamentals of Water System Design I-P
Figure 8-24
265
Differential pressure control curve above piping friction loss.
minimum speeds are recommended to be 30% of the maximum speed. The controller/drive is set for minimum pump speed to protect the pump motor and provide enough circulation to flush and cool the seals. In addition, as pumps become larger where couplers are used between the motor and the pump shaft, the pump manufacturer must be notified and the pump specified for variable-speed operation so that the appropriate coupler is applied. Constant-speed couplers tend to be hard rubbers or plastic materials. Variable-speed pump couplers incorporate a more flexible rubber material that is designed to absorb the speed changes. Typically, constant-speed couplers are not tolerant of constant changes in speed, and the coupler can be ground up as each individual shaft’s mating piece impacts the coupler material. The more flexible coupler is noticeably larger and more expensive and can absorb the twists that occur when the pump changes speed. Many pump controllers are of microprocessor design and are available with the following features: • • •
Staging multiple pumps to maintain the head-flow conditions to prevent a pump from becoming overloaded before extending beyond its power rating Automatic bypass controls for mechanical or electrical failure or power consumption/efficiency Operator alarm indications and warnings
Some of these functions are integrated with the building DDC system, and some are stand-alone dedicated pump control systems.
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Figure 8-25
Pump curve showing head reduction with change in pump speed.
Source Distribution Pumping Reviewing the basic concept of primary/secondary pumping (see Figure 8-27), we can see how this principle is applied to the source (such as a group of chillers or boilers) to meet the partial loads and bring on additional units to meet high loads. The primary/secondary concept allows the distribution pumping of the source supply from a central pumping facility (see Figure 8-28) or distributing the pumping to remote buildings or zones of a large facility (see Figure 8-29). Each case must be studied by the designer to determine the most economical operating cost versus the installed cost and the maintenance staff capabilities. In some cases, the staff may be limited to a central boiler or chiller plant operation. In other cases, remote locations of pumping stations may fit very well if the individual buildings have adequate mechanical equipment rooms and staff to maintain them. The primary/secondary concept allows continual flow through the source and still permits two-way valve control in the loads. There is flexibility in dedicating a pump to a chiller or boiler or manifolding the pumps. This might simplify the need for having backup pumps for every system. Pumps, manifolds, accessories, and associated pumping control may be assembled to
Fundamentals of Water System Design I-P
Figure 8-26
Pump curve showing pumping power reduction with change in pump speed.
Figure 8-27
Primary/secondary pumping concept.
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Figure 8-28
Main-source primary/secondary variable-speed pumping.
Figure 8-29
Distributed variable-speed pumping.
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match installation constraints or can be factory prepackaged as an assembly for a designated mounting location. The designer must weigh the pros and cons of cost, flexibility, and installation requirements of the various concepts to determine the best arrangement and cost for the project.
The Next Step The next chapter covers water chillers and load control.
Summary The pumps covered in this chapter include the following: • • • • • • • • •
Single pump: selected for a simple application Single pump with trimmed impeller: optimizing pump capacity for a specific application Single pump with backup pump: in addition to a selected application, provides 100% backup Two-speed pump: provides limited variable-flow steps with an added investment Parallel pumps: flexible capacity control without increasing system head, good for two-way valve control Series pumps: steep head change with limited flow change, two-way valves would require high differential pressure operation and capability Primary/secondary pumping: flexible zoning approach with minimum pumping energy Distributed pumps: special application of primary/secondary pumping Variable-speed pumps: applied to pumping systems to reduce power by lowering pump speed to meet control differential pressure in selected locations, usually applied to parallel pumping distribution systems using primary/secondary or distributed pumping with two-way control valves
In this chapter, we covered the following: • • • • • • •
The methods to match pump operation to a system That the parallel pump curve is constructed by adding flow capacity at each value of the pump head That the series pump curve is constructed by adding pump head at each value of pump flow How to provide emergency flow in case of a pump failure How to take advantage of capacity and head combinations provided by two-speed pumps What the pump controller measures to control pump speed The methods to provide flow to buildings that are remote from a chiller plant
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References and Bibliography ASHRAE. 2012. ASHRAE Handbook—HVAC Systems and Equipment. Atlanta: ASHRAE. Bell & Gossett. 1985. VIS Pumping Fundamentals. Technical Report 685. Morton Grove, IL: Xylem Inc. Hegberg, R. 1991. Converting constant-speed hydronic pumping systems to variable speed pumping. ASHRAE Transactions (97)1:739–45. Rishel, J.B. 1991. Control of variable-speed pumps on hot- and chilled-water systems. ASHRAE Transactions (97)1:746–50. Stethem, W. 1988. Application of constant speed pumps to variable volume systems. ASHRAE Transactions 94(2):1458–68.
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Skill Development Exercises for Chapter 8 Complete these questions by writing your answers on the worksheets at the back of this book. 8-1
Assume two pumps (6 × 8 × 9.5 in., see Figure 8-14) are to be applied in parallel for a 70 ft head system. What flow will result at 70 ft?
8-2
Assume the same two pumps (6 × 8 × 9.5 in., with a 9.6 in. impeller) are to be piped in series. What will the resulting head be at 1600 gpm?
8-3
Assume the 6 × 8 × 9.5 in. pump is being considered for variable-speed operation of a system designed at 1600 gpm at 70 ft head and the pump manufacturer advised that the NEMA type B electric motor to be furnished can be run down to 40% of its furnished speed (1750 rpm). What is the minimum flow and head the pump can be run at following the same system curve?
8-4
In Exercise 8-3, what is the theoretical brake horsepower reduction if the pump is run at 40% speed?
8-5
What methods are used to provide flow to buildings that are remote from a chiller plant?
8-6
What does the pump controller measure to control pump speed?
8-7
Explain how combinations of two-speed pumps can be used to provide required head and flow capacity with improved efficiency at part-load conditions.
8-8
What is a good way to provide emergency flow in case of a pump failure?
8-9
What methods are used to match pump operation to a system?
Water Chillers and Load Control
Study Objectives After completing this chapter, you should be able to • • • • •
• •
•
identify a water chiller’s capability to reduce water temperature, dependent on the tonnage rating and water flow rate; understand water temperature rise across a chilled-water (CHW) coil; understand temperature drop in condenser water temperature; understand the many types of refrigeration compressors used in chillers, depending on the manufacturer and the tonnage size; know that the theoretical horsepower used in a chiller is directly related to the heat (enthalpy) absorbed in the evaporator, the weight rate of refrigerant flow, the enthalpy change in the compressor, and the chiller tonnage; understand that chillers can be piped in series or parallel but are commonly in parallel to provide for expansion; understand piping of multiple chillers, using primary/secondary principles, and loading chillers evenly when locating the common bridge between the load and the chillers; and know how to optimize chiller operation versus CHW supply temperature, taking into account the compressor power and the pumping power.
Instructions Read Chapter 9 and answer all of the questions at the end.
Basic Water Chiller Components Water chillers provide cooling for water, brines, or other secondary coolants for air conditioning or refrigeration (ASHRAE 2014). The systems can be either factory assembled and wired packages or shipped as component sections and built up in the field. The basic components of a liquid chilling system include the compressor, evaporator (liquid cooler), condenser, refrigerant flow
274
Chapter 9 Water Chillers and Load Control control device, and a control center, as well as other auxiliary devices (receiver, intercooler, oil separator, etc.) within the chiller package (see Figure 9-1). Externally, the chiller must have a condensing water system (cooling tower, pumps, diverting valve, etc.) to transfer the heat properly. Figure 9-2 shows a schematic of a simple liquid chiller cooling a water system from 54°F to 44°F and transferring that energy to a condenser water system operating from 85°F to 95°F.
Refrigeration Cycle A pressure-enthalpy (p-h) chart (also called a Mollier diagram) is a portrayal of the pressure enthalpy values of the refrigerant. Figure 9-3 shows the lowpressure (evaporation) side versus the high-pressure (condensing) side of a basic system and is useful for visualizing the refrigeration cycle of the water chiller (Trane 2002). Exact values of the refrigerant pressures, temperatures, enthalpy, density, and specific volumes are found in manufacturers’ tables or in ASHRAE Handbook—Fundamentals (ASHRAE 2013). To understand the power used in a water chiller system, it is important to understand the basic refrigeration cycle and how compressor power and pumping power interrelate. In the p-h chart (Figures 9-3 to 9-5), temperature lines are constant with the pressure line between saturated liquid and the saturated vapor line, or the wet region. If the refrigerant is at point A in Figure 9-5, it
Figure 9-1
Basic components of liquid chilling system.
Fundamentals of Water System Design I-P
Figure 9-2
Schematic of simple liquid chilling system.
Figure 9-3
Pressure-enthalpy diagram for a refrigerant.
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Chapter 9 Water Chillers and Load Control
Figure 9-4
Simplified pressure-enthalpy diagram for a refrigerant.
Figure 9-5
Refrigeration cycle shown on simplified pressure-enthalpy diagram.
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absorbs heat with no change in pressure and will begin to boil, and evaporation will occur with no change in temperature. As heat is added at point D, the refrigerant’s enthalpy increases until it becomes saturated at point B. Further addition of heat at constant pressure moves the refrigerant condition into the superheat region to point C. Tracing the simple chiller shown in Figure 9-2 onto the p-h chart (Figure 9-3), it is shown that liquid is throttled by the expansion valve from the condenser pressure at point 4 in Figure 9-5 to the evaporator low pressure at point 1 at a constant enthalpy. As the refrigerant flows through the evaporator, it absorbs the heat necessary to completely vaporize it by point 2 (water chilling) and in practice is superheated. The compressor raises the vapor from low pressure at point 2 to high pressure at point 3. The high-pressure vapor is then condensed from point 3 to point 4, where the heat picked up in the evaporator plus the superheat and the heat of compression are transferred to the cooling-tower water. The Air-Conditioning, Heating, and Refrigeration Institute (AHRI) is the industry trade association that rates chillers at their operating conditions. The applicable standard is AHRI Standard 550 (2011). In this standard, operating points for evaluation are designated so that different chillers and vendors may be examined. AHRI rating points are as follows: • • • • •
Leaving CHW temperature: 44°F CHW flow rate: 2.4 gpm/ton Entering condenser water temperature: 85°F Condenser water flow rate: 3 gpm/ton Evaporator fouling factor and 0.00025 condensing fouling factor
Heat Transfer Chiller The size of the chiller is rated in tonnage or tons of refrigeration. The historic definition of a ton comes from making 1 ton of ice in 24 hours. 2000 lb 144 Btu 1 ton = ------------------ ------------------- = 12,000 Btu/h 24 h lb
(9-1)
Reviewing Chapter 1, we remember that the heat transferred to or from water is Q = q 500 t ent – t lvg
(9-2)
tons = 12,000 Btu/h = 500 q t
(9-3)
then
q t q t tons = ---------------------- = --------------24 12,000 ---------------- 500
(9-4)
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Chapter 9 Water Chillers and Load Control This may also be referred to in the following format, showing flow in gallons per minute per ton of cooling: q gpm ---------- = 24 -----ton t
(9-5)
This is a good formula for estimating flows versus t. If the liquid is a brine or glycol mixture and not water only, then the formula must include specific heat and specific gravity of the solution: q t c p sg tons = ------------------------------------24
(9-6)
In the condenser, the heat transferred to the condenser water includes the heat from the evaporator plus the heat of compression. For most practical comfort air-conditioning applications, a value of 14,400 Btu/h is used as the total heat transferred to the condenser water (Trane 2002). This may be referred to as a cooling tower ton. Following the same logic as for the evaporator, the short form for the condenser is gpm q ---------- = 28.8 ---------ton t
(9-7)
Example 9-1 What CHW flow will a 100 ton chiller handle for a 12°F rise in water temperature and a 8°F drop in tower water?
Solution
First, determine the flow in the evaporator (chiller flow): Flow q gpm ---------- = 24 -----ton t 100 tons 24 gpm = --------------------------------- = 200 gpm 12F Next, find the flow in the condenser: gpm 28.8 Flow q ---------- = ---------ton t tons 28.8 gpm = 100 ------------------------------------- = 360 gpm 8F
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Using these formulas, if you revisit AHRI rating points, you will find that the evaporator is rated for a 10°F t with entering water temperature of 54°F. Similarly for the condenser, a temperature difference of 9.6°F is calculated, leaving a leaving condenser water temperature of 94.6°F. The temperature difference is often rounded to 10°F t.
Refrigeration Power The theoretical power required by a water chiller compressor is a function of the refrigerant flow rate and the change in enthalpy during compression. The heat absorbed in the evaporator is Q = W RE Btu/min where RE = = = Q = W =
refrigerating effect heat absorbed in evaporator (h2 – h1) Btu/lb of refrigerant (see Figure 9-3) heat absorbed in the evaporator, Btu/min weight rate of refrigerant flow, lb/min
or tons 200 W = ------------------------- lb/min h2 – h1 The theoretical compressor power (THP) required is h2 – h1 W THP = --------------------------------HP 42.4 The compressor brake horsepower (BHP) required is THP BHP = -------------------------------------------------- HP overall efficiency* * Efficiency from manufacturer’s data.
Chiller Types and Control There are several types of chillers available for air conditioning: centrifugal, screw, scroll, and reciprocating. A general guideline for choosing the type depending on the chiller’s tonnage size is shown in Figure 9-6. Figure 9-7 shows a cross section of a centrifugal compressor. A section of a single-screw
280
Chapter 9 Water Chillers and Load Control refrigeration compressor is shown in Figure 9-8, and its compression process sequence is shown in Figure 9-9. Figure 9-10 shows the components of a specific type of twin-screw compressor. The scroll compression process is shown in Figure 9-11, and its components are shown in Figure 9-12. Figure 9-13 shows a reciprocating compressor refrigeration system.
Figure 9-6
General guideline of the types of chillers available for air conditioning. Source: ASHRAE Handbook—Refrigeration (2014).
Figure 9-7
Centrifugal compressor cross section.
Fundamentals of Water System Design I-P
Figure 9-8
Section of single-screw refrigeration compressor.
Figure 9-9
Sequence of compression process in single-screw compressor.
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Absorption refrigeration machines are available for water chilling in capacities of 3 to 5 tons (ammonia), 3 to 30 tons (lithium bromide), and 50 to 1500 tons (lithium bromide) (see Figures 9-14 and 9-15). Reciprocating compressors are used in smaller systems of up to 100 tons. An interesting discussion is presented in the 2012 ASHRAE Handbook—HVAC Systems and Equipment about the performance characteristics of reciprocating compressors compared to those of centrifugal and screw compressors (see Figure 9-16) (ASHRAE 2012). A distinguishing feature of the reciprocating compressor is its pressure rise versus capacity characteristic.
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Chapter 9 Water Chillers and Load Control
Figure 9-10
Vertical, discharge-cooled, hermetic twin-screw compressor.
Pressure rise has only a slight influence on the volume flow rate of the compressor; therefore, a reciprocating liquid chiller retains nearly full cooling capacity even on days above design wet bulb. It is well suited for air-cooled condenser applications and low-temperature refrigeration. Methods of capacity control are furnished by • • • • •
unloading compressor cylinders, on/off cycling of compressors, hot-gas bypass, compressor speed control, and a combination of the above.
Figure 9-17 shows the relationship between system demand and compressor performance with three equal steps of cylinder unloading. This illustrates the relationship between the system demand and performance of a
Fundamentals of Water System Design I-P
Figure 9-11
283
Scroll compression process. Source: Purvis (1987).
compressor with three steps of unloading. As cooling load drops to the left of fully loaded compressor line A (Figure 9-17), compressor capacity is reduced to that shown by line B, which produces the required refrigerant flow. Because cooling load varies continuously, whereas machine capacity is available in fixed increments, some compressor on/off cycling or successive loading and unloading of cylinders is required to maintain fairly constant liquid temperature. In practice, a good control system minimizes load/unload or on/off cycling frequency while maintaining satisfactory temperature control. Two basic piping arrangements for multiple chiller systems are parallel and series CHW flow (as described in the 2014 ASHRAE Handbook—Refrigeration). In the parallel arrangement, the water to be chilled is divided among the liquid chillers and combined again in a common header after chilling (see
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Chapter 9 Water Chillers and Load Control
Figure 9-12
Bearings and Other Components of Scroll Compressor Source: Elson et al. (1990).
Figure 9-18). As the cooling load decreases, one unit may be shut down, but the remaining units must then provide colder-than-design CHW so that when the streams combine, the design water supply is provided. Usually the idling chiller’s pump is shut down when the chiller is stopped and a check valve closes in this pump’s discharge to prevent a bypass flow. In the case of water chilling designs above 45°F, all units should be controlled by the combined leaving water temperature or by the return water temperature, because overchilling will not cause a dangerously low water temperature in the operating machine. In the case of water-chilling designs below 45°F (see Figure 9-19), each machine should be controlled by its own CHW temperature, both to prevent dangerously low evaporator temperatures and to avoid frequent shutdowns by the low-temperature cutout.
Fundamentals of Water System Design I-P
Figure 9-13
Reciprocating compressor refrigeration system.
Figure 9-14
Two-shell lithium bromide cycle water chiller.
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Chapter 9 Water Chillers and Load Control
Figure 9-15 Schematic of double-effect, direct-fired absorption chiller with reverse parallel flow cycle.
Figure 9-16
Comparison of single-stage centrifugal, reciprocating, and screw compressor performance. Source: ASHRAE Handbook—HVAC Systems and Equipment (2012).
Fundamentals of Water System Design I-P
Figure 9-17
287
Reciprocating liquid chiller performance with three equal steps of unloading.
Figure 9-18 Parallel arrangement: water to be chilled is divided among liquid chillers and combined again in common header after chilling.
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Chapter 9 Water Chillers and Load Control
Figure 9-19
Series arrangement.
In this case, the temperature differential setting of the return water temperature must be carefully adjusted to prevent short cycling caused by the step increase in CHW temperature when one chiller is cycled off. No overchilling is required and compressor power consumption is lower than it is for the parallel method at part loads. In the series method, a valved piping bypass is suggested around the lead and lag chillers to facilitate future servicing. This piping design philosophy also applies to parallel chillers; it again gives the maintenance staff an opportunity for service without a complete shutdown.
Chiller Piping Arrangements In general, we see multiple chiller designs applied in parallel instead of series because this permits adding chillers in the future due to single or multiple building additions, such as for campus-type site plans. The designer must allow space in the chiller plant for the additions and the distribution mains that must be sized for the future flows (see Figure 9-20). Until the mid 1990s, chiller controllers were mainly pneumatic or electric proportional operation, with switched interlocks. To protect the operating components from freezing due to load changes, chillers were operated with constant water flow, typically represented by the three-way valve. These applications usually used three-way valves on unit terminals to permit continual flow in the chiller (see Figure 9-21) and to permit part-load terminal control by shifting coil flow to bypass flow around the coil. The result is that the CHW pumping power was theoretically constant and did not reduce with load. As the size of the system increased, the distribution system pumping increased in size, pumping a constant volume, and was not energy efficient.
Fundamentals of Water System Design I-P
Figure 9-20
Chillers piped in parallel in primary production loop.
Figure 9-21
Schematic showing constant- and variable-flow arrangements.
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Chapter 9 Water Chillers and Load Control During the first energy crisis era of the early 1970s, application of the twoway valve was given new opportunity to reduce the distribution pumping flow rate with the load in large campus-oriented systems. Popular primary/secondary pumping systems were installed with greater vigor, keeping flow to the chillers constant through the primary pump while distribution flow to the system was made variable through the secondary pumps by the closing of the valves riding the pump curve for reduced horsepower. System flow is controlled through engineering decisions that allow the valves to perform properly, reducing excess flow due to high valve differential pressures by staging parallel pumps and judicious pump selection and layout. In the late 1980s, speed drives became cost-effective enough for large systems to apply variable-speed pumping control. The chillers continued to be piped in parallel in a primary production loop (see Figure 9-20) with a common bridge to hydraulically decouple the chiller pumps from the distribution pumping. Distribution pumping was kept at constant speed and constant flow to maintain the integrity of the chiller controller. Chiller staging was determined from recirculation and at part-load if chillers should be shut down. The location of the common bridge determines how the chillers will be loaded or unloaded (Coad 2011). Since the 1980s, several simultaneous advances in the industry have occurred, shifting system design thoughts. Direct digital control (DDC) replaced pneumatic and electric proportional controls, ushering in new capabilities in advanced control strategies and applied monitoring of system operating data. Adjustable-speed drive prices became much more economical due to rising energy costs and manufacturing economies of scale. In this era a renewed focus on energy efficiency took hold, and with it new capabilities were added to chillers and sources in general, allowing for better control at partial loads, which allowed for variable flow to the source. With measurable data, known temperature operating problems in the primary/secondary system occurring from primary and secondary mixing are now catching the attention of the industry. Some of the difficulties can be seen in Figure 9-22, which represents a system consisting of three chillers. Each chiller is fed by a constant-speed chiller pump (with a check valve) that operates only when the chiller is on. Each chiller pump is sized to achieve a t of 16°F, and the chiller is controlled by a discharge thermostat at 42°F. With the common bridge located between the load and the production sections (see Figure 9-23) and the system operating at part load (1200 gpm, 800 ton load) due to the hydraulics inherent in the design, chillers 2 and 3 will load proportionally and chiller 1 will shut down. Chillers 2 and 3 receive the same-temperature water from the return main. The chillers load in proportion to the ratio of their flow rates to the total load flow. Chiller 2 loads to 320 tons and chiller 3 loads to 480 tons (each is 64% of their full load output, because the load is 800/1250 = 64% of the combined chiller flow rate). Because the chiller pumps are constant speed, chiller 2 delivers 750 gpm and chiller 3 delivers 1125 gpm, for a total of 1875 gpm. Because the load is calling for 1200 gpm, 675 gpm must flow in the common bridge.
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Figure 9-22
Three chillers in parallel, each with a dedicated constant-speed pump.
Figure 9-23
Parallel arrangement with common bridge between load and production sections.
Mixing occurs at the return tee, creating a temperature of 675 42F + 1200 58F -------------------------------------------------------------------------- = 52.24F 1875 The load on chiller 2 is 750 52.24F – 42F tons = ---------------------------------------------------------- = 320 tons 24
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Chapter 9 Water Chillers and Load Control and on chiller 3 is 1125 52.24F – 42F tons = ------------------------------------------------------------- = 480 tons 24 If the common bridge is located on the opposite end of the production section (see Figure 9-24), the chillers will load unevenly due to the hydraulics. Looking at this part-load example, chiller 3 loads fully: 58F – 42F tons = 1125 ------------------------------------------------------ = 750 tons 24 Chiller 2 receives 75 gpm (1200 to 1125) from the return main, which is mixed with 675 gpm (750 to 75) from the common bridge at 42°F. This water enters chiller 2 at 43.6°F: 75 58F + 675 42F - = 43.6F ------------------------------------------------------------------750 The load on chiller 2 is only 750 43.6F – 42F tons = ------------------------------------------------------- = 50 tons 24
Figure 9-24
Parallel arrangement with common bridge at opposite end of production section.
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Chiller 1 shuts down due to the 42°F return water. So you can see the uneven loading that can occur and why the designer should consider the common bridge between the load section and the production section. What is the best way to handle this? This is an introductory course, so know that there are exceptionally sophisticated models and great debate about system design and optimization, and there will be many fast-moving changes in the industry that this text cannot keep up with. But the fundamentals of system design are established here, and even in new and sophisticated plant design, all of the techniques discussed are quite applicable. Sources, and particularly chillers, still have some limits to their flexibility with respect to operation. Where previous designs required constant flow, today’s systems allow variable fluid flow, although it tends to be in a limited range. In many cases, chiller controls are able to control with chiller tube velocities down to a range of 1 to 2 fps; however, concerns over tube fouling often increase this low limit to around 3 fps to reduce the possibility of fouling. Maximum velocity can often be stated as high as 10 fps; however, this is for limited hours of operation. Chiller heat exchanger tubes are sophisticated pieces of manufacturing engineering, often being rifled on the interior and exterior to increase the heat transfer surface area. The result is great performance; the proviso is that while 10 fps tube velocity for limited time periods is acceptable, a more common operating limit of 8 fps is recommended to reduce tube erosion, performance degradation, and system maintenance. Still, if chiller tubes are selected and operated in a range of 3 to 8 fps, there is a serious capability to reduce energy consumption in the pumping system and the chiller operating energy. A point will be reached where minimum flow must be maintained for the chiller to operate properly. The simplest implementation to create a more efficient system, then, is to consider the benefits of the primary/secondary system with the new technological twist, making it variable-speed primary/variable-speed secondary. Primary/secondary pumping techniques do increase the cost of new systems slightly because there are more pumps. On the other hand, primary/secondary techniques allow for the designer to manage system constraints due to loads and hydraulics normally attributed to adapting the HVAC system to the facility to be controlled better than any other options. As previously mentioned, primary/secondary systems allow various system components to be hydraulically isolated from each other, and controlling system pressure differences is the first step in attaining flow control performance at part load that allows the chillers or other sources to operate at reduced loads and energy consumption. There are a lot of plain systems that may not require very sophisticated techniques; on the other hand it is often the case that the larger systems become, the more techniques such as primary/secondary become important tools in a designer’s tool box. These are particularly noticeable whenever there is a possibility that system size will be expanded. As seen in Figure 9-25, the primary/secondary allows for simple addition of extra chillers and loads through proper up-front planning. Another option to this implementation is to manifold the pumping system
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Chapter 9 Water Chillers and Load Control
Figure 9-25
Primary/secondary allows for simple addition of extra chillers and loads.
to the sources, giving additional flexibility for maintenance and failure operation modes. Often, pumps have the capability to provide more system flow, and in the manifolded system the benefits of parallel pumping can be taken advantage of, as diagrammed in Figure 9-26. In these applications, the designer should determine the sequence of operation as a combined temperature control and hydraulic problem as the staging of chillers and pumps is considered. Properly maintained (and selected) hydraulic relationships allow for zone controls to act independently and for the source system to react to the loads as indicated by the demanded flow from the system. Not all control points are diagrammed. Temperatures are shown in Figure 9-25 to specifically monitor and control against system degradation due to secondary remix and resupply back to the system. Varying of chiller flow and minimum flow control can be handled by the addition of flow, temperature, differential pressure data, and closed-loop feedback control strategies, or open-loop control tied in with the chiller control panel. In smaller systems of one to two chillers, or systems where expansion will not be required, the variable primary system is also an option.
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Figure 9-26 Pumping system manifolded to sources gives flexibility for maintenance and failure operation modes as well as capability to provide more system flow.
Dedicated operating pumps for the chillers are eliminated, and the distribution pumps are selected for a larger head to account for the chillers. The common bridge is maintained in the system, with a bypass control valve to ensure that there is the capability to maintain the minimum flow required by the chiller. There are several ways that minimum flow and chiller staging can be accomplished, either through addition to the diagrammed schematic of flow or differential pressure sensors. It is also recommended that minimal chiller loading be monitored in the control package to properly sequence pumps and chillers.
Chiller Energy Performance The commonly applied evaluation of a chiller’s performance is kilowatt of energy per ton of refrigeration (kW/ton). Plotted against the percentage of design capacity of the chiller, as shown in Figure 9-27, which shows a realworld example, this curve shows the improvement in efficiency that a particular manufacturer accomplished in redesign to make the chiller more energy efficient over time. Older chiller systems have a challenge operating at less than 1 kW/ton. Manufactured equipment today has embraced the application of variable-speed drives on compressors and sophisticated control strategies, in part because of competition and in part due to legislated requirements for
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Chapter 9 Water Chillers and Load Control
Figure 9-27
Chiller performance improvement versus percent load.
energy efficiency improvement. Today’s chillers now normally achieve efficiencies on the order of 0.56 kW/ton and, depending on their sophistication of application, can operate as low as 0.2 kW/ton at reduced loads. In a simple form, Figure 9-28 (Mannion 1988) shows a control and design strategy for higher ts, various control valves, pumping methods, and chiller staging options for a hypothetical design. These options always require a great deal of study and due diligence. It is not at all uncommon to see detailed and complex modeling and mapping of the total system cost and operating expense to determine the lowest life-cycle cost for system selection. As one might imagine, the competition among those desiring this type of work can be quite ferocious and lead to many claims and counterclaims. It is also important while performing due diligence to be aware of proprietary features and design strategies that may require the designer to consider patent licensing. It is not at all uncommon to find specific operating strategies and control sequences granted patent protection. What might seem like a well-meaning educational article in a trade journal about what might be considered common practice or a new idea could often be a veiled advertisement for a protected concept.
Fundamentals of Water System Design I-P
Figure 9-28
297
Chiller, control, and pumping alternatives versus design temperature rises. Source: Figure 10, Mannion (1988).
Energy performance of a CHW system should not be limited to the most efficient chillers, piping, valves, and variable-speed pumping arrangements. There is a need for skilled, trained operators working with some form of DDC system to optimize the overall chiller system by optimizing the following (Cascia 1988): •
The staging on or off of multiple chillers
• •
Condenser water temperature and cooling-tower operation CHW temperature when variable-speed CHW pumps are used
Resetting CHW supply temperature upward under lighter load conditions saves energy at the chiller due to lower refrigerant head requirements of the compressor. This reset can be done through cascade control of the supply influenced by the return water temperature or incorporation of outdoor temperature reset (or both). However, increasing the CHW temperature can cause the variable-speed pumps to increase in speed due to the two-way valves opening to satisfy the load. How much that happens depends on the speed of response of the system and how far the reset influence extends into the system. For
298
Chapter 9 Water Chillers and Load Control instance, under lighter loads, the designer might also have the zone controller setpoint reset slightly higher (1°F to 2°F), which is barely noticeable to the occupant, or might apply that criteria as a high load day as occupancy begins to come to a close. Figure 9-29 shows the relationship between chiller power consumption and the variable-speed pump power consumption and how the optimal CHW temperature can be found.
Thermal Storage Another important design consideration is to provide a thermal storage facility as part of the CHW system to reduce construction costs and operating costs. The principles of thermal storage can be applied with the primary/secondary chiller piping and variable-speed pumping discussed. It is also suggested that ASHRAE publications such as Cool Storage Modeling and Design (1989) and “Achieving Energy Conservation with Ice-Based Thermal Storage” (Brady 1994) be studied for further design considerations. Overall improvement in water chiller performance and selection for a particular design requires careful study by the designer, not only in various chiller manufacturers’ performance data, pumping and piping design arrangements, control valve selections, and pump speed control, but also in how they perform together from part load to full load.
Figure 9-29 Relationship between chiller power consumption and variable-speed pump power consumption.
Fundamentals of Water System Design I-P
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An often overlooked thermal storage concept that requires little if any additional cost is application of the building mass, or fabric, as the storage vessel. Thermal storage often does not save energy in the form of kilowatt-hours. It does, however, reduce energy cost by shifting use from peak electrical demand points to off peak and, in the case of a chiller, a usually lighter operating load condition (e.g., hottest time of the day with solar loading to night and about 20°F cooler). In the building mass concept, the idea is to run the chiller and HVAC plant off peak (usually at night) to precool the building and lower the mass temperature. The discharge of the stored energy is controlled by manipulation of the zone controller’s setpoint, operating cooler in the morning and slowly raising the zone setpoint going into the loaded portion of the day (afternoon). In the facility that the author is familiar with and where this was implemented, occupants were made aware of the strategy and told that if they were too cool in the morning they could wear a sweater. The facility was built with two side-by-side, equal but separate buildings that shared a connecting interstitial space. To reduce costs, no excess or safety capacity was installed in the CHW plant. Changing facility uses and requirements made the owner consider installation of a full-sized backup CHW plant to be shared by the two buildings. The ultimate result of the storage implementation was that during an extended one-week period when design conditions were exceeded by greater than 20°F, the tested building was able to operate at peak operating load with 75% of the plant’s installed capacity while maintaining occupancy setpoint. The owner was able to revise the backup plan to a simpler piping cross-connection so that each individual operating plant could back up the other in case of required backup. Figure 9-30 shows a more typical example of a thermal storage concept using an ice builder with a CHW system. This is often the way most designers think of thermal storage, and it too helps reduce installed system costs and operating expense. Thermal storage, regardless of the style employed, should always be considered as part of the base system design.
Summary In this chapter, we covered the following: • • • • •
•
Identification of a water chiller’s capability to reduce water temperature, dependent on the tonnage rating and water flow rate Temperature rise in water temperature across a CHW coil Temperature drop in condenser water temperature The types of refrigeration compressors used in chillers, depending on the manufacturer and the tonnage size That the theoretical horsepower used in a chiller is directly related to the heat (enthalpy) absorbed in the evaporator, the weight rate of refrigerant flow, the enthalpy change in the compressor, and the chiller tonnage That chillers can be piped in series or parallel but are commonly in parallel to provide for expansion
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Chapter 9 Water Chillers and Load Control
Figure 9-30
Schematic of thermal storage concept using ice builder with a chilled-water system.
•
•
Piping of multiple chillers, using primary/secondary principles and loading chillers evenly when locating the common bridge between the load and the chillers How to optimize chiller operation versus CHW supply temperature, taking into account the compressor power and the pumping power
References AHRI. 2011. AHRI Standard 550/590, Standard for Performance Rating of Water-Chilling and Heat Pump Water-Heating Packages Using the Vapor Compression Cycle. Arlington, VA: Air-Conditioning, Heating, and Refrigeration Institute. ASHRAE. 1989. Cool Storage Modeling and Design. Technical Data Bulletin. Atlanta: ASHRAE. ASHRAE. 2012. ASHRAE Handbook—HVAC Systems and Equipment. Atlanta: ASHRAE. ASHRAE. 2013. ASHRAE Handbook—Fundamentals. Atlanta: ASHRAE.
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ASHRAE. 2014. ASHRAE Handbook—Refrigeration. Atlanta: ASHRAE. Brady, T. 1994. Achieving energy conservation with ice-based thermal storage. ASHRAE Transactions 100(1):1735–45. Cascia, M. 1988. Optimizing chiller plant energy savings using adaptive DDC algorithms. ASHRAE Transactions 94(2):1937–46. Coad, W. 2011. Hydronic Systems. Technical data bulletin. Atlanta: ASHRAE. Elson, J., G. Hundy, and K. Monnier. 1990. Scroll compressor design and application characteristics for air conditioning, heat pump, and refrigeration applications. Proceedings of the Institute of Refrigeration, 2.1–2.10. Mannion, G. 1988. High temperature rise piping design for variable volume systems: Key to chiller energy management. ASHRAE Transactions 94(2):1424–43. Purvis, E. 1987. Scroll compressor technology. Heat Pump Conference, New Orleans. Trane. 2002. Air Conditioning Manual. Atlanta: Trane Company.
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Skill Development Exercises for Chapter 9 Complete these questions by writing your answers on the worksheets at the back of this book. 9-1
What pumping flow is required for a 400 ton CHW system using 20% ethylene glycol and water at 44°F supply and 12°F rise? (Per Table 6 of Chapter 31 of the 2013 Handbook—Fundamentals, the density of a 20% ethylene glycol/ water solution at 44°F is 64.6 lb/ft3.)
9-2
What is the pumping flow in Exercise 9-1, except using propylene glycol and water at the same freezing conditions? (Per Table 9 of Chapter 29 of the 2013 ASHRAE Handbook—Fundamentals, the density of a 20% propylene glycol/ water solution at 44°F is 64.0 lb/ft3.)
9-3
Estimate the volumetric flow rate of condensing water to be pumped for the condenser of an R-22 water-cooled unit operating at a condensing temperature of 110°F and an evaporating temperature of 40°F with a 10°F liquid subcooling and 10°F suction superheat. Water enters the condenser at 85°F and exits at 95°F; the load is 400 tons (use the figure below).
9-4
In Figure 9-17, at what load does chiller 1 shut down?
9-5
The capacity of a chiller is dependent on what two basic load factors?
9-6
What are three advantages of a DDC system for a multiple-chiller plant?
Fundamentals of Water System Design I-P
303
9-7
What type of control valves should be used with variable-speed CHW pumps?
9-8
Name several methods that manufacturers furnish for refrigeration capacity control.
9-9
Name at least four types of chillers.
Skill Development Exercises To receive full continuing education credit, all questions must be answered and submitted at www.ashrae.org/sdlonline. Please log in using your student ID number and the SDL number. Your student ID number is composed of the last five digits of your Social Security Number or another unique five-digit number you create when first registering online. The SDL number for this course can be located near the top of the copyright page of this book.
Fundamentals of Water System Design I-P
Total number of questions: 11 1-1
Water systems that convey heat to or from a conditioned space or process with hot or chilled water are frequently called _____________________________.
1-2
What is the fundamental difference between closed and open types of water systems?
1-3
A cooling tower has at least two points of interface with air. Where are they?
1-4
What is the maximum working pressure for LTW boiler systems?
1-5
What is a CHW system? How is it different from a CW system?
1-6
What are the fundamental components of a closed hydronic system?
Chapter 1 Skill Development Exercises
Skill Development Exercises for Chapter 1
Chapter 1 Skill Development Exercises
Skill Development Exercises Chapter 1 1-7
Explain the most common source devices for heating and cooling systems.
1-8
Explain what load means.
1-9
What factors influence the heating and cooling load requirements?
1-10
Define sensible heat transfer.
1-11
Name five heating load devices and describe how each is used in system applications.
Fundamentals of Water System Design I-P
Total number of questions: 9 2-1
What causes unbalanced flow rates in direct-return piping arrangements?
2-2
What is the most often used equation that relates to pressure drop?
2-3
Fluids can flow through a pipe under two different conditions. Name them. Explain the differences between these flow conditions.
2-4
Explain friction factor.
Chapter 2 Skill Development Exercises
Skill Development Exercises for Chapter 2
Chapter 2 Skill Development Exercises
Skill Development Exercises Chapter 2 2-5
In commercial installations, it is suggested that _____% should be added to the friction loss to allow for aging.
2-6
What is the Bernoulli principle?
2-7
What factors determine pressure drop in piping?
2-8
What methods allow thermal expansion?
2-9
What is the minimum distance upstream and downstream for a water flowmeasuring device (in pipe diameters)?
Fundamentals of Water System Design I-P
Total number of questions: 10 3-1
What US organizations issue codes and standards for piping systems and components?
3-2
What is the allowable working pressure (ASTM A53B to 4000°F) for each of the following steel pipe diameters: Nominal Size and Pipe OD
Schedule
10 in., D = 10.75
40XS
20 in., D = 20
30XS
4 in., D = 4.5
40ST
8 in., D = 8.625
80XS
Working Pressure, psig
3-3
List three joining methods used with copper tubing.
3-4
List five methods of joining metal pipe.
Chapter 3 Skill Development Exercises
Skill Development Exercises for Chapter 3
Chapter 3 Skill Development Exercises
Skill Development Exercises Chapter 3 3-5
Name and briefly describe the two main categories of plastic piping materials, and list at least three of each type of plastic pipe.
3-6
List the ASME B31 HDS (psi at 73°F) for each of the following plastic pipe materials: Material
ASME B31 Hydrostatic Design Stress (psi at 73°F)
PB 2110, Type II, Gl ABS 1210 TI, G2
3-07
3-8
3-9
List the five methods of corrosion control.
List the k values for each of the following screwed pipe fittings: Nominal Pipe Diameter, in.
Fitting Type
1.25
90° ell long
0.5
Globe valve
3
Tee branch
What is the function of a backflow prevention device?
k Value
Fundamentals of Water System Design I-P 3-10
• 1 heating coil rated at 35 gpm and 3 ft drop • 1 gate valve: 2 in. • 1 control valve: 2 in. rated at 35 gpm at 9 ft PD • 1 balance valve: 2 in. rated at 40 gpm at 2.0 ft PD • 2 tee branches • 4 elbows Total pipe length = 200 ft (schedule 40 screwed pipe)
Diagram for exercise 3-10.
Chapter 3 Skill Development Exercises
Determine the pipe sizing and total PD for the piping system shown below. You are given that the system consists of:
Fundamentals of Water System Design I-P
Total number of questions: 13 4-1
List three factors that influence the type of pump selected for a particular application.
4-2
Label the components of the centrifugal pump shown below. Describe the function of each component.
4-3
The suction flange gage of a pump reads 10 psi. If the temperature of water being pumped is 220°F and the atmospheric pressure is 14.7 psia, what is the available NPSH? At what temperature will the pump cavitate? (Water at 220°F vaporizes at 17.2 psia, per ASHRAE Handbook—Fundamentals [2013], Table 3, p. 6.10.)
Centrifugal Pump.
Chapter 4 Skill Development Exercises
Skill Development Exercises for Chapter 4
Chapter 4 Skill Development Exercises
Skill Development Exercises Chapter 4 4-4
What is the NPSH on the inlet to a 2 hp pump rated at 140 gpm at 32 ft for a cooling tower application? The centerline of a pump inlet is to be 2.5 ft below the tower sump water surface; assume tower water at 120°F and piping equivalent to 60 ft of 2 in. pipe on pump suction. Assume atmospheric pressure is 14.7 psia or 34.0 ft; assume friction head in 2 in. pipe is 20 ft (according to Figure 2-15, 140 gpm flow in 2 in. pipe yields 33.2 ft/100 ft of pipe: 60/100 × 33.2 = 19.92, or 20 ft) and vapor pressure of water at 120°F is 1.69 psia (per ASHRAE Handbook—Fundamentals, Table 3, Chapter 1) or 3.9 ft abs (1.69 psia × 34 ft/14.7 psi = 3.9 ft abs). Pump curve shows 10 ft NPSHR.
4-5
What is radial thrust?
4-6
Explain what pump cavitation is and how it can be avoided.
4-7
Write the NPSHA formula for a proposed design and explain what each variable represents.
Fundamentals of Water System Design I-P Write the pump affinity laws and explain how they are applied.
4-9
Speed Change Diameter Change:
Flow:
Head:
Horsepower:
4-10
Explain how to determine the horsepower for a centrifugal pump.
4-11
How does the capacity of a centrifugal pump change?
Chapter 4 Skill Development Exercises
4-8
Chapter 4 Skill Development Exercises
Skill Development Exercises Chapter 4 4-12
Name six types of centrifugal pumps and their mounting arrangements.
4-13
A pump is rated at 500 gpm at 60 ft of head. What are the flow and head if the impeller size is changed to 85% of its original diameter? Assume there is no static head.
Fundamentals of Water System Design I-P
Total number of questions: 14 5-1
A typical fan-coil terminal requires 4 gpm. What valve Cv should be selected if a control valve is specified for a 9 ft drop?
5-2
What piping components should be specified at the terminal?
5-3
What type of control action should be considered to reduce discharge temperature cycle?
5-4
A control valve is to be selected for a 44 gpm terminal coil; coil drop is 18 ft. Select the correct size of control valve if the valve is specified for 50% of the coil drop, and the typical valve sizes and Cv available (Cv is in parentheses) are • •
0.5 in. (2.5), 0.75 in. (6), 1 in. (10) 1.25 in. (16), 1.5 in. (20), 2 in. (36)
Chapter 5 Skill Development Exercises
Skill Development Exercises for Chapter 5
Chapter 5 Skill Development Exercises
Skill Development Exercises Chapter 5 5-5
What control valve flow characteristic should be specified for proportional control of a hot water heating control?
5-6
An on/off thermostat controls a cabinet heater in a hallway. What valve flow characteristic should be specified?
5-7
A three-way valve is to be applied to a refrigeration condenser and cooling tower to maintain a 95°F condensing temperature. What type of three-way valve arrangement should be applied?
5-8
It is desirable to control flow in a chilled-water coil down to a minimum of 5% of design flow before close off. In addition to proper valve sizing for design flow capacity and proportional control, what else should be specified?
5-9
What should be specified in the bypass circuit of a three-way valve?
Fundamentals of Water System Design I-P What type of control method varies airflow through a terminal coil?
5-11
Explain the difference between primary and secondary pumping systems.
5-12
Define valve authority, rangeability, and selection.
5-13
How is terminal control valve size selected?
5-14
What are the three types of terminal control action?
Chapter 5 Skill Development Exercises
5-10
Fundamentals of Water System Design I-P
Total number of questions: 13 6-1
What maintains the maximum and minimum pressure limits of a hydronic system?
6-2
What must a closed water system have to permit the expansion and contraction of the water volume?
6-3
What should be specified for terminal coil returns and high points in the piping system to provide service for air in the hydronic system?
6-4
Where does air come from in a hydronic system?
6-5
What routine tasks should be performed by the building maintenance staff to the hydronic systems prior to the heating and the cooling seasons?
Chapter 6 Skill Development Exercises
Skill Development Exercises for Chapter 6
Chapter 6 Skill Development Exercises
Skill Development Exercises Chapter 6 6-6
How much air can be present in water at 120°F and 30 psig?
6-7
A steel pipe system has 1000 gal total volume and will operate periodically in the cooling season with 40°F water, but when it is off it will reach 100°F ambient; minimum pressure is 10 psig (24.7 psia) and maximum is 25 psig (39.7 psia). What size diaphragm expansion tank is required?
6-8
A pump is selected for 100 gpm at 50 ft head, the system has 20 psig static pressure when off, and the expansion tank is improperly located on the pump discharge. What will the gages read on the pump suction and discharge when the pump is started? What will happen at the float-type air vents in the boiler room?
6-9
Explain what factors are needed to size and select an expansion tank.
6-10
What are the differences between open and closed systems?
Fundamentals of Water System Design I-P What are the three types of expansion tanks?
6-12
Give three examples of where expansion tanks are used.
6-13
Why is it important to have the correct size expansion tank?
Chapter 6 Skill Development Exercises
6-11
Fundamentals of Water System Design I-P
Total number of questions: 17 7-1
In the direct-return example (see Figure 7-8), which unit’s piping path dictates the pump head from A to F? How much head is required?
7-2
What size pump capacity and head would be required to handle the four AHUs in Figure 7-8 for the conditions shown from A to F?
7-3
What size pump capacity and head would be required in Figure 7-8 (direct return for supplying four identical floors), assuming 10 ft between floors and allowing a 25% head drop in the main for fittings and a 20 ft drop across the chiller?
7-4
In the reverse-return example in Figure 7-9, which unit’s piping path dictates the pump head from A to F? How much head is required?
Chapter 7 Skill Development Exercises
Skill Development Exercises for Chapter 7
Chapter 7 Skill Development Exercises
Skill Development Exercises Chapter 7 7-5
What size pump capacity and head would be required to handle the four AHUs in Figure 7-9 for the conditions shown from A to F?
7-6
What size pump capacity and head would be required in Figure 7-9, assuming 10 ft between floors and allowing a 25% head drop in the main for fittings and a 20 ft drop across the chiller?
7-7
In the primary/secondary examples in Figures 7-10 and 7-11, which unit’s piping path dictates the pump head from A to F? How much head is required?
7-8
What size distribution pump capacity and head would be required to handle the four AHUs in the Figure 7-11 primary/secondary example for the conditions shown from A to F?
Fundamentals of Water System Design I-P What size distribution pump capacity and head would be required for supplying four identical floors (similar to the Figure 7-11 primary/secondary example), assuming 10 ft between floors and allowing a 25% head drop in the main for fittings and a 20 ft drop across the chiller?
7-10
What is the cooling capacity of 100 gpm of water flow with 50% propyleneglycol mixture at an average temperature of 50°F and a 10°F rise?
7-11
What is the cooling capacity of 100 gpm of water flow at a 10°F rise? How many tons of cooling?
7-12
What is the pumping horsepower for the propylene-glycol mixture (specific gravity = 1.05) in Exercise 7-10, compared to the plain water in Exercise 7-11, if the coil pressure drop is 20 ft, assuming a pump efficiency of 75% and motor efficiency of 85%?
Chapter 7 Skill Development Exercises
7-9
Chapter 7 Skill Development Exercises
Skill Development Exercises Chapter 7 7-13
How can the possibility of frozen coils and piping in an HVAC system be reduced?
7-14
What are two advantages of primary and secondary hydronic systems?
7-15
What type of pump curves should be selected if two-way valves are to be used?
7-16
How can flow be determined in a given loop?
7-17
Which piping system design gives more uniform pressure drop? Why?
Fundamentals of Water System Design I-P
Total number of questions: 9 8-1
Assume two pumps (6 × 8 × 9.5 in., see Figure 8-14) are to be applied in parallel for a 70 ft head system. What flow will result at 70 ft?
8-2
Assume the same two pumps (6 × 8 × 9.5 in., with a 9.6 in. impeller) are to be piped in series. What will the resulting head be at 1600 gpm?
8-3
Assume the 6 × 8 × 9.5 in. pump is being considered for variable-speed operation of a system designed at 1600 gpm at 70 ft head and the pump manufacturer advised that the NEMA type B electric motor to be furnished can be run down to 40% of its furnished speed (1750 rpm). What is the minimum flow and head the pump can be run at following the same system curve?
8-4
In Exercise 8-3, what is the theoretical brake horsepower reduction if the pump is run at 40% speed?
Chapter 8 Skill Development Exercises
Skill Development Exercises for Chapter 8
Chapter 8 Skill Development Exercises
Skill Development Exercises Chapter 8 8-5
What methods are used to provide flow to buildings that are remote from a chiller plant?
8-6
What does the pump controller measure to control pump speed?
8-7
Explain how combinations of two-speed pumps can be used to provide required head and flow capacity with improved efficiency at part-load conditions.
8-8
What is a good way to provide emergency flow in case of a pump failure?
8-9
What methods are used to match pump operation to a system?
Fundamentals of Water System Design I-P
Total number of questions: 9 9-1
What pumping flow is required for a 400 ton CHW system using 20% ethylene glycol and water at 44°F supply and 12°F rise? (Per Table 6 of Chapter 31 of the 2013 Handbook—Fundamentals, the density of a 20% ethylene glycol/ water solution at 44°F is 64.6 lb/ft3.)
9-2
What is the pumping flow in Exercise 9-1, except using propylene glycol and water at the same freezing conditions? (Per Table 9 of Chapter 29 of the 2013 ASHRAE Handbook—Fundamentals, the density of a 20% propylene glycol/ water solution at 44°F is 64.0 lb/ft3.)
9-3
Estimate the volumetric flow rate of condensing water to be pumped for the condenser of an R-22 water-cooled unit operating at a condensing temperature of 110°F and an evaporating temperature of 40°F with a 10°F liquid subcooling and 10°F suction superheat. Water enters the condenser at 85°F and exits at 95°F; the load is 400 tons (use the figure below).
9-4
In Figure 9-17, at what load does chiller 1 shut down?
Chapter 9 Skill Development Exercises
Skill Development Exercises for Chapter 9
Chapter 9 Skill Development Exercises
Skill Development Exercises Chapter 9
9-5
Figure 9-1 Diagram for Exercise 9-3. The capacity of a chiller is dependent on what two basic load factors?
9-6
What are three advantages of a DDC system for a multiple-chiller plant?
9-7
What type of control valves should be used with variable-speed CHW pumps?
Fundamentals of Water System Design I-P Name several methods that manufacturers furnish for refrigeration capacity control.
9-9
Name at least four types of chillers.
Chapter 9 Skill Development Exercises
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ASHRAE LEARNING INSTITUTE Self-Directed Learning Course Evaluation Form Course Title: Fundamentals of Water System Design (I-P Edition), Second Edition On a scale of 1 to 5, circle the number that corresponds to your feeling about the statements below. (1 = strongly agree, 5 = strongly disagree, 3 = undecided) Strongly Agree
Course Content 1. The objectives of the course were clearly stated. 2. The course content supported the stated objectives. 3. The content quality and format of the course material make it valuable as a future reference. 4. The quality and clarity of the charts and diagrams enhanced your ability to understand the course concepts. 5. The organization of course material supported effective mastery of the topic. 6. The material presented will be of practical use to you in your work. 7. The degree of difficulty (level) of this course was correct to meet your needs and expectations.
Strongly Disagree
Undecided
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General 1. Which description best characterizes your primary job function? _____Architect*
_____Developer
_____Manufacturer
_____Sales
_____Code Agency
_____Educator/Research
_____Marketing
_____Specifier
_____Consultant
_____Energy Conservation
_____Owner
_____Student
_____Contractor/Installer
_____Facilities Engineer
_____Plant Engineer
_____Utilities
_____Government
_____Policy Maker/Regulator
_____Consumer/User
_____Other (please specify) _______________________________________________________________________ *Are you a registered architect? ___No ___Yes, AIA Membership Number (required): _____________________ 2. Which describes your educational background? _____High School
_____Master's Degree—Engineering
_____Associates Degree/Certificate Program
_____Master's Degree—Other Than Engineering
_____Bachelor's Degree—Engineering Technology
_____Doctoral Degree—Engineering
_____Bachelor's Degree—Engineering
_____Doctoral Degree—Other Than Engineering
_____Bachelor's Degree—Other Than Engineering _____Other (please specify) _______________________________________________________________________
3. Approximately how many hours did it take you to complete this course? _____10 hours
_____20 hours
_____30 hours
_____40 hours
_____Other (please specify)___________
4. What topics would you suggest for future courses? ______________________________________________________________
General comments regarding any aspect of the course, including suggestions for improvement: _________________________________________________________________________________________________________ _________________________________________________________________________________________________________ _________________________________________________________________________________________________________ _________________________________________________________________________________________________________ _________________________________________________________________________________________________________ _________________________________________________________________________________________________________ _________________________________________________________________________________________________________ _________________________________________________________________________________________________________ _________________________________________________________________________________________________________ _________________________________________________________________________________________________________ _________________________________________________________________________________________________________ _________________________________________________________________________________________________________ _________________________________________________________________________________________________________ _________________________________________________________________________________________________________ _________________________________________________________________________________________________________ _________________________________________________________________________________________________________ _________________________________________________________________________________________________________ _________________________________________________________________________________________________________ _________________________________________________________________________________________________________ _________________________________________________________________________________________________________ _________________________________________________________________________________________________________
Name (optional) __________________________________________________________________________________________ Phone (optional) __________________________________________________________________________________________ E-mail (optional) __________________________________________________________________________________________ Return to: ASHRAE, Education Department, 1791 Tullie Circle NE, Atlanta, GA 30329 Fax: 404-321-5478
Flexible and Effective Continuing Education for HVAC&R Professionals Key Knowledge for Successful Hydronic System Design Water, or hydronic, systems are a key foundation technology of HVAC system design. They represent an important tool in the designer’s toolbox for providing effective and efficient HVAC systems to buildings around the world. Fundamentals of Water System Design, Second Edition, is an entry-level text on the principles of hydronic HVAC system design and calculations. The fundamentals presented in this text can be used to successfully design the modern hydronic system, and they provide an equally sound footing for work in hydronic system commissioning, operation, and troubleshooting. This course book’s nine chapters provide detailed information on the following topics: • • • • • • •
Components of closed and open hydronic systems Pipe materials and fittings Piping system design Pumps and their operation Matching pumps to systems Control of terminal units Optimizing water-chiller operation
Each chapter includes detailed diagrams, examples, and calculations, plus exercises at the end of each chapter to develop skills in using the knowledge provided. This book is part of the ASHRAE Learning Institute (ALI) Fundamentals of HVAC&R Series of self-directed or group learning courses. For information on topics covered in the series, visit www.ashrae.org. 1791 Tullie Circle Atlanta, GA 30329-2305 Telephone: 404/636-8400 Fax: 404/321-5478 E-mail: [email protected] www.ashrae.org/ali 978-1-936504-66-4
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Product Code: 98020 11/15
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