ASHRAE Fundamentals of Air System Design SI 2010 9781933742878

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Fundamentals of Air System Design Robert McDowall

SI

International System

A Course Book for Self-Directed or Group Learning

Includes Skill Development Exercises for PDH, CEU, or LU Credits



Fundamentals of Air System Design

Robert McDowall

A Course Book for Self-Directed or Group Learning ASHRAE

ASHRAE Fundamentals of

HVAC&R Series

Fundamentals of Air System Design SI A Course Book for Self-Directed or Group Learning ISBN 978-1-933742-87-8 SDL Number: 42878 © 2010 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.

Print Fundamentals of... Psychrometrics Air System Design Steam System Design

No part of this publication may be reproduced without permission in writing from ASHRAE, except by a reviewer who may quote brief passages or reproduce illustrations in a review with appropriate credit; nor may any part of this book be reproduced, stored in a retrieval system, or transmitted in any way or by any means (electronic, photocopying, recording or other) without permission in writing from ASHRAE. Requests for permission should be submitted at www.ashrae.org/permissions. 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.

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Fundamentals of... HVAC Control Systems HVAC Systems Refrigeration Standard 62.1-2010, Ventilation for Acceptable Indoor Air Quality Standard 90.1-2004, Energy Efficiency in New Buildings

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Karen M. Murray

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Manager  Professional Development

Dear Student,  Welcome to an ASHRAE Learning Institute (ALI) self-directed or group learning course. 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 via the ALI online-monitoring system, please complete the exercises in the workbook, then submit your answers at www.ashrae.org/sdlonline. To log in, please enter your student ID number and the SDL number. Your student ID number can be the last five digits of your Social Security number or another unique 5-digit number you create when first registering online. The course number is located near the top of the copyright page of this book. Please keep copies of your completed exercises for your own records. When you finish all exercises, you will receive a link to submit a course evaluation. Once the evaluation is completed, you will be sent 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 ASHRAE AN INTERNATIONAL ORGANIZATION

Robert McDowall is a professional engineer who has been involved in building services for over forty years and has taught courses on heating, ventilating, and air conditioning for thirty years. He has been responsible for the construction and maintenance of buildings for IBM and the University of Manitoba. An active member of ASHRAE, Robert served as a member of the board of directors from 1997–2000, as a member of the Standard 62 committee, and as author of the online courses Fundamentals of HVAC Systems (with course book) and Fundamentals of HVAC Control Systems.

Table of Contents Preface . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .vii Acknowledgments . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . ix Chapter 1: Fundamentals of Airflow . . . . . . . . . . . . . . . . . . . . . . . . . . 1 Static and Dynamic Compressible Fluid (Air) Laws . . . . . . . . . . . . . . 1 Friction Effects . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 15 The Friction Chart . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 19 Density and Altitude Effects . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 20 Chapter 2: Air Distribution System Components . . . . . . . . . . . . . . . 27 Air Distribution System Overview . . . . . . . . . . . . . . . . . . . . . . . . . 27 Air-Handling Units . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 27 Ducts . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 32 Controls . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 32 Air Distribution Devices . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 32 Sound Absorbers. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 33 Chapter 3: Human Comfort and Air Distribution . . . . . . . . . . . . . . . 41 Principles of Human Comfort . . . . . . . . . . . . . . . . . . . . . . . . . . . . 41 Principles of Space Air Distribution . . . . . . . . . . . . . . . . . . . . . . . . . 50 Types of Air Distribution Devices. . . . . . . . . . . . . . . . . . . . . . . . . . 61 Chapter 4: Relationship of Air Systems to Load and Occupancy Demands . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 73 Operating System Selection Criteria. . . . . . . . . . . . . . . . . . . . . . . . 73 System Types by Heating/Cooling Equipment Type . . . . . . . . . . . . 75 System Type by Duct Configuration. . . . . . . . . . . . . . . . . . . . . . . . 91 Air-Side Economizers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 92 Outdoor-Air Intake . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 94 Chapter 5: Exhaust and Ventilation Systems . . . . . . . . . . . . . . . . . . 101 Design Considerations. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 101 Ventilation and Exhaust Systems. . . . . . . . . . . . . . . . . . . . . . . . . . 105 Energy Recovery . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 110 Chapter 6: Air Movers and Fan Technology . . . . . . . . . . . . . . . . . . 119 Fan Principles. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 119 Fan Drives. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 130 Fan Selection . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 130 Fan Installation Design . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 139 Fan Controls . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 148 Effect of Variable-Resistance Devices . . . . . . . . . . . . . . . . . . . . . . 151

vi

Table of Contents Chapter 7: Duct System Design . . . . . . . . . . . . . . . . . . . . . . . . . . . Duct System Design Overview . . . . . . . . . . . . . . . . . . . . . . . . . . Duct Materials . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Duct Construction. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Duct Design and Sizing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Sample Systems. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Chapter 8: Codes and Standards . . . . . . . . . . . . . . . . . . . . . . . . . . . Building Code Requirements . . . . . . . . . . . . . . . . . . . . . . . . . . . . ASHRAE/IES Standard 90.1 . . . . . . . . . . . . . . . . . . . . . . . . . . . . . ASHRAE Standard 62.1 . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Other Codes and Standards . . . . . . . . . . . . . . . . . . . . . . . . . . . . Sources of Information . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Chapter 9: Air System Auxiliary Components. . . . . . . . . . . . . . . . . Dampers. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Air Filters. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Humidifiers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Duct Heaters . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Duct Insulation . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Chapter 10: Sound and Vibration in Air Systems. . . . . . . . . . . . . . . Fundamentals of Sound . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Sound and Vibration Sources . . . . . . . . . . . . . . . . . . . . . . . . . . . . Sound Attenuation. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Vibration Control. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Chapter 11: Air System Startup and Diagnostics. . . . . . . . . . . . . . . Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Design Considerations. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Air Volumetric Measurement Methods. . . . . . . . . . . . . . . . . . . . . Balancing Procedures for Air Distribution Systems . . . . . . . . . . . . Noise and Vibration Diagnostics. . . . . . . . . . . . . . . . . . . . . . . . . . Chapter 12: A Duct Design Problem. . . . . . . . . . . . . . . . . . . . . . . . Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Duct Design Procedure . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . The Building and System . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Working through the Problem . . . . . . . . . . . . . . . . . . . . . . . . . . . Skill Development Exercises

159 159 159 161 167 182 193 193 195 197 198 201 205 205 210 221 225 226 233 233 239 242 255 262 263 264 264 267 275 283 283 283 284 287

Preface Fundamentals of Air System Design SI is an introductory guide to designing systems for the transfer of air in buildings and is intended for an audience of professional engineers or engineering students. For this SI version, standards and guidelines have been updated, damper information has been revised based on research project information, and Chapter 10, “Sound and Vibrations in Air Systems,” has been completely rewritten. The course begins with a presentation of airflow fundamentals and then proceeds to an overview of various system components that may be included, such as fans, filters, and coils. Human comfort and the sizing and location of air supply outlets are addressed to enable the designer to develop a system for effective thermal, ventilation, and comfort control. Methods of duct sizing are discussed, and detailed design calculations are demonstrated using the constant-pressure-drop method. Design issues for sound and vibration control are introduced, followed by a discussion of the main industry standards governing air systems. Finally, a system design is presented in a step-by-step format that allows the reader to make design decisions and calculations and then check them against those provided in the text.

Acknowledgments Production of this SI version of the course has been a joint effort. I am particularly grateful to Tino Mendez for reviewing the changes, catching mistakes, and making helpful suggestions. ASHRAE associate editor, Matt Walker, did an excellent job as editor and in encouraging the project to completion.

Robert McDowall Winnipeg, MB December 6, 2010

Fundamentals of Air System Design

Fundamentals of Airflow Study Objectives After completing this chapter, you should be able to T T T T T

explain static pressure, velocity pressure, and total pressure and the relationship between them; calculate change in volume of air with change in temperature at constant pressure; calculate the approximate volume and temperature resulting from mixing airstreams; sketch and explain the psychrometric chart parameters of temperature, moisture, relative humidity, and specific volume; and explain duct frictional losses.

Instructions Read the material in Chapter 1. Verify the examples presented in the chapter with your own calculations. At the end of the chapter, complete the skill development exercises without referring to the text. Review those sections of the chapter as needed to complete the exercises.

Static and Dynamic Compressible Fluid (Air) Laws Because this course is designed to address the needs of people with varying backgrounds and experience, it is necessary to review the fundamental principles of fluid mechanics. Your understanding of these principles is essential to learning the applied system design concepts that follow in later chapters. The concepts are presented in the context of heating, ventilating, and airconditioning (HVAC) applications, and terms are defined as they are used in that field. This course makes use of three of the four basic principles of fluid mechanics: • • •

Fluid statics The Continuity Equation The Energy Equation

2

Chapter 1 Fundamentals of Airflow

Difference between Mass and Weight A fundamental and often confused point must be addressed here: the difference between mass and weight. Mass is a property of matter that is invariant with location. For example, the mass of the astronauts remained essentially the same during their trip to and landing on the moon. However, their weight changed dramatically. The relationship of mass to weight is given by Newton’s Law of Motion: Force  Mass  Acceleration In the context of this course, the force is the weight required to hold up the matter (that is, to keep it from falling), and the acceleration is the acceleration due to gravity. This acceleration due to gravity, g, is virtually constant at sea level on Earth at 9.81 m/s2. Note that the law is stated as a proportionality, so we must insert a proportionality constant to make it an operationally useful equation. The symbol for the proportionality constant that has been in use for generations is 1/gc. This is unfortunate because of its similarity with the symbol for acceleration due to gravity, g. The proportionality constant gc is actually used to convert the units of mass × acceleration to units of force. In the system currently used extensively in the HVAC industry, the value of gc is 9.81 kgm·m/ kgf ·s2. The reason for this choice is that the weight of a kilogram mass (kgm) is numerically equal to a kilogram force (kgf) at sea level. A kilogram force is equal to 9.81 newtons, the basic SI unit of force. A newton (N) is the force required to give a mass of one kilogram an acceleration of one metre per second squared. Two important rules result from this: •

Mass and weight are inherently different but related, and kilogram-mass (kgm) is completely different from kilogram-force (kgf).

•

The symbol gc is a units conversion factor.

As a thermodynamic property, density is the ratio of mass to volume. It has the units of kgm/m3 and is denoted by the symbol  (lower case Greek letter rho). At times, it is convenient to express the density as weight density, , (the lower case Greek letter gamma), and by this we mean the force of gravity on a unit volume of mass. The conversion is accomplished by multiplying by g (the acceleration due to gravity) and by taking into account the need for units conversion with the constant gc. Thus,  = g/gc. The units of  are

Fundamentals of Air System Design SI

3

⎛ kg m⎞ ⎛ m⎞ ρ ⎜ --------3-⎟ g ⎜ ----2⎟ kg f ⎝ m ⎠ ⎝s ⎠ --------------------------------- = ------3 m m⎞ ⎛ kg m ⎞ ⎛ ---g c ⎜ --------- ⎟ ⎜ 2⎟ ⎝ kg f ⎠ ⎝ s ⎠ Also, numerically g/gc = 1, because both have the value 9.81. Specific volume is the reciprocal of density and is defined as the volume of a unit mass of material. It is expressed in cubic meters per kilogram: v = m3/ kgm.

Fluid Statics Hydrostatic pressure is something we all experience in a swimming pool; recall how the pressure on your eardrums increases as you dive deeper in the pool. The pressure is due to two factors: • •

Atmospheric air pressure on the surface of the water The weight of a column of water equal to the depth below the surface

Imagine a column of water as shown in Figure 1-1. The weight is equal to the volume times the weight density, W = γhA. The force required to hold the column of fluid plus oppose the pressure of the atmosphere is W + pa. The force, F, is also the pressure at the point of application of force, F, times the area. In equilibrium, we write F = pA = paA + W = paA + γhA Therefore, p = pa+ γh. The pressure difference between A and B in Figure 1-1 is Δp = γΔh

(1-1)

Note that the density is the weight density, γ = (g/gc)ρ.

The Continuity Equation The Continuity Equation expresses the idea that all mass is accounted for; none is lost or created. Mass that enters a space also leaves the space, provided no change in the stored amount occurs. Filling a tank, or releasing gas from a compressed gas bottle, obviously are cases where stored mass changes. However, when air flows through a duct, or into and out of a fan, the amount of air flowing per unit time is the same at the inlet as at the outlet. Figure 1-2 depicts such a situation. The volume of air that passes through a cross section of the duct is given as VA, where V is the velocity and A is the area. Rationalize this by imagining that

4

Chapter 1 Fundamentals of Airflow

Figure 1-1

Fluid static system.

Figure 1-2

Continuity equation.

the flow rate would double if the area doubles or if the velocity doubles. The mass associated with a unit of volume is the density, ρ. Therefore, the mass flow rate is given by ρAV. The conservation of mass idea states that no change in mass flow rate occurs under steady conditions when there is no storage change. Considering two locations, 1 and 2, on the same duct, we can write ( ρAV ) 1 = ( ρAV ) 2 Under normal conditions in a short length of duct with no heating or cooling coils, the pressure and temperature changes are so small that the density is virtually constant. The Continuity Equation can then be written as

Fundamentals of Air System Design SI ( AV ) 1 = ( AV ) 2

5 (1-2)

In the HVAC industry, flow rate almost invariably means volume flow rate, not mass flow rate. The most common units are cubic meters per second, m3/s, and liters per second, L/s. Air behaves as a perfect gas, and the change in density is proportional to the absolute temperature. Absolute temperature in kelvins (K) is the temperature above absolute zero, which is –273°C. To convert from degrees Celsius to kelvins, add 273°. Thus if outdoor air is heated from 2°C to 22°C as it comes in through the air-conditioning system, the density, γ, changes from γ 2 to 2 + 273 γ 2 × --------------------- = γ 2 × 0.932 22 + 273 which is a 6.7% decrease in density. Similarly, on a hot day, cooling 40°C outdoor air down to 13°C will increase the density by 40 + 273-------------------= 1.094 or 9.4% 13 + 273 The difference is even more significant in a cold climate. For example, suppose it is January and the outside temperature is –35°C. The outdoor air is brought in over a heating coil and supplied at 24°C. The drop in density as the air is heated is from γ–35 to – 35 + 273 γ – 35 × ------------------------- = γ – 35 × 0.801 24 + 273 This means that the air will be approximately 20% less dense. Being 20% less dense, the air occupies 20% more space at our constant pressure. Thus 1000 L/s at –35°C when heated to 24°C becomes 1248 L/s at 24°C. The mass of air stays the same. However, with rising temperature, the volume increases as density drops. Remember, this is in the normal commercial and institutional HVAC system with very small pressure changes. A very common process in air conditioning is mixing airstreams. Suppose we have a situation where 11,000 L/s of air at 24°C from a space is being mixed with 3000 L/s of outdoor air at 5°C. We want to know what the resulting mixed-air temperature and volume will be. The industry practice, which works well for estimating and small temperature differences, is to assume that the volume, L/s, is equivalent to mass and use Equation 1-3: ( L1 ⁄ s × T1 ) + ( L2 ⁄ s × T2 ) = [ ( L1 ⁄ s + L2 ⁄ s ) × Tm ] For example, based on Equation 1-3, the resulting volume is 11, 000 + 3000 = 14, 000 L/s

(1-3)

6

Chapter 1 Fundamentals of Airflow ( 11, 000 × 24 ) + ( 3000 × 5 ) = [ ( 11, 000 + 3000 ) × T m ] Tm = 19.9°C, which is very close to the correct answer of 19.7°C. For wide temperature differences, the inaccuracy can be almost eliminated by adjusting the incoming L/s values to be the L/s at the initially calculated mixed temperature and recalculating the mixed temperature using the adjusted flow values.

The Energy Equation (First Law of Thermodynamics) The third principle we will use in this course is the Energy Equation, which is based on the idea that energy, like mass, is neither created nor destroyed. A major consequence of this idea is that the forms that energy takes are interchangeable; that one form can be converted into another. There is one caveat to this idea, based on the Second Law of Thermodynamics: heat cannot be completely converted into work in a cyclic process. The units of the forms of energy are many and varied. Authors, industries, and nations all seem to have a unique set of preferences and biases. However, because most studies of energy begin with a definition of mechanical work (force × distance), it is appropriate to say that the fundamental unit of energy is force × distance, or newton-metre, also called a joule (J). The joule is the energy equivalent of one amp flowing through one ohm for one second. One amp flowing through one ohm is a watt, so a watt per second is a joule. This unit is too small to be useful, so we normally talk about kilowatt hours. The kilowatt-hour (kWh) is the same energy as 1000 for 60 seconds × 60 minutes = 3,600,000 J, or 3.6 MJ. It is common to write energy conversion in terms of a unit of mass flowing, for example, kWh/kg, MJ/kg. Whatever the units used, all values must be expressed in the same units when comparing or adding. Conversion factors are available in many texts and reference books, and you are expected to obtain and use conversion tables competently. Following is a brief listing and discussion of the forms of energy: •

Work (W) results from a force applied through a distance in the direction of the force. It is also the result of a torque applied through an angular displacement. Work can be either internal or external. Machines such as pumps, fans, and compressors do mechanical work, Wm, on the fluid. Machines such as turbines produce mechanical work, Wm, performed by the fluid on an external machine, such as an electric generator. Fluid friction, Wf, can be considered to be work done by the fluid on the duct or obstruction in much the same way as aircraft engines do work to overcome air drag in a flying airplane. In the HVAC industry, frictional forces exist in ductwork as the air passes down a straight section, as it makes a turn, or as it passes through louvers or a heat exchanger. This work results in a loss of pressure that must always be compensated for by the fan.

Fundamentals of Air System Design SI

7

•

Flow work (pv) is energy supplied to the system when fluid crosses the boundary entering the system. It is also done by the system as fluid leaves. Consider the situation where air discharges from a compressed air tank and makes room for itself in the atmosphere. The air pushes away the atmosphere. The air occupies space, and the atmosphere must make way for it by becoming a little bit higher, therefore increasing its own potential energy. Flow work is always present, even though the amount done on the system at the inlet may be very nearly equal to the amount done by the system at the outlet. Flow work is always given by the product of pressure and specific volume, pv, for one kilogram of fluid.

•

Heat (Q) is the result of energy transfer due to a temperature difference. That heat can be transformed into work, and that work can be dissipated into internal energy and transferred as heat, constitute the main business of thermodynamics. Note that heat is not stored and it is not “contained” by a fluid. Heat is thermal energy in transit. It is defined only at the boundary of a system.

•

Internal energy (u) is often confused with heat, but they are different concepts. Internal energy is associated with molecular motion, molecular bonding, and other forms of molecular activity, such as spinning or rotation of molecules. Internal energy can have units of kWh/kg or MJ/kg. In an ideal gas and a liquid, u is directly related to temperature. For example, a 1 kWh increase in internal energy is represented by a 86.1°C rise in temperature for 10 kg of water.

•

Potential energy is energy that represents the work done on a mass in moving it in the Earth’s gravitational field. For example, if a 1 kgm book is elevated 1 m above a desk, work in the amount of 1 m·kgf has been done on it. This work can be recovered by lowering the book and raising a mass someplace else through linkages or pulleys. Or the force of gravity will accelerate the book if it is allowed to drop, and the potential energy will be converted to kinetic energy associated with the velocity. Potential energy is always measured relative to some datum of zero elevation: PE = (mg/gc)z, where z is measured relative to some assigned datum in the system. Kinetic energy results from motion. For example, an automobile traveling at 100 km/h has kinetic energy, as does a baseball thrown at 140 km/h. The kinetic energy is derived from a steady force applied through a distance required to accelerate the body from rest to a velocity V:

•

Kinetic Energy = mV2/2 J (m in kg, V in m/s) The Energy Equation is simply a balance of the forms of energy described above. It is assumed that the system is in steady state. If another form was found to be important (such as chemical energy in combustion), it could be added to the list. If one or more of the forms is not present or important, it can be dropped. If we include those discussed above, the Energy Equation is written as follows:

8

Chapter 1 Fundamentals of Airflow

2 g V W m – W f + Q + m u + pv + ----- z + -------gc 2g c

in

2 g V = m u + pv + ----- z + -------gc 2g c

out

(1-4)

In air systems, the mass is that of flowing air; heat is added (or removed at a specified rate) and work is done at a certain rate, such as in 10 kW motor driving a fan. The Energy Equation can be turned into a Rate Equation by considering the following: · • Heat as a rate, kJ/h or kW; Q · • Work as a rate, kW; W · • Mass as a mass flow rate, kg/h; m A dot over the symbol commonly indicates a rate. In practice, always put the time in seconds, minutes, or hours to ensure consistent units. Although various units for the energy terms have been suggested above, the units for all terms in the equation must be the same: 2 g · · · V · W m – W f + Q + m u + pv + ----- z + -------gc 2g c

in

2 g V · = m u + pv + ----- z + -------gc 2g c

out

(1-5a)

If this form of the Energy Equation is divided through by the mass flow rate, m· , the following form results, where heat and work are on a unit mass flowing basis: 2 g V---------W m – W f + q + u + pv + z + gc 2g c

in

2 g V---------= u + pv + z + gc 2g c

out

(1-5b)

The terms of the Energy Equation are depicted in Figure 1-3.

Static Pressure, Velocity Pressure, and Total Pressure Now let's discuss a run of ductwork with the following conditions: • • • •

No machines exist, so all work terms are zero. No heat transfer exists, because the duct air is the same temperature as the surrounding air. There are no significant changes in elevation, so z is constant. The internal energy, u, is essentially constant.

Fundamentals of Air System Design SI

Figure 1-3

9

The energy equation applied to a flow system.

In this case, we have the simpler form of the Energy Equation: 2

2

V pv + -------2g c

in

V = pv + -------2g c

(1-6a) out

The above circumstances exist for a pitot tube, as shown in Figure 1-4, where the flow comes to zero velocity (where the arrow indicates total pressure in direction of flow). The Energy Equation becomes 2

V pv + -------2g c

= [ pv ] total

(1-6b)

duct or static

Suppose further that the specific volume is constant because of the small pressure changes involved and that we change the specific volume to the mass density using v = 1/ρ, multiply through by gc/g, and replace (g/gc)ρ by the weight density, γ. Equation 1-6b then becomes 2

p--- V + ------γ 2g

duct

p = --γ

total

(1-6c)

The location duct could be anyplace, and we can say that the total duct pressure is constant in the absence of friction and significant heat transfer. The following is known as Bernoulli’s Equation: 2

p--- V + ------γ 2g

duct

= constant

(1-6d)

10

Chapter 1 Fundamentals of Airflow

Figure 1-4

Static and total pressures.

Among other things, Bernoulli’s Equation says that as the velocity goes up or down (perhaps due to area changes or takeoff air), the static pressure changes. Note that the units for Equation 1-6d as written are meters. These are pressure equivalents to the weight of a column of the fluid on a unit area. Thus the units are meters of air or meters of water, depending on the fluid actually flowing and not the instrument that is used for measuring. Returning to Equation 1-6c and multiplying through by the weight density, g, we define the velocity pressure and obtain 2

V p +  -------2g

duct

= p total

or p static + p velocity = p total Examining the units of the velocity pressure term, we find that

2 kg f -------- = --------V 3 2g m

2 2 kg f m  sec --------------------- = --------- = Pa 2 2 m  sec m

(1-6e)

Fundamentals of Air System Design SI

11

where Pa is the symbol for pascals, kilogram force per square metre. Here again, the density is for the fluid flowing. Note that the relationship between velocity and velocity pressure can be used both ways, to find pressure or velocity. Two equations commonly used in practice are as follows: pv V = 1.41 -------- m/s ρ air 2

P v = 0.602V Pa

(1-7)

(1-8)

where the numbers 1.41 and 0.602 contain the conversion factors appropriate for pv in pascals; density, ρ, in kg/m3 (standard air density is 1.204 kg/m3); and velocity, V, in m/s. Standard air, for the HVAC industry, is dry air at 20°C and 101.325 kPa, with a mass density of 1.204 kg/m3. Sea level pressure is 101.325 kPa, so standard air can be considered typical dry air at sea level. For this reason, most airflow tables and charts are based on standard air. Note that defining an airflow in terms of standard air also defines the weight and mass flow. Thus, 10 m3/s of standard air is also 10 × 1.204 = 12.04 kg/s. As elevation increases, air density decreases, and above 1000 meters, density corrections should be considered. Because most projects are located at altitudes from sea level to 1000 m, most designs can use standard air without correction. Air expands as it is heated and density drops. For many air-conditioning systems, this can be ignored, but be careful. As shown in the cold climate example, outdoor air at –35°C has a density about 20% lower than it does at 24°C. Standard air is dry air with no moisture vapor, but the air we experience is never dry; atmospheric air always includes water in the form of moisture vapor. Also, the quantity of moisture vapor varies. It is typically under 2% by weight, and it influences the density and thermal properties of air. The addition and removal of moisture are common processes in air systems and can be conveniently shown on a chart called a psychrometric chart. The main axes on a psychrometric chart are temperature along the bottom x-axis and moisture weight compared to dry air weight, kg/kg or g/kg, on the y-axis. There is a maximum proportion of moisture vapor with the air at any given temperature, so the chart has the characteristic form shown in Figure 1-5. Shown are • • •

vertical temperature lines, °C. horizontal moisture content (humidity ratio) lines, kg of moisture/kg of dry air. sloping down left to right specific-volume lines, m3/kg. For example, air at 25°C and 25% relative humidity (RH) has a specific volume of 0.85 m3/kg. At this specific volume, 1 kg of air occupies 0.85 m3.

Psychrometric chart.

Chapter 1 Fundamentals of Airflow

Figure 1-5

12

Fundamentals of Air System Design SI

Figure 1-6

13

Conversion of static pressure to velocity pressure.

•

curved RH lines, %. The highest of these lines, labeled 100% RH, is the maximum moisture that can be in gaseous form at that temperature.

When the air is saturated with moisture, we say the humidity is 100%. When the same volume of air holds only half the weight of water vapor that it has the capacity to hold at that temperature, we call it 50% RH. The chart shows the 25%, 50%, 75%, and 100% RH lines. The saturation line is 100% and the horizontal line along the x-axis is 0%. Note that on the chart, the RH lines are not linearly related. Thus, at a particular temperature, the 50% RH curve is not at half the height of the saturation (100% humidity) line. Psychrometric charts are based on standard air, and humidity ratio may be labeled kg/kg or g/kg. The advantage of using grams is that the scale is in whole numbers; for example, 8 rather than 0.008. For most above-ground terrestrial systems, the kgm and kgw issue can be ignored; but be careful with units when dealing with substantial pressure changes as those that occur in mines, submarines, planes, and space vehicles. We will return to the psychrometric chart in future chapters.

Air Handling—A Practical Application How these basic principles apply to air system design is illustrated in Figure 1-6, which shows a duct with the air coming in the left and going out the right. For this example, we assume this to be a frictionless process. Notice that the duct reduces in cross section, with area A1 greater than A2. There is one velocity at A1 and another at A2. This process can be analyzed using the Continuity Equation.

14

Chapter 1 Fundamentals of Airflow The Continuity Equation says that for a given mass flow, and by the law of conservation of matter or mass, air entered on the left side must come exit the right side, because we can neither destroy nor create air in the duct between the two points. The Continuity Equation says that the cubic meters in is equal to the cubic meters out, ignoring any kind of compressibility or temperature change. In other words, the quantity of air in (m31) is equal to the quantity of air out (m32), giving m31 = m32. Because m3 = AV, then A1V1 = A2V2. If we measure the duct, we know what A1 and A2 are. If we know V1, we can solve for V2, and we know that because A1 is bigger than A2, V2 must be bigger than V1. This relationship can be explained by the Continuity Equation: ⎛ A 1-⎞ V 2 = V 1 ⎜ ----A ⎟ ⎝ 2⎠

(1-9)

So as the cross-sectional area is reduced, the velocity is increased, as predicted by the Continuity Equation. Let’s return to the Energy Equation and the relationship in which the total pressure is equal to the velocity pressure plus the static pressure (pt = pv + ps). If the velocity increases, the velocity pressure has to increase, because velocity pressure is pv = V2/2g. As the air flows from left to right in Figure 1-6, both velocity and kinetic energy increase. The simple device shown in Figure 1-6 converts potential energy into kinetic energy. But how does this happen? In this example, there are two forms of energy: static pressure (the flow work) and velocity pressure (kinetic energy). If the kinetic energy increases, then the static pressure must decrease by the same amount. If A2 is one-half as big as A1, then V2 is twice as big as V1. Because the velocity pressure is proportional to the square of the velocity (V2/2g), pV2 is four times pV1, and static pressure ps is smaller by an equal amount. This is not too difficult to understand because it is expected that the static pressure will be less at A2 than at A1. Figure 1-6 shows an accelerator. Velocity is increased by making the duct area smaller. Suppose Figure 1-6 is reversed, as in Figure 1-7, which shows a decelerator. Because the air comes in at a higher velocity through the smaller section and goes out at a lower velocity through the larger section, the kinetic energy is reduced. If the velocity is reduced by a factor of two, the kinetic energy level (and the velocity pressure) is reduced by four. Consequently, the static pressure increases by an equal amount. Static-pressure probe manometers placed at A1 and at A2 in Figure 1-7 would show that the static pressure at A2 is greater than at A1. This phenomenon is called static pressure regain, and it is a very important principle of air system design. One method of designing ducts is called the static pressure regain method, which is applied to a duct with a series of out-

Fundamentals of Air System Design SI

Figure 1-7

15

A decelerator.

lets. After each outlet, the velocity is reduced and the duct size is reduced so that the static pressure at the next outlet is approximately the same.

Friction Effects Until now, we have considered frictionless systems. But in the real world of air system design, friction must be taken into account. Viscosity is the property responsible for dissipation of the fluid’s kinetic energy into intrinsic internal energy. In air ducts, the amount of energy transferred is small, but the effect on pressure drop is major. The frictional pressure drop is commonly characterized by the Darcy-Weisbach equation: 2

L V Δp f = f ---- γ ------- Pa D 2g

(1-10)

In terms of head as meters of fluid flowing, the Darcy-Weisbach equation can be written 2 Δp f V L -------- = f ---- ------- m γ D 2g

(1-11)

For standard air, the Darcy-Weisbach equation can be written 1000fL Δp f = ----------------Dh

2

ρV ----------- Pa 2

(1-12)

16

Chapter 1 Fundamentals of Airflow where f =

dimensionless friction factor

L

=

duct length, m

D

=

diameter, mm

Dh

=

hydraulic diameter, mm. Actual diameter for round ducts but a calculated value for oval and rectangular ducts.

This is a purely empirical formula which states that the frictional pressure drop is proportional to length, L, inversely proportional to the effective diameter, Dh, and proportional to velocity pressure or velocity head. One would hope that the proportionality constant, f (dimensionless), would be truly constant, and that turns out to be partially true. When the flow is fast, f (called the friction factor) is fairly constant and depends only on the duct roughness. When the flow is slow, f is inversely proportional to velocity, but the wall roughness is unimportant. The terms fast and slow must be explained. Consider all of the properties and characteristics involved in fluid friction: velocity, diameter, viscosity, and density. Consider also the variety of motions that we observe: slow, such as the streamline flow of water out of a hose, or fast, such as the turbulent flow of water out of the same hose when the faucet is fully open. We are fortunate that these phenomena can all be related through a single parameter known as the Reynolds number, which is defined as VρD Re = -----------μ

(1-13)

where µ is the absolute viscosity in mPa·s, V is the velocity in m/s, D is the diameter in m, and ρ is the density in kg/m3. The Reynolds number is the ratio of the momentum of the flow, Vρ, to the viscosity, µ. If the viscosity is high relative to the momentum, the flow is laminar or streamline (like maple syrup). But if the viscosity is low (as for air), the flow will be turbulent for any realistic duct size. Laminar airflow occurs in laminar flow filters where the pore size, D, is very small. So there are two distinct regimes of flow (laminar and turbulent) that depend on the Reynolds number. The effect of these distinctions is manifest in the behavior of the friction factor, as shown on the Moody chart (Figure 1-8) (Moody 1944). This chart shows the friction factor as a function of the Reynolds number. Note that both axes have logarithmic scales. Several interesting features are present on the Moody chart: •

The laminar flow region is shown for Reynolds numbers smaller than about 2000. The dashed extension of the solid line indicates that under some circumstances, the relationship can be extended up to 4000. This part of the line is not to be trusted. In the laminar flow region, the friction factor is inversely proportional to the Reynolds number:

Fundamentals of Air System Design SI

Figure 1-8

17

Moody chart (ASHRAE 2009).

f = 64 ⁄ Re •

(1-14)

If this value is substituted into Equation 1-12, we get the following:

vLVΔp f = n --------2 Pa D

(1-15)

where n is a numeric constant and V, the kinematic viscosity, is μ/ρ with units of m2/s. Note that the frictional pressure drop varies with the first power of velocity.

18

Chapter 1 Fundamentals of Airflow •

•

There is also a dashed line labeled “fully rough.” To the right of this line, the friction factor is constant for a particular value of roughness, ε. The relative roughness values are shown. For example, the roughness of commercial steel pipe is 152 µm. Relative to a 100 mm pipe, ε/D is about 0.00152. In this fully rough region, a constant value of f can be used regardless of flow rate or velocity, and the frictional pressure drop varies with the second power of velocity. Between laminar and fully turbulent flow, the friction factor depends on the Reynolds number and the relative roughness, and an iterative solution to a problem may be necessary. In this region, pressure drop varies with a power of velocity between 1 and 2. Unfortunately, many air duct flows occur in this transition region and HVAC air velocities are neither known accurately nor constant.

Test work performed by ASHRAE and its predecessor organization ASHVE (American Society of Heating and Ventilating Engineers) indicated prior to Moody’s work (Moody 1944) that ε for galvanized sheet metal ductwork was about 0.15 mm. This is based on transverse joints spaced at 750 mm intervals. When joints are spaced at 1150 mm intervals, the value is reduced to 0.09 mm.

The System Constant Form of the Darcy-Weisbach Equation HVAC air and piping systems usually use a simplified form of the DarcyWeisbach equation, where it is assumed that the friction factor is constant and that L and D do not change (although the system may be made up of various L and D and fittings). Grouping all of the constants together, we can rewrite Equation 1-15 in two forms. The first form includes the loss coefficient, K, and the second for includes the system constant, Cs. Note that Cs2 equals the reciprocal of K, Cs2 = 1/K. 3

m Δp f = K ------s

2

(1-16)

or 3

m ------- = C s Δp f s

(1-17)

Equation 1-16 is the system constant form of the Darcy-Weisbach equation. It is used extensively in HVAC systems work. A system curve, as shown in Figure 1-9, portrays the frictional pressure drop for a particular system. The curve is a parabola that can be generated with one known experimental or calculated value for a particular system. One pair of m3/s and Δpf is required to determine K or Cs. Consider a complex air-handling system where we want to move 10 m3/s through the system. The pressure drop in the system is calculated to be 1 kPa.

Fundamentals of Air System Design SI

Figure 1-9

19

Typical system curve.

The system constant form of the Darcy-Weisbach equation can be used to find the system constant: 10 C s = ---------------- = 0.316 1000 Similarly, we find that K = 10. Values of Δpf and m3/s can be plotted on a graph. Other values can be determined by using Δp f = K ( m 3 ⁄ s )

2

As long as the system is unchanged, it will operate on this curve.

The Friction Chart In 1945, D.K. Wright published a paper titled, “A new friction chart for round ductwork” in the ASHVE Transactions (1945). A graph from this article has become essentially the standard for HVAC work. This graph, often known as the Wright Friction Factor Chart, takes the Darcy-Weisbach relationship and the Moody chart (Moody 1944) and converts them into a graphical presentation that lets us determine frictional pressure drops at various diameters of round ductwork and at various velocities based on an ε value for galvanized sheet metal ductwork of 0.15 mm.

20

Chapter 1 Fundamentals of Airflow Since that time, ASHRAE, assisted by the Sheet Metal and Air Conditioning National Contractors' Association (SMACNA), has conducted a series of tests and obtained slightly different numbers than those used by Wright. The new data have been included in the ASHRAE Handbook—Fundamentals since 1993 (ASHRAE 2009) The friction factor chart (see Figure 1-10) was revised based on standard galvanized sheet metal ductwork with an absolute ε roughness of 0.09 mm instead of 0.15 mm. Other factors, including the shape of the duct, the roughness of the material of construction, and fittings used must be considered. These are discussed in Chapter 7, “Duct System Design.”

Density and Altitude Effects Standard psychrometric charts and performance data published by manufacturers generally assume equipment operation at sea level with standard air. However, when the project is located at a significantly higher altitude, allowances must be made for lower pressure. Factors by which the usual data must be multiplied when operating at higher altitudes are summarized in Table 1-1. For items not listed, consult appropriate sources, such as Carrier’s Altitude Effects (1967).

The Next Step This chapter introduced the fundamentals of airflow theory and included discussion of air flowing in ducts. Chapter 2 introduces the other common components of air systems that condition air and deliver it to the occupied space. Included are the systems’ function and main operating characteristics. More detail on choosing components and their detailed operation are explained in later chapters.

Summary The difference between mass and weight was explained. Mass is a property of matter that is invariant with location, but weight changes depending on the local gravitation. Conveniently for designers of most buildings, gravity is constant, with kg and kgw being numerically the same. Hydrostatic pressure, commonly referred to as static pressure, is the pressure at one place due to the column of fluid above. The pressure is the same in all directions at any point. In a duct with air flowing under pressure, the static pressure around the duct is the same on all sides. The Continuity Equation states that mass is neither created nor destroyed. Thus, under steady conditions with no storage, the mass flow into a system must equal the mass flow out of the system. The volume of air at constant pressure is proportional to the absolute temperature. Thus, while the mass into a system equals the mass out, the volume in can be different from the volume out if the temperature is changed.

Figure 1-10

Friction chart (ASHRAE 2009).

Fundamentals of Air System Design SI 21

22

Chapter 1 Fundamentals of Airflow Table 1-1

Typical Altitude Correction Factors Altitude, m (above sea level)

Item

750

1500

2000

3000

Compressors

1.00

1.00

1.00

1.00

Condensers, air-cooled

0.95

0.90

0.85

0.80

Condensers, evaporative

1.00

1.01

1.02

1.03

Chillers

1.00

1.00

1.00

1.00

Induction room terminals (chilled water)

0.93

0.86

0.80

0.74

Total capacity (*SHF = 0.40 – 0.95)

0.97

0.95

0.93

0.91

Sensible capacity (SHF = 0.40 – 0.95)

0.92

0.85

0.78

0.71

Total capacity (SHF = 0.95 – 1.00)

0.93

0.86

0.79

0.73

Total capacity (*SHF = 0.40 – 0.95)

0.98

0.96

0.94

0.92

Sensible capacity (SHF = 0.40 – 0.95)

0.92

0.85

0.78

0.71

Total capacity (SHF = 0.95 – 1.00)

0.96

0.82

0.88

0.84

Fan-coil units

Packaged air-conditioning units, air-cooled condenser

*SHF = Sensible Heat Factor

The useful but not absolutely correct formula for calculating the result of mixing airstreams was introduced: (V1 × T1) + (V2 × T2) = [(V1 + V2) × T3]. Remember to keep the volume units consistent, m3/s or L/s. Energy, like mass, is neither created nor destroyed. It can be converted from one form to another and measured in different units. However, in consistent units in any process energy in = energy out – energy stored in the system Energy can be in a number of forms: work, done by a force over distance or torque through an angle; heat, energy transfer due to a temperature difference; internal energy, due to thermal energy relative to some datum; potential energy, work done by movement in the Earth’s gravitational field; and kinetic energy from motion. Static pressure is the pressure exerted by a fluid at rest. Velocity pressure is the pressure exerted by a fluid by virtue of its motion. Typically, measuring the pressure at a tapping in the side of a duct provides the static pressure. The pressure on the open end of a tube facing the flow of air measures both the static and the velocity pressure, called total pressure. The difference between the static pressure and total pressure is the velocity pressure.

Fundamentals of Air System Design SI

23

Bernoulli’s Equation states that in a system without energy losses or gains, the sum of static pressure and velocity pressure is constant: p V2 --- + ------γ 2g

= constant duct

The valuable concept in Bernoulli’s Equation is that if the velocity is reduced due to a wider duct, the drop in velocity pressure (V2 reduces) is exactly matched by an increase in static pressure with friction ignored. Velocity pressure in pascals for standard air = 0.602V 2 Pa. Standard air and the psychrometric chart were introduced to raise the issue of decreasing density with increasing temperature and the issue of moisture in the air. Friction effects occur in ducts for several reasons, including surface roughness, duct joints, fittings, equipment, and outlets. The theory behind duct friction was discussed, including Reynolds number, the Moody chart (Moody 1944), and the Darcy-Weisbach equation. The critical point to remember is that in a fixed system, the pressure drop through the system is about proportional to the square of the flow: Δp = K(m3/s)2. Thus, doubling the flow creates four times the pressure drop. The pressure drops through ducts can be calculated, but the simplest method is to use a friction chart, such as that shown in Figure 1-10 for standard air. Standard psychrometric charts and performance data generally assume equipment operation with standard air. The lower air density at high altitudes significantly affects some equipment but not all. Reference tables can be used and manufacturers contacted for assistance in these cases.

References ASHRAE. 2009. ASHRAE Handbook—Fundamentals. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. Carrier. 1967. Altitude Effects. Syracuse, NY: Carrier Corp. Moody, L. 1944. Friction factors for pipe flow. ASME Transactions. New York: American Society of Mechanical Engineers. Wright, D. 1945. A new friction chart for round ductwork. ASHVE Transactions. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc.

24

Chapter 1 Fundamentals of Airflow

Skill Development Exercises for Chapter 1 Complete these questions by writing your answers on the worksheets at the back of this book. 1-1

In Figure 1-6, area A1 = 2 m2, area A2 = 1.25 m2, and velocity V1 = 10 m/s. Calculate V2 (m/s). a) 16 m/s

1-2

b) 25.6 m/s

c) 7.5 m/s

d) 12.6 m/s

The total pressure at a certain point in a system is determined to be 1.25 kPa, and the static pressure at that point is determined to be 500 Pa. What is the velocity pressure (Pa) at that point? a) 1.75

b) 1

c) 750

d) 250

1-3

Which of the following is the most correct definition of static pressure regain? a) as the velocity of an airstream decreases, the static pressure increases. b) as the velocity of an enclosed airstream decreases due to friction, the static pressure increases. c) friction reduces static pressure, while velocity pressure increases with reduction in duct size.

1-4

An air-handling system is determined to have a 1500 Pa pressure drop through the system at a flow of 4000 L/s. What is the system constant? a) 2.7

1-5

b) 6000

c) temperature

d) all of the above

b) kg/m2

c) N/m

d) none of the above

If the cross-sectional area of a duct decreases in size, the velocity of an airstream passing through the duct increases. a) True

1-9

b) pressure

Fan pressures are typically indicated in what units? a) Pa

1-8

b) Reynolds number (Re) d) flow work

A water manometer measures _______________. a) velocity

1-7

d) 103

The product of fluid pressure and specific volume is _______________? a) internal energy c) kinetic energy e) viscosity

1-6

c) 77.4

b) False

Air passes through a length of inaccessible duct with a constant cross-sectional area. You suspect that there is a serious leak in the duct. The velocity pressure drops from 200 to 160 Pa along the suspect section of duct. Approximately what percentage of air is being lost through the leak? a) 5%

b) 10%

c) 20%

d) 41%

Fundamentals of Air System Design SI 1-10

2000 L/s of outdoor air at 2°C is drawn in over a heater and delivered into the building at 24°C. What volume of air is delivered? a) 1851 L/s

1-11

25

b) 2000 L/s

c) 2079 L/s

d) 2160 L/s

In an air-conditioning system, 15 m3/s of return air at 25°C is mixing with 2.3 m3/s of outdoor air at 35°C. What is the approximate resulting volume and temperature? a) 17.3 m3/s, 22.7°C

b) 17.3 m3/s, 29.9°C

c) 17.3 m3/s, 26.3°C

d) 17.3 m3/s, 33.7°C

Air Distribution System Components Study Objectives The goal of this chapter is to give you an overview of air distribution system components and their schematic symbols, which will serve as a foundation of knowledge as each component is discussed in detail in later chapters. After completing this chapter, you should be able to list and explain the functions of the components of an air distribution system and T identify the schematic symbols of air distribution system components. T

Instructions Read the material in Chapter 2. Verify the examples presented in the chapter with your own calculations. At the end of the chapter, complete the skill development exercises without referring to the text. Review those sections of the chapter as needed to complete the exercises.

Air Distribution System Overview An air distribution system is used to maintain desired environmental conditions within a space. In almost every application, many options are available to the designer to satisfy that goal. Air distribution systems are categorized in many ways including by how they control the conditioned area, by special equipment arrangement, and by duct configuration. This chapter provides an overview of the basic components of an air distribution system: air-handling units; fans, fan motors, and fan drives; coils; filters; ducts; controls; air distribution devices; intake and exhaust louvers; and sound absorbers.

Air-Handling Units An air-handling unit (AHU) combines fans, coils, filters, dampers, connections to supply and return ducts, and other components into a device used to move air. It may also be used to clean, heat, cool, humidify, dehumidify, and

Chapter 2 Air Distribution System Components

Figure 2-1 Large air-handling unit.

28

Fundamentals of Air System Design SI

29

mix the air. Figure 2-1 shows a large typical central AHU. Types of AHUs include the following: •

• • • •

A central-station unit is a factory-made encased assembly consisting of the fan and other necessary equipment. It does not include a source of heating or cooling, but it may include heating and/or cooling coils. A cooling unit that includes the means for cooling. It may also perform other AHU functions. A heating unit that includes the means for heating. It may also perform other AHU functions. A makeup air unit is a factory-assembled fan heater or cooler used to supply tempered fresh air to replace the air that is exhausted. A ventilating unit has the means to provide ventilation, and may also perform other AHU functions.

Fans, Fan Motors, and Fan Drives A fan is an air pump that creates a pressure difference and causes airflow. The fan impeller performs work on the air, imparting to it both static and kinetic energy, which varies in proportion depending on the fan type. Fans are generally classified as centrifugal fans or axial flow fans according to the direction of airflow through the impeller. Figure 2-2 shows the general configuration and schematic symbol for a centrifugal fan. Figure 2-3 shows the configuration and schematic symbol for an axial flow fan. All fans must have some type of power source, usually an electric motor. On packaged fans, the motor is furnished and mounted by the manufacturer. On larger units, the motor is mounted separately and coupled directly to the fan or indirectly by a drive mechanism. The schematic symbol for a motor is also shown in Figure 2–3. Two standard fan drive arrangements are available: •

•

Direct drive, where the fan is mounted directly on the motor shaft or an extension of the motor shaft, offers a more compact assembly and ensures constant fan speed. Fan speeds used to be limited to available motor speeds, an economical solution when practical. Today, at additional cost, the motor speed can be adjusted over a wide range by supplying the motor through a variable-frequency controller. Capacity is set during construction by variations in fan impeller geometry and motor speed. Belt drive offers flexibility in that the fan speed can be changed by altering the drive ratio. This allows initial adjustment to match the fan output with the system actually installed. In some applications, this flexibility allows for changes in system capacity or pressure requirements due to changes in process, hood design, equipment location, or air-cleaning equipment.

30

Chapter 2 Air Distribution System Components

Figure 2-2

Centrifugal fan configuration.

Figure 2-3

Axial fan configuration.

Fundamentals of Air System Design SI

31

Coils A coil is a cooling or heating element made of pipe or tube. Coils are sometimes finned and are found in a number of shapes (serpentine, helical, etc.). Some coils commonly encountered in air systems include the following: • • • •

A cooling coil uses refrigerant or secondary coolant to provide cooling or cooling with dehumidification. A heating coil provides heat. Electric heating coils use a resistance element instead of a fluid to create a heating effect. A preheat coil is a heating coil installed upstream of a cooling coil or at the front of an air-handling system to preheat air. A reheat coil is a heating coil installed downstream of a cooling coil.

Cooling and heating coils are often seen as labeled boxes, as shown in Figure 2-1.

Filters A filter is a device used to remove solids from an airstream. Filter performance is based on the ability to collect a particular size or type of dust and is stated for each filter as a rating. The rating may denote air cleaning efficiency as a percentage of dust removal or as the ability to remove dust particles of certain size ranges. These efficiencies are defined by standardized ASHRAE test methods that we discuss in Chapter 9. A filter used to remove gases is correctly called an adsorber, as the gas is chemically adsorbed onto the filter material rather than mechanically collected on the filter surface. Filters encountered in air system design include the following: • •

•

• •

A disposable filter has elements that are discarded after use. Efficiencies range from very low to relatively high, depending on the construction. A pleated filter provides a high ratio of media area to face area, thus allowing reasonable pressure drop. The filter media may be self supporting because of inherent rigidity or because the airflow inflates it into an extended form, such as with bag filters. A roll filter (moving curtain filter) has a filter medium on a continuous belt on movable rolls that brings clean filter area into the airstream, either automatically or manually. Efficiencies are usually fairly low. A viscous impingement filter has a medium made from materials that have been impregnated with a viscous oil to increase dust retention. An absolute filter has an efficiency of 99.9% or higher and can filter particles down to 0.01 µm (microns) in size. It is also known as a high-efficiency par-

32

Chapter 2 Air Distribution System Components

•

• •

ticulate air (HEPA) filter and is tested and rated to an American Society for Testing and Materials standard. An active electrostatic filter has the airstream passing through a highvoltage ionizing field to impart a positive electrical charge to the particles, which are then collected on electrically negative plates. A static electrostatic filter consists of plastic media that generate an electrostatic attraction by the airflow over the plastic. A carbon filter (adsorber) uses a mass of granulated activated carbon to adsorb certain gases.

Labeled boxes are often used to indicate filters in diagrams (see Figure 2-1). Filters are discussed in more detail in Chapter 9 of this course.

Ducts A duct is a tube or conduit for conveying air. Ducts are classified in terms of application and pressure. HVAC systems in public assembly, business, educational, general factory, and mercantile buildings are usually designed as commercial systems. Air-pollution control systems, industrial exhaust systems, and systems outside the pressure range of commercial system standards are classified as industrial systems. Ducts may be round, oval, or rectangular. They may be made of galvanized steel, aluminum, fibrous glass, and other materials. They may be rigid or flexible. Schematic symbols for ducts are shown in Figure 2-4.

Controls A control is a device that regulates a function such as the airstream. Controls may be manual or automatic. If automatic, the implication is that the control is responsive to a change in pressure, temperature, or some other variable to be regulated. Two common and important controls are dampers and thermostats. A damper is a device used to vary the volume of air passing through an outlet, inlet, or duct. A thermostat is an automatic device that is responsive to temperature. Thermostats are used to maintain a constant temperature in a regulated space, to permit the passage of control air when the temperature of the controlled air is within the limits at which the thermostat is set, and for other temperature control purposes. Schematic symbols for these controls are shown in Figure 2-5.

Air Distribution Devices Air distribution devices are devices or openings through which air is discharged into a conditioned space. Included in this category of devices are registers, grilles, and diffusers. Registers and grilles are also used to withdraw air from a conditioned space. Schematic symbols for these devices are shown in Figure 2-6, and they are discussed fully in the next chapter.

Fundamentals of Air System Design SI

Figure 2-4

33

Duct symbols.

Intake and Exhaust Louvers A louver is a device consisting of multiple blades that, when mounted in an opening, permits the flow of air but inhibits the entrance of other elements. An intake louver is used at the entrance to an air system. An exhaust, or relief, louver is used at an exit. Schematic symbols for louvers are shown in Figure 2-7.

Sound Absorbers A proper acoustical environment can be as important for human comfort as are other environmental factors controlled by air-conditioning systems. The objective of sound control is to achieve an appropriate sound level for all activities

34

Chapter 2 Air Distribution System Components

Figure 2-5

Controls symbols.

and people involved. Sound absorbers diminish the intensity of sound energy from fans, ducts, and other sources. Chapter 10 in this course provides additional information on acoustical environments. Sound and vibration isolation are required for most central-system fan installations. Mountings of fiberglass, ribbed rubber, neoprene, and springs are available for most fans and prefabricated units. Noise transmitted through ductwork can be reduced by sound-absorbing units, acoustical lining, and other means. The schematic symbol for a sound absorber in ductwork is shown in Figure 2-8.

The Next Step The primary task of commercial and institutional HVAC systems is to keep the building occupants comfortable. To achieve this, the system designer requires knowledge of the factors affecting comfort and how air can be distributed in occupied spaces to achieve comfort conditions, which is the subject of Chapter 3.

Figure 2-6

Air distribution devices.

Fundamentals of Air System Design SI

35

36

Chapter 2 Air Distribution System Components

Figure 2-7

Louvers.

Figure 2-8

Sound absorber.

Summary This chapter briefly introduced the main components of an air-conditioning system. More details of their construction and operation are included in later chapters. Air-handling units (AHU) are a combination of fans, coils, filters, controls, louvers, and dampers, which together provide a supply of conditioned air. Depending on the particular requirements, the air may be filtered, mixed, cooled, dehumidified, heated, or humidified, and fans provide the necessary static pressure and velocity to the airflow. A fan is an air pump. The fan creates a pressure difference (static pressure) and causes airflow (kinetic energy). The first main type is the centrifugal fan, where the air enters the center of the drum-shaped impellor and is thrown radially into the fan outlet casing. The second main fan type is the axial fan, where the air flows axially, or parallel, to the fan shaft. Most fans are driven by an electric motor. The simplest arrangement is mounting the impellor directly on an extended motor shaft. This arrangement

Fundamentals of Air System Design SI

37

works for smaller sizes but is limited to the few available motor speeds. Belt drives are a popular mechanical method of connecting the fan and impellor shaft, and they can be adjusted to change speeds. In addition, electrical speed controllers are available to provide variable-speed drive. A coil is an array of finned pipes containing a flow of cooling or heating fluid. The fins greatly extend the heat transfer area of the pipes. Coils used for cooling are often cool enough for condensation to occur and thereby dehumidify the air. Filters remove dirt from an airstream. Their performance is rated on the basis of particle removal based on quantity or particle size. Filters are available in a large range of designs, each aimed at a specific market segment. Filters are discussed in more detail in Chapter 9. Units that remove gases are called adsorbers, although the most common type, made of activated carbon granules, is called a carbon filter. A duct is a tube or conduit for conveying air. Ducts are most commonly made of light galvanized steel with round or rectangular sections. They can be made in many other materials for particular duties. The main criteria for choosing ducts are pressure and contaminants. Controls regulate the performance of a system. Manual controls are preset, such as a damper preset to restrict flow through a duct. Automatic controls regulate some functions continuously, such as a thermostat controlling a heater. Some air distribution devices distribute air into occupied spaces, while others allow air out of the spaces. A louver allows air into or out of the building while restricting the entrance of unwanted rain, snow, animals, and birds. A mechanical plant is inherently noisy. The noise can be distributed either by direct transfer into the building structure or as airborne noise along the ducts. A variety of materials are used to isolate the vibration and to attenuate the noise distributed through the ductwork. Chapter 10 provides additional information on sound absorbtion.

Bibliography ASHRAE. 2008. ASHRAE Handbook—Systems and Equipment. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. ASHRAE. 2009. ASHRAE Handbook—Fundamentals. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc.

38

Chapter 2 Air Distribution System Components

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

The symbol in the figure below represents a(n) _______________.

a) centrifugal fan

b) axial fan

c) diffuser

d) none of the above

In the figure below, this ductwork is _______________, and the dimension of the side shown is _______________.

a) dropping, 500

b) dropping, 300

c) rising, 500

d) none of the above

The symbol in the figure below is for a flexible duct.

a) True 2-4

b) False

The symbol in the figure below shows a(n) _______________.

a) blanked-off duct, with a top dimension of 300 b) return air duct, with a side dimension of 450 c) supply air duct, with a side dimension of 450 d) none of the above

Fundamentals of Air System Design SI 2-5

The dimension of the duct shown in the figure below is 600.

a) True 2-6

2-7

2-8

b) False

A filter that uses a liquid as an adhesive is a(n) _______________. a) carbon filter

b) electrostatic filter

c) viscous filter

d) all of the above

An air-handling unit may be used to _______________. a) move air

b) mix air

c) heat air

d) all of the above

The symbol in the figure below is for a(n) _______________.

a) manually operated damper b) electrically controlled damper c) manual damper d) none of the above 2-9

The symbol in the figure below is for a(n) _______________.

a) pneumatically operated damper b) inline psychrometric observation device c) fire damper d) all of the above

39

40

Chapter 2 Air Distribution System Components 2-10

The symbol in the figure below is for a(n) _______________.

a) temperature relay b) test station c) remote bulb thermostat d) all of the above

Human Comfort and Air Distribution Study Objectives After completing this chapter, you should be able to list and explain the main issues involved in thermal comfort; T principles of air distribution as they relate to human comfort; T principles of space air distribution; and T functions of the different types of air distribution devices. T

Instructions Read the material in Chapter 3. Verify the examples presented in the chapter with your own calculations. At the end of the chapter, complete the skill development exercises without referring to the text. Review those sections of the chapter as needed to complete the exercises.

Principles of Human Comfort Human comfort is dependent on a variety of factors relating to the individual, the individual’s current activity, and the space the individual inhabits. The space temperature, humidity, air quality, and acoustics are controlled or influenced by air conditioning. However, other space factors, such as lighting, are not controlled by air conditioning. Individuals vary. For example, one person may have a much higher metabolic rate and be comfortable in a much cooler environment than someone else. In contrast, the elderly are often more comfortable with a significantly higher temperature than younger people. Finally, the activity level of the individual and their clothing influence their comfort. In this chapter, we begin with thermal comfort and continue on to air quality before discussing air delivery and movement in the occupied space.

Thermal Interchange between People and Environment One of the first steps in designing an air distribution system for human comfort is to establish comfort criteria for the intended service. These criteria should include space temperature and humidity, ventilation rate, indoor air quality, and

42

Chapter 3 Human Comfort and Air Distribution sound level. The selection of these criteria is influenced by many conditions, including the ages and activities of the occupants, the occupant density, and the contaminants present in the space. The human body can be thought of as a power plant. The body takes in raw materials and uses them to generate energy for daily life and activities. One major function of the body is the heat rejection that occurs in the thermal processes that the body goes through to produce mechanical energy. As shown in Figure 3-1, the body uses three of the major heat transfer mechanisms to reject this heat: radiation, convection, and evaporation. Radiation is important occasionally. We feel radiation when we sit next to a window with the sun shining in or in the winter when we are too close to a cold window or wall. However, in general, the basic modes of heat transfer the body uses are convection and evaporation. They are similar in magnitude in most cases, although when we begin adjusting dry-bulb temperature or humidity, a mechanism in the body reacts to that change and shifts more of the heat transfer to one mode or the other as needed. The problem is that both convection and evaporation depend on the same phenomenon: air motion over the skin surface.

Figure 3-1

Heat mechanisms.

Fundamentals of Air System Design SI

43

Evaporation from the skin surface is based on two driving forces: •

•

The difference between the partial pressure of the water vapor at the skin temperature and the partial pressure of the water vapor at the dewpoint temperature in the room (how humid it is) The velocity of the air past the occupant

If there is no air velocity, the mechanism of moisture diffusion is not very good. Also, the more humid the room, the lower the mechanism to evaporate water, and consequently, the lower the evaporative heat transfer. Similarly, convection is driven by the difference between the skin temperature and the space temperature. As the space temperature increases, the heat transfer decreases. As the space temperature decreases, the heat transfer increases. Because the body tries to maintain the skin temperature at a relatively constant level, room temperature is quite important.

Temperature and Humidity Comfort Zone Comfort is a complex, subjective response to several interacting variables. Not everyone perceives a given temperature and humidity level with the same degree of satisfaction. The perception of comfort relates to individual physical conditions, body heat exchange with the surroundings, and physiological characteristics. The heat exchange between the individual and the surroundings is influenced by several factors, including the following: • • • • • • •

Dry-bulb temperature, °C Relative humidity, RH Thermal radiation Air movement, m/s Insulation value of clothing, clo Activity level, met Direct contact with surfaces not at body temperature

Two units—clo and met—are probably new for you. Clothing has an insulating value and, in general, the greater the insulating value, the lower the ambient temperature for the same comfort level. Typical indoor winter clothing is 1 clo: a person with shoes, socks, pants or full-length skirt, underwear, shirt, and jacket. Typical light summer clothing, including shorts or knee-length skirt and short-sleeved shirt, is 0.5 clo. The met is a unit of metabolic activity resulting in a heat loss of about 58.2 W/m2. A resting adult typically produces 1 met; light office work produces 1 to 1.3 met; and walking at 3.2 km/h produces 2 met. Figure 3-2, which is adapted from ANSI/ASHRAE Standard 55–2004, Thermal Environmental Conditions for Human Occupancy (ASHRAE 2004),

44

Chapter 3 Human Comfort and Air Distribution

Figure 3-2

Acceptable range of temperature and humidity.

specifies conditions likely to be thermally acceptable to at least 80% of the adult occupants in a mechanically conditioned space where •

activity levels are between 1 and 1.3 met

• •

clothing is near 0.5 or 1 clo air speeds are below 0.2 m/s

The design space temperature and humidity for both heating and cooling seasons should be based on Figure 3-2 for most applications. The comfort zone is defined for people in winter clothing (1 clo) and summer clothing (0.5 clo), primarily engaged in sedentary activities. As a practical matter, the higher the conditioned-space relative humidity (RH), the cooler the space needs to be to provide the same thermal comfort for the occupants. This has been given as a reason for increasing the humidity indoors in cold dry climates in winter. The sales pitch is that not having to keep the indoor temperature so high saves on the heating bill. Unfortunately, the proponents

Fundamentals of Air System Design SI

45

conveniently do not assess the real cost of humidification, which is higher than the heating saving. In a hot humid climate, dehumidification is costly in plant and operating costs. So allowing the humidity to rise saves in air-conditioning operating costs. However, allowing the humidity to rise enough to allow mold growth can make the building uninhabitable until very expensive remedial work has been completed. The comfort chart indicates that RH does not have a very significant bearing on comfort as long as the space dry-bulb temperature is in the comfort range. The upper moisture level shown as humidity ratio of 0.012 kgmoisture/ kgdry air is far higher than acceptable in a building in a moist climate. Because mold can grow above 60% RH, it is prudent to maintain buildings in hot, humid climates with a humidity ratio significantly lower, at about 0.010 kgmoisture/kgdry air. In addition, RH affects odor perceptibility and respiratory health. Because of these considerations, 40% to 50% RH is the preferred design range. However, maintaining humidity within this range during winter is complicated by the following: • • •

Energy costs for humidification The risk of condensation on windows and window frames during cold weather The need to provide and maintain humidifying equipment incorporated into the air-conditioning system

Where winter humidification is provided for comfort, a minimum of 20% RH is generally acceptable in cold climates. If a higher level of humidity is acceptable under summer conditions, considerable energy savings can be realized, as shown in Figure 3-3. To determine an approximate value of the energy used for dehumidification at a constant 25°C dry-bulb temperature, enter the annual wet-bulb degree hours above 19°C in the occupied space at the bottom left, intersect this value with the indoor RH chosen, then draw a vertical line to the weekly hours of cooling system operation and read the energy used (in 1000 mJ per 1000 L/s) on the upper-left scale. Repeating this procedure for a different value of RH yields the energy savings obtainable by raising RH. However, be cautious about choosing an excessively high humidity. Computer rooms (particularly computer printers and drafting rooms) are two applications for which more than 50% to 55% RH is undesirable or unacceptable due to effects of moisture on the paper products.

Indoor Air Quality Air Contaminants Indoor air contaminants can be solid or liquid particles, gases, or vapors. Some can be irritants or odorous, thus affecting occupant comfort. The same

46

Chapter 3 Human Comfort and Air Distribution

Figure 3-3

Summer dehumidification energy requirements.

contaminants at higher concentrations, as well as others of which occupants may be unaware, can be health risks. People vary in their sensitivity to contaminants. Even very small concentrations of certain fungi and other impurities can cause serious discomfort and impairment of sensitive individuals while not affecting most occupants.

Fundamentals of Air System Design SI

47

Standards for vapors and gases specify a quantity of pollutant per unit volume in parts per million (ppm) of air. Standards for particles often specify the mass concentration of particles expressed as micrograms per cubic metre (µg/m3). They include all particle sizes, or the total suspended particulate concentration. Large particles are filtered by the nasal passages and cause no adverse physiological response unless they are allergenic or pathogenic. Smaller respirable suspended particles are important because they can lodge in the lungs. Respirable particles range in size up to 5 µm. Particles of specific interest include • • • • •

respirable particulates as a group; tobacco smoke (solid and liquid droplets), which also contains many gases, asbestos fibers; allergens (pollen, fungi, mold spores, and insect feces and parts); and pathogens (bacteria and viruses), which are almost always contained in or on other particulate matter. Vapors and gases of interest include

• • • • •

carbon dioxide (CO2), carbon monoxide (CO), radon (decay products become attached to solids), formaldehyde (HCHO), and other volatile organic compounds (VOCs).

Although some contaminants (such as sulfur dioxide) are brought in with outside air by mechanical ventilation or uncontrolled infiltration, most indoor contaminants come from inside sources. People are sources of carbon dioxide, biomatter, and other contaminants characterized as body odors. People’s activities (such as smoking, cleaning, cooking, gluing, and refinishing furniture) also cause pollution, and building materials and finishes can outgas pollutants. Furnishings, business machines, and appliances (particularly unvented or poorly vented wood- and fossil-fueled heaters and ranges) can be contaminant sources. The soil surrounding a building can be a source of radon and pesticides that enter the building through cracks or drains or by diffusion. HVAC systems, drains, plumbing systems, and poor construction or maintenance practices can have environmental niches where pathogenic or allergenic organisms collect and multiply to be reintroduced into the air. Many microorganisms (such as molds) have accelerated growth rates at RH levels above 60%. An additional complicating factor in the buildup of contaminants is the variation in dilution rates and effectiveness of the ventilation delivery systems often found within buildings. Concentrations vary spatially as well as over time. These variations add further nonuniformity to the pollutant concentration. ANSI/ASHRAE Standard 62.1-2010, Ventilation for Acceptable Indoor

48

Chapter 3 Human Comfort and Air Distribution Air Quality tabulates enforceable and guideline maximum concentration levels of common indoor contaminants (ASHRAE 2010). It also includes the National Primary and Secondary Ambient Air Quality Standards (EPA 1987) used for building ventilation. If the outdoor-air source exceeds the contaminant parameters, it may be cleaned or purified prior to introduction into occupied spaces.

Outdoor-Air Requirements AHRAE Standard 62.1 provides HVAC designers a means of determining ventilation rates necessary to achieve acceptable indoor air quality, which is defined as air in which there are no known contaminants at harmful concentrations as determined by cognizant authorities and with which a substantial majority (80% or more) of the people exposed do not express dissatisfaction.

Two procedures for determining the required ventilation rate are offered to the designer: the Ventilation Rate Procedure and the Indoor Air Quality (IAQ) Procedure. The Ventilation Rate Procedure sets forth prescriptive rates for a variety of applications. Unless unusual pollutants are present, these rates are intended to produce acceptable IAQ. The basis for most of the rates specified is an underlying minimum of 2.5 L/s per sedentary occupant plus a minimum of 0.3 L/s·m2 to deal with pollutants from the space. These minimums are increased for more active occupants, for example 10 L/s per person in a health club aerobics room. Similarly, the space ventilation rate is increased where there are anticipated contaminants, for example 0.6 L/s·m2 in a library. The IAQ Procedure offers an analytical alternative, allowing the designer to determine the ventilation rate based on knowledge of the contaminants being generated within the space and the capability of the ventilation air supply to limit these contaminants to acceptable levels.

Exhaust Requirements Exhaust air systems are either general systems that remove air from large spaces, or local systems that capture aerosols, heat, or gases at specific locations within a space and transport them to where they can be collected, filtered, made inactive, or safely discharged to the atmosphere. The air in local exhaust systems can sometimes be dispersed safely to the atmosphere, but sometimes contaminants must be removed so the emitted air meets air quality standards. ASHRAE Standard 62.1 specifies the exhaust rate for many spaces in terms of L/s·m2. Examples are 2.5. L/s⋅m2 for barber shops, arenas, locker rooms, and copy/print rooms. Twice the exhaust, 5 L/s⋅m2, is required for darkrooms, janitor, trash, recycling, and science classrooms due to the higher anticipated pollution to be removed.

Fundamentals of Air System Design SI

49

Air-Movement Effect ASHRAE Standard 55-2004 includes no minimum air velocity past the occupant for comfort. In the private residential environment, comfort and negligible air movement are the norm. However, the experiences of many commercial building operators have shown air motion is a significant benefit to comfort in mechanically ventilated spaces. The standard further prescribes a maximum rate of air movement of 0.2 m/s to avoid drafts. Higher air speeds (up to 0.8 m/s) may be used to enhance cooling if the air speed is under the occupant’s control.

Minimum Air Changes Low air velocity may affect the ability to maintain uniformity of a comfortable temperature throughout the occupied zone and the dilution of contaminants generated within that zone. Occupant comfort has been reported to suffer as a consequence of low total supply airflow in the space, even when the space temperature is within the comfort envelope. Often, this dissatisfaction is not due to air change but due to a source of warm or cool radiation, poor temperature/ humidity control, or occupant expectations. However, to ensure adequate air changes, many designers have adopted a minimum total supply airflow of 3 to 4 L/s·m2 for office applications. These values are based on an all-air system with conventional mixing supply outlets. They can be reduced when outlets with high induction ratios are employed, because they increase the average room air motion.

Terminal Air Velocity Terminal air velocity is the airstream velocity at the end of the throw (the horizontal or vertical axial distance an airstream travels before the stream velocity is reduced to a specified terminal velocity). The specified terminal velocity must be high enough to maintain the desired level of comfort.

Drafts A draft is a localized effect caused by one or more factors of high air velocity, low ambient temperature, or direction of airflow, where more heat is withdrawn from a person’s skin than is normally dissipated. It can be thought of as any air motion that causes discomfort. Air movement in excess of 0.2 m/s may well be considered a draft. The location of the draft has considerable effect. The back of the neck and the ankle are the most sensitive exposed locations.

Stratification Stratification in a space (such as an atrium or other high-ceiling room) is the division of air into a series of temperature layers. If conditioned air is introduced at about the 3 m level or below, the space close to the floor is conditioned. The cooling requirements of the elements above the 3 m level may be reduced.

50

Chapter 3 Human Comfort and Air Distribution

Principles of Space Air Distribution Room air distribution systems can be classified as mixing, displacement, and local systems.

Mixing Systems Conditioned air is normally supplied to air outlets at velocities much greater than those acceptable in the occupied zone. This relatively high velocity jet of air stimulates mixing and air movement to create relatively uniform air conditions in the occupied zone. The exception is underfloor systems, which supply air from below the floor. Underfloor systems are considered in additional detail in the next section. Mixing air outlets (underfloor excepted) are classified into five groups: •

•

•

Group A outlets are mounted in or near the ceiling and discharge air horizontally (see Figure 3-4). Because these outlets discharge horizontally near the ceiling, the warmest air in the room is mixed immediately with the cool primary supply air above the occupied zone. Consequently, these outlets are capable of handling relatively large quantities of air at large temperature differentials when cooling. During heating, warm supply air introduced at the ceiling can cause stratification in the space if there is insufficient induction of room air at the outlet. Selecting diffusers properly, limiting the room supply temperature differential, and maintaining air supply rates at a level high enough to ensure air mixing by induction can provide adequate air diffusion and minimize stratification. Group B outlets are mounted in or near the floor and discharge air vertically in a nonspreading jet. Figure 3-5 shows that a stagnant zone forms outside the conditioned-air region above its terminal point. Judgment is needed to determine the acceptable size of the space outside the conditioned-air zone. A distance of 4.5 to 6 m between the drop region and the exposed wall is a conservative design value. A comparison of Figures 3-4 and 3-5 for heating shows that the stagnant region is smaller for Group B than for Group A outlets because the air entrained in the immediate vicinity of the outlet is taken mainly from the stagnant region, which is the coolest air in the room. This results in greater temperature equalization and less buoyancy in the total air than occurs with Group A outlets. Cooling effectiveness of Group B is inferior to Group A for the same reasons. Group C outlets are mounted in or near the floor and discharge air in a vertical spreading jet (Figure 3-6). Although outlets of this group are related to Group B, they are characterized by wide-spreading jets and diffusing action. Conditioned air and room air characteristics are similar to those of Group B, but the stagnant zone formed is larger during cooling and smaller during heating.

Figure 3-4

Air motion characteristics of Group A outlets.

Fundamentals of Air System Design SI 51

52

Chapter 3 Human Comfort and Air Distribution

Figure 3-5

Air motion characteristics of Group B outlets.

Diffusion of the primary air usually causes the conditioned air space to fold back on the primary air during cooling, instead of following the ceiling. This diffusing action of the outlets makes it more difficult to project the cool air, but it also provides a greater area for induction of room air. This action is beneficial during heating, because the induced air comes from the lower regions of the room. •

Group D outlets are mounted in or near the floor and discharge air horizontally (see Figure 3-7). This group includes baseboard and low sidewall registers and similar outlets that discharge the primary air in single or multiple jets. However, because the air is discharged horizontally across the floor, the total air during cooling remains near the floor, and a

Fundamentals of Air System Design SI

Figure 3-6

53

Air motion characteristics of Group C outlets.

•

large stagnant zone forms in the entire upper region of the room. During heating, the conditioned air rises toward the ceiling because of the buoyant effect of warm air. The temperature variations are uniform except in the conditioned-air region. Group E outlets are mounted in or near the ceiling and project primary air vertically (see Figure 3-8). During cooling, the conditioned air projects toward and follows the floor, producing a stagnant region near the ceiling. During heating, the conditioned airflow reaches the floor and folds back toward the ceiling. If projected air does not reach the floor, a stagnant zone results.

54

Chapter 3 Human Comfort and Air Distribution

Figure 3-7

Air motion characteristics of Group D outlets.

The principles of air diffusion found in these five groups are as follows: •

The primary air from the outlet down to a velocity of about 0.75 m/s can be treated analytically. The heating or cooling load has a strong effect on the characteristics of the primary air.

•

The conditioned air (shown by lightly shaded envelopes in Figures 3-4 through 3-8) is influenced by the primary air and is of relatively high velocity (but less than 0.75 m/s), with air temperatures generally within 0.5°C of

Fundamentals of Air System Design SI

Figure 3-8

55

Air motion characteristics of Group E outlets.

room temperature. The conditioned air is also influenced by the environment and drops during cooling or rises during heating; it is not subject to precise analytical treatment. •

Natural convection currents form stagnant zones from the ceiling down during cooling and from the floor up during heating. This zone forms below the terminal point of the conditioned air during heating and above the terminal point during cooling. Because this zone results from natural convection currents, the air velocities within it are usually low (approximately 0.1 m/s), and the air stratifies in layers of increasing temperatures. The concept of a stagnant zone is important in properly applying and selecting outlets, because it considers the natural convection currents from warm and cold surfaces and internal loads.

•

A return inlet affects the room air motion only within its immediate vicinity. The intake should be located in the stagnant zone to return the warmest room air during cooling or the coolest room air during heating. The importance of the location depends on the relative size of the stagnant zone that results from various types of outlets.

56

Chapter 3 Human Comfort and Air Distribution •

The general room air motion (shown by clear areas in Figures 3-4 through 3-8) is a gentle drifting of air. Room conditions are maintained by the entrainment of the room air into the conditioned airstream. The room air motion between the stagnant zone and the conditioned air is relatively slow and uniform. The highest air motion occurs in and near the conditioned airstreams.

This review of outlets and their resulting airflows indicates that the air velocity and temperature vary substantially through the occupied space. The airflows are also different in cooling and heating modes. For cooling mode, a standard method for rating diffusers has been developed called the Air Diffusion Performance Index (ADPI). The ADPI for an outlet is the percentage of points within the occupied space where the draft temperature, θ, is between –1.5°C and +1°C and the air velocity is below 0.35 m/s. The draft temperature is a measure of perceived temperature difference at one location compared to the average temperature and air velocity of 0.15 m/s. The draft temperature is calculated as Draft temperature = θ = (tactual – taverage) – 8(local air velocity – 0.15) For example, a location temperature 1°C cooler than room average with air velocity of 0.2 m/s has a draft temperature of θ = ( – 1 ) – 8 ( 0.2 – 0.15 ) = – 1 – 0.4 = – 1.4°C This reduction in perceived temperature due to the draft helps reduce the effect of the space being too warm. It also makes the space feel even colder when the air temperature is cool. Thus, people may complain about drafts when the space is cool, although the air velocity is the same as when the space was warm. To calculate the ADPI, a test room with air supplied 11°C cooler than room average is checked at an array of points within the occupied zone, and the percentage within the draft temperature range is the ADPI. Full details of the ADPI methodology are given in ANSI/ASHRAE Standard 113-2009, Method of Testing for Room Air Diffusion (ASHRAE 2009). Outlet performance selection data from numerous tests is shown for high sidewall outlets in Table 3-1. Now consider a 6 × 6 m office with a cooling load of 70 W/m2 that is cooled by a sidewall diffuser. We see for a room load of 65 W/m2, the maximum ADPI of 85 is obtained with a T0.25/L of 1.5 and the ADPI range is 1.0–1.9 for over 80%. Aiming for the maximum, we would choose a grille with a 0.25 m/s terminal velocity throw of 1.5 times the room length: T0.25/6 = 1.5, so T0.25 = 9 m.

Underfloor Air Distribution Systems Underfloor air distribution (UFAD) is supplied from a raised floor through numerous small floor grilles. The floor typically consists of 600 mm2 metal

Fundamentals of Air System Design SI Table 3-1 Terminal Device

High sidewall grilles

57

Air Diffusion Performance Index (ADPI) Selection Guide

Room Load, W/m2

T0.25/L for Maximum ADPI

Maximum ADPI

250

1.8

68

—

—

190

1.8

72

70

1.5–2.2

125

1.6

78

70

1.2–2.3

65

1.5

85

80

1.0–1.9

For ADPI Range of T0.25/L Greater Than

plates, or tiles, supported by a 250−500 mm high supporting leg, or column, at each corner. Some of the tiles have outlet grilles installed in them. The tiles can be lifted and moved around, making grille relocation, addition, or removal a simple task (see Figure 3-9). Typically, the floor is covered with carpet tiles, and laying these with their joints not aligned with the tile joints substantially reduces uncontrolled leakage from the floor plenum. Air at 14°C to 18°C is supplied to the cavity and discharges through the floor grilles. The floor grilles are designed to create mixing so that the velocity is below 0.3 m/s within 1.2 m of the floor. You can think of the air as turbulent columns spreading out as they flow toward the ceiling. Return air is taken from the ceiling or high on the wall. The rising column of air takes contaminants with it up and out of the breathing zone. This sweep-away action is considered more effective rather than mix-and-dilute. As a result, the ventilation requirements of ASHRAE Standard 62.1 can be satisfied with 10% less outside air. There are numerous outlets because the individual outlet volume is typically limited to 50 L/s. The entering air does not sweep past the occupants as occurs in displacement ventilation, so there is no restriction on cooling capacity. However, there is a limit on how well the system works with rapidly changing loads. For spaces with high solar cooling loads or high winter perimeter heating loads, thermostatically controlled fan coils or other methods are required to modulate the capacity to match the changing load. Because the air rises toward the ceiling, the convection heat loads above the occupied zone do not influence the occupied zone temperature. Therefore, the return air temperature can be warmer than the occupied zone and a return air temperature sensor is a poor indicator of occupied zone temperature. The cool plenum air flows continuously over the structural floor that acts somewhat as a passive thermal storage unit. This storage can be used to reduce peak loads and means that the system is slow to respond to change. Night setback of temperature is not advisable, and many systems are run continuously but without outside air during unoccupied hours. For perimeter heating, small fan-coil units can be installed under the floor using finned hot-water pipes or electric coils. The tempering of the plenum air as it flows over the structure often makes it necessary to duct the plenum air some 3−4 m to the perimeter fan coils to maintain an adequately low supply

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Chapter 3 Human Comfort and Air Distribution

Figure 3-9

Underfloor air distribution.

temperature. Similarly, conference rooms that have a highly variable load can have a thermostatically controlled fan to boost the flow into the room when it is occupied. A modification of the underfloor system with individual grilles is the use of a porous floor. The floor tiles are perforated with an array of small holes, and a porous carpet tile allows air to flow upwards over the entire tile area. This is a modification of the standard grill and has yet to gain popularity. The underfloor air delivery system has the following advantages: •

Changing the layouts of partitions and electrical and communications cables is easy. For buildings with high churn (frequent layout changes), this flexibility may in itself make the added cost of the floor economically justified.

•

The flow of air across the concrete structural floor provides passive thermal storage.

•

When the main supply duct and branches to the floor plenums are part of a well-integrated architectural design, the air-supply pressure drop can be very low, resulting in fan power savings.

•

Less ventilation outside air can potentially be used.

Fundamentals of Air System Design SI

59

Disadvantages include • • •

a significant cost per square metre for the floor system supply, installation, and maintenance; a tendency to require a greater floor-to-floor height, because space for lights and return air ducts is still required at the ceiling level; and a need for specific and detailed knowledge and skills on the part of the designer and installers. Examples include coordination between floor layout and duct layout to avoid floor pedestals through the duct, and sealing the structure and other service penetrations into the plenum to minimize uncontrolled leakage.

Displacement Systems In displacement systems, conditioned air with a temperature slightly lower than the desired room air temperature in the occupied zone is supplied from air outlets at low air velocities of 0.5 m/s or less. The outlets are at or near the floor level for comfort conditioning, and the supply air is directly introduced to the occupied zone. Returns are located at or close to the ceiling through which the warm room air is exhausted from the room. The supply air is spread over the floor and then rises as it is heated by the heat sources in the occupied zone. Heat sources (such as people, computers, etc.) in the occupied zone create upward convective flows in the form of thermal plumes. These plumes remove heat and contaminants, as they are less dense than the surrounding air (see Figure 3-10). In contrast to mixing ventilation, displacement ventilation is designed to minimize mixing of air within the occupied zone.

Unidirectional Airflow Systems In a unidirectional airflow system, air is either supplied from the ceiling and exhausted through the floor, typical of many hospital operating theatre systems (or vice versa), or supplied through the wall and exhausted through returns at the opposite wall, typical of many industrial cleanroom systems. The outlets are uniformly distributed to provide a low turbulent airflow across the entire room. This type of system is mainly used for ventilating cleanrooms or high air-change areas in which the main objective is to remove contaminant particles within the room. It is also used in areas where a unidirectional airflow is desired (such as computer rooms, paint booths, etc.).

Local Systems Air is supplied locally for occupied regions, such as desks in offices or working places in industrial buildings. Conditioned air is supplied toward the breathing zone of the occupants to create comfortable conditions and/or to reduce the concentration of pollutants. Several special air diffusers are available. Figure 3-11

60

Chapter 3 Human Comfort and Air Distribution

Figure 3-10

Schematic of displacement ventilation.

Figure 3-11

Localized ventilation.

Fundamentals of Air System Design SI

61

shows one such arrangement, with diffusers placed on the desks in front of the occupants and the supply air coming from a raised floor plenum.

Exhaust and Return Air Pickup Return and exhaust air openings should be located to minimize short circuiting of supply air into the return air opening. If air is supplied by the jets attached to the ceiling, exhaust openings should be located between the jets or at the other side of the room away from the supply air jets. In a room with temperature stratification along its height, exhaust openings should be located near the ceiling to collect warm air, odors, and fumes. For industrial rooms with gas release, selection of exhaust opening locations depends on the specific weight of the released gases and their temperatures. The locations should be specified for each application. Exhaust outlets located in walls, depending on their elevation, have the characteristics of either floor or ceiling returns. In large buildings with many small rooms, return air should not be brought through door grilles or undercuts into the corridors and then to a common return or exhaust, because smoke would accumulate in the main egress pathway in the case of a fire. Most building codes restrict the application.

Room System Balancing Room system balancing is adjusting the airflows within the room so they are in accordance with specified design quantities. In designing a system, ducts and diffusers should be sized so that the supply of air is distributed properly. However, for flexibility and cost considerations, standard sizes are typically used. Consequently, the room as designed may not be self balancing. The results of an unbalanced room system can be drafts, doors slamming shut or open, and other undesirable effects. Air system balancing is discussed in Chapter 11.

Types of Air Distribution Devices Supply air outlets and diffusing equipment introduce air into a conditioned space to obtain a desired environment. Return and exhaust air are removed from a space through return and exhaust inlets. This section discusses some common types of diffusing equipment.

Supply Air Outlets The following basic supply outlet types are commonly available: grille outlets, slot diffuser outlets, ceiling diffuser outlets, and perforated ceiling panels. These differ in their construction features, physical configurations, and the way they diffuse or disperse supply air and induce or entrain room air into a primary airstream.

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Chapter 3 Human Comfort and Air Distribution

Grille Outlets A grille outlet may be louvered or perforated and located in a sidewall, ceiling, or floor. Several types of grilles are available: •

•

• •

Adjustable bar grille. This is the most common type of grille, used as a supply outlet. It is available as either a single-deflection grille (with a single set of vanes) or double-deflection grille (with two sets of vanes, one in front of the other at right angles to each other). Vertical vanes deflect the airstream in the horizontal plane; horizontal vanes deflect the airstream in the vertical plane. Fixed bar grille. This type of grille is similar to the adjustable singledeflection grille except that the vanes are not adjustable. The vanes may be straight or set at an angle. The angle at which the air is discharged from this grille depends on the type of deflection vanes. Stamped grille. This grille is stamped from a single sheet of metal to form openings through which air can pass. Variable area grille. This type of grille is similar to the adjustable doubledeflection grille but can vary the discharge area to achieve an air volume change (variable volume outlet) at constant pressure, so that the variation in throw is minimized for a given change in supply air volume.

Properly selected grilles operate satisfactorily from high side wall and perimeter locations in the sill, curb, or floor. Ceiling-mounted grilles, which discharge the airstream down, are generally unacceptable in comfort airconditioning installations in interior zones and may cause drafts in perimeter applications. Accessories available for grille outlets include the following: •

•

•

•

Opposed-blade dampers. These can be attached to the backs of grilles (the combination of a grille and a damper is called a register) or installed as separate units in the duct (see Figure 3–12a). Adjacent blades of this damper rotate in opposite directions and may receive air from any direction, discharging it in a series of jets without adversely deflecting the airstream to one side of the duct. Parallel-blade dampers. These have a series of gang-operated blades that rotate in the same direction (see Figure 3–12b). This uniform rotation deflects the airstream when the damper is partially open. Gang-operated turning vanes (extractors). These are sometimes installed in collar connections to grilles near the main ducts. The device shown in Figure 3-12c has vanes that pivot and remain parallel to the duct airflow, regardless of the setting. This allows for field adjustment, which the fixed set of vanes shown in Figure 3-12d do not allow. Dual blade collector. Figure 3-12e shows a dual blade collector and turning vane allowing directional control of the air as it enters the outlet.

Fundamentals of Air System Design SI

Figure 3-12

63

Grille and register outlet accessory controls.

Slot Diffuser Outlets A slot diffuser is an elongated outlet consisting of single or multiple slots. It is usually installed in long, continuous lengths. Outlets with dimensional aspect ratios of 25:1 or greater and a maximum opening of approximately 75 mm generally meet the performance criteria for slot diffusers. Slot diffusers are usually equipped with accessory devices for uniform supply air discharge along the entire length of the slot. While accessory devices help correct the airflow pattern, proper approach conditions for the airstream are also important for satisfactory performance. When the plenum supplying a slot diffuser is designed, the transverse velocity in the plenum should be less than the discharge velocity of the jet, as recommended by the manufacturer and as shown by experience. If tapered ducts are used for introducing supply air into the diffuser, they should be sized to maintain a velocity of about 2.5 m/s and tapered to maintain constant static pressure. Slot diffusers that have a single-slot discharge are available for use in conjunction with recessed fluorescent light troffers. A plenum mates with a light fixture and is concealed from the room. It discharges air through openings in the fixture and is available with fixed or adjustable air discharge patterns, air distribution plenum, inlet dampers for balancing, and inlet collars suitable for

64

Chapter 3 Human Comfort and Air Distribution flexible duct connections. Accessories available for slot diffuser outlets include dampers and flow equalizing vanes.

Ceiling Diffuser Outlets A ceiling diffuser is a supply air diffuser designed for ceiling mounting. A number of designs are available: •

•

•

•

•

Multipassage ceiling diffusers. These diffusers consist of a series of flaring rings or louvers that form a series of concentric air passages. They may be round, square, or rectangular. For easy installation, these diffusers are often made in two parts: an outer shell with a duct collar, and a removable inner assembly. Flush and stepped-down diffusers. In the flush unit, all rings or louvers project to a plane surface. In the stepped-down unit, the rings project beyond the surface of the outer shell. Perforated diffusers. These meet architectural demands for air outlets that blend into ceilings. Each has a perforated metal face with an open area of 10% to 50% that determines its capacity. Units are usually equipped with deflection devices to obtain multipattern horizontal air discharge. Large perforated diffusers are used to provide laminar flow in laboratories, hospital operating rooms, and other spaces having high air-change rates. Designers are cautioned to thoroughly investigate the airflow and induction characteristics under both cooling and heating conditions for this type of diffuser, particularly in applications with varying airflows, such as variable-air-volume systems. Variable-area diffusers. These feature a means of varying the discharge area to achieve an air volume change (variable volume outlet) at a constant pressure so that the variation in throw is minimized for a given change in supply air volume. Antismudge rings. These are round or square metal frames attached to and extending approximately 100 to 300 mm beyond the outer edge of the diffuser. Their purpose is to minimize ceiling smudging.

Dampering and accessories of various types are available for ceiling diffusers: •

•

Multilouver dampers. Consisting of a series of parallel blades mounted inside a frame, multilouver dampers are installed in the diffuser collar or the duct system branch. The blades are usually arranged in two groups rotating in opposite directions and are key operated from the face of the diffuser (see Figure 3-13a). Opposed-blade dampers. These usually consist of a series of pie-shaped vanes mounted inside a round frame installed in the diffuser collar or the duct system branch. The vanes pivot about a horizontal axis and are arranged in two groups, with adjacent vanes rotating counter to each other

Fundamentals of Air System Design SI

Figure 3-13

65

Ceiling diffuser outlet accessory controls.

(see Figure 3-13b). The vanes are key operated from the diffuser face. Another opposed-blade design is similar in construction to the damper shown in Figure 3-12a and has either a round or square frame. Designers should note that volume-control devices near outlets can generate objectionable noise. •

Blank-off baffles. These baffles are used for minor adjustments of the airflow from a diffuser. They blank off a section of the diffuser and prevent the supply air from striking an obstruction such as a column, partition or the wall of the conditioned space by reducing flow in a given direction. Blankoff baffles generally reduce the area and increase supply air velocity, which must be considered when selecting diffuser size. Pattern control in diffusers having removable directional cores may be accomplished by rearranging the cores, generally without a change in area or increase in velocity.

Due to noise considerations, dampening in the branch duct to the diffuser is preferable to a damper in the diffuser, as long as there is easy access to the damper.

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Chapter 3 Human Comfort and Air Distribution

Procedure for Outlet Selection The following procedure is generally used in selecting outlet locations and types: • •

•

•

Based on system design and heating and cooling load calculations, determine the amount of air to be supplied to each room. Select the type and quantity of outlets for each room, considering such factors as air quantity required, distance available for throw or radius of diffusion, structural characteristics, and architectural concepts. Table 3-2 is based on experience and typical ratings of various outlets. It may be used as a guide for the outlets applicable for use with various room air loadings. Manufacturers’ ratings should be consulted to determine the suitability of the outlets used. Locate outlets in the room to distribute the air as uniformly as possible. Outlets may be sized and located to distribute air in proportion to the heat gain or loss in various parts of the room. Select proper outlet size from manufacturers’ ratings according to air quantities, discharge velocities, distribution patterns and sound levels. Obstructions to the primary air distribution pattern require special consideration.

Other Supply Air Outlet Considerations Other supply air outlet considerations include the surface effect, smudging, and sound level. The induction or entrainment characteristics of a moving airstream cause a surface effect. An airstream moving adjacent to or in contact with a wall or ceiling surface creates a low-pressure area immediately adjacent to that surface, causing the air to remain in contact with the surface substantially throughout the length of throw. The surface effect counteracts the drop of horizontally projected cool airstreams. Smudging occurs with the use of ceiling and slot diffusers. Dirt particles held in suspension in the room air are subjected to turbulence at the outlet face. Table 3-2

Outlet Usage Guide

Type of Outlet

Air Loading of Floor Space, L/s·m2

Approximate Maximum ach for a Three-Meter Ceiling

Grille

3–6

7

Slot

4–10

12

Perforated panel

4.5–15

18

Ceiling diffuser

4–25

30

Perforated ceiling

5–50

60

Fundamentals of Air System Design SI

67

This turbulence is primarily responsible for smudging. The cleanliness of the room affects when the smudging becomes visible. An outlet’s sound level is a function of the damper arrangement, discharge velocity, and transmission of systemic noise, all of which are influenced by the size of the outlet and the duct velocity. Higher-frequency sounds can be the result of excessive outlet velocity but may also be generated in the duct by the moving airstream. Lower-pitched sounds are generally the result of mechanical equipment noise transmitted through the duct system and outlet. The cause of higher-frequency sounds can be pinpointed as outlet or systemic sounds by removing the outlet during operation. A reduction in sound level with the outlet removed indicates the portion of the noise caused by the outlet. If the sound level remains essentially unchanged, the system is at fault. Sound is covered in greater detail in Chapter 10. Suggested duct velocities, where takeoffs to grilles or diffusers are close to the outlet are as follows: •

Acceptable high noise levels: 7.5 m/s maximum

• •

General office of classroom: 5 m/s maximum Noise-sensitive areas: 4 m/s maximum

Return inlets may either be connected to a duct or be simple vents that transfer air from one area to another. Exhaust inlets remove air directly from a building and are always connected to a duct or directly to outside. Whatever the arrangement, inlet size and configuration determine velocity and pressure requirements for the required airflow. In general, the same types of equipment (for example, grilles, slot diffusers, and ceiling diffusers) used for supplying air may also be used for air return and exhaust. Return and exhaust inlets do not require the accessory devices used in supply outlets. However, dampers are necessary when it is desirable to balance the airflow in the return duct system. Return and exhaust inlets may be mounted in almost any location, including ceilings, high or low side walls, and floors, when using mixing systems for supply. When using displacement and underfloor supply, distributed ceiling exhaust is required. The opposed-blade dampers shown in Figure 3-12a are used in conjunction with grille return and exhaust inlets. The type of damper does not affect the performance of the inlet. Usually, no other accessory devices are required.

The Next Step This chapter discussed human comfort and air distribution within the occupied space. This air is supplied by a distribution system. Chapter 4 introduces the various air systems that provide the conditioned air.

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Chapter 3 Human Comfort and Air Distribution

Summary The space, the individual, and the individual’s current activity affect human comfort. Thermal comfort depends on individual characteristics and the ability of the body to reject heat primarily by convection and evaporation. Radiation is usually less important. The main thermal factors affecting comfort are: dry-bulb temperature, relative humidity, thermal radiation, air movement, insulation value of clothing, activity level, and direct contact with warmer or cooler surfaces. ASHRAE Standard 55 details requirements for thermal comfort. ASHRAE Standard 62.1 prescribes supply ventilation rates, requirements for contaminant removal, and exhaust rates for human satisfaction with air quality. Room air distribution systems are classified as mixing, underfloor, displacement, and local systems. In mixing systems, the air enters the occupied zone at a fairly high velocity and mixes with zone air to be at an acceptable velocity and average temperature in the occupied space. Inlets are divided into five groups for air movement classification. With mixing systems, the profile of the primary air jet can be forecast with some certainty. But as the primary jet mixes with room air, its behavior is modified by the temperature difference between primary air and room air and the shape of the space. When the primary air is cooler, there is a tendency for the air to drop and for stagnant areas to occur near the floor. In most comfort situations, a general drift of air occurs through the space. This general drift is not influenced significantly by the location of the return air outlet. The performances of various types of mixing outlets were analyzed for their relative ability to maintain comfort conditions throughout a space. This data is presented as the Air Diffusion Performance Index and can be used to make choices about outlets. Underfloor air distribution (UFAD) uses a plenum created above the structural floor using 600 mm2 metal panels on support columns. The air at 14°C to 18°C is supplied up through diffusers distributed among the floor panels. The system uses the vertical supply and convection to lift the air toward ceiling outlets. As the air flows across the structure, the structure acts as a thermal buffer, and the system is slow to change. In UFAD systems, perimeter heating and cooling can be challenging. The use of perimeter fan coils and some ducting may be required to provide adequate capacity. The UFAD system has advantages in layout flexibility, structural thermal storage, lower fan power in some cases, and a 10% lower requirement for outside ventilation air. However, these advantages must be balanced with the cost of the floor, possibly greater floor-to-floor height, and a need for competent design and construction. Displacement systems, for comfort, supply a large volume of low-velocity air near room temperature. Outlets are close to or at floor level, so the air sweeps across the space and convection lifts the contaminated air to high-level return outlets. The system minimizes mixing.

Fundamentals of Air System Design SI

69

A wide range of grille outlets with fixed or/and adjustable vanes provides a supply of air shaped from a narrow jet perpendicular to the room surface for a long throw to a wide spreading, short-throw jet. The flow and throw may be adjusted by using the grille blades or adjustable damper and turning vane accessories. Slot diffuser outlets are long, narrow (75 mm or less), grille-like outlets designed for ceiling installation on their own or as part of the long side of fluorescent light fixtures. Their long, narrow supply of air mixes quickly with room air to avoid primary air drafts in the occupied zone. Due to their length, the air supply must be carefully designed to obtain consistent performance. Supply air ceiling diffusers have flaring vanes with an open or perforated face. They spread the air across the ceiling, entraining room air to produce a large volume of well-mixed circulating air. Having established the quantity of supply air, a preliminary choice of outlet style and layout can be made. Using manufacturers’ data on airflow and sound generation, the final airflows and layout are determined. The diffuser choice is often significantly influenced by the available duct space for bringing air to the space and the room aesthetics. Return air outlets can use the same grilles or diffusers but no direction control is needed, although a damper for balancing may be required. The location is not critical for mixing systems but must be high in the room for floor supply and displacement systems.

References and Bibliography ASHRAE. 2004. ANSI/ASHRAE Standard 55-2004, Thermal Environmental Conditions for Human Occupancy. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. ASHRAE. 2007. Air-Conditioning Systems Design Manual, Second edition. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. ASHRAE. 2007. ASHRAE Handbook—HVAC Applications. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. ASHRAE. 2008. ASHRAE Handbook—Systems and Equipment. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. ASHRAE. 2009. ASHRAE Handbook—Fundamentals. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. ASHRAE. 2009. ANSI/ASHRAE Standard 113-2009, Method of Testing for Room Air Diffusion. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. ASHRAE. 2010. ANSI/ASHRAE Standard 62.1-2010, Ventilation for Acceptable Indoor Air Quality. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. EPA. 1987. National Primary and Secondary Ambient Air Quality Standards. Code of Federal Regulations, Title 40 Part 50 (40 CFR 50). United States Environmental Protection Agency, Washington, D.C.

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Chapter 3 Human Comfort and Air Distribution

Skill Development Exercises for Chapter 3 Complete these questions by writing your answers on the worksheets at the back of this book. 3-1

The human body uses which of the following heat transfer mechanisms? a) radiation

3-2

b) convection

c) evaporation

d) all of the above

The perception of comfort relates to which of the following? a) individual physical condition b) body heat exchange with the surroundings c) physiological characteristics d) all of the above

3-3

Based on the comfort chart from ASHRAE Standard 55 (Figure 3-2), which of the following would be within the acceptable range of temperature and humidity for human comfort when wearing light summer clothing? a) 27°C, 30% RH c) 23°C, 40% RH

3-4

In a system with 7000 annual wet-bulb degree hours above 19°C, with a maximum 55% RH indoor desired, and 72 h of cooling system operation per week, the energy used will be _______________ × 1000 mJ per year per 1000 L/s. a) 87

3-5

b) 25°C, 50% RH d) both a and b

b) 130

c) 155

d) none of the above

The __________________ procedure for determining the required ventilation rate is based on knowledge of the contaminants being generated within the space and the capability of the ventilation air supply to limit them to acceptable levels. a) indoor air quality c) contaminant mitigation

3-6

b) ventilation rate d) all of the above

Many designers have adopted a minimum total supply airflow of ___________ for office applications. a) 0.6 to 1.8 L/s·m2

b) 3.6 to 4.8 L/s·m2

c) 1.2 to 12.0 L/s·m2 d) all of the above 3-7

The airstream velocity at the end of the throw is called _______________. a) terminal velocity c) airstream velocity

3-8

b) primary velocity d) all of the above

_________________ air distribution systems create relatively uniform air conditions in the occupied zone. a) Unidirectional

b) Local

c) Mixing

d) All of the above

Fundamentals of Air System Design SI 3-9

The stagnant region of a Group B mixing outlet in a heating-only system is _______________ the stagnant region of a Group A mixing outlet. a) larger than c) smaller than

3-10

b) the same as d) all of the above

In displacement systems, the outlets are frequently located _______________. a) at or near the floor level c) in the ceiling

3-11

b) in the walls d) all of the above

Smudging is most likely to occur from dirt particles held in suspension in _______________ a) room air

3-12

71

b) supply air

c) return air

d) all of the above

The fan power for underfloor supply systems can often be less than required for a ceiling supply mixing system due to which of the following? a) much cooler supply air b) the low resistance to airflow in the plenum c) the insulating value of the floor and carpet

3-13

Underfloor supply systems work well for large open areas and the most effective control is a thermostat in the return duct. a) True

b) False

Relationship of Air Systems to Load and Occupancy Demands Study Objectives After completing this chapter, you should be able to describe operating system criteria; T air systems by heating/cooling equipment type; T air systems by duct configurations; and T considerations for outdoor-air intake. T

Instructions Read the material in Chapter 4. Verify the examples presented in the chapter with your own calculations. At the end of the chapter, complete the skill development exercises without referring to the text. Review those sections of the chapter as needed to complete the exercises.

Operating System Selection Criteria To select an operating system, detailed building design and use information and weather data at selected design conditions are required. Although a detailed discussion of load calculations is outside the scope of this course, the air system designer should be aware that generally all of the following are considered when performing load calculations: • •

Building characteristics. Determine building materials, areas, external surface colors, and shapes from building plans and specifications. Building configuration. Determine building location, orientation, and external shading from building plans and specifications. Shading from adjacent buildings should be carefully evaluated to assess its probable permanence before including it in the calculation. The possibility of abnormally high ground-reflected solar radiation (for example from adjacent water, sand, or parking lots) or solar load from adjacent reflective buildings should be considered. The thermal zones within the building should be identified. For example, external offices with windows have different thermal characteristics than windowless rooms in the interior of the building. Additionally, some

74

Chapter 4 Relationship of Air Systems to Load and Occupancy Demands

•

•

•

•

areas of the building may have to be kept at different temperatures than others. Room pressure relationships should be considered. For example, in a building with a natatorium (swimming pool), the air-pressure gradients within the building should draw air into the natatorium from the rest of the building rather than vice versa. This prevents the rest of the building from smelling like a swimming pool. The same concept applies in buildings with laboratories or areas where noxious smells may be generated. The uses to which the building will be put affect the levels of noise permissible in the building. For example, an office environment is typically less tolerant of noise from the HVAC system than a warehouse. Outdoor design conditions. Obtain appropriate weather data (wet- and drybulb temperatures, daily range, heating and cooling degree days, elevation, etc.) and select outdoor design conditions from local weather stations. The 2009 ASHRAE Handbook—Fundamentals also lists outdoor design conditions for a large number of weather stations across the world. The National Climatic Center in Asheville, North Carolina, has additional data. Space psychrometric requirements. Select indoor design conditions, such as indoor dry-bulb temperature range and indoor wet-bulb temperature (or relative humidity) range. Note that ANSI/ASHRAE Standard 55-2004,Thermal Environmental Conditions for Human Occupancy (ASHRAE 2004) deals only with comfort. The maximum and minimum levels specified for comfort are often excessively wide for ensuring no mold growth in the building fabric or occupant complaints about low humidity in cold climates. Include permissible variations and control limits. Different areas within a building may have different psychrometric requirements (for example a facility having a cleanroom, temperature controlled laboratory, and general office space). Outdoor-air ventilation requirements for each space. ASHRAE Standard 62.1 specifies the methods of calculating the required supply ventilation rates and exhaust rates for polluted areas, such as toilets. The proper design and sizing of central heating and air-conditioning systems requires more than calculation of the cooling load in the space to be conditioned. The type of heating and air-conditioning system, fan energy, fan location, duct heat loss and gain, duct leakage, heat extraction lighting systems, and type of return air system all affect system load and component sizing. Adequate system design and component sizing require system performance be analyzed as a series of psychrometric processes. The 2008 ASHRAE Handbook—HVAC Systems and Equipment (ASHRAE 2008) and the 2009 ASHRAE Handbook—Fundamentals (ASHRAE 2009) describe elements of this technique in detail. Operating schedule. Obtain a proposed schedule of lighting, occupants, internal equipment, appliances, and processes that contribute to the internal thermal load. Determine the probability that the cooling equipment will be operated continuously or shut OFF during unoccupied periods (such as

Fundamentals of Air System Design SI

•

•

75

nights and weekends). Performance of the system at part-load conditions must be considered. Date and time. Frequently, several different times of day and several different months must be analyzed to determine the peak load time. For example, Table 4-1 shows peak loads in buildings having a large amount of glass located at 32°N latitude. Owning and operating costs. The total cost of a facility includes the cost of the HVAC system. The cost of an HVAC system is customarily broken down into owning costs and operating costs. Owning costs include the initial cost of the system and annual fixed charges that are present whether the system is used at all (taxes, insurance, etc.). Operating costs are what it costs to run the system, including energy and maintenance costs. The 2007 ASHRAE Handbook—HVAC Applications (ASHRAE 2007b) provides a detailed discussion of this subject.

System Types by Heating/Cooling Equipment Type Unitary Equipment Systems Unitary equipment systems are factory-assembled into an integrated package, including fans, filters, heating coil, cooling coil, refrigerant compressors, refrigerant-side controls, air-side controls, and condenser. This equipment is manufactured in various configurations to meet a wide range of applications. Window air conditioners, through-the-wall room air conditioners, rooftop packaged units, air source heat pumps, and water source heat pumps are examples. This equipment can be applied in single units and as multiple units to form a complete air-conditioning system for a building. Single-space applications. Window-mounted and through-the-wall mounted air conditioners and heat pumps are designed to cool or heat individual room spaces. They include a complete system in an individual package. Each room is an individually controlled zone. Air conditioners and heat pumps are installed in buildings that require many temperature control zones (such as motels, apartments, and dormitories). These systems are applicable for renovation of existing Table 4-1

Peak Load Times

Perimeter Zone

Peak Load Time

Month

East

8:00 a.m.

August

West

4:00 p.m.

August

South

12:00 noon

December

North

12:00 noon

June

Northeast and Southeast

10:00 a.m.

March and October

Northwest and Southwest

2:00 p.m.

March and October

Note: Interior zones at times of peak occupancy.

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Chapter 4 Relationship of Air Systems to Load and Occupancy Demands buildings because existing systems can still be used. However, the user should be cautioned that these systems do not dehumidify and tend to be noisy and cause drafts. Entire-building applications. Unitary equipment is used in both outdoor and indoor locations to cool and heat entire buildings. The complete system consists of a unit with a condenser, air distribution system, and temperature controls. The equipment may be single or multizone, installed outdoors on the roof or at grade level, or indoors in service areas adjacent to the conditioned space. Totally indoor condenser installations require the unit be water cooled. Multiple-unit systems generally use single-zone units with a unit for each zone (see Figure 4-1). Zoning is determined by cooling and heating loads,

Figure 4-1

Multiple packaged units.

Fundamentals of Air System Design SI

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occupancy considerations, flexibility requirements, and thermal zones. Appearance considerations, costs, and equipment and duct space availability may dictate compromises in selecting the ideal zoning. Designers are also cautioned to evaluate carefully the use of unitary equipment in cases of more than 25% outside air. Many unitary systems do not remove sufficient moisture at high outdoor-air quantities. For adequate part-load performance at high outdoor-air quantities, direct expansion systems may require a hot-gas bypass to prevent coil freezing. In both all-air systems and air-and-water systems, air is used to perform the heating and cooling function within the occupied space. Unitary systems are discussed in detail in the 2008 ASHRAE Handbook— HVAC Systems and Equipment (ASHRAE 2008).

All-Air Systems An all-air system provides complete sensible and latent cooling, preheating, and humidification capacity in the air supplied by the system. No additional cooling or humidification is required at the zone, except in special cases. Heating may be accomplished by the same airstream, either in the central system or at a particular zone. All-air systems may be adapted to many applications for comfort or process work. They are used in buildings that require individual control of multiple zones (such as office buildings, schools and universities, laboratories, hospitals, stores, hotels, and ships). All-air systems are also commonly used in special applications for close control of temperature and humidity (including clean rooms, computer rooms, hospital operating rooms, and research and development facilities) as well as in many industrial/manufacturing facilities. All-air systems have the following advantages: •

•

• • •

The central mechanical equipment room location for major equipment allows operation and maintenance to be performed in unoccupied areas and permits the maximum range of choices of filtration equipment and vibration and noise control. The complete absence within the conditioned area of piping, electrical equipment, wiring, filters, and vibration- and noise-producing equipment reduces potential harm to occupants, furnishings, and processes, thereby minimizing service needs. Systems have the greatest potential for the use of outside air and “free” cooling systems to augment the use of mechanical refrigeration for cooling. Seasonal changeover is simple and readily adaptable to automatic control. A wide choice of zoning, flexibility, and humidity control under all operating conditions is available with the option of simultaneous heating and cooling, even during off-season periods.

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Chapter 4 Relationship of Air Systems to Load and Occupancy Demands • • • • • •

Air-to-air and other heat recovery systems may be readily incorporated. Systems permit good design flexibility for optimum air distribution, draft control, and adaptability to varying load requirements. Systems are well suited to applications requiring unusual exhaust or makeup air quantities (negative or positive pressurization, etc.). Systems adapt well to winter humidification. The primary system may be used to introduce outside air required for ventilation without the need for supplemental systems. By increasing the air change rate, these systems are able to maintain operating conditions of ±0.5°C dry-bulb and ±5% relative humidity fairly simply. There are systems that can essentially maintain constant space conditions. All-air systems have the following disadvantages:

• • • • •

They require additional duct clearance, which reduces usable floor space and increases the height of the building. Depending on layout, vertical shaft space may be needed for distribution, thereby requiring larger floor planes. The accessibility of terminal devices requires close cooperation between architectural, mechanical, and structural designers. Air balancing, particularly in large systems, can be more difficult. Heating systems are not always available for use in providing temporary heat during construction.

Heating and Cooling Calculations Basic calculations for airflow, temperatures, relative humidity, loads, and psychrometrics are covered in the 2009 ASHRAE Handbook—Fundamentals (ASHRAE 2009). It is important that the designer understand the operation of the various components of a system, their relationship to the psychrometric chart, and their interaction under various operating conditions and system configurations.

Categories of All-Air Systems All-air systems are classified in two basic categories: single-duct and dualduct. These classifications may be further divided as follows: • • • • •

Constant volume: single zone, multiple zoned reheat, bypass Variable air volume (VAV): reheat, induction, fan powered, dual conduit, variable diffusers Dual-duct: constant volume, variable volume. Multizone: constant volume, variable volume, three deck, Texas multizone Combinations of the above systems

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Constant-Volume Single-Duct Systems Single-duct systems contain the main heating and cooling coils in a series flow air path. A common duct distribution system at a common air temperature feeds all terminal apparatus. These systems change the supply air temperature in response to the space load. Variations of the constant-volume single-duct system include: single-zone systems, zoned reheat systems, and bypass systems. The single-zone system is the simplest all-air system using a supply unit to serve a single temperature control zone (see Figure 4-2). The unit may be installed within or outside of the space it serves, and it may operate with or without distributing ductwork. In Figure 4-2, heatflows and airflows are indicated by arrows, and temperatures are indicated by t. The subscripts in the sequence of the airflow are: R

=

room

rp

=

return plenum

o

=

outside air

m

=

mixed air

r

=

return air

cc

=

cooling coil

hc

=

heating coil

sf

=

supply fan

s

=

supply

In the psychrometric chart in Figure 4-2, all pertinent points are identified by the same subscripts. The room sensible and latent loads are denoted by qSR and qLR, respectively, and the outdoor-air sensible and latent loads are denoted by qSo and qLo, respectively. The cooling load qcc is the difference in enthalpies between states m and cc. Note that the cooling coil discharge air draws heat from the supply air fan and the supply air ducts, accounting for the difference in dry-bulb temperatures between points cc and s in Figure 4-2 before entering the room. Room sensible and latent loads (due to occupants, lights, machinery, solar radiation, transmission, etc.) are picked up and carried to the return air plenum. Additional heat may be picked up from recessed ceiling lights, floors above, the roof, and the return air fan, accounting for the increase in temperature between points R and r. Some of the air is exhausted, while outside (ventilation) air o is taken in, resulting in a mixed airstream m, which is cooled and dehumidified by the cooling coils, producing the state of air at cc. A heating coil is provided immediately downstream of the cooling coil to raise the air temperature in winter when required. Properly designed systems can maintain temperature and humidity closely and efficiently and can be shut down when desired without affecting

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Chapter 4 Relationship of Air Systems to Load and Occupancy Demands

Figure 4-2

Single-zone schematic and psychrometric chart.

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81

the operation of adjacent areas. They are energy efficient, easy to control, and easily adaptable to economizers. The disadvantage of these systems is that they respond to only one set of space conditions. Therefore, their use is limited to situations where variations occur approximately uniformly throughout the zone served or where the load is stable. Single-zone systems are applicable to small department stores, small individual stores in a shopping center, individual classrooms in a school, computer rooms, hospital operating rooms, and large open areas, such as gymnasiums. For example, a rooftop unit complete with refrigeration system and serving an individual space is considered a single-zone system. However, the refrigeration system may be remote and may serve several single-zone units in a larger installation. A return fan may be necessary to maintain proper space pressure in relation to the outdoor-air inlet pressure and other adjacent spaces. The designer should consider relief-air fans in place of return air fans if the relief path has high pressure losses. A system using multiple fan-coil units is a collection of single-zone systems put together to control different zones. A single-zone system can be controlled by varying the quantity and/or the temperature of the supply air, by providing reheat, by face and bypass dampers, or by a combination of these. The multiple-zoned reheat system is a modification of the single-zone system. It provides zone or space control for areas of unequal load, simultaneous heating or cooling of perimeter areas with different exposures, and close tolerance of control for process or comfort applications and better performance for dehumidification. As the word reheat implies, heat is added as a secondary simultaneous process to preconditioned primary air. Single-duct systems without reheat offer cooling flexibility but cannot control summer humidity independent of temperature requirements. Single-duct systems with reheat provide flexibility for both temperature and humidity control; the cooling coil cools the air to the desired humidity level, and the reheat coil raises the dry-bulb temperature to the desired value. However, ANSI/ASHRAE/IESNA Standard 90.1-2007, Energy Efficient Design of New Buildings Except Low-Rise Residential Buildings (ASHRAE 2007a) severely restricts the application of reheat, limiting this option to special cases because of the high energy consumption of the system. If high humidity and low dry-bulb temperatures are desired, a humidifier may have to be included in the system. The bypass system is a variation of the constant-volume reheat system using face and bypass dampers in place of reheat. This system is essentially a constant-volume primary system and may have a VAV secondary system.

Variable-Air-Volume Single-Duct Systems A VAV system (as shown in Figure 4-3) controls temperature within a space by varying the quantity of supply air rather than varying the supply air temperature.

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Chapter 4 Relationship of Air Systems to Load and Occupancy Demands

Figure 4-3

Variable-air-volume system schematic and psychrometric chart.

A VAV terminal device is used at the zone to vary the quantity of supply air to the space. The supply air temperature is held relatively constant, depending on the season. VAV systems are easy to control, are highly energy efficient, allow fairly good room control, and are easily adaptable to economizers. A potential

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drawback includes the possibility of poor ventilation, particularly under low zone loads. They are suitable for offices, classrooms, and many other applications, and are currently widely used for commercial and institutional buildings despite the fact that humidity control under widely varying latent loads is difficult. With the current concern for indoor air quality, care should be exercised to provide minimum ventilation in any occupied space and required outdoor-air quantities under all operating conditions. The pressure relationships of the system change when the supply fan is throttled. Means such as outdoor-air injection fans with capacity control may be required. The typical return air fan generally should not be used because it is difficult to control supply and return fans in tandem. If relief is necessary, a relief fan with capacity control may be used. VAV systems are available in a number of configurations, including the following: •

Simple VAV. This system applies to cooling-only service with no requirement for simultaneous heating and cooling in different zones; a typical application is the interior of an office building. To permit system volume variations without fan volume control, on chilled-water systems the air supply can ride the fan curve down to the lowest acceptable airflow, usually at least 50% of the full airflow. Care must be exercised in the selection of air outlets to maintain the desired mixing and throw conditions. Avoid varying zone air volume while keeping fan and system volume substantially constant by dumping excess air into a return air ceiling plenum or directly into the return air duct system. Dumping cold air into the return air plenum wastes energy and can cause overcooling under low load conditions due to radiation from the cool ceiling surface to the zone below. Dumping can also cause a shortage of system volume if it is used for system balancing as well as temperature control. Dumping and bypassing are generally undesirable. Fan speed control is preferred. Figure 4-4a shows the simplest of three VAV box arrangements. It is a pressure-independent box, which means that it adjusts to allow for variations in supply duct pressure. The unit has a velocity sensor that is used to control for constant velocity and, hence, volume. The room thermostat requests more or less flow to maintain the room temperature. The box is lined with acoustically absorbent material, typically protected fiberglass, to reduce any noise from the higher-pressure air going over the control damper. Figure 4-4b shows the VAV box with a reheat coil. Typically, the supply volume is throttled to minimum flow before the coil is operated to provide heating. The final diagram shows a series fan box. This type of box can be used to maintain the air distribution within the space by keeping a constant volume flowing into the space. The fan capacity meets, or exceeds, the maximum primary supply airflow. When the primary airflow is reduced, the fan draws more air from the ceiling plenum, maintaining the

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Chapter 4 Relationship of Air Systems to Load and Occupancy Demands

Figure 4-4

Variable-air-volume box arrangements.

constant flow. A heating coil may also be included so that the fan and coil can run as a fan-coil heater unit with the primary air system OFF during unoccupied hours. •

VAV reheat or VAV dual duct. Full heating/cooling flexibility can be achieved more energy efficiently after throttling the cold air supply to the zone.

•

VAV perimeter system. All-air cooling and heating can be accomplished by a constant-volume system serving interior spaces in connection with a VAV perimeter system. The constant-volume system provides cooling yearround, taking care of all variations in all zone internal heat gains. The perimeter system can use an outdoor/indoor temperature schedule VAV air supply, which simply offsets the skin transmission gains or losses. The

Fundamentals of Air System Design SI

•

•

•

•

•

85

perimeter system requires individual zone control based on solar exposure. If a hydronic perimeter heating system is provided, the air system accomplishes all cooling in all zones year round, while the perimeter heating system offsets the winter transmission heat losses, but not the summer transmission heat gains. VAV with constant zone volume. Individual zone fans may be used to maintain minimum or constant supply air to the zone while the system primary air fed to the zone is throttled. Terminals in these systems are commonly referred to as fan-powered terminals. The load is satisfied by recirculating return air, thus keeping the sum of the throttled system air and the recirculated return air substantially constant. This technique is particularly useful for zones with large variations of internal loads (such as conference rooms), and it may be combined with terminal reheat. Fan-powered terminals can be used to ensure good air circulation in occupied spaces during periods of reduced cooling load. Care should be taken to ensure that proper outside air is still delivered to the occupied zone when the primary air is throttled. A distributed outdoor-air duct system may be required. VAV with economizer. When the enthalpy of the outside air is lower than that of the return air, chiller power can be reduced by taking in more outside air than is required for ventilation and relieving the excess return air. Under favorable conditions, all of the return air can be relieved and replaced by outside air. This mode of operation is called an economizer cycle. While this cycle requires large outdoor-air intakes and exhausts, it improves the economy of operation except in areas such as the southeastern United States, where these favorable conditions occur so rarely that the additional first cost of providing for economizer operation is not justified. Even so, some Southern states have adopted energy codes that require the use of an economizer. VAV with induction terminal. The VAV induction system uses a terminal unit to reduce cooling capacity by simultaneously reducing primary air and inducing room air or air from the ceiling return plenum to maintain a relatively constant room supply volume. Dual-conduit VAV. The dual-conduit system is designed to provide two air supply paths: one to offset exterior transmission cooling or heating loads and the other where cooling is required throughout the year. The typical terminal device (box) has two inlets, one for cold air and one for hot air or bypass air. Each inlet has a throttling damper and actuator. Typically, the cold damper is throttled to a preset minimum condition before the hot damper is opened. VAV with variable diffusers. These devices reduce the discharge aperture of the diffuser. This keeps the discharge velocity relatively constant while reducing the conditioned supply airflow. Under these conditions, the induction effect of the diffuser is kept high, and cold air mixes in the space.

One of the important and difficult issues with VAV systems is providing enough ventilation air to each space all the time. Consider a simplified example

86

Chapter 4 Relationship of Air Systems to Load and Occupancy Demands where 20% outside air is required at full system flow. If the system, as a whole, throttles back to 80% capacity, the proportion of outside air will rise to 25% (20 out of 80). However, if one of the zones is throttled back to 60% flow, it will only receive 0.6× 0.25 = 0.15, or 15% outside air. ASHRAE Standard 62.1 provides rules for dealing with this issue. A second issue is ensuring adequate air distribution in the space when the volume is throttled back. Diffusers that maintain their performance at reduced flows must be chosen to ensure the ventilation effectiveness is maintained even at times of low airflow.

Constant-Volume Dual-Duct Systems Dual-duct systems contain the main heating and cooling coils in parallel flow or series-parallel flow air paths with either a separate cold and warm-air duct distribution system that blends the air at the terminal apparatus (dual duct systems) or a separate supply air duct to each zone, with the supply air blended to the required temperature at the main unit mixing dampers (multizone). There are two types of constant-volume dual-duct systems: •

•

Single fan, no reheat. This is similar to a single-duct system except that it contains a face-and-bypass damper at the cooling coil arranged to bypass a mixture of outdoor and recirculated air as the latent heat load fluctuates in response to a zone thermostat. Single fan, reheat. This is similar to a conventional reheat system. The difference is that reheat is applied at a central point instead of at individual zones (see Figure 4-5).

Variable-Air-Volume Dual-Duct Systems Dual-duct VAV systems blend cold and warm air in various volume combinations. These systems include: •

•

Single fan. A single supply fan is sized for the coincident peak of the hot and cold decks. Control of the fan is by two static pressure controllers: one located in the hot deck and the other in the cold deck. The duct requiring the highest pressure governs the fan airflow. Dual fan. The volume of each supply fan is controlled independently by the static pressure in its respective duct. The return fan is controlled based on the sum of the hot and cold fan volumes using flow-measuring stations (see Figure 4-6).

Multizone Dual-Duct Systems Multizone systems supply several zones from a single centrally located air-handling unit. Different zone requirements are met by mixing cold and warm air through zone dampers at the central air handler in response to zone thermostats. The mixed, conditioned air is distributed throughout the building by a system of single-zone ducts. The return air is handled in a conventional manner. A Texas

Fundamentals of Air System Design SI

Figure 4-5

Dual-duct system.

Figure 4-6

Dual-duct dual-fan system.

87

88

Chapter 4 Relationship of Air Systems to Load and Occupancy Demands multizone system has a heating coil in each mixed-air zone, which is energized only when the cooling damper is closed.

Air-and-Water Systems Air-and-water systems condition spaces by distributing air and water sources to terminal units installed in habitable space throughout the building. The air and water are cooled or heated in central mechanical rooms. Sometimes a separate electric heating coil is included instead of a hot-water coil. The room terminal may be an induction unit, a fan-coil unit, or a conventional supply air outlet combined with a radiant panel. Generally, the air supply has a constant volume and is called primary air to distinguish it from room air or secondary air that has been induced.

Induction System Figure 4-7 shows a basic arrangement for an air-water induction terminal. Centrally conditioned primary air is supplied to the unit plenum at mediumto-high pressure. The acoustically treated plenum attenuates part of the noise

Figure 4-7

Air-water induction terminal.

Fundamentals of Air System Design SI

Figure 4-8

Fan-coil unit.

Figure 4-9

Ceiling panel example.

89

90

Chapter 4 Relationship of Air Systems to Load and Occupancy Demands generated in the unit and duct system. A balancing damper adjusts the primary air quantity within design limits. These systems are not used very often anymore.

Fan-Coil Systems Figure 4-8 shows a typical fan-coil unit. The basic elements of fan-coil units are a finned-tube coil, filter, and fan section. The fan recirculates air continuously from the space through the coil or coils. The unit may contain an additional electric resistance, steam, or hot-water heating coil.

Panel Heating and Cooling Systems The sensible heating and cooling loads in a zone can be met by using ceiling panels. An example of one type is shown in Figure 4-9. If the panels are used for cooling, the panel temperature must not go below the air dewpoint to avoid any possibility of condensation. The proportion of load is thus limited in cooling applications, less so in heating applications. One very effective system is to use ceiling panels with a dedicated outdoor-air system (DOAS). The DOAS provides a constant volume of conditioned outdoor air for ventilation, humidity control, and some cooling. The balance of the cooling load is absorbed by the ceiling panels. This DOAS with panel cooling For heating, the floor may also be used as the heating panel. Pipes cast into a concrete floor with warm water pumped through provide a large area for low-temperature heating of the space. For wooden floors, the pipes can be run on the underside of the floor with insulation below to maximize the upward heat flow.

Evaporative Cooling Evaporative coolers exchange sensible heat for latent heat. Evaporative air cooling evaporates water into an airstream. Figure 4-10 illustrates thermodynamic changes that occur between the air and water in direct contact in a moving airstream. The continuously recirculated water reaches an equilibrium temperature equal to the wet-bulb temperature of the entering air. The heat and mass transfer between the air and water lowers the air dry-bulb temperature and increases the humidity ratio at a constant wet-bulb temperature. The extent to which the leaving air temperature approaches the thermodynamic wet-bulb temperature of the entering air is expressed as a percentage of evaporative cooling (or saturation effectiveness) and is defined as ( t1 – t2 ) e c = 100 -----------------′ ( t1 – t )

(4-1)

Fundamentals of Air System Design SI

Figure 4-10

91

Thermodynamic interaction of water and air.

where ec =

evaporative cooling or saturation effectiveness, %

t1

=

dry-bulb temperature of the entering air

t2

=

dry-bulb temperature of the leaving air

t′

=

thermodynamic wet-bulb temperature of the entering air

Evaporative air-cooling equipment can be classified as either direct or indirect. Direct evaporative equipment cools air by direct contact with the water, either by an extended wetted-surface material (as in packaged air coolers) or with a series of sprays (as in an air washer). Indirect systems cool air in a heat exchanger, which transfers heat to either a secondary airstream that has been evaporatively cooled (air-to-air) or to water that has been evaporatively cooled (by a cooling tower).

System Type by Duct Configuration Duct construction is classified in terms of application and pressure. HVAC systems in public assembly, business, educational, general factory, and mercantile buildings are usually designed as commercial systems. Air-pollution control systems, industrial exhaust systems and systems outside the pressure range of commercial system standards are classified as industrial systems.

92

Chapter 4 Relationship of Air Systems to Load and Occupancy Demands The designer must select a numerical static pressure class or classes that satisfy the requirements of the particular system. Duct pressure classification and duct construction are discussed in Chapter 7.

Air-Side Economizers Air-handling systems that have access to 100% outside air can provide full cooling without the assistance of mechanical refrigeration whenever the outside temperature is lower than the required supply air temperature. This socalled air-side economizer (see Figure 4-11) is progressively more effective in northern latitudes, saving up to 70% of mechanical refrigeration energy. In southern areas (such as Florida), the air-side economizer is seldom used. This is because the number of hours during which the outside enthalpy falls below the controlled space temperature is insufficient to justify the investment in the return air fan, air-mixing chambers, and louvers necessary to dissipate the air pressure caused by supplying 100% outside air. More energy savings are achieved with an economizer when •

•

•

the outdoor-air enthalpy is lower than the supply air enthalpy required to meet the space-cooling load; compressors and chilled water pumps are turned OFF; and outdoor-air, return air, and exhaust air dampers are positioned to attain the required space temperature. the outdoor-air enthalpy is higher than the supply air enthalpy but is lower than the return air enthalpy, compressor and chilled-water pumps are energized, and the dampers are positioned for 100% outside air. the outdoor-air enthalpy exceeds the return air enthalpy and the dampers are positioned to bring in the minimum outdoor air required for ventilation.

As a simple rule of thumb, air-side economizers can be based on dry-bulb temperature (Figure 4-12). But to be truly effective, economizer operation should be based on enthalpy, as shown in Figure 4-13. Compartmented air-handling systems that lack the potential for 100% outside air may adopt a winter “free cooling” concept by adding a heat exchanger in the supply airstream to circulate the cooling tower water for cooling rather than the chilled water. This adds capital cost for the heat exchanger. Waterside free cooling is less energy conserving than air-side free cooling, depending on climate. Another form of free cooling involves purging the conditioned areas with cool night air. Cool night air is passed through the building to cool the entire structure. Specifically, lights and the structure above them become warm throughout the day. The cool night purge removes this stored energy, which reduces the air-conditioning load the following day. This purging cycle is highly effective in dry climates with low nighttime temperatures, such as in the southwestern United States, but should not be used in humid climates because of the potential moisture buildup.

Fundamentals of Air System Design SI

Figure 4-11

Airside economizer.

Figure 4-12

Airside temperature economizer cycle.

93

94

Chapter 4 Relationship of Air Systems to Load and Occupancy Demands

Figure 4-13

Enthalpy economizer cycle.

Outdoor-Air Intake Outdoor air is air outside a building or air taken from outdoors and not previously circulated through the system. Outdoor air that flows through a building, either intentionally as ventilation air or unintentionally as infiltration, is important for two reasons: • •

Outdoor air is used to dilute indoor air contaminants. The energy associated with heating or cooling outdoor air is a significant space-conditioning load.

In large buildings, the effect of infiltration and ventilation on distribution and interzone airflow patterns, which include smoke circulation patterns in the event of fire, should be determined; for more information, see “Fire and Smoke Management” in the ASHRAE Handbook—HVAC Applications. Outside air can be used to pressurize the building and minimize infiltration. Outdoor-air intakes should be located so that cross-contamination from exhaust fans to the intake louver does not occur. Outdoor air is typically drawn in through louvers designed to minimize the entry of snow, water, birds, trash, and other foreign matter into the equipment. Figure 4-14 depicts a typical outdoor-air louver design. The screen and louver are located sufficiently above the roof to minimize the pickup of roof dust and the probability of snow accumulating. This height is determined by the annual snowfall. However, a minimum of 0.75 m is recommended for most areas. In some locations, doors are added outside the louver for closure during very bad weather (such as hurricanes and blizzards).

Fundamentals of Air System Design SI

Figure 4-14

95

Outdoor-air louver and screen.

When outdoor air must be drawn in through the roof, a gooseneck outdoorair intake as shown in Figure 4-15 may be used. Codes also restrict the location of inlets to minimize drawing in contaminated air. ASHRAE Standard 62.1 requires the following to minimize rain entrainment: Use rain hoods sized for no more than 2.5 m/s face velocity with a downward-facing intake so that all intake air passes upward through a horizontal plane that intersects the solid surfaces of the hood before entering the system.

The Next Step This chapter considered supply air systems. Chapter 5 considers exhaust systems to remove excess air and contaminants from a building.

Gooseneck outdoor-air intake.

Chapter 4 Relationship of Air Systems to Load and Occupancy Demands

Figure 4-15

96

Fundamentals of Air System Design SI

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Summary System selection depends on many factors, including the following: • • • • • •

Building construction Building layout Schedule of operation and use of spaces Summer and winter external design conditions Internal design requirements and limits for ventilation, filtration, temperature, humidity, and pressure Owning and operating cost requirements

Once the requirements are known, the most appropriate system can be selected. Complete, factory assembled units range from the small window airconditioner serving a single room to large packages serving a whole building. Larger units may be supplied as a set of bolt-together parts. These range from the economical mass-produced window unit up to the best one-off designed unit. An all-air system provides complete sensible and latent cooling, heating, humidification, ventilation, and filtration through the air supplied to each space. Their main advantage is that the equipment is located outside the occupied space, which is particularly important in many clean spaces in manufacturing and medical facilities. These systems allow for free cooling with outside air and heat recovery from the exhaust; they provide air for processes with high exhaust needs as well as flexibility in zoning and control performance. Disadvantages of all-air systems include requiring space for ducting to each zone from the mechanical room, careful integration with the architectural layout, and other services. Systems provide temperature control by varying the air volume and/or temperature to each zone. For a system serving a single zone, this can be achieved at the main unit. For multiple zones, the varying loads in each zone can be served by one of the following main system types or a modification of such: • • • •

Multizone: Mixing of warm and cool air at the main unit to provide a separately ducted supply to each zone. VAV: Single supply duct supplying cool air to a variable-air-volume damper on the branch to each zone (with a reheat coil if required). Reheat: Single supply duct supplying a constant volume of cool air with a reheat at each zone branch. Dual duct: Two ducts, one with warm air and one with cold air, run through the building. At each zone, air from each duct is connected to a dual-duct box, which chooses the proportion of warm and cool air to deliver to the zone to maintain temperature control.

98

Chapter 4 Relationship of Air Systems to Load and Occupancy Demands Air-and-water systems provide ventilation and humidity control by supplying air to each zone, while the majority of the cooling and heating loads are dealt with by water coils in the zone. The ventilation air may be used as a power source for inducing room air over the coil, as in induction system, or fan-coil units may be used. Evaporative coolers evaporate water into the air. The water absorbs latent heat to evaporate. This heat comes from the air, which lowers the air temperature. In direct evaporative coolers, cooler, wetter air is produced. In indirect evaporative coolers, water is cooled by evaporation and used in coils to cool the air with no increase in air moisture content. Duct construction is classified in terms of application and pressure and is discussed in detail in Chapter 7. Mechanical cooling can be minimized by using outside air whenever the outdoor-air enthalpy is lower than the return air enthalpy. Depending on the climate, this may occur most of the year or almost never. The saving in mechanical cooling operating cost is somewhat offset by the additional first cost of larger intake, exhaust, and control dampers. Outdoor air is normally drawn in through louvers designed to minimize the entry of rain, snow, water, birds, trash, and other foreign matter into the equipment. The intake should be located to minimize drawing in pollutants.

References and Bibliography ASHRAE. 2004. ANSI/ASHRAE Standard 55-2004, Thermal Environmental Conditions for Human Occupancy. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. ASHRAE. 2007a. ANSI/ASHRAE/IESNA Standard 90.1-2007, Energy Efficient Design of New Buildings Except Low-Rise Residential Buildings. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc ASHRAE. 2007b. ASHRAE Handbook—HVAC Applications. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. ASHRAE. 2008. ASHRAE Handbook—Systems and Equipment. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. ASHRAE. 2009. ASHRAE Handbook—Fundamentals. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. ASHRAE. 2010. ANSI/ASHRAE Standard 62.1-2010, Ventilation for Acceptable Indoor Air Quality. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. Carrier. 1967. Altitude Effects. Syracuse, NY: Carrier Corp.

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Skill Development Exercises for Chapter 4 Complete these questions by writing your answers on the worksheets at the back of this book. 4-1

External offices with windows will have different thermal characteristics than windowless rooms in the interior of the building. a) True

b) False

4-2

In a building with a natatorium, the air pressure gradients within the building should _______________. a) draw air from the natatorium into the rest of the building b) draw air into the natatorium from the rest of the building c) relieve the natatorium air intake d) all of the above

4-3

Which of the following is an advantage of an all-air system? a) additional duct clearance is not required b) air balancing in large systems is less difficult c) vertical shaft space is not required d) none of the above

4-4

Single-duct, single-zone systems can respond simultaneously to more than one set of space conditions, in more than one area at a time. a) True

4-5

b) False

In air-and-water systems, the air supply generally has a constant volume. a) True

b) False

4-6

Evaporative coolers _______________. a) evaporate water into an airstream b) exchange sensible heat for latent heat c) can be either direct or indirect d) all of the above

4-7

An economizer can achieve energy savings when _______________. a) the outdoor-air enthalpy is lower than the supply air enthalpy b) the outdoor-air enthalpy is higher than the supply air enthalpy, but lower than the return air enthalpy c) both of the above d) none of the above

4-8

A minimum height of _______________ above the roof surface is recommended for locating outdoor-air louvers where light snowfall is expected. a) 0.25 m

b) 0.50 m

c) 0.75 m

d) 1.00 m

Exhaust and Ventilation Systems Study Objectives After completing this chapter, you should be able to describe design considerations for exhaust and ventilation systems and some energy recovery systems.

Instructions Read the material in Chapter 5. Verify the examples presented in the chapter with your own calculations. At the end of the chapter, complete the skill development exercises without referring to the text. Review those sections of the chapter as needed to complete the exercises.

Design Considerations Ventilation and exhaust systems control heat, odors, and contaminants. The two types of exhaust systems are • •

general exhaust, in which an entire workspace is exhausted without considering specific operations, and local exhaust, which is applied to specific areas. Local exhaust offers better control with minimum air volumes, thereby lowering the cost of air cleaning and replacement air equipment. Local exhaust is required for hazardous contaminant exhaust.

Ventilation may be provided by natural draft, by a combination of general supply and exhaust air fan and duct systems, by exhaust fans only (with makeup air through inlet louvers and doors), or by supply fans only (exhaust through relief louvers and doors).

Ventilation System Selection and Design Some factors to consider in ventilation system selection and design include the following: • •

Local exhaust systems provide general ventilation for the work area. A balance of the supply and exhaust systems is required for either system to function as designed.

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Chapter 5 Exhaust and Ventilation Systems •

•

•

• • •

•

•

Natural ventilation systems are most applicable when internal heat loads are high and the building is tall enough to produce a significant stack effect (such as steelmaking plants and glass-melting furnaces). To provide effective general ventilation for heat relief by either natural or mechanical supply, the air must be delivered low in the work zones. A sufficient exhaust volume is necessary to remove the heat liberated in the space. Local relief systems may require supplemental supply air for heat removal. Supply and exhaust air cannot be used interchangeably. Supply air can be delivered where it is wanted at controlled velocities, temperature, and humidity. Exhaust systems should be used to capture heat and fumes at the source. General building exhaust may be required in addition to local exhaust systems. The exhaust discharge should not be located where it will be recirculated into the outdoor-air intake. The inlet air quantity of the exhaust is established by the volume and velocity required to contain and remove heat and contaminants. For human occupancy, ANSI/ASHRAE Standard 62.1-2010, Ventilation for Acceptable Indoor Air Quality (ASHRAE 2010) has requirements for ventilation air and exhaust, as described in Chapter 3. For industrial applications, minimum values are prescribed for local exhaust systems in Industrial Ventilation: A Manual of Recommended Practice (ACGIH 2010) and sometimes by code. Properly sized ductwork keeps contaminants flowing. This requires high velocities for heavy materials. The selection of materials and the construction of exhaust ductwork and fans depend on the nature of the contaminant, the ambient temperature, the lengths and arrangement of duct runs, the method of fan operation, and the flame and smoke spread rating. Care must be taken to minimize the following: ° ° ° °

•

Corrosion, or destruction by chemical or electrochemical action Dissolution, a dissolving action; coatings and plastics are subject to dissolution, particularly by solvent fumes Melting, which can occur in certain plastics and coatings at such elevated temperatures as may be found in an exhaust system Abrasion from conveyed particles impacting the duct, particularly at fittings

Low temperatures that cause condensation in ferrous metal ducts may increase corrosive attack. Ductwork is less subject to attack when the runs are short and direct to the terminal discharge point. The longer the runs are, the longer are the period of exposure to fumes and the greater the degree of condensation. Horizontal runs allow moisture to remain longer than it can on vertical surfaces. Intermittent fan operation can contribute to longer periods of wetness (because of condensation) than can continuous operation. Exhaust ducts from high-moisture areas (such as shower rooms) must have drains and watertight bottoms. Corrosion-resistant material should be considered.

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National and local Clean Air Acts have requirements for controlling the discharge of contaminants to the atmosphere.

Makeup Air For safe, effective operation, most industrial plants require makeup air to replace the large volumes of air exhausted. If makeup air is provided consistently with good air distribution, more effective cooling can be provided in the summer and more efficient and effective heating will result in the winter. Using windows or other inlets that cannot be used in stormy weather should be discouraged. The needs for makeup air include • • •

• •

replacing air exhausted from combustion processes, local exhausts, and general exhaust systems. eliminating cross drafts through proper arrangement of supply air and exhausts. preventing infiltration through doors, windows, and similar openings that may cause discomfort or adversely affect processes through cross drafts or uncontrolled temperature variations. obtaining clean air. Supply air can be filtered; infiltration air cannot. controlling building pressure and airflow from space to space. Such control is necessary ° ° °

Table 5-1

to avoid positive or negative pressures that make it difficult or unsafe to open doors and to avoid the conditions that are detailed in Table 5-1. to confine contaminants, reduce their concentration, and to control temperature, humidity and air movement positively. to recover heat and conserve energy.

Negative Pressures That May Cause Unsatisfactory Building Conditions

Negative Pressure, Pa

Adverse Conditions

2.5 to 5

Worker Draft Complaints: High-velocity drafts through doors and windows

2.5 to 12

Natural Draft Stacks Ineffective: Ventilation through roof exhaust ventilators, flow through stacks with natural draft greatly reduced

5 to 12

Carbon Monoxide Hazard: Backdrafting takes place in hot water heaters, unit heaters, furnaces, and other combustion equipment not provided with induced draft

7 to 25

General Mechanical Ventilation Reduced: Airflows reduced in propeller fans and low pressure supply and exhaust systems

12 to 25

Doors Difficult to Open: Serious injury may result from nonchecked, slamming doors

25 to 50

Local Exhaust Ventilation Impaired: Centrifugal fan exhaust flow reduced

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Chapter 5 Exhaust and Ventilation Systems

Stack Effect Temperature differences between indoors and outdoors cause density differences and, therefore, pressure differences that drive infiltration. During the heating season, warmer air rises and flows out of the building near its top. It is replaced by colder outdoor air that enters the building near its base. During the cooling season, the neutral level is elevated, because the indoor-outdoor temperature differences are smaller or reversed. Qualitatively, the pressure distribution over the building in the heating season due to the stack effect takes the form shown in Figure 5-1. The height at which the interior and exterior pressures are equal is called the neutral pressure level. Above this point (during the heating season), the interior pressure is greater than the exterior pressure; below this point, the greater exterior pressure causes airflow into the building. The pressure difference due to the stack effect at height h is ⎛ T o – T i⎞ Δp s = ( ρ o – ρ i )g ( H NPL – H ) = ρ o ⎜ ----------------⎟ g ( H NPL – H ) ⎝ Ti ⎠

Figure 5-1

Pressure differences due to stack effect (heating season).

(5-1)

Fundamentals of Air System Design SI where Δps ρ g H

= = = =

105

pressure difference due to stack effect, Pa air density, kg/m3 (about 1.2 kg/m3 for indoor conditions) gravitational constant, 9.81 m/s2 height of observation, m

HNPL =

height of neutral pressure level above reference plane with no other driving forces, m To, Ti = outside and inside absolute temperature (°C + 273), K

Ventilation and Exhaust Systems This section describes some of the more common ventilation and exhaust systems.

Ventilation for Heat Relief Many situations involve processes that release heat and moisture to the environment. Ventilation is one of many controls that may be used to mitigate heat stress conditions. The HVAC designer must distinguish between the control needs for hot-dry and warm-moist conditions. In the first case, the process gives off only sensible and radiant heat without adding moisture to the air. The heat load on exposed workers is increased, and the rate of cooling by evaporation of sweat is increased. Heat balance may be maintained, although possibly at the expense of excessive sweating. In the warm-moist situation, the wet process gives off mainly latent heat. The rise in the heat load on workers may be small, but the increase in moisture content of the air reduces heat loss by evaporation of sweat by the workers. Hot-dry work situations occur around hot furnaces, forges, metal-extruding and rolling mills, glass-forming machines, and so forth. Typical warm-moist operations are found in many textile mills, laundries, dye houses, and deep mines where water is used extensively for dust control. However, these industrial applications are outside the scope of this course. Where appropriate, local exhaust ventilation can remove the natural convection column of heated air rising from a hot process with a minimum of air from the surrounding space.

Toilet Exhaust The ventilation of locker rooms, toilets, and shower spaces is important to remove odor and reduce humidity. Supply air may be introduced through door or wall grilles. In some cases, plant air may be so contaminated that filtration or mechanical ventilation may be required. When mechanical ventilation is used, the supply system should have supply fixtures, such as wall

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Chapter 5 Exhaust and Ventilation Systems grilles, ceiling diffusers, or supply plenums, to distribute the air adequately throughout the area. Pressure relationships must be carefully considered to prevent airflow from locker rooms, toilets, and shower spaces to other occupied spaces. In the absence of specific codes, Table 5-2 provides a guide for ventilation of these spaces. Remember that some codes prohibit combining toilet exhaust with other exhaust systems.

Kitchen Exhaust Kitchens typically have a great concentration of noise, sensible and latent heat load, smoke, and odors. Ventilation is the chief means of removing and preventing these elements from entering other occupied spaces. Kitchen air pressure should be kept negative relative to other areas to ensure odor control. Maintenance of reasonably comfortable working conditions is important. Kitchens present common load problems encountered in other occupied spaces, with additional factors that include • • • •

extremely variable loads with high peaks, in many cases occurring twice daily; high sensible and latent heat gains because of appliances, people, and food; heavy infiltration of outdoor air through doors during rush hours in commercial establishments; and grease in the ductwork.

Codes require exhaust hoods with grease filters for cooking equipment where grease is generated, and hoods over all gas-fired appliances. Other equipment that generates a lot of heat or moisture should be located under hoods. Whether through the use of hoods or otherwise, ASHRAE Standard 62 requires a minimum exhaust rate from commercial kitchens of 3.5 L/s⋅m2. Table 5-2

Ventilation for Locker Rooms, Ancillary and Toilet Spaces Description

Units

Coat hanging or clean change room for nonlaboring shift employees with clean work clothes

1.25 L/s·m2

Change room for laboring employees with wet or sweaty clothes*

2.5 and 3.5 L/s·m2

Change room for laborers or workers assigned to heavy work and where clothes will be wet or pick up odors*

4 and 5 L/s·m2

Toilets, public (per ASHRAE Standard 62-2007)

23/35 L/s·unit

Toilets, private (per ASHRAE Standard 62-2007)

12.5/25 L/s·unit

Shower spaces (at least 25 L/s per shower head)

10 L/s·m2

*This ventilation rate is to be exhausted from each locker.

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Smoke Control When a fire occurs in a building, smoke often flows to locations remote from the fire, threatening life and damaging property. Stairwells and elevators frequently become smoke filled, blocking or inhibiting evacuation. Smoke causes the most deaths in fires. Smoke control describes systems that use pressurization produced by mechanical fans to limit smoke movement in fire situations. A smoke-control system must be designed so that it is not overpowered by the driving forces that cause smoke movement, including the following: •

•

•

Stack effect. As discussed earlier, when the air outside a building is colder than the air inside a building, the building air moves upward within building shafts (such as stairwells, mechanical shafts, and elevator shafts). This is the normal stack effect. When the outside air is warmer than the building air, a downward, or reverse stack effect occurs. Smoke movement from a building fire can be dominated by stack effect. In a building with normal stack effect, the existing air currents can move smoke considerable distances from the fire origin. Buoyancy. High-temperature smoke from a fire has a buoyancy force due to its reduced density. As smoke travels away from the fire, its temperature drops due to heat transfer and dilution. Therefore, the effect of buoyancy generally decreases with distance from the fire. Expansion. In addition to buoyancy, the energy released by a fire can move smoke by expansion. The ratio of volumetric flows can be expressed as a ratio of absolute temperatures: Q out T out ---------- = --------Q in T in where Qout Qin Tout Tin

•

(5-2)

= volumetric flow rate of smoke out of the fire compartment, m3/s or L/s = volumetric flow rate of smoke into the fire compartment, m3/s or L/s, consistent with Qout = absolute temperature of smoke leaving the fire compartment, K (°C + 273) = absolute temperature of smoke entering the fire compartment, K (°C + 273)

Wind. Frequently in fire situations, a window breaks in the fire compartment. If the window is on the leeward side of the building, the negative pressure caused by the wind vents the smoke from the fire compartment. This reduces smoke movement throughout the building. However, if the broken window is on the windward side, the wind forces the smoke

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•

throughout the fire floor and to other floors, which endangers the lives of building occupants and hampers firefighting. Pressures induced by the wind in this situation can be large and can dominate air movement throughout the building. HVAC system. The HVAC system frequently transports smoke during fires. Before the concept of using the HVAC system for smoke control, systems were shut down when fires were discovered. Although shutting the system down prevents it from supplying air to the fire, it does not prevent smoke movement through the supply and return air ducts, air shafts, and other building openings due to stack effect, buoyancy, or wind.

Additional information on smoke control can be found in the 2007 ASHRAE Handbook—HVAC Applications (ASHRAE 2007b).

Stair Pressurization Systems Many pressurized stairwells have been designed and built to provide a tenable escape route in the event of a building fire. They also provide a tenable staging area for firefighters. On the fire floor, a pressurized stairwell must maintain a positive pressure difference across a closed stairwell door so that smoke does not enter the stairwell. During building fire situations, some stairwell doors are opened intermittently during evacuation and firefighting, and some doors may even be blocked open. Ideally, when the stairwell door is opened on the fire floor, airflow through the door should be sufficient to prevent smoke backflow. Designing such a system is difficult because of the many combinations of open stairwell doors and weather conditions affecting airflow. The stairwell pressurization fan must be sized to allow for doors to be open to floors and often to the outside during the fire. If no doors are open, the static pressure could easily rise high enough to make opening doors very difficult. To avoid this overpressurization, some form of pressure control is often provided. A simple barometric relief damper with wind shield can be used to relieve any excess pressure to the atmosphere. Alternatively, pressure sensors measuring the pressure between a floor and the stairwell can control a damper on a shortcircuit duct around the fan. When the pressure rises above the setpoint pressure, the damper opens to let air short circuit around the fan, thereby lowering its capacity. The maximum allowed design pressure difference across a door is typically 50–75 Pa so that it can be opened. The minimum pressure to hold back smoke is about 20 Pa, so the pressure control should be designed to hold the pressure from floor to stairwell in that range. Controls to limit differential pressures at the doors are very complicated and difficult to maintain. Stairwell pressurization systems may be single and multiple injection systems. A single injection system has pressurized air supplied to the stairwell at one location, usually at the top. Associated with this system is the potential of

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smoke entering the stairwell through the pressurization fan intake. Therefore, automatic shutdown during such an event should be considered. For tall stairwells, single-injection systems can fail when a few doors are open near the airsupply injection point. Such a failure is especially likely when a ground-level stairwell door is open in bottom-injection systems. Multiple injection points are recommended no more than 14 m apart (see Figures 5-2a and 5-2b). Compartmentation of a stairwell is depicted in Figure 5-3. Additional information on stair pressurization can be found in the 2007 ASHRAE Handbook—HVAC Applications.

Healthcare Facilities The application of air conditioning to healthcare facilities presents many problems not encountered in the usual comfort conditioning system. The basic differences between air conditioning for medical facilities and other types of facilities stem from • •

the need to restrict air movement in and between the various departments; specific requirements for ventilation and filtration to dilute and remove contamination in the form of odor, airborne microorganisms and viruses, and hazardous chemical and radioactive substances;

(a)

Figure 5-2

(b)

(a) Stairwell pressurization, ground-level fan and (b) stairwell pressurization, roof-mounted fan.

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Chapter 5 Exhaust and Ventilation Systems

Figure 5-3

Compartmentation of pressurized stairwell.

• •

the need for different temperature and humidity requirements for various areas; and the need for sophistication in design to permit accurate control of environmental conditions.

The specific environmental conditions required by a particular medical facility can be complex and vary depending on the planned use of the facility and the agency responsible for the facility environmental standard. Healthcare facilities are discussed in greater detail in the 2007 ASHRAE Handbook— HVAC Applications.

Energy Recovery Much of this chapter focuses on exhaust. Air that has been cooled or heated before exhausting takes energy from the building. In many situations, it is both possible and practical to recover some of that energy. Energy recovery may be of sensible heat or sensible and latent heat.

Fundamentals of Air System Design SI

Figure 5-4

111

Runaround energy recovery coils.

Energy Recovery Coils Runaround Coils One way to achieve energy recovery is with runaround energy recovery coils. A typical runaround coil arrangement is shown in Figure 5-4. In summer, the conditioned exhaust air cools the fluid in the exhaust air coil. This fluid is then pumped over to the supply air coil to precool the incoming outside air. In winter, the heat transfer works the other way; the warm exhaust air heats the fluid in the exhaust air coil, which is then pumped over to the supply air coil to heat the cold incoming air. At intermediate temperatures, the system is shut OFF, because it is not useful. When outside temperatures are below freezing, the three-way valve is used with a glycol antifreeze mixture in the coils. In cold weather, some of the fluid bypasses the supply air coil to avoid overcooling. The combination of very cold fluid from the supply air coil and diverted fluid mix to a temperature that is high enough to avoid causing frost on the exhaust air coil. The maximum

112

Chapter 5 Exhaust and Ventilation Systems amount of cooling that can be achieved with the exhaust air coil is limited by the temperature at which frost starts to form in the coil. This frosting of the exhaust coil effectively sets a limit to the transfer possible at low temperatures. The runaround coil system has three particular advantages: •

•

•

There is no possibility of cross contamination between the two airstreams. This factor makes it suitable for hospital or fume hood exhaust heat recovery. The exhaust coil must be resistant to corrosion from any chemicals in the exhaust. The two coils do not have to be adjacent to one another. A laboratory building could have the outdoor-air intake low in the building and the fume hood exhaust on the roof, with the runaround pipes connecting the two coils. The runaround coils transfer sensible heat and, under favorable conditions, condense the water in the exhaust to recover latent heat. This makes them particularly suitable for natatoriums in some climates.

Heat Pipes A heat pipe is a length of pipe with an interior wick that contains a charge of refrigerant, as shown in Figure 5-5. The type and quantity of refrigerant installed is chosen for the particular temperature requirements. In operation, the pipe is approximately horizontal, and one end is warmed, which evaporates refrigerant. The refrigerant vapor fills the tube. If the other half of the tube is cooled, the refrigerant will condense and flow along the wick to the heated end to be evaporated once more. This heat-driven refrigeration cycle is surprisingly efficient.

Figure 5-5

Cutaway section of a heat pipe.

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The normal heat pipe unit consists of a bundle of pipes with external fins and a central divider plate. Figure 5-6 shows a view down onto a unit that is mounted in the relief and intake airstreams to an air-handling unit. Flexible connections are shown that facilitate the tipping. To adjust the heat transfer, one or the other end of the tubes would be lifted. The outside air is cold as it comes in over the warm coil. This warms the air, and the tube is cooled. The cooled refrigerant inside condenses and gives up its latent heat, which heats the air. The recondensed refrigerant wicks across to the exhaust side and then absorbs heat from the exhaust air. This heat evaporates the refrigerant back into a vapor that fills the pipe and is again available to warm the cold outside air. The usual heat pipe unit must be approximately horizontal to work well. A standard way to reduce the heat transfer is to tilt the evaporator (cold) end up a few degrees. This tilt control first reduces and then halts the flow of refrigerant to the evaporator end, and the process stops. Figure 5-6 was based on winter operation. In summer, the unit only has to be tilted to work the other way and cool the incoming outside air as it heats the outgoing exhaust air. The unit is designed as a sensible heat-transfer device, although allowing condensation to occur on the cold end can transfer worthwhile latent heat. Effectiveness ratings range up to 80% with 14 rows of tubes. However, each

Figure 5-6

Heat pipe assembly in exhaust and outdoor-air entry.

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Chapter 5 Exhaust and Ventilation Systems additional row contributes proportionally less to the overall performance. As a result, the economic choice is ten or fewer rows. A major advantage of the unit is very low cross contamination.

Desiccant Wheels Desiccants are chemicals that are quick to pick up heat and moisture and quick to give them up again if exposed to a cooler, drier atmosphere. A matrix, as shown on the left of Figure 5-7, may be coated with such a chemical and made up into a wheel several centimeters thick. In use, the supply air is ducted through one half of the wheel and the exhaust air is ducted through the other half. On a hot summer day the exhaust is cooler and drier than the supply of outside air. The chemical coating in the coil section that’s in the exhaust stream becomes relatively cool and dry. Now the wheel is slowly rotated, and the cool, dry section moves into the incoming hot, humid air, drying and cooling the air. Similarly, another section moves from hot and humid into cool and dry, where it gives up moisture and becomes cooler. The wheel speed—a few revolutions per minute—is adjusted to maximize the transfer of heat and moisture. Control of wheel speed to truly maximize savings is a complex issue, because the transfers of sensible and latent heat do not vary in direct relation to each other. The depth of the wheel is filled with exhaust air as it passes into the supply airstream, so there is some cross contamination. There are ways of minimizing this cross contamination, but it cannot be eliminated. In most comfort situations, the cross contamination in a well-made unit is quite acceptable.

Figure 5-7

Desiccant-wheel matrix and operation.

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115

The use of heat recovery is required in many energy codes, particularly for larger systems and systems with a high proportion of outside air. ANSI/ ASHRAE/IESNA Standard 90.1-2007, Energy Standard for Buildings Except Low-Rise Residential Buildings (ASHRAE 2007a) has several mandatory requirements for the use of heat recovery equipment.

The Next Step This chapter covered ventilation and exhaust. Chapter 6 covers fans and the movement of air through systems.

Summary Ventilation and exhaust systems control heat, odors, and contaminants. Exhaust can be local, removing the contaminant before it mixes with the air in the space or general, changing the air in the space on a regular basis to keep the concentration of contaminants down to an acceptable level All the air exhausted must enter the building so that there is a continuous balance. Failure to provide adequate supply air makeup can create problems of pressure difference. Therefore, the supply and exhaust systems, though physically separate, must be designed as a total system. Usually human comfort dictates supply quantities, but for commercial and industrial processes exhaust is usually the criteria. Temperature, corrosion, and erosion may all dictate the design and construction of the exhaust system. In natural exhausts, stack effect can be used as the motive power where there are sufficient and reliable temperature differences. Ventilation for heat relief under hot working conditions is used in many industries. It is less effective in moist conditions because sweating is less effective. Locker rooms and toilets should be kept at a slightly negative pressure relative to surrounding areas to contain smells. Building codes usually dictate the minimum exhaust per fixture. Kitchen exhaust fumes are typically warm, aromatic, and grease laden. Most codes require the use of grease filters to reduce the quantity of grease (which deposits in the ducts) and to reduce the likelihood of fire entering the ducts. The large quantity of exhaust makes kitchens a challenge for supplying adequate makeup air at a reasonable operating cost. Smoke-control systems are designed to provide a small pressure difference between the fire zone and other zones. Maintaining this difference (less than 25 Pa) can be very difficult due to • • •

stack effect where the difference between inside and outside temperatures causes pressure differences, buoyancy of the hot gases from a fire, expansion of the air due to temperature around the fire,

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Chapter 5 Exhaust and Ventilation Systems • •

wind blowing past the building, creating a higher pressure on the windward side and a lower pressure on the leeward side, and the HVAC system if it is left running.

Stairwell pressurization is provided to keep smoke out of the means of escape and firefighter access routes. Design is a challenge because the pressure must be maintained even with doors open but be limited to prevent doors being held shut by the pressure. Barometric dampers and short-circuit dusts on fans are used to regulate the effective supply fan capacity. Energy recovery from large exhausts is often economically very attractive and is mandated in energy codes. With energy recovery coils, one coil in the exhaust piped to another coil in the makeup air system allows the energy to literally be pumped from exhaust to intake. In freezing climates, an antifreeze mixture is used. The system has the advantages of enabling the intake and exhaust to be separated, and there is zero cross contamination. Heat pipes use the boiling and condensation of refrigerant in sealed lengths of pipe to transfer heat from one end of the tube to the other. Capacity is controlled by tilting the pipes. Some cross contamination may occur. Desiccant wheels are deep porous wheels coated in a chemical to collect heat and moisture. The wheel slowly rotates in the two airstreams, collecting moisture and energy from one airstream and giving up energy and moisture to the other airstream. Their value is in very high recovery rates. However, there is some cross contamination, which is an issue in processes with toxic exhaust contaminants.

Bibliography ACGIH. 2010. Industrial Ventilation: A Manual of Recommended Practice, 27th ed. Cincinnati, OH: American Conference of Governmental Industrial Hygienists. ASHRAE. 2007a. ANSI/ASHRAE/IESNA Standard 90.1-2007, Energy Standard for Buildings Except Low-Rise Residential Buildings. American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. ASHRAE. 2007b. ASHRAE Handbook—HVAC Applications. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. ASHRAE. 2008. ASHRAE Handbook—Systems and Equipment. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. ASHRAE. 2009. ASHRAE Handbook—Fundamentals. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. ASHRAE. 2010. ANSI/ASHRAE Standard 62.1-2010, Ventilation for Acceptable Indoor Air Quality. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc.

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Skill Development Exercises for Chapter 5 Complete these questions by writing your answers on the worksheets at the back of this book. 5-1

Natural ventilation systems are most applicable when the building produces a significant stack effect. a) True

5-2

b) False

Care must be taken in exhaust systems to minimize _______________. a) corrosion

5-3

5-4

b) short and vertical

c) direct to the terminal discharge

d) all of the above

Kitchen air pressure should be kept _______________ relative to other areas. b) neutral

c) negative

d) all of the above

Smoke movement is driven by _______________. b) buoyancy

c) expansion

d) all of the above

To prevent smoke infiltration on a fire floor, a pressurized stairwell must maintain a _______________ pressure difference across a closed stairwell door. a) positive

5-7

d) all of the above

a) long and horizontal

a) stack effect 5-6

c) melting

All other things being equal, ductwork is least subject to condensation corrosion when the runs are _______________.

a) positive 5-5

b) dissolution

b) neutral

c) negative

d) all of the above

Health facility ventilation requires _______________. a) little need for accurate control of temperature and humidity b) free movement of air between departments c) removal of airborne microorganisms d) all of the above

Air Movers and Fan Technology Study Objectives After completing this chapter, you should be able to list and explain fan principles; T list and describe the main types of HVAC fans, fan drives, and fan controls; T explain factors to be considered when selecting an appropriate fan for a given set of conditions; and T explain factors to be considered when installing a fan once it has been selected. T

Instructions Read the material in Chapter 6. Verify the examples presented in the chapter with your own calculations. At the end of the chapter, complete the skill development exercises without referring to the text. Review those sections of the chapter as needed to complete the exercises.

Fan Principles A fan is an air pump that creates a pressure difference and causes airflow. The impeller performs work on the air, imparting both static and kinetic energy, which varies in proportion depending on the fan type. Symbols and definitions commonly encountered when working with fans include the following: A

=

fan outlet area, m2

D

=

fan size or impeller diameter

N

=

rotational speed, rpm (sometimes revolutions per second)

Q

=

volume flow rate moved by fan at fan inlet conditions, m3/s or L/s

ptf

=

fan total pressure rise, or the fan total pressure at outlet minus fan total pressure at inlet, Pa

pvf

=

fan velocity pressure, or the pressure corresponding to average velocity determined from the volume flow rate and fan outlet area, Pa

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Chapter 6 Air Movers and Fan Technology psf

=

fan static pressure rise, or fan total pressure rise diminished by fan velocity pressure, Pa (The fan inlet velocity head is assumed equal to zero because the inlet is not connected to ductwork and is unobstructed for fan rating purposes.)

V

=

fan inlet or outlet velocity, m/s

Wo

=

power output of fan based on fan volume flow rate and fan total pressure, kW

Wi

=

power input to fan measured by power delivered to fan shaft, kW

ht

=

mechanical efficiency of fan (or fan total efficiency), or the ratio of power output to power input (ht = Wo /Wi)

hs

=

static efficiency of fan, or mechanical efficiency multiplied by the ratio of static pressure to fan total pressure, hs = (ps /pt)ht

ρ

=

gas density, kg/m3

Principles of Fan Operation Fans produce pressure by altering the velocity vector of the flow. Fans produce pressure and/or flow because the rotating blades of the impeller impart kinetic energy to the air by changing its velocity. This velocity change is the result of tangential and radial velocity components in the case of centrifugal fans and of axial and tangential velocity components in the case of axial flow fans. Axial flow fans produce pressure from the change in velocity passing through the impeller, with no pressure being produced by centrifugal force. The basic fan types can be further subdivided and characterized as follows: • • • •

Centrifugal fans: airfoil, backward inclined/backward curved, forward curved, and radial Axial fans: propeller, tubeaxial, and vaneaxial Special designs: tubular centrifugal, centrifugal power roof ventilator, and axial power roof ventilator Plug fans

Figure 6-1 depicts most of these fan types and provides details of the impeller design, housing design, performance characteristics, and typical applications. The plug, or plenum, fan is not shown. A single inlet impeller, similar to one for a centrifugal fan is mounted on the end of the drive shaft. The impeller is mounted between two walls. It draws the air into the center of the impeller and blows it out evenly in all directions. This fan design can be particularly useful in compact air-handling units and in industrial situations, such as ovens,

Fundamentals of Air System Design SI

Figure 6-1

Types of fans.

121

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Chapter 6 Air Movers and Fan Technology

Figure 6-1

Types of fans (continued).

Fundamentals of Air System Design SI

123

when the only components subjected to the high temperature are the impeller and drive shaft.

Fan Laws The fan laws (see Table 6-1) relate the performance variables for any dynamically similar series of fans. Fan Law 1 shows the effect on volume flow, pressure, and power of changing size, speed, or density. Fan Law 2 shows the effect on volume flow rate, speed, and power of changing size, pressure, or density. Fan Law 3 shows the effect on speed, pressure, and power of changing size, volume flow, or density. The fan laws simplify analyzing a given fan because there is no change in fan size (D) or density (ρ). Another way to remember these relationships is that the air quantity is directly proportional to fan speed. Static pressure varies as the square of the speed change. Power input to the fan varies as the cube of the speed. Table 6-1

Fan Laws

Fan Law 1 1a

Q1 = Q2

(D1/D2)3 (N1/N2)

1b

p1 = p2

(D1/D2)2 (N1/N2)2 ρ1/ρ2

1c

W1 = W2

(D1/D2)5 (N1/N2)3 ρ1/ρ2

2a

Q1 = Q2

(D1/D2)2 (p1/p2)1/2 (ρ1/ρ2)1/2

2b

N1 = N2

(D2/D1) (p1/p2)1/2 (ρ1/ρ2)1/2

2c

W1 = W2

(D1/D2)2 (p1/p2)3/2 (ρ1/ρ2)1/2

3a

N1 = N2

(D2/D1)3 (Q1/Q2)

3b

p1 = p2

(D2/D1)4 (Q1/Q2)2 ρ1/ρ2

3c

W1 = W2

(D2/D1)4 (Q1/Q2)3 ρ1/ρ2

Fan Law 2

Fan Law 3

Notes: 1. Subscript 1 denotes the variable for the fan under consideration. Subscript 2 is the variable for the tested fan. 2. For all fan laws 3. p equals either ptf or psf.

124

Chapter 6 Air Movers and Fan Technology Figure 6-2 illustrates the application of the fan laws for a change in fan speed N for a specific size fan. The computed ptf curve is derived from the base ptf curve. For example, Point E (N1 = 650) is computed from Point D (N2 = 600) as follows: At Point D, Q2 = 6000 L/s and ptf2 = 280 Pa Using Fan Law 1a at Point E, Q1 = 6000 × 650/600 = 6500 L/s Using Fan Law 1b, ptf1 = 280 (650/600)2 = 329 Pa The completed total pressure curve, the ptf1 at N = 650 curve, may be generated by computing additional points from data on the base curve, such as Point G from Point F. If equivalent points of rating are joined (as shown by the dotted lines in Figure 6-2), these points form parabolas that are defined by the relationship expressed in the following equation:

Figure 6-2

Sample application of the fan laws.

Fundamentals of Air System Design SI

125

2

⎛Q ⎞ ( Δp 2 ) -------------- = ⎜ ------2⎟ ( Δp 1 ) ⎝ Q 1⎠

(6-1)

Each point on the base curve ptf determines only one point on the computed curve. For example, Point H cannot be calculated from either Point D or Point F. However, Point H is related to some point between these two points on the base curve, and only that point can be used to locate Point H. Furthermore, Point D cannot be used to calculate Point F on the base curve. The entire base curve must be defined by testing. Finally, the power required by a fan is related to both the volume flow rate and the pressure. The relationship can be expressed in several ways: kW ~ Q3 or kW ~ N3 This is an important observation because when dealing with an existing system where all of the components are fixed in place, if the amount of air moving can be reduced by changing the speed, the power requirement is reduced by the cube of the reduction in volume flow. For example, if the airflow rate is reduced by 20% to 80% of the previous rate, Q2/Q1 = 0.80. Therefore, kW 2 ⎛Q ⎞ -------------- = ⎜ ------2⎟ kW 1( ) ⎝ Q 1⎠

3 3

= 0.8 = 0.512

(6-2)

The fan power is reduced to 51% of the original amount. Another way to express this is that a 20% reduction in airflow results in a 49% reduction in power.

Fan and System Pressure Relationships As previously stated, a fan impeller imparts static and kinetic energy to the air. This energy is represented in the increase in total pressure and can be converted to static or velocity pressure. These two quantities are interdependent; fan performance cannot be evaluated by considering one or the other alone. The conversion of energy, indicated by changes in velocity pressure to static pressure and vice versa, depends on the efficiency of conversion. Energy conversion occurs in the discharge duct connected to a fan being tested in accordance with the joint standards ANSI/AMCA Standard 210, Laboratory Methods of Testing Fans for Certified Aerodynamic Performance Rating (AMCA 2007) and ANSI/ ASHRAE Standard 51-2007, Laboratory Methods of Testing Fans for Certified Aerodynamic Performance Rating (ASHRAE 2007), and the efficiency is reflected in the rating.

126

Chapter 6 Air Movers and Fan Technology Fan total pressure (ptf) is a true indication of the energy imparted to the airstream by the fan. System pressure loss (Δp) is the sum of all individual total pressure losses plus system effects imposed by the arrangement of duct elements on both the inlet and outlet sides of the fan. An energy loss in a duct system can be defined only as a total pressure loss. The measured static pressure loss in a duct element equals the total pressure loss only in the special case where air velocities are the same at both the entrance and exit of the duct element. By using total pressure for both fan selection and air distribution system design, the design engineer is assured of proper design. These fundamental principles apply to both high- and low-velocity systems. For additional information refer to the 2009 ASHRAE Handbook—Fundamentals (ASHRAE 2009). A very important relationship is p v = 0.602V

2

(6-3)

where pv is in Pa, and V is in m/s. To specify the pressure performance of a fan, the relationship of total pressure ptf , static pressure psf , and velocity pressure pvf must be understood, especially when negative pressures are involved. Most importantly, psf is a term defined in AMCA Standard 210 and ASHRAE Standard 51 as psf = ptf – pvf . Except in special cases, psf is not necessarily the measured difference between static pressure on the inlet side and static pressure on the outlet side of the fan. Figures 6-3 through 6-6 depict the relationships among these various pressures. Note that, as defined, ptf = pt2 – pt1. Figure 6-3 depicts a fan with an outlet

Figure 6-3

Pressure relationships of fan with outlet system only.

Fundamentals of Air System Design SI

127

system but no connected inlet system. In this case, the fan static pressure psf equals the static pressure rise across the fan. Figure 6-4 shows a fan with an inlet system but no outlet system. Figure 6-5 shows a fan with both an inlet system and an outlet system. In both cases, the measured difference in static pressure across the fan (ps2 – ps1) is not equal to the fan static pressure. All of the systems depicted in Figures 6-3 to 6-5 have inlet or outlet ducts that match the fan connections in size. Usually the duct size desired is not identical to the fan outlet or the fan inlet, so a further complication is introduced. To illustrate the pressure relationships in this case, Figure 6-6 shows a diverging outlet cone, which is a commonly used type of fan connection. In this case, the pressure relationships at the fan outlet do not match the pressure relationships in the flow section. Furthermore, the static pressure in the cone increases in the

Figure 6-4

Pressure relationships of fan with inlet system only.

Figure 6-5

Pressure relationships of fan with equal-sized inlet and outlet systems.

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Chapter 6 Air Movers and Fan Technology

Figure 6-6

Pressure relationships of fan with diverging cone outlet.

direction of flow because the velocity pressure is decreased. The static pressure changes throughout the system, depending on velocity. The total pressure (which, as noted in Figure 6-6, decreases in the direction of flow) more truly represents the loss introduced by the cone or by flow in the duct. Only the fan changes this trend (that is, the decrease of total pressure in the direction of flow). Therefore, total pressure is a better indication of fan and duct system performance. In the rather normal fan situation depicted in Figure 6-6, the static pressure across the fan (ps2 – ps1) does not equal the fan static pressure (psf). This phenomenon is known as system effect, which is discussed later in this chapter.

Fan Testing and Rating Fan efficiency ratings are based on ideal conditions. Some fans are rated at more than 90% total efficiency. However, necessary inlet and outlet arrangements often make it impossible to achieve ideal efficiencies in the field. Fans are tested in accordance with the strict requirements of ASHRAE Standard 51 and AMCA Standard 210. These joint standards specify the procedures and test setups to be used in testing the various types of fans and other air-moving devices. Figure 6-7 depicts one of the most common procedures for developing the characteristics of a fan. The fan is tested from shutoff conditions to nearly free delivery conditions. At shutoff, the duct is completely blanked off; at free delivery, the outlet resistance is reduced to zero. Between these two conditions, various flow restrictions are placed on the end of the duct to simulate various conditions on the fan. Sufficient points are obtained to define the curve between shutoff point and free delivery conditions. A nozzle chamber is often

Fundamentals of Air System Design SI

Figure 6-7

129

Pressure relationships of fan with diverging cone outlet.

used to determine the airflow rate. The point of rating may be any point on the fan performance curve. For each case, the specific point on the curve must be defined by referring to the flow rate and the corresponding total pressure. Other test setups, also described in AMCA Standard 210 and ASHRAE Standard 51, should produce the same performance curve. Fans designed for use with duct systems are tested with a length of straight duct between the fan discharge and the measuring station on a flow-through test setup. This length of duct smooths the flow from the fan and provides stable, uniform flow conditions at the plane of measurement. This allows the eddying around a centrifugal fan outlet with a cutoff, and allows the radially unequal air velocities from vaneaxial and propeller fans to become uniform along this length of straight duct. The equalization of velocity is usually reflected in an increase in static pressure. In the case of free discharge or duct fittings near the fan outlet or inlet, some of the pressure conversion is not realized. The measured pressures are corrected back to the fan outlet. Fans designed for use without ducts (including almost all propeller fans and power roof ventilators) are tested without ductwork. Not all sizes are tested for rating. Test information may be used to calculate the performance of larger fans that are geometrically similar, but such information should not be extrapolated to smaller fans. For the performance of one fan to be determined from the known performance of another, the two fans must be

130

Chapter 6 Air Movers and Fan Technology dynamically similar. Strict dynamic similarity requires that the important nondimensional parameters vary in only insignificant ways. These nondimensional parameters include those that affect aerodynamic characteristics, such as Mach number, Reynolds number, surface roughness, and gap size. (For more specific information, consult the manufacturer.)

Fan Drives A proper motor and drive selection aids in long life and minimum service requirements. Direct-drive fans are normally used in applications where exact air quantities are not required (such as with small fan-coil units), because ample heat transfer surface is available at more than enough temperature difference to compensate for any lack of air quantity that may exist. For example, this could apply to a unit heater application. Direct-drive fans are also used on applications where system resistance can be accurately determined. However, most air-conditioning applications use belt drives. V-belts must be applied in matched sets and used on balanced sheaves to minimize vibration problems and to ensure long life. They are particularly useful on applications where adjustments may be required to obtain more exact air quantities. These adjustments can be accomplished by varying the pitch diameter on adjustable sheaves or by changing one or both sheaves on a fixed-sheave drive system. Belt guards are required for safety on all V-belt drives, and coupling guards are required for direct-drive coupling equipment. The fan motor must be selected for the maximum anticipated power requirements of the fan plus drive losses. The motor must be large enough to operate within its rated power capacity, including drive losses and reductions in line voltages and short-term conditions. Normal torque motors are generally used for fan duty.

Fan Selection Figure 6-8 shows two fan characteristic curves for the same fan. They are constant-speed curves. Curve 1 is run at one speed and curve 2 at a lower speed. In terms of fan selection, the objective is always to keep the operating point somewhere in the optimum selection zone as illustrated in Figure 6-8. If the fan is to operate in zone A, a larger fan is more efficient. Conversely, if the fan is to operate in zone B, a smaller fan is more efficient. Keep in mind that a fan is a constant-volume device. There is no magic number to defining the optimum zone, although it should include maximum efficiency. The application also dictates the appropriate width of the optimum zone. Some HVAC applications allow a fairly wide optimum zone. In areas where big fans requiring a lot of energy are needed (such as mills or power plants), the optimum zone is much narrower because it is more important to operate near peak efficiency.

Fundamentals of Air System Design SI

Figure 6-8

131

Optimum fan selection zone.

Figure 6-9 shows a series of maximum efficiency curves for various fan sizes. It is plotted on log-log paper to show the exponential curve as a straight line. The fan sizes shown are a standard range, where 365 is a fan with a 927 mm diameter impeller or wheel. The value of a chart like that shown in Figure 6-9 is that, once it is prepared for a given type of fan, you can enter the x-axis with the flow in L/s × 2.12 and the y-axis with total pressure, defining a point in the graph. The fan represented by the curve closest to that point is the most efficient fan on the chart for that volume/pressure combination. In theory, the chart indicates the best fan. However, both the next smaller and the next larger fans should be evaluated for the particular application, even though they are both less efficient and possibly noisier. In practice, the AMCA sizes are so close together that it is quite likely that the next larger or smaller size will probably be acceptable. For example, suppose the chart suggests a 927 mm fan. It is quite likely that you can go down to 838 mm fan. While it is less efficient, it is down only a few points, it will not be

Figure 6-9

Maximum efficiency lines for various fan sizes.

132 Chapter 6 Air Movers and Fan Technology

Fundamentals of Air System Design SI

133

that much noisier, and the first costs will be lower. For variable volume applications, the next smaller size fan should always be evaluated. Note that the curves in Figure 6-9 are for one type of fan. If you have another type, a series of curves must be obtained from the manufacturer for that type. Another important point is that you cannot satisfy all applications simply by speeding up the fan. Recall from the fan laws earlier in this chapter that the input power goes up as the cube of the speed ratio [kW = (L/s)3]. Suppose you have 100% of design air in a system and it is determined that an additional 10% is required. If the fan is sped up by 10%, the pressure goes up by the square of the speed increase, to about 1.12 = 1.21, or 121%. However, the power requirements go up from 100% to 133% (1.13 = 1.33), and few systems can easily tolerate that large an increase. Happily, this works in reverse. That is, if the flow can be reduced, the whole process is reversed. Instead of being worried about overloading the system and requiring new equipment, the power bill is reduced appreciably because of the cube ratio.

Density, Temperature, and Altitude Unless otherwise identified, fan performance data are based on dry air at standard conditions: 101.325 kPa and 20°C, with a mass density of 0.204 kg/m3. In actual applications, the fan may be required to handle air or gas at some other density. The change in density may be a result of temperature, composition of the gas, or altitude. As indicated by the fan laws, fan performance is affected by gas density. With constant size and speed, the power and pressure vary in accordance with the ratio of gas density to the standard air density. Most of the time, air-handling systems are operated at or near sea level, so altitude is not a consideration. However, at higher altitudes, atmospheric density becomes a factor. At higher altitudes, or when handling gases lighter than standard air, the pressure is lowered. When working with a gas of lower density than air at sea level, the air cannot build up the pressure that the original standard air could. However, flow rate does not change. If a fan produces 5000 L/s at sea level, it will produce 5000 L/s at 3000 m above sea level but not at the same pressure. The flow rate remains the same no matter what the density. The change is strictly in pressure. Happily, a reduction in power also occurs with the reduction in density. The point to remember is that catalog information is developed at standard density, it has to be converted to lower density, and lower-density air will not transfer as much heat at higher altitudes as it will at sea level. Consequently, the required airflow for a given energy delivered may need to be increased, which results in higher pressures and higher power requirements.

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Chapter 6 Air Movers and Fan Technology

Static Pressure Versus Total Pressure Fan data in catalogs for unitary equipment are usually specified in static pressure not total pressure. This can cause errors in selection. The objective of this section is to show the problem and alert you to the prospective difficulties you may encounter. For example, assume the duct system for two systems has a static pressure loss of 250 Pa, and assume a fan delivers 4000 L/s across an outlet area of 0.2 m2, giving a velocity of (4000/1000)/0.2 = 20 m/s. From the equation for velocity pressure we have kW 2 ⎛Q ⎞ ----------- = ⎜ ------2⎟ kW 1 ⎝ Q 1⎠

3 3

= 0.8 = 0.512

(6-4)

With a static pressure of 250 Pa at the fan outlet, the total pressure becomes 250 + 241 = 491 Pa. Now consider another fan. Here is the same 4000 L/s, but this fan has 0.4 m2 of output area. Therefore, the velocity is 10 m/s. Again using Equation 6-1, the velocity pressure equals 0.602 × 102, or 60 Pa. To make things equal, we want 491 Pa total pressure just as before. Subtracting the 60 Pa velocity pressure from the total pressure leaves 431 Pa of static pressure. The trouble starts when we consider efficiency. The equation for efficiency, ε, is ( L/s ) ( kPa ) ε = -----------------------------( 1007 ) ( kW )

(6-5)

If we use static pressure in this equation, we get static efficiency. If we use total pressure, we get total efficiency. Some computations demonstrate the effect of output area on static efficiency. Assume from tests it has been determined that input kW = 2.48: ( L/s ) ( kPa ) ε = ------------------------------ × 100% ( 1007 ) ( kW )

(6-6)

Multiply total efficiency in both cases by 100 to convert to percent: ( 4000 ) ( 0.5 ) ε t = -------------------------------- × 100 = 80% ( 1007 ) ( 2.48 ) Static efficiency in Case 1 is ( 4000 ) ( 0.25 ) ε s = -------------------------------- × 100 = 40% ( 1007 ) ( 2.48 )

(6-7)

Fundamentals of Air System Design SI

135

Static efficiency in Case 2 is ( 4000 ) ( 0.431 ) ε s = ----------------------------------- × 100 = 69% ( 1007 ) ( 2.48 )

(6-8)

So here are two fans with the same input power (2.48 kW), the same total pressure (0.491 kPa), the same 4000 L/s flow, and the same total efficiency (80%). However, by doubling the outlet area, the static efficiency is increased from 40% to 69%—all because the outlet areas and, consequently, the outlet velocities of the two fans are different. By using total efficiency, you can avoid costly mistakes that can occur by looking at static efficiencies. The reason that fans are sometimes specified in terms of static pressure is that, particularly in older systems, when velocities are low, the difference between static pressure and total pressure is relatively small. However, particularly in systems with higher velocities (>7.5 m/s), it is important to deal with total pressure, not static pressure. A fan introducing unheated outside air discharges a larger volume of air after the air is heated. The fan motor should be selected for this added power requirement. The density of air varies as the absolute temperature difference. The absolute temperature in K is °C + 273. So a temperature of 20°C is 20 + 273 = 293 K. Therefore, a 4000 L/s rated fan whose discharge air is then heated to 75°C (75 + 273 = 348 K) introduces 4750 L/s (4000 × 348/293) into the duct system.

Fan Performance Under Installed Conditions It is not unusual for a fan and system combination to operate at a volume flow rate and pressure different from those for which the system was designed. There are two basic reasons why this may occur. First, if a system is not the same as specified in the design, the point of operation will not be at the design point on the fan curve. Referring to Figure 6-10, Point B is the specified point of operation, but the system actually operates at Point A. The different point of operation produces a different combination of capacity and pressure; in the case shown, it produces a higher pressure and a lower flow rate. If the original design volume flow rate must be retained, the situation can be corrected by changing the fan speed until the fan curve and the system curve pass through the required capacity point. Another way of correcting this situation is to reduce the pressure loss in the system so that the point of rating moves out on the curve to point B, as shown in Figure 6-10. This change in the system characteristics may be accomplished by a change in damper setting, a change in outlet grille setting, or an actual change in the duct design to achieve the lower pressure characteristic. The important note in this case is that the difference between the specified point of rating and the actual point of rating is due to a change in the system characteristic curve and not a difference in the fan. The fan curve is in its original

136

Chapter 6 Air Movers and Fan Technology

Figure 6-10

Operating points.

position; the challenge is simply to get the system characteristic curve to cross the fan curve at the desired point. Second, an entirely different change in the operation between the fan and the fan system can occur by an actual change in the fan performance curve. Remember, all fans impart energy to the air by some form of rotational motion. Fans are designed so they depend on uniform, straight airflow into the fan inlet. If this flow is upset in any way, the fan will not perform on the original performance curve but rather will work on a new one. Why this happens is a system effect.

System Effect Factors Figure 6-11 illustrates deficient fan/system performance resulting from one or more undesirable flow conditions (improper outlet connections, nonuniform inlet flow and/or swirl at the fan inlet). It is assumed that the system pressure losses have been accurately determined (Point 1, Curve A) and a suitable fan selected for operation at that point. However, no allowance has been made for the effect of the system connections on the fan’s performance. To compensate for this system effect, a system effect factor must be added to the calculated system pressure losses to determine the actual system curve. The system effect

Fundamentals of Air System Design SI

Figure 6-11

137

Deficient duct system performance.

is treated as a pressure loss even though it cannot be accurately measured as such in the field. The system effect factor for any given configuration is velocity dependent and, therefore, varies across the range of flow volumes of the fan (see Figure 6-12). In Figure 6-11, the point of intersection between the fan performance curve and the actual system curve is Point 4. Therefore, the actual flow volume will be deficient by the difference from 1– 4. To achieve design flow volume, a system effect factor equal to the pressure difference between Points 1 and 2 should have been added to calculate system pressure losses, and the fan selected to operate at Point 2. Note that because the system effect is velocity related, the difference represented between Points 1 and 2 is greater than the difference between Points 3 and 4. Figure 6-12 shows a series of 24 system effect curves (labeled A through X); determination of which curve to use is discussed later in the section, “Computing the Effect of Inlet Obstructions.” By entering the chart at the appropriate air velocity (on the abscissa), it is possible to read across from any curve (to the ordinate) to find the system effect factor for a particular configuration. The system

138

Chapter 6 Air Movers and Fan Technology

Figure 6-12

System effect curves.

Fundamentals of Air System Design SI

139

effect factor is given in Pa and must be added to the total system pressure losses, as shown in Figure 6-11. The velocity figure used in entering the chart is either the inlet or the outlet velocity of the fan. This depends on whether the configuration in question is related to the fan inlet or the outlet. Most catalog ratings include outlet velocity figures, but for centrifugal fans it may be necessary to calculate the inlet velocity. A more detailed discussion of system effects and tables detailing system effects for a wide range of equipment and configurations can be found in AMCA 201-90, Fans and Systems (AMCA 1990).

Fan Installation Design Computing the Effect of Fan Outlet Conditions Imagine an ideal uniform flow downstream from the fan. The reality is, in fact, quite different than the ideal. Figure 6-13 shows the flow patterns of a centrifugal fan and an axial fan. In either case, the flow is nonuniform at the fan discharge. Ideally, the outlet duct should be the same size as the fan outlet. To best use the energy developed by the fan, the length of duct known as the 100% effective duct length should be provided at the fan outlet. Acceptable flow can be obtained if the duct is not greater in area than 110% nor less in area than 85% of the fan outlet, and system effects can usually be tolerated at fan outlet velocities below 10 m/s. The slope of transition elements should not be greater than 15° for the converging elements nor greater than 7° for the diverging elements. There is a system effect for most fans at effective duct lengths of less than 100% of straight duct. Any closer and there is an effect such as that illustrated in Table 6-2, and the losses at other duct components (elbows, tees, etc.) will be higher than the standard fitting losses. One way to calculate effective duct length for round duct is as follows: •

If the duct velocity is greater than 13 m/s V A   o o L e = ----------------------4500

•

If the duct velocity is less than 13 m/s A L e = ---------o350

where = Vo Le = Ao =

duct velocity, m/s effective duct length, m duct area, mm2

(6-9)

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Chapter 6 Air Movers and Fan Technology

Figure 6-13

Blast areas for centrifugal and axial fans.

If the duct is rectangular, the equivalent duct diameter is calculated by 4HW D h = ------------------------2 H + W  where Dh =

equivalent duct diameter, mm

H

=

rectangular duct height, mm

W

=

rectangular duct width, mm

(6-10)

Fundamentals of Air System Design SI Table 6-2

141

Blast Area Ratios for Various Fan Types Fan Type

Blast Area Ratio

Centrifugal Airfoil

0.70

Backward curved

0.70

Backward inclined

0.70

Modified radial

0.60

Radial

0.80

Forward curved

0.50

Propeller

0.90

Axial Hub ratio: 0.3

0.90

0.4

0.85

0.5

0.75

0.6

0.65

0.7

0.50

Note: Use actual manufacturer’s data when available.

In cases where you use a shorter discharge length than one effective duct length, an additional pressure loss results. This additional pressure must be added to the fan total pressure requirements. The additional pressure loss may be calculated by Equation 6-11, which is also used to calculate additional pressure losses for other inlet and outlet configurations: Δp = 0.502K 1 ρ ( V )

2

where Δp =

pressure loss, Pa

V

=

velocity at outlet plane, m/s

K1

=

factor from Table 6-3

ρ

=

density, kg/m3 (about 1.2 kg/m3 for indoor conditions)

The blast area ratio is calculated by Blast area Blast area ratio = -------------------------Outlet area

(6-11)

142

Chapter 6 Air Movers and Fan Technology Table 6-3 Blast Area Ratio

No Duct

0.5

K1 Factors* Effective Duct Length 12%

25%

50%

2.0

1.0

0.40

0.18

0.6

1.0

0.66

0.33

0.18

0.7

0.8

0.40

0.14

0.14

0.8

0.47

0.22

0.01

0.9

0.22

0.14

*For calculating added pressure requirements due to short discharge length.

Typical blast areas for centrifugal and axial fans are identified in Figure 6-13. The blast area for centrifugal fans is the outlet area minus the area of the cutoff plate. The blast area for axial fans is the area of the annular space between the hub and the fan housing. The blast area should be obtained from the fan manufacturer for the particular fan being considered. For estimating purposes, the values of blast area ratio shown in Table 6-2 may be used if actual areas cannot be determined. Elbows can contribute to additional pressure loss. To obtain the rated performance from a fan, the first elbow fitting should be at least one effective duct length from the fan outlet (see Figure 6-14). If this length cannot be provided, an additional pressure loss results, and this additional pressure must be added to the fan total pressure requirements using the curve letter designation shown in Figure 6-12 and Table 6-4. The additional pressure loss may also be determined by using Equation 6-11.

Computing the Effect of Fan Inlet Conditions If an elbow must be installed on the fan inlet, a straight run of duct should be put between the elbow and the fan, and a long radius elbow should be used. Inlet elbows without the straight duct run create an additional loss that must be added to the fan total pressure requirements. The additional loss may also be calculated by using Equation 6-11.

Computing the Effect of Inlet Obstructions For obvious reasons, every effort should be made to keep the fan inlet free of obstructions. The fan inlet should be located so it is not obstructed (by other equipment, walls, pipes, beams, columns, etc.), because such obstructions will degrade the fan’s performance. Where obstructions are unavoidable, the resulting pressure losses can be estimated using Equation 6-11. The K factors for inlet area obstructions are shown in Table 6-5.

Fundamentals of Air System Design SI

Figure 6-14

143

Outlet duct elbows.

When you estimate the percentage of inlet area remaining obstructed, use that part of the projected area of the obstruction perpendicular to the airflow and subtract this area from the area of the inlet plane to obtain the net area. Divide the flow rate by this net area to determine the flow for V in the above equation.

Inlet Spin Figure 6-15a shows top and front views of two inlet duct combinations. Fans are normally tested with open inlets and uniform flow to the wheel. When

144

Chapter 6 Air Movers and Fan Technology Table 6-4

System Effect Curves for Outlet Elbows

No Outlet Duct

12% Effective Duct

25% Effective Duct

50% Effective Duct

A

N

O

P–Q

S

B

M

M–N

O

R

C

L–M

M

N

Q

D

L–M

M

N

Q

A

P

Q

R

T

B

N–O

O–P

P–Q

S

C

M–N

N–O

O–P

R–S

D

M–N

N–O

O–P

R–S

A

Q

Q–R

R–S

U

B

B

Q

R

T

C

N–O

O–P

P–Q

S

D

O

P

Q–R

S–T

A

S–T

T

U

W

B

R–S

S

T

V

C

Q–R

R

S

U–V

D

R

R–S

S–T

U–V

A

S

S–T

T–U

V–W

B

R

R–S

S–T

U–V

C

Q

Q-R

R-S

U

D

Q–R

R

S

U–V

Blast Area Outlet Elbow Outlet Area Position

100% Effective Duct

0.4

0.5

0.6

0.7

No System Effect Factor

0.8

These factors are for single-width single-inlet (SWSI) fans. For double-width double-inlet (DWDI) fans, apply the following multipliers: Elbow position B = P × 1.25 Elbow position D = P × 0.85 Positions A & C = P × 1.00

Fundamentals of Air System Design SI Table 6-4

System Effect Curves for Outlet Elbows (Continued) No Outlet Duct

12% Effective Duct

25% Effective Duct

50% Effective Duct

A

S–T

T

U

W

B

R–S

S

T

V

C

R

R–S

S–T

U–V

D

R–S

S

T

V

A

R–S

S

T

V

B

S–T

T

U

W

C

R–S

S

T

V

D

R–S

S

T

V

Blast Area Outlet Elbow Outlet Area Position

145

100% Effective Duct

0.9

1.0

These factors are for single-width single-inlet (SWSI) fans. For double-width double-inlet (DWDI) fans, apply the following multipliers: Elbow position B = P × 1.25 Elbow position D = P × 0.85 Positions A & C = P × 1.00

Table 6-5 % Inlet Area Obstructed K Factor

K Factor for Inlet Area Obstructions

5

10

15

25

50

75

0.22

0.40

0.53

0.80

1.20

1.60

angled ductwork is too close to the fan inlet (as shown in the figure), a spin component is imparted to the air. The flow is no longer uniform and nonstandard fan performance results. This means that the fan is no longer operating along the expected curve, and the fan performance is different than specified. It does not matter if it is spun in the direction of wheel rotation or against the direction of wheel rotation. If there is uncontrolled spin in the direction of the wheel, pressure is lost and the flow rate is reduced. A good indication of this is that the motor load goes down. If you are getting lower motor load than the manufacturer’s data indicates, this may be the cause. If there is uncontrolled spin against the direction of the wheel, the results are slightly higher pressure, lower flow, and higher than expected power draw. If there is enough spin, enough power is drawn to blow the circuit breakers or heaters on your system. If you know your fan is overloaded, but you cannot figure out why because the system appears in good order otherwise, look for uncontrolled spin. The best remedy is to enlarge the duct approaching the fan

146

Chapter 6 Air Movers and Fan Technology

(a)

(b)

Figure 6-15

(a) Inlet duct connections causing inlet spin and (b) corrections for inlet spin.

Fundamentals of Air System Design SI

147

to reduce the velocity and, therefore, the loss. Another remedy for this condition may be turning vanes, as shown in Figure 6-15b.

Computing the Effect of Enclosure Restrictions In cases where a fan (or several fans) is built into a fan cabinet construction or installed in a plenum, the walls should be at least one inlet diameter from the fan housing, and a space of at least two inlet diameters should be provided between fan inlets. If these recommendations cannot be met, additional pressure losses will result. These additional losses must be added to the fan total pressure requirements, as shown in Table 6-6. The additional pressure losses may also be calculated using Equation 6-11.

Computing the Effect of Inlet and Outlet Restrictions Normally, fan performance data do not include the effects of any accessories supplied with the fan. The loss caused by fan accessories (such as bearings, bearing pedestals, inlet vanes, inlet dampers, belt guards and motors) should be determined from tests by the fan manufacturer. The losses should be subtracted from the original fan performance and the resulting fan curve presented as the installed performance curve. If such data are not available, the losses due to accessories may be estimated as explained above in the section, “Computing the Effect of Inlet Obstructions.”

Parallel Fan Operation The combined performance curve for two fans operating in parallel may be plotted by using the appropriate pressure for the ordinates and the sum of the volumes for the abscissas. When two fans having a pressure reduction to the left of the peak pressure point are operated in parallel, a fluctuating load condition may result if one fan operates to the left of the peak static point on its performance curve. This problem may be reduced using two fans on a single shaft. The pressure curves (pt) of a single fan and two identical fans operating in parallel are shown in Figure 6-16. Curve A-A shows the pressure characteristics of a single fan. Curve C-C is the combined performance of the two fans. The unique figure-eight shape is a plot of all possible combinations of volume flow at each pressure value for the individual fans. All points to the right of CD are the result of each fan operating at the right of its peak point of rating. Stable performance results for all systems with less obstruction to airflow than is shown on the Δp curve D-D. At points of operation to the left of CD, system requirements may be satisfied with one fan operating at one rating point, while the other fan is at a different rating point. For example, consider Δp curve E-E, which requires a pressure of 200 Pa and a volume of 2.5 m3/s. The requirements of this system can be satisfied with each fan delivering 1.25 m3/s at 200 Pa at Point CE. The system

148

Chapter 6 Air Movers and Fan Technology

Figure 6-16

Two forward curve centrifugal fans in parallel operation.

Table 6-6 Length (L) 0.75 × inlet diameter 0.50 × inlet diameter 0.40 × inlet diameter 0.30 × inlet diameter

K Factor for Enclosure Restrictions System Effect Curves* V–W U T S

where D1 = diameter of the fan inlet *See Table 6-3

can also be satisfied at Point CE, with one fan operating at 0.66 m3/s at 180 Pa, while the second fan delivers 1.63 m3/s at the same 180 Pa. Note that system curve E-E passes through the combined performance curve at two points. Under such conditions, unstable operation can result. Under conditions of CE, one fan is underloaded and operating at poor efficiency. The other fan delivers most of the system requirements and uses substantially more power than the underloaded fan. This imbalance may reverse and shift the load from one fan to the other.

Fan Controls In many heating and ventilating systems, the volume of air handled by the fan varies. The choice of the proper method for varying flow for any particular case is influenced by two basic considerations: the frequency with which

Fundamentals of Air System Design SI

Figure 6-17

149

System total pressure loss curves.

changes must be made and the balancing of reduced power consumption against increases in first cost. To control flow, the characteristic of either the system or the fan must be changed. The system characteristic curve may be altered by installing dampers or orifice plates. This technique reduces flow by increasing the system pressure required and, therefore, increasing power consumption. Figure 6-17 shows three different system curves (A, B, and C), such as would be obtained by changing the damper setting or orifice diameter. Dampers are usually the lowest first-cost method of achieving flow control; they can be used even in cases where essentially continuous control is needed. However, there is a system effect loss created even at the full-open position. Changing the fan characteristic (pt curve) for control can reduce power consumption. From the standpoint of power consumption, the most desirable

150

Chapter 6 Air Movers and Fan Technology control method is to vary the fan speed to produce the desired performance. If the change is infrequent, belt-driven units may be adjusted by changing the pulley on the fan’s drive motor. Variable-speed motors or variable-speed drives (whether electrical or hydraulic) may be used when frequent or essentially continuous variations are desired. When speed control is used, the revised pt curve can be calculated by the fan laws. Inlet vane control is frequently used. Figure 6-18 illustrates the change in fan performance with inlet vane control. Curves A, B, C, D, and E are the pressure and power curves for various vane settings between wide open (A) and nearly closed (E). Tubeaxial and vaneaxial fans offer adjustable pitch blades to permit balancing of the fan against the system or to make infrequent adjustments. Vaneaxial fans are also produced with controllable pitch blades (pitch that can be varied while the fan is in operation) for frequent or continuous adjustment. Varying pitch angle retains high efficiencies over a wide range of conditions. Figure 6-19 shows the performance of a typical fan with variable pitch blades. From the standpoint of noise, variable speed is somewhat better than variable blade pitch. However, both control methods give high operating efficiency control and generate much less noise than inlet vane or damper control.

Figure 6-18

Effect of inlet vane control on backward curve centrifugal fan performance.

Fundamentals of Air System Design SI

151

Effect of Variable-Resistance Devices Variable-resistance devices (such as dampers and louvers) can have significant effects on a system. As discussed earlier, the system curve is a composite of several components in series with each other. If one component varies, the system curve also changes. Some system components are truly fixed, such as the ductwork. Others are variable, either by design or operationally over time. Components that vary by design are referred to as varying with a purpose, and the others are referred to as varying without a purpose. Examples of components that vary without a purpose are filters and coils. As shown in Figure 6-20, dirty filters push the system curve to the left, while dry coils push it to the right. If a coil is not dehumidifying and becomes dry, the pressure drop is less, the system curve slides to the right, and more air is delivered. As the coil begins to dehumidify, or remove moisture, the pressure drop is greater and the system curve slides back up to its original range. Figure 6-21 shows two system curves for a variable-volume system. In this case, the volume is varied with dampers at the terminal devices. The original operating point was Point X on system Curve A. The thermostat in this system is activated and causes the damper to close down. The operating point now shifts to Point Y on system Curve B, which gives 75% of the previous volume flow and a higher operating pressure. Therefore, the damper has to function to reduce the pressure to Point Z on the original system curve.

Figure 6-19

Effect of blade pitch on controllable pitch vaneaxial fan performance.

152

Chapter 6 Air Movers and Fan Technology

Figure 6-20

Fan curve and system curve.

Figure 6-21

Variable-volume system at three-quarters flow.

Fundamentals of Air System Design SI

Figure 6-22

153

Variable-volume system at one-half flow.

If the flow rate is halved, as shown in Figure 6-22, these dampers continue to close down. The damper pressure differential is now quite large and can contribute to both noise and operating flow instabilities. Consequently, it is usually necessary to provide some type of capacity control at the fan. This reduces the effective pressure available at the fan, and keeps the available system pressure at or near the original system curve. On systems with a minor variation between maximum and minimum flow, designs may be based on riding the fan curve. Note that duct leakage is based on the pressure of the system operating at Point Y.

The Next Step Chapter 6 discussed how fans work and produce airstreams with static and dynamic pressure. Chapter 7 discusses ducts that distribute air around a facility.

Summary A fan is an air pump with rotating blades that creates an increase in static and velocity pressure. The two main types are centrifugal fans, where the air enters the eye of a barrel and is thrown out radially into the spiral scroll, and axial fans.

154

Chapter 6 Air Movers and Fan Technology For a specific fan connected to a system, the volume is proportional to the rpm, the static pressure is proportional to rpm squared, and power input is proportional to rpm cubed. A fan creates a velocity pressure and a rise in static pressure in a system. Because the system can and usually does influence fan performance, both velocity pressure and static pressure must be addressed. The easy estimation of system static pressure loss and choosing a fan with that static pressure rise may produce acceptable results in a low-velocity system but probably will not in a higher-velocity system. The reason for this is that the inlet and outlet conditions can significantly influence fan performance—a phenomenon known as system effect. Fan efficiency ratings are based on ideal conditions, a new fan, unobstructed inlet, and same-size duct outlet. However, in real installations, the designed inlet and outlet conditions are often not ideal. The reduction in fan performance due to inlet and outlet conditions can greatly reduce effective fan performance. To minimize risk of error, designs should be based on total fan pressure and not just on fan static pressure. Direct drives are used in smaller systems where oversizing the fan is easier than matching the fan to the load. In larger systems, belt drives are commonly used to adjust from the motor speed to the required fan speed. Motors and drives must be sized for the maximum anticipated load. Selecting a fan involves finding one that provides the required flow and total pressure at a good efficiency and noise level. The type of fan may be influenced by system operation, so a very flat characteristic might be sought in a system where variations in flow are required without any fan adjustment. The motor power rises as the cube of the flow (fan law), so the fan must not be significantly undersized. Equally, if a fan is significantly oversized, changing pulleys to reduce capacity to what is actually required will save substantial motor power and operating costs. Fan performance data are normally given for standard air. At constant speed, the power and pressure vary with gas density, proportional to absolute temperature, K, or °C + 273. At altitudes significantly above sea level, the air density drops. A fan provides the same volume flow, but the same volume transfers less thermal energy due to the lower air density. The design volume is typically higher at higher altitudes, so a design for sea level operation should be reevaluated before being built at higher altitude. Fan data are usually specified in terms of static pressure not total pressure. However, due to velocity changes at inlet and outlet connections, the use of static pressure alone can cause errors. Particularly with system velocities over 7.5 m/s, it is important to work with total pressure not just static pressure. Remember, velocity pressure equals (0.602V 2) Pa. When installed, a fan may not provide the expected flow. This may be a result of system pressure losses being different from design calculations or the fan being affected by the way the inlet and outlet are configured. A difference

Fundamentals of Air System Design SI

155

in system pressure loss results in the fan riding its curve until the system curve and fan curve meet. Correction may be possible by adjusting fan speed. Remember, if the reason is inlet or outlet configuration, this is a system effect. System effects are due to one or more of the following: outlet connection geometry, uneven flow across the inlet, and swirl at the inlet. The outlet effects are due to the uneven velocity profile coming out of the fan and the difference between the fan air outlet size (blast area) and the fan connection size. At the inlet, an uneven flow across the fan effectively overloads some blade positions and underloads other blade conditions, causing loss of efficiency. Swirling of the entering air effectively changes the velocity of the air as it meets the blade, again jeopardizing efficiency. The flow at a fan discharge is very uneven and takes a length of duct to balance out. The required length and static pressure loss can be calculated based on manufacturers’ data of blast area versus outlet area. If this duct length is not available, the pressure drop is increased. Care must be taken to transition from fan outlet to duct system with minimal losses. For centrifugal fans, a bend close to the fan outlet causes an additional static pressure loss that must also be factored into total static pressure losses. The way the air enters a fan can significantly influence fan performance and power consumption. If the entering airstream is biased to one side of the inlet or is swirling the fan performance is reduced and power may be increased. Careful analysis of these inlet effects can be very important in ensuring that the system performs as required and that energy is not wasted due to poor design.

References and Bibliography AMCA. 1990. AMCA 201-90, Fans and Systems. Arlington Heights, IL: Air Movement and Control Association Inc. AMCA. 2007. Standard 210, Laboratory Methods of Testing Fans for Certified Aerodynamic Performance Rating. Arlington Heights, IL: Air Movement and Control Association Inc. ASHRAE. 2007. ASHRAE Standard 51-2007, Laboratory Methods of Testing Fans for Certified Aerodynamic Performance Rating. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. ASHRAE. 2008. ASHRAE Handbook—Systems and Equipment. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. ASHRAE. 2009. ASHRAE Handbook—Fundamentals. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. SMACNA. 1990. HVAC Systems—Duct Design. Chantilly, VA: Sheet Metal and Air Conditioning Contractors’ National Association.

156

Chapter 6 Air Movers and Fan Technology

Skill Development Exercises for Chapter 6 Complete these questions by writing your answers on the worksheets at the back of this book. 6-1

6-2

A fan delivers 6000 L/s at a pressure of 300 Pa at a rotational speed of 750 rpm. If the fan speed is reduced to 600 rpm, how much air will the fan deliver and at what pressure? a) 4800 L/s, 240 Pa

b) 4800 L/s, 192 Pa

c) 3840 L/s, 240 Pa

d) 3840 L/s, 192 Pa

Given a fan operating at 4000 L/s, 750 Pa total pressure, and 3.75 kW, what is the fan total efficiency? a) 84%

6-3

b) 75%

d) none of the above

b) 2.18 m

c) 4.33

d) none of the above

b) 400 mm

c) 460 mm

d) none of the above

For any given system, the system effect factor is constant across the range of flow volumes of the fan. b) False

A fixed-fan system draws 2.2 kW to deliver 5000 L/s. If the airflow requirement can be reduced to 3500 L/s by decreasing the fan speed, the motor power requirement will be reduced to _______________. a) 2.1 kW

6-9

c) 1.57 m

A rectangular duct is 250 mm high and 500 mm wide. What is the equivalent duct diameter of this duct?

a) True 6-8

d) none of the above

What is one effective duct length for a duct with a duct velocity of 4.1 m/s and an area of 140,000 mm2?

a) 550 mm 6-7

c) 65%

b) 4.3 m

a) 1.07 m 6-6

d) none of the above

What is one effective duct length for a duct with a duct velocity of 20 m/s and an area of 125,000 mm2? a) 6.79 m

6-5

c) 74%

Given a fan operating at 4000 L/s and using 2.3 kW, what is the fan total efficiency? a) 85%

6-4

b) 79%

b) 1.0 kW

c) 0.76 kW

d) none of the above

The _______________ is the highest efficiency centrifugal fan design. a) radial b) forward curved c) backward inclined, backward curved d) none of the above

Fundamentals of Air System Design SI 6-10

Power roof ventilators _______________. a) usually operate without discharge ductwork b) operate at low pressure c) operate at high volume d) none of the above

157

Duct System Design

Study Objectives After completing this chapter, you should be able to lay out and size a simple duct system that transports the required quantity of air from the fan to the conditioned space using appropriate methods and materials; and T calculate the pressure losses in a duct system. T

Instructions Read the material in Chapter 7. Verify the examples presented in the chapter with your own calculations. At the end of the chapter, complete the skill development exercises. Review those sections of the chapter as needed to complete the exercises.

Duct System Design Overview Air duct system design must consider space availability, space air diffusion, noise, duct leakage, duct heat gains and losses, balancing, fire and smoke control, and initial investment and system operating cost. This chapter covers duct construction, methods of duct sizing, and duct sizing examples. The ancillary effects of air leakage and potential heat gain or loss along the duct run are also addressed.

Duct Materials A variety of materials are used in the construction of ducts. Duct materials should receive the same careful selection consideration as other system components. The material used in a duct system can substantially affect the overall system performance. The advantages and disadvantages of the available materials should be weighed. Materials used for ducts include galvanized steel, carbon (black) steel, stainless steel, aluminum, copper, fiberglass reinforced plastic (FRP), polyvinyl chloride (PVC), polyvinyl steel (PVS), concrete, fibrous glass (duct board), and gypsum board. These materials are compared in Table 7-1.

160

Chapter 7 Duct System Design Table 7-1

Material

Duct Materials (SMACNA 2004)

Applications

Limitations

High strength, rigidity, durability, rust resistance in ordinary conditions, availability, nonporous, workability

Weldability, paintability, weight, corrosion resistance

High strength, rigidity, durability, availability, paintability, weldability, non-porous

Corrosion resistance, weight

Aluminum

Duct systems for moisture-laden air, louvers, special exhaust systems, ornamental duct systems. Often substituted for galvanized steel in HVAC duct systems.

Weight, resistance to some forms of corrosion, availability

Low strength, material cost, weldability, thermal expansion

Stainless Steel

Duct systems for kitchen exhaust, moisture-laden air fume exhaust

High resistance to many common forms of corrosion (but care is definitely required in alloy selection)

Material cost, workability, availability

Copper

Duct systems for exposure to outside elements and moisture-laden air

Accepts solder readily, durable, resists corrosion, nonmagnetic

Cost, electrolytic action if in contact with galvanized steel, thermal expansion, stains

Fiberglass Chemical exhaust, scrubbers, Reinforced underground duct systems Plastic (FRP)

Corrosion resistant, ease of modification

Cost, weight, range of chemical and physical properties, brittleness, fabrication, code acceptance

Polyvinyl Chloride (PVC)

Exhaust systems for chemical fumes and hospitals, underground duct systems

Corrosion resistance, weight, weldability, ease of modification

Cost, fabrication, code acceptance

Polyvinyl Steel (PVS)

Underground duct systems, moisture-laden air, corrosive air systems

Corrosion resistance, weight, workability fabrication, rigidity

Susceptible to coating damage, temperature limitations (120°C max.), weldability, code acceptance

Concrete

Underground ducts, air shafts

Compressive strength, corrosion Cost, weight, porous, resistance (steel reinforcement in fabrication (requires concrete must be properly treated) forming processes)

Galvanized Steel

Widely used for most air handling applications. Not recommended for corrosive product handling or temperatures above 200°C.

Advantages

Breechings, flues, stacks, hoods, Carbon Steel other high temperature duct systems, (Black Iron) kitchen exhaust systems, ducts requiring paint or special coatings

Rigid Fibrous Interior HVAC low-pressure Glass duct systems

Weight, thermal insulation and vapor barrier, acoustical qualities, ease of modification, inexpensive tooling for fabrication

Cost, susceptible to damage, system pressure, code acceptance, questionable cleanability

Gypsum Board

Cost, availability

Weight, code acceptance, leakage, deterioration when damp

Ceiling plenums, corridor ducts, airshafts

Fundamentals of Air System Design SI

161

Duct sizing and construction specifications are generally given for galvanized steel, and correction factors for other materials must be used. Unless otherwise noted, this chapter considers galvanized steel exclusively. Consideration must also be given to selection of duct construction components other than those materials used for the duct walls, including flexible ducts, duct liners, pressure sensitive tapes, sealants, reinforcements, and hangers. Lined ducts must be sized to include the lining. The duct drawing must clearly state that the duct dimension is the metal size or the airway size.

Duct Construction Rectangular Metal Ducts Rectangular metal ducts must be designed to avoid vibration and structural failure. Vibration can occur when a flat sheet acts as a drum, with the center of the sheet vibrating back and forth. This behavior becomes more pronounced in both the scale of the vibration and the noise generated as the size of the flat metal area increases. This vibration is reduced by breaking or, less commonly, beading the duct sheets as shown in Figure 7-1. The structural requirements for ducts are determined by duct sheet thickness, duct dimension, and duct pressure. The joined sheets may need reinforcement at the joints as well as additional reinforcing across the sheet and tie rods. For commercial and institutional ductwork, HVAC Duct Construction Standards—Metal and Flexible (SMACNA 2005) provides full details for duct design, specification, and construction. The following tables are samples from the third edition. Table 7-2 indicates what reinforcing is required for a rectangular sheet steel duct under a 1000 Pa working pressure. Similar tables are available for higher and lower pressures. The minimum required sheet thickness for no joint reinforcement is provided in the second column: Min, mm. Where the value is not designed, read horizontally to the right to get the material thickness and required joint reinforcement. For example, a 2000 mm duct requires 1.61 mm sheet, joint tie rod (JTR), or joint reinforcement 2(I), plus intermediate reinforcement type K. For small ducts, the seams joining sections provide adequate stiffening of the duct. For larger sizes, modified transverse jointing pieces are used as shown in Table 7-3 to produce adequate stiffness at the joint. For intermediate stiffening, Table 7-4 shows material sections to provide the stiffness. Note the use of letter designations for stiffness grades. For example, the letters in the right-hand column in Table 7-2 are for matching in the left-hand column of Table 7-4. Transverse joints and, when necessary, intermediate structural members are designed to reinforce the duct system. Ducts larger than 2.5 m require internal tie rods to maintain their structural integrity. Tie rods allow the use of smaller reinforcements than would otherwise be required. Fittings must be reinforced similarly to sections of straight duct. On size change fittings, the greater fitting dimension determines material thickness.

Figure 7-1

Cross broken and beaded ducts (SMACNA 2005).

162 Chapter 7 Duct System Design

Fundamentals of Air System Design SI Table 7-2

1000 Pa Duct Reinforcement (SMACNA 2005)

1.80 m Joints

100 Pa Static Pos. or Neg.

163

1.80 m Joints w/0.90 m Reinf. Spacing Joints/Reinf.

Int. Reinf.

Duct Dimension, mm

Min, mm

Joint Reinf.

Alt. Joint Reinf.

200 and under

0.55

N/R

N/A

230–250

0.70

N/R

N/A

0.55

N/R

N/A

MPT

B

251–300

0.70

N/R

N/A

0.55

N/R

N/A

MPT

C

301–350

0.85

N/R

N/A

0.55

N/R

N/A

MPT

C

351–400

0.85

N/R

N/A

0.55

N/R

N/A

MPT

C

401–450

0.85

N/R

N/A

0.55

N/R

N/A

MPT

C

451–500

0.85

N/R

N/A

0.55

N/R

N/A

MPT

D

501–550

1.00

N/R

N/A

0.55

N/R

N/A

MPT

D

551–600

1.00

N/R

N/A

0.70

N/R

N/A

MPT

E

601–650

1.00

N/R

N/A

0.70

N/R

N/A

MPT

E

651–700

1.00

N/R

N/A

0.70

N/R

N/A

MPT

E

701–750

1.31

N/R

N/A

0.70

N/R

N/A

MPT

E

751–900

1.31

N/R

N/A

0.85

N/R

N/A

MPT

F

1.61

JTR

(2) H

0.85

JTR

(2)C

MPT

G

1.00

N/R

N/A

MPT

G

0.85

JTR

(2)E

MPT

H

1.31

N/R

N/A

MPT

H

1.00

JTR

(2)H

MPT

I

1301–1500

1.00

JTR

(2)H

2 MPT

I

1501–1800

1.31

JTR

(2)H

2 MPT

J

1.61

JTR

(2) I

2 MPT

K

2101–2400

1.61

JTR

(2) K

N/A

L

2401–2700

1.61

JTR

(2) K

—

L

2701–3000

1.61

JTR

(2) K

—

L

Min, mm

Joint Reinf.

Alt. Joint Reinf.

Tie Rod

Alt. Reinf.

Use 1.80 m Joints

901–1000 1.61

JTR

(2) H

1001–1200 1201–1300

1.61

JTR

(2) H

1801–2100 Not designed

N/R = Not required N/A = Not applicable JTR = Joint tie rod

MPT = Mid panel tie rod(s) (2) (X) Indicates two external reinforcements of class (X) to be used in lieu of joint tie rods

Chapter 7 Duct System Design

164

Table 7-3

Transverse Joint Reinforcement (SMACNA 2005)

Reinf. Class T-2 Standing Drive Slip EI*

H×T (mm)

KG LM

Use B

T-10 Standing S H×T (mm)

A

0.12

B

0.29 28.6 × 0.55

0.6

25 × 0.55

C

0.55 28.6 × 0.85

0.9

D

0.78 28.6 × 1.31

1.2

E

1.9

F

3.7

Use G

G

4.5

41.3 × 1.31

H

7.6

KG LM

Use B

T-11 Standing S H×T (mm)

KG LM

T-12 Standing S H×T (mm)

KG LM

T-14 Standing S H × T + HR (mm)

12.7 × 0.55

0.74

Use B

0.9

12.7 × 0.85 25 × 0.55

0.9

25 × 0.55

1.0

Use D

25 × 0.85

1.2

25 × 0.85

1.2

25 × 0.70

1.2

Use D

28.6 × 1.00 25 × 0.85 (+)

1.3

25 × 1.00 25 × 0.85 (+)

1.3

Use E

1.5

41.3 × 0.70 38.1 × 3.2 Bar

28.6 ×

1.5

25 × 1.31 (+)

1.3

25 × 1.31 38.1 × 1.00

1.8

Use F

1.31

38.1 × 1.00

LM

Use D

Use G

1.9

KG

1.9

2.1

41.3 × 0.85 38.1 × 3.2 Bar

2.2

41.3 × 1.00 38.1 × 3.2 Bar

2.6

41.3 × 1.31 38.1 × 3.2 Bar

3.0

54 × 1.00 51 × 51 × 3.2 Angle

4.3

54 × 1.00 51 × 51 × 4.76 Angle

5.5

Not given I

Not given

20 Not given

J

23

K

30

L

60

Not given

Not given *Effective EI is number listed times 105 before adjustment for bending moment capacity. T-2 and T-10 through T-14 are restricted to 750 mm length at 1000 Pa, to 914 mm length at 750 Pa, and are not recommended for service above 1000 Pa. (+) indicates positive pressure use only.

Fundamentals of Air System Design SI Table 7-4

Reinf. Class E1*

Intermediate Reinforcement (SMACNA 2005)

Angle H × T (MIN) (mm)

165

Channel or Zee KG LM

H × B × T (MIN) (mm)

Hat Section KG LM

H × B × D × T × (MIN) (mm)

KG LM

A

0.12

Use C

Use B

Use F

B

0.29

Use C

19.1 × 12. 7 × 1.00

0.36

Use F

C

0.55

C 25 × 1.61 C 19.1 × 3.2

0.60 0.85

19.1 × 12.7 × 1.31 25 × 19.1 × 1.00

0.46

Use F

D

0.78

H 19.1 × 3.2 C 25 × 3.2

0.85 1.19

25 × 19.1 × 1.31

0.67

Use F

E

1.9

C 31.8 × 2.75 H 25 × 3.2

1.34

51 × 28.6 × 1.00

0.89

Use F

F

3.7

H 31.8 × 3.2

1.52

38.1 × 19.1 × 1.31

0.80

38.1 × 19.1 × 15.9 × 1.31 38.1 × 38.1 × 19.1 × 1.00

1.34 1.24

G

4.5

38.1 × 3.2

1.83

38.1 × 19.1 × 1.61

0.98

38.1 × 19.1 × 15.9 × 1.31

1.19

H

7.6

38.1 × 4.8 51 × 3.2

2.64 2.46

38.1 × 19.1 × 3.2

1.95

38.1 × 38.1 × 19.1 × 1.31 51 × 25 × 19.1 × 1.00

1.61 1.34

I

20

C 51 × 4.8 63.5 × 3.2

3.63 3.13

51 × 28.6 × 2.5 76 × 28.6 × 1.61

2.38 1.56

51 × 25 × 19.1 × 1.61

2.14

J

23

H 51 × 4.8 C 51 × 6.4 63.5 × 3.2 (+)

3.63 4.76 3.13

51 × 28.6 × 3.2

2.75

51 × 25 × 19.1 × 2.5 63.5 × 51 × 19.1 × 1.31

3.65 2.28

K

30

63.5 × 4.8

4.61

76 × 28.6 × 2.5

2.98

63.5 × 51 × 19.1 × 1.61 76 × 38.1 × 19.1 × 1.61

2.80 2.98

L

60

H 63.5 × 6.4

6.10

76 × 28.6 × 3.2

3.40

63.5 × 51 × 19.1 × 3.2 76 × 38.1 × 19.1 × 2.75

5.51 5.06

*Effective EI is number listed times 105 before adjustment for bending moment capacity. C and H denote cold-formed and hotrolled ratings; when neither is listed, either may be used. See tie rod option elsewhere. (+) indicates positive pressure use only. Hat Section dimension B may be equal to two times dimension H with the same reinforcement class rating.

166

Chapter 7 Duct System Design Where fitting curvature or internal member attachments provide equivalent rigidity, such features may be credited as reinforcement. Pressure classification in relation to the fan curve must be considered, especially with variable-air-volume (VAV) systems, where the dampers may throttle the airflow, raising the duct pressure. Fire dampers and manual balancing dampers may be inadvertently closed, with a resulting rise in system pressure. Supply ducts sometimes blow apart and return ducts sometimes collapse as a result of these effects. Table 7-5 shows the SMACNA duct pressure classification scheme.

Round Metal Ducts Round ducts are inherently strong and rigid, and are generally the most efficient and economical ducts for air systems. The dominant factor in round-duct construction is the ability of the material to withstand the physical abuse of installation and negative pressure requirements. Construction requirements are a function of static pressure, type of seam (spiral or longitudinal), and diameter.

Flat-Oval Ducts Hanger designs and installation details for rectangular ducts generally apply to flat-oval ducts.

Fibrous Glass Ducts Fibrous glass ducts are a composite of rigid fiberglass and a factory-applied facing (typically aluminum or reinforced aluminum), which serves as a finish and vapor barrier. This material is available in molded round sections or in board form for fabrication. Duct systems of round and rectangular fibrous glass are generally limited to 12.5 m/s and ±500 Pa. Molded round ducts are available in higher pressure ratings. Flexible ducts connect mixing boxes, light troffers, diffusers, and other terminals to the air distribution system. Because unnecessary length, offsetting, Table 7-5

SMACNA Duct Pressure Classifications (SMACNA 2005) Static Pressure

Pressure Class

Operating Pressure

125 Pa

Up to 125 Pa

250 Pa

Over 125 Pa, up to 250 Pa

500 Pa

Over 250 Pa, up to 500 Pa

750 Pa

Over 500 Pa, up to 750 Pa

1 kPa

Over 750 Pa, up to 1 kPa

1.5 kPa

Over 1 kPa, up to 1.5 kPa

2.5 kPa

Over 1.5 kPa, up to 2.5 kPa

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and compression of these ducts significantly increases airflow resistance, they should be kept as short as possible and fully extended. For further information on fibrous glass ducts, consult Fibrous Glass Duct Construction Standards (SMACNA 2003) or manufacturers’ construction standards.

Duct Sealing Ducts must be sufficiently airtight to ensure economical and quiet performance of the system. Both leakage and the noise of leaks increase with increasing duct pressure. A variety of materials and techniques have been developed for duct sealing, including liquids, mastics, gaskets, pressure sensitive tapes, heat-applied materials, and embedded fabric. Surfaces to receive sealant should be free from oil, dust, dirt, rust, moisture, ice crystals, and any other substances that would inhibit or prevent bonding. No sealant system is recognized as a substitute for mechanical joining. Also, the designer should carefully evaluate proposed duct sealants. Some use solvents that are toxic to workers applying the sealant. Some deteriorate or crystallize as they dry, and do not provide adequate sealing only a few months after being installed.

Duct Design and Sizing Heating, ventilating, and air-conditioning system duct design follows the determination of room loads and desired air quantities. Consider the type of duct system needed, based on an economic analysis of the building design and use, unless the owner or architect specifies a preference for a particular type. In any event, the specific type of system affects the type of air-handling apparatus selected.

Air Distribution Before duct design and sizing can begin, one must locate supply air outlets and select the size and type required for proper air distribution in each conditioned space (see Chapter 3, “Human Comfort and Air Distribution”). Air distribution in the conditioned space is highly important in influencing the comfort of the occupants. Good air distribution is ensured by proper consideration of the basic factors in the selection of the outlet terminal devices. Drafts caused by too much air or physical flow disturbances within the room should be avoided. The outlet terminal devices should provide the proper air velocities within the room occupied zone (floor to 1.8 m above the floor) and the proper temperature equalization. Entrainment of the room air by the primary (or supply) airstream at the outlet terminal to attain the required temperature

168

Chapter 7 Duct System Design equalization and to counteract the effects of natural room air convection is very important. Select air distribution terminal devices from industry standard types or configurations so that they can be obtained from many sources. Most terminal device manufacturers’ catalogs furnish data on airflow throw, drop, air pattern, terminal velocities, acoustics, ceiling heights, etc. Supply outlets on the same branch should be chosen with approximately the same pressure loss (no more than 13 Pa variation) through the outlet. Mixing ceiling supply diffusers with sidewall supply grilles on the same branch should be avoided unless there is no significant difference in pressure drops between the different types. For a comprehensive review of considerations in the selection of air distribution equipment, refer to the 2008 ASHRAE Handbook—Systems and Equipment (ASHRAE 2008b) and to air distribution equipment manufacturers’ application engineering data. Some of the basic procedures used in the selection of air distribution equipment are as follows: •

Consider the ambient conditions that could affect comfort.

•

Decide on the location of air supply outlets (such as in the floor, sill, sidewall, exposed duct, or ceiling), taking into account the type of system serving them. Locate return and exhaust air devices.

•

Consider the special requirements affecting outlets used with systems such as a VAV system.

•

Place balancing dampers to be used with outlet devices at a convenient location, preferably well upstream from the outlet as long as access is available. Refer to manufacturer’s data regarding throw, spread, drop, noise level, etc.

•

Zoning With the outlet devices selected and before duct layout and duct sizing can begin, the designer must determine how many zones of temperature control are required for both perimeter zones and interior zones. In general, spaces with exterior walls are grouped into perimeter zones determined by building exposure (north, east, south, or west exposure). These perimeter zones may be further subdivided into smaller control zones, depending on variations in internal load or a requirement for individual occupant control. Typical situations include private executive offices, where the owner may want individual control, or areas of high heat gain or loss, such as computer rooms, conference rooms, or corner rooms with two exposed walls. Similarly, the interior zones may also be divided into control zones to satisfy individual room requirements or variations created by internal loads, such as lights, people, or equipment.

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Preliminary Layout The next step is to draw a preliminary schematic diagram for the ductwork that conveys the design air quantity to the selected zones and outlets by the most efficient and economical path. It is suggested that this layout be made on a reproducible tracing of the architectural floor plans. By doing this, the designer has a better feel for the final relationship of air terminals, branch ducts, main ducts, risers, and apparatus. This procedure helps the designer coordinate the ductwork with the structural limitations of the building and other building systems and services. On this preliminary layout, the designer should indicate the design airflows throughout the system. If a constant-volume system is chosen, it will be the arithmetic sum of the airflow of each terminal (including branches) working back from the end of the longest run to the fan. However, if a variable-air-volume system is chosen, the designer must apply the proper diversity factors to allow a summarization of the peak design airflows to determine their impact on branch and main duct sizes coming from the supply fan. The same procedure must also be followed for return air and exhaust air systems. In addition to sizing the ductwork properly, this also permits the designer to evaluate the effect of the total HVAC system design, balancing the proper proportions of supply air to return air and exhaust air and outside makeup air. Pressure losses due to fittings and transitions must also be included in the calculation.

Duct Sizing After completing the preliminary HVAC system duct layout, the designer proceeds to use one of the methods for sizing the duct system discussed in the section “Design Methods.” Generally, these give the equivalent round-duct sizes and the pressure losses for the various elements of the duct system. The designer then incorporates this information into the preliminary duct layout. If round ductwork is to be used throughout, the duct sizing efforts are completed, providing the ductwork physically fits into the building. If rectangular or flat oval ductwork is chosen, the proper conversions must be made from the equivalent round-duct sizes to rectangular or flat-oval sizes. Applying the appropriate duct friction loss correction factors and using the duct fitting loss coefficients, the duct system total pressure loss can be calculated. With HVAC system duct sizes now selected and the total pressure or static pressure losses calculated, the designer must determine if the ductwork fits into the building. At this point, the designer must consider the additional space required beyond the bare sheet metal sizes for reinforcing and circumferential joints. In addition, consideration must be given to external insulation or duct liner that may be required, clearance for piping, conduit, light fixtures, etc., where applicable, and clearance for the removal of ceiling tiles. Further considerations in the sizing and routing of a ductwork system include space

170

Chapter 7 Duct System Design and access requirements for air terminals, mixing boxes, VAV boxes, fire and smoke dampers, balancing dampers, reheat coils, and other accessories.

Design Methods There is no single design method that automatically provides the most economical duct system for all conditions. A careful evaluation of all cost variables for a duct system should be made with each design method or combination of methods. Cost variables to consider include the cost of the duct material (the aspect ratios are a large factor), duct insulation or lining (duct heat gain or loss), type of fittings, space requirements, fan power, balancing requirements, sound attenuation, air distribution terminal devices, and heat recovery equipment. Slightly different duct system pressure losses can be obtained using the different design methods. Some require a broad background of design knowledge and experience. Careful use of these methods allows the designer to efficiently size HVAC duct systems for larger residences and institutional and commercial buildings, including some light industrial process ducts. Traditionally used duct design methods include the following: • • • • • •

Equal friction Static regain T-method Extended plenums Velocity reduction Constant velocity

Equal Friction (Equal Friction Rate) The equal-friction method of duct sizing (where the pressure loss per metre of duct is the same for the entire system) is probably the most universally used means of sizing lower-pressure supply air, return air, and exhaust air duct systems. It normally is not used for higher-pressure systems. With supply air duct systems, this design method automatically reduces air velocities in the direction of the airflow, thus reducing the possibility of generating noise (against the airflow in return or exhaust duct systems). The major disadvantage of the equal-friction method is that there is no provision for equalizing pressure drops in duct branches (except in symmetrical layouts). A manual balance of short runs, to achieve the same pressure drop as a long branch run, is required. The friction chart (Figure 7-2) gives the pressure drop in Pa/m, where the shaded area indicates the suggested design limits. Many designers use 0.8 Pa/m for ductwork with no acoustic treatment. For systems with VAV boxes, which provide a measure of sound attenuation, 1.6 Pa/m might be used from the supply from fan to VAV boxes, dropping to 0.8 Pa/m from VAV box to outlet. Whatever equal-friction choices are made, the data can be extracted from the friction chart and tabulated to provide a quick reference to the data needed.

Figure 7-2

Friction chart for round duct.

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Chapter 7 Duct System Design Table 7-6

Data for Duct Sizing at 0.8 Pa/m

Flow, L/s

Diameter, mm

Velocity, m/s

Velocity Pressure, Pa

30

125

2.5

3.8

60

160

3.0

5.4

110

200

3.5

7.4

The beginning of such a table is shown in Table 7-6. The reason for including velocity and velocity pressure becomes obvious when calculating the pressure drop through fittings in the ductwork. The Friction Chart (Figure 7-2) provides round duct sizes for a given flow and pressure drop (or velocity). Equivalent rectangular duct sizes are found in Table 7-7. For example, the required round duct size is 700 mm, but it is restricted to a dimension of 350 mm. Choosing 350 mm along the top of Table 7-7 drop down until a round equivalent diameter of at least 700 is reached. In this case it is 701 mm diameter and going to the left column one reads off the required dimension 1300 mm. Reading data from the friction chart and then another table to obtain equivalent sizes can be done with a simple cardboard device called a ductulator. By rotating one sheet of the ductulator, you can input two variables from volume, velocity, round-duct diameter, or rectangular-duct sides, pressure drop per m and read the other corresponding variables. For example, setting the pressure drop at 1 Pa/m and volume at 4000 L/s, you can read duct diameter, the combinations of equivalent rectangular-duct sides, and duct velocity.

Static Regain The static regain method of duct sizing may be used to design supply air systems of any velocity or pressure. It normally is not used for return air systems, where the airflow is toward the HVAC unit fan. This method is more complex than the equal friction method, but it is a theoretically sound method that meets the requirements of maintaining uniform static pressure at all branches and outlets. Duct velocities are systematically reduced, which allows a large portion of the velocity pressure to convert to static pressure that offsets the friction loss in the succeeding section of duct. The duct system stays in balance because the losses and gains are proportional to a function of the velocities. This static regain, which is often assumed at 75% for average duct systems, could be as low as 50%, or as high as 100+% under ideal conditions. The assumed regain factors can create installed systems that are quite different than the design requirements. The classical static regain method should not be used without a computer program to make actual mass flow calculations at branches, due to the unpredictable regain factor. A disadvantage of the static regain method is the oversized ducts that can occur at the ends of long branches, especially if one duct run is unusually long. Often, the resultant very low velocities require the installation of additional thermal insulation on that portion of the duct system to prevent unreasonable duct heat gains or losses.

Fundamentals of Air System Design SI Table 7-7

173

Equivalent Round and Rectangular Duct Sizes

Lgth Adj.

100

125

150

175

200

100 125 150 175 200 225 250 275 300 350 400 450 500 550 600 650 700 750 800 900 1000 1100 1200 1300 1400 1500 1600 1700 1800 1900 2000 2100 2200 2300 2400 2500 2600 2700 2800 2900

109 122 133 143 152 161 169 176 183 195 207 217 227 236 245 253 261 268 275 289 301 313 324 334 344 353 362 371 379 387 395 402 410 417 424 430 437 443 450 456

137 150 161 172 181 190 199 207 222 235 247 258 269 279 289 298 306 314 330 344 358 370 382 394 404 415 425 434 444 453 461 470 478 486 494 501 509 516 523

164 177 189 200 210 220 229 245 260 274 287 299 310 321 331 341 350 367 384 399 413 426 439 452 463 475 485 496 506 516 525 534 543 552 560 569 577 585

191 204 216 228 238 248 267 283 299 313 326 339 351 362 373 383 402 420 437 453 468 482 495 508 521 533 544 555 566 577 587 597 606 616 625 634 643

219 232 244 256 266 286 305 321 337 352 365 378 391 402 414 435 454 473 490 506 522 536 551 564 577 590 602 614 625 636 647 658 668 678 688 697

Length of One Side of Rectangular Duct (a), mm 225 250 275 300 350 400 450 500 Circular Duct Diameter, mm

246 259 272 283 305 325 343 360 375 390 404 418 430 442 465 486 506 525 543 559 575 591 605 619 663 646 659 671 683 695 706 717 728 738 749

273 287 299 322 343 363 381 398 414 429 443 457 470 494 517 538 558 577 595 612 629 644 660 674 688 702 715 728 740 753 764 776 787 798

301 314 339 361 382 401 419 436 452 467 482 496 522 546 569 590 610 629 648 665 682 698 713 728 743 757 771 784 797 810 822 834 845

328 354 378 400 420 439 457 474 490 506 520 548 574 598 620 642 662 681 700 718 735 751 767 782 797 812 826 840 853 866 879 891

383 409 433 455 477 496 515 533 550 567 597 626 652 677 701 724 745 766 785 804 823 840 857 874 890 905 920 935 950 964 977

437 464 488 511 533 553 573 592 609 643 674 703 731 757 781 805 827 849 869 889 908 927 945 963 980 996 1012 1028 1043 1058

492 518 543 567 589 610 630 649 686 719 751 780 808 835 860 885 908 930 952 973 993 1013 1031 1050 1068 1085 1102 1119 1135

547 573 598 622 644 666 687 726 762 795 827 857 886 913 939 964 988 1012 1034 1055 1076 1097 1116 1136 1154 1173 1190 1208

550

600

650

700

601 628 653 677 700 722 763 802 838 872 904 934 963 991 1018 1043 1068 1092 1115 1137 1159 1180 1200 1220 1240 1259 1277

656 683 708 732 755 799 840 878 914 948 980 1011 1041 1069 1096 1122 1147 1172 1195 1218 1241 1262 1283 1304 1324 1344

711 737 763 787 833 876 916 954 990 1024 1057 1088 1118 1146 1174 1200 1226 1251 1275 1299 1322 1344 1366 1387 1408

765 792 818 866 911 953 993 1031 1066 1100 1133 1164 1195 1224 1252 1279 1305 1330 1355 1379 1402 1425 1447 1469

Note that the loss coefficients for duct fittings found in the 2009 ASHRAE Handbook—Fundamentals (ASHRAE 2009) include static pressure regain or loss for the velocity condition changes that occur at divided flow or change-of-size duct fittings. Additional duct static pressure regain (or loss) must not be calculated and added to (or subtracted from) the total duct system pressure losses when those fitting losses are used. The Total Pressure Method is a further refinement of the static regain method that allows the designer to determine the actual friction and dynamic losses at each section of the duct system. The advantage is having the actual pressure losses of the duct sections and the fan total pressure requirements provided.

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Chapter 7 Duct System Design

T-Method The T-method of duct sizing is a comprehensive duct design optimization procedure that includes system initial costs and operating costs, energy costs, hours of operation, annual escalation, interest rates, etc. A description of the method and main procedures and equations may be found in the 2009 ASHRAE Handbook—Fundamentals chapter on duct design. The method requires computer software and an extensive evaluation of acoustic results.

Extended Plenums An extended plenum is a trunk duct (usually at the discharge of a fan, fan-coil unit, mixing box, VAV box, etc.) extended as a plenum to serve multiple outlets and/or branch ducts with essentially equal pressure. In a similar way, one can choose to maintain the duct size beyond more than one branch in what is known as a semiextended plenum. For example, a duct with four branches could be the same size until after the second branch and a smaller size until after the fourth branch rather than reduced after each branch. Advantages may include lower first-costs, lower operating costs, ease of balancing, and adaptability to branch duct or outlet changes. A disadvantage is that heat gins or losses will be slightly increased.

Velocity Reduction In this method, a system velocity is selected at the section next to the fan and arbitrary reductions in velocity are made after each branch or outlet. The resultant pressure loss differences in the various sections of the duct system are not taken into account, and balancing is attempted mainly by the use of good dampers at strategic locations. An experienced designer who uses sound judgment in selecting arbitrary velocities may design a relatively simple duct system using the velocity reduction method. Other practitioners should not attempt to use this method except for estimating purposes, unless the system has only a few outlets and can be easily balanced.

Constant Velocity With adequate experience, many designers are able to select an optimum velocity that is used throughout the design of a duct system. This method is best adapted to the higher pressure systems that use attenuated terminal boxes to reduce the velocity and noise before distribution of the air to the occupied spaces. Industrial exhaust systems often use the constant-velocity method to ensure particulate movement along with the exhaust airstream.

Other Design Considerations The amount of duct leakage in an HVAC system may be determined by the system designer using data from HVAC Duct Construction Standards—Metal and Flexible (SMACNA 2005) and HVAC Air Duct Leakage Test Manual (SMACNA 1985). Leakage in ducts varies with the fabricating machinery

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used, the methods of assembly, and the quality of the installation workmanship, plus the effectiveness of any sealants, if used, and the workmanship in their application. A variety of sealed and unsealed duct leakage tests have confirmed that longitudinal seam, transverse joint, and assembled duct leakage can be represented by Equation 7-1, and that for the same construction, leakage is not significantly different in the negative and positive modes: N

Q = CΔp S where Q = C =

leakage rate, L/s per m2 constant reflecting area characteristics of leakage path

Δps N

static pressure differential from duct interior to exterior, Pa exponent relating turbulence or laminar flow in leakage path

= =

(7-1)

Analysis of the data resulted in the categorization of duct systems into a leakage class, CL, the accepted value of N = 0.65, and Q now defined in terms of surface area of the duct:

Figure 7-3

Duct leakage vs. static pressure (SMACNA 2004).

176

Chapter 7 Duct System Design 0.65

C L = 1000Q ⁄ Δp s where Q =

leakage rate, L/s per m2 (surface area)

CL

leakage class, mL/(s⋅m2) at 1 Pa

=

(7-2)

Figure 7-3 shows how duct pressure affects the leakage rate for each leakage class. Table 7-8 is a forecast of the leakage class attainable for commonly used duct construction and sealing practices. Note that connections of ducts to grilles, diffusers, and registers are not represented in the test data. The HVAC system designer is responsible for assigning acceptable leakage rates. Table 7-8

Duct Leakage Classification Sealed

Duct Type

Unsealed

Predicted Leakage Class CL

Leakage Rate*

Predicted Leakage Class CL

Leakage Rate*

4

0.14

42

1.5

(8 to 99)

(0.3 to 3.6)

68

2.5

(17 to 155)

(0.6 to 5.6)

68

2.5

(17 to 155)

(0.6 to 5.6)

42

1.5

(17 to 76)

(0.6 to 2.8)

30

1.5

(6 to 76)

(0.2 to 2.8)

Metal (flexible excluded) Round and flat oval

Rectangular ≤500 Pa

17

0.62

(both positive and negative pressures) >500 and ≤2500 Pa

8

0.29

(both positive and negative pressures)

Flexible Metal, aluminum

Nonmetal

11

17

0.40

0.62

Fibrous glass Round

4

0.14

N/A

N/A

Rectangular

8

0.29

N/A

N/A

* Leakage rate, L/(s⋅m2) at 250 Pa

Fundamentals of Air System Design SI

177

Example 7-1 Problem Solution

How much leakage is likely from an unsealed 24 m section of 1 m square duct (Figure 7-4) if the internal pressure is 600 Pa? From Table 7-8 it is determined that the predicted leakage rate for an unsealed rectangular duct with pressure ≥500 Pa is typically 68. Using Figure 7-2, for a pressure difference of 600 Pa and CL of 68, Q is about 4.3 L/(s⋅m2). The duct has an area of 4 × 1 × 24 = 96 m2. Total leakage is thus 4.3 × 96 = 413 L/s, depending on construction and installation quality. Note that in Table 7-8, poor construction and installation of unsealed ductwork can double the typical leakage rate. Even more significant is the reduction in leakage achieved by sealing the duct. In the example case, the classification drops from 68 to 8; leakage drops by a factor of 8. ANSI/ASHRAE/IESNA Standard 90.1-2007, Energy Efficient Design of New Buildings Except Low-Rise Residential Buildings (ASHRAE 2007) prescribes minimum sealing requirements for supply, return, and exhaust ducts run outside in conditioned spaces and unconditioned spaces.

Duct Heat Gain or Loss At the beginning of this chapter, it was stated that duct design follows building load calculations. An often overlooked factor in load calculations is duct heat gain or loss. The method of calculating this load is well described in other texts, such as the 2009 ASHRAE Handbook—Fundamentals. In this section, some practical considerations in duct design that affect duct heat gain or loss are noted. Consider first a conditioned-air supply system with the air-handling apparatus and ductwork in the conditioned space and with no additional load imposed on the system. However, if the ductwork is long and the velocities are low, the designer should check that airflows are proportioned properly. The air in the ductwork still becomes warmer or cooler as it passes through the conditioned space, thus decreasing the temperature difference. As a result, less air is

Figure 7-4

Unsealed 24 m section of 1 m square duct (see Example 7-1).

178

Chapter 7 Duct System Design required to supply the outlets at the start of the supply run and more is required at the end. Naturally, when a duct or plenum carrying conditioned air is located outside the conditioned space, the heat gain or loss must be accounted for in both the design air quantity and total sensible load. This system load must be calculated by the designer when running conditioned-air ductwork through boiler rooms, attics, outdoors, or other unconditioned spaces. Alternate routing might be more desirable than increasing the system load. With certain exceptions, ASHRAE/IES Standard 90.1 requires thermal insulation of all duct systems and their components (such as ducts, plenums, and enclosures) installed in or on buildings. To estimate duct heat transfer and entering or leaving air temperatures, use Equations 7-3, 7-4, and 7-5. Duct air exit temperatures can then be estimated using the following equations:

where y y A V D L Ql U p ρ te tl ta

UPL t e + t 1 Q 1 = ------------ ⎛ --------------⎞ – t a 12 ⎝ 2 ⎠

(7-3)

t 1 ( y + 1 ) – 2t a t e = -------------------------------(y – 1)

(7-4)

t 1 ( y – 1 ) + 2t a t 1 = ---------------------------------(y + 1)

(7-5)

= 2.4 AVρ/UPL for rectangular ducts = 0.6 DVρ/UL for round ducts = cross-sectional area of duct, mm2 = average velocity, m/s = diameter of duct, m = duct length, m = heat loss/gain through duct walls = overall heat transfer coefficient of duct wall, W/(m2⋅K) = perimeter of bare or insulated duct, mm = density, kg/m3 = temperature of air entering duct, °C = temperature of air leaving duct, °C = temperature of air surrounding duct, °C

Use Figure 7-5 to determine the U-factors for insulated and uninsulated ducts. For a 50 mm thick, 12 kg/m3 fibrous glass blanket compressed 50% during installation, the heat transfer rate increases about 20%, as shown in Figure 7-5. Pervious flexible duct liners also influence heat transfer significantly, as shown in Figure 7-5. At 12.5 m/s, the pervious liner U-factor is 1.1 W/(m2·K). For an impervious liner, the U-factor is 1.9 W/(m2·K).

Fundamentals of Air System Design SI

Figure 7-5

Heat transfer coefficients.

179

180

Chapter 7 Duct System Design

Example 7-2 Problem

Solution

A 20 m length of 0.6 × 0.9 m uninsulated sheet metal duct, freely suspended, conveys heated air through a space maintained at 4°C. The ASTM C680-08, Standard Practice for Estimate of the Heat Gain or Loss and the Surface Temperatures of Insulated Flat, Cylindrical, and Spherical Systems by Use of Computer Programs (ASTM 2008) heat loss calculation gives a heat transfer rate of 444 W/m2. Based on heat loss calculations for the heated zone, 8.1 m3/s of standard air (cp = 1006 J/[kg·K]) at a supply air temperature of 50°C is required. The duct is connected directly to the heated zone. Determine the temperature of the air entering the duct. The area of the duct is 0.6 × 0.9 = 0.54 m2 Air velocity V is calculated as flow/area = 8.1/0.54 = 15 m/s Duct perimeter P is (0.6 + 0.9) × 2 = 3 m Temperature drop tdrop is 444 × 3 × 20 -----------------------------------------------------= 2.7K 15 × 1006 × 1.2 × 0.54 Temperature of air entering the duct is 50 + 2.7 = 52.7°C.

Fitting Losses Pressure loss at fittings is a critical element of duct system design. The simplest way to incorporate fitting losses into the design is to use loss coefficients taken from the ASHRAE Duct Fitting Database (ASHRAE 2008a) tables such as the ones found in the 2009 ASHRAE Handbook—Fundamentals. These loss coefficients represent the ratio of total pressure loss to the dynamic pressure (in terms of velocity pressure). They do not include duct friction loss (which is picked up by measuring the length of duct sections to fitting center lines). However, the loss coefficients do include static regain (or loss) where there is a change in velocity. The total pressure (Δpt) loss of a fitting is determined using the loss coefficient in the following equation: Δp t = C o × p v where Δpt = Co = Pv =

total pressure loss, Pa Dimensionless local loss coefficient Velocity pressure, Pa

(7-6)

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181

By using duct fitting loss coefficients that include static pressure regain or loss, accurate duct system fitting pressure losses are obtained. When combined with the friction losses of the straight duct sections sized by the modified equal friction method, the result is the closest possible approximation of the actual system total pressure requirements for the fan.

Example 7-3 Problem

Solution

To demonstrate the use of the loss coefficient tables, assume a velocity of 12 m/s in a 300 mm 7 Gore (segment), 90° elbow, as shown in Figure 7-6. According to Table 7-9, the Co for this fitting is 0.08 for D = 300 mm. Using Equation 1-8, Pv = 0.602 V2 Pa, we determine that the velocity pressure is 87 Pa. Using Equation 7-6, we determine the total pressure loss: Δp t = C 0 × p v = 0.08 × 87 = 7 Pa

Figure 7-6

CD3-10 elbow.

Table 7-9

Co Values for CD3-10 Elbow*

D, mm

75

150

230

300

380

450

690

1500

Co

0.16

0.12

0.10

0.08

0.07

0.06

0.05

0.03

*7 Gore, 90 degree, r/D = 2.5

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Sample Systems The following simplified sample systems illustrate calculating pressure drop in “real” systems.

System 1 Problem

Solution

As illustrated in Figure 7-7, this system consists of a fan, a straight length of galvanized steel duct and an outlet. The duct is the same size as the fan outlet, so no system effect factor needs to be added. The outlet is a VAV box with a 220 Pa pressure drop at 2000 L/s. The fan speed is adjusted to deliver 2000 L/s. What is the pressure drop in the system? 1. Find the circular equivalent of the rectangular duct using Table 7-7. Find the column headed 250 mm, drop down to the row labeled 350, and read the 322 mm equivalent circular duct diameter. It can also be calculated using the following formula: 0.625

( ab ) - = 322 mm D e = 1.30 -------------------------0.250 (a + b) where De = circular equivalent of rectangular duct for equal length, fluid resistance, and airflow, mm

Figure 7-7

System 1.

Fundamentals of Air System Design SI a b

183

= length of one side of duct, mm = length of adjacent side of duct, mm

2. Calculate the velocity of the airstream in the duct: 2000 -----------Q 1000 V = ----- = ------------------------------------ = 22.86 m/s Ae 250- ----------350-⎞ ⎛ ----------× ⎝ 1000 1000⎠ 3. Calculate the fan outlet velocity pressure from Equation 1-8: 2

2

p v = 0.602V = 0.602 × 22.86 = 315 Pa 4. Find the pressure drop in the duct per metre using the friction chart shown in Figure 7-2: 2000 L/s in a 322 mm duct gives approximately 20 Pa/m 5. Calculate the pressure drop in the 9 m run of duct: 9 m × 20 Pa = 180 Pa 6. Add the pressure drops: Pressure drop in 9 m duct run

180 Pa

Fan outlet velocity pressure

315 Pa

Required VAV outlet pressure drop

200 Pa

Total pressure required at fan

695 Pa

System 2 Problem

Solution

Figure 7-8 shows the same system in Figure 7-7, except that the outlet at the end of the duct run was removed and a 20° 350 × 250 mm to 600 × 250 mm rectangular transition was added. Attached to the transition outlet are two 300 × 250 mm elbows with r = 1.5 (Co = 0.2). Attached to each elbow is a branch duct. One branch is 6 m long, the other branch is 4 m long. At each end of the duct extension is a VAV box with a 220 Pa pressure drop and capacity of 1000 L/s. What is the pressure drop in the system? 1. Calculate the pressure drop, velocity, and pressure at the end of the 9 m run. Because this is the same configuration as System 1, on the main 9 m

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Figure 7-8

System 2.

run the pressure drop is 180 Pa, the velocity is 22.86 m/s, and the velocity pressure is 315 Pa (same as in System 1). 2. Calculate the pressure drop in the transition: a. Calculate the ratio of the inlet area to the outlet area: Outlet area600 × 250 600 ------------------------= ------------------------ = --------- = 1.7 Inlet area 350 × 250 350 b. Refer to Table 7-10. This table gives Co values for rectangular transitions (Figure 7-9). Because the table does not give an exact value for an outlet/inlet ratio of 1.7, by interpolation, the Co value is estimated to be 0.5. Multiply the velocity pressure at the inlet by the Co value to calculate the pressure drop across the transition: 315 Pa × 0.5 = 157.5 Pa

Fundamentals of Air System Design SI Table 7-10

185

SR4-1 Rectangular Transition* (ASHRAE 2009) Co Values Θ

Ao/A1

0

3

5

10

15

20

30

45

60

90

120

150

180

0.10

0.0

0.12

0.09

0.05

0.05

0.05

0.05

0.06

0.08

0.19

0.29

0.37

0.43

0.167

0.0

0.11

0.09

0.05

0.04

0.04

0.04

0.06

0.07

0.19

0.28

0.36

0.42

0.25

0.0

0.10

0.08

0.05

0.04

0.04

0.04

0.06

0.07

0.18

0.27

0.36

0.41

0.50

0.0

0.08

0.09

0.06

0.04

0.04

0.04

0.06

0.07

0.12

0.17

0.20

0.27

1.00

0.0

0.00

0.00

0.00

0.00

0.00

0.00

0.00

0.00

0.00

0.00

0.00

0.00

2.00

0.0

0.64

0.96

0.54

0.52

0.62

0.94

1.40

1.48

1.52

1.48

1.44

1.40

4.00

0.0

4.16

4.64

2.72

3.09

4.00

6.72

9.60

10.88 11.20 11.20 10.88 10.56

6.00

0.0

12.24 10.08

7.38

8.10

10.80 17.28 23.40 27.36 29.88 29.88 29.34 28.80

10.00

0.0

40.50 27.20 23.30 25.10 34.00 52.84 69.00 82.50 93.50 93.50 92.40 91.30

16.00

0.0

112.64 68.35 63.74 67.84 92.93 142.13 182.53 220.16 254.21 254.21 251.90 249.60

*Two sides parallel, symmetrical.

Figure 7-9

SR4-1 rectangular transition, two sides parallel, symmetrical, supply air systems.

3. Calculate the pressure drop in the elbows. The calculation for each elbow is the same. The average velocity of the airstream entering each elbow is: 2000 -----------Q 1000 V = ---- = ------------------------------------ = 13.3 m/s A 600 250-⎞ ⎛ ------------ × ----------⎝ 1000 1000⎠

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Chapter 7 Duct System Design a. Convert the 300 × 250 mm rectangular duct to circular measurement using Table 7-7. Find the column headed 250 mm, drop down to the row labeled 300, and read off 299 mm; use 300 mm equivalent circular duct diameter. b. Calculate the velocity pressure in the elbows: 2

p v = 0.602 × 13.3 = 107 Pa c. Multiply the pressure at the inlet by the Co value (given in the problem statement as 0.2) to calculate the pressure drop across each elbow: 107 Pa × 0.2 = 21.4 Pa 4. Calculate the pressure drop in the 6 m branch, the longest branch. a. Find the pressure drop in the branch duct per metre using the friction chart in Figure 7-2, based on the equivalent diameter of 300 mm and volume flow of 1000 L/s: 6 Pa per m b. Calculate the pressure drop for the 6 m branch run: 6 m × 6 Pa/m = 36 Pa 5. The loss of the VAV box is given as 220 Pa (which is assumed to include the pressure losses downstream of the box). Also note that the duct size and the box inlet are the same size. If this is not the case then there would be losses or gains, depending on whether the inlet is smaller or larger than the branch duct. If the inlet is smaller, there would be an additional loss due to increasing the velocity, which is equal to the difference in velocity pressures, which must be included. 6. The pressure drop in the 4 m branch is slightly less than in the 6 m branch. The difference in length is 2 m, so the 4 m branch drop is 2 m × 6 Pa = 12 Pa less than in the 6 m branch. As the flows are being controlled by VAV boxes, this variation does not matter, as the boxes control their flows. 7. Add the pressure requirements: Pressure drop in 9 m duct run

180 Pa

Pressure drop at transition

157 Pa

300 × 250 mm elbow

21 Pa

6 m branch duct

36 Pa

Required VAV outlet pressure

220 Pa

Fan outlet velocity pressure

315 Pa

Total pressure drop

929 Pa

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The Next Step Chapter 8 deals with codes and standards that are relevant for air system design and energy usage.

Summary Air duct system design must consider space availability, space air diffusion, noise, duct leakage, duct heat gains and losses, balancing, fire and smoke control, and initial costs plus operating costs. Many materials are used for ductwork, but the vast majority uses galvanized steel. For this reason, duct design information is for galvanized steel with corrections for other materials. Other materials offer better chemical, moisture, acoustic, and high temperature performance, typically at a premium cost. Lined duct must be sized to include the lining. The duct drawing must clearly state that the duct dimension is the metal size or the airway size. Rectangular metal ducts are manufactured to standard specifications for size, static pressure (positive or negative), material thickness, jointing, reinforcing, and supports produced by SMACNA. When choosing the pressure rating, take care to allow for probable maximum and minimum pressures on all but the smallest systems. Round and oval ducts are inherently strong and rigid and are generally the most efficient and economical ducts for air systems. However, their shape may not fit the available route through the building. Fibrous glass ducts are a composite of faced rigid fiberglass available in molded round sections or in board form for fabrication. Duty is generally limited to 14 m/s and 500 Pa. Flexible ducts are typically manufactured from a coiled wire and fabric and are used for connecting ducts to diffusers. Ducts must be sufficiently airtight to ensure economical and quiet performance of the system. Leakage classifications are given in L/s⋅m2 surface area at 1 Pa. Actual leakage is Q = CLΔ p0.65. Many materials, gaskets, and tapes are available, but many have an unfortunately short life. ASHRAE/IES Standard 90.1 prescribes minimum sealing requirements for many duct situations. Once the room loads have been calculated and temperature difference chosen, the air volumes to each room can be calculated. Depending on the duct insulation and the temperature of the space the duct runs through, there will be some heat gain or loss that should be included at this stage. For designs to meet ASHRAE/IES Standard 90.1, the minimum insulation values for energy conservation must be met. Once the air volume to the room, room layout, and architectural features and requirements are known, a preliminary layout for outlets is made (as discussed in Chapter 3). Generally, all outlets on the same branch duct should have the same pressure drop, particularly if they are of different types.

188

Chapter 7 Duct System Design Due to variations in loads, the HVAC system will be zoned. Typically, interior and exterior spaces are on separate zones, and the duct layout must accommodate the zoning and associated air-control devices. The preliminary layout is drawn, ideally over the architectural layout for supply, return, and exhaust ducts. Some very preliminary sizing is done at this stage to ensure there is space for the main duct runs. It is important at this stage to remember to include space for duct joints, supports, insulation, and outlets. Allowance must also be made for other services that may have required slopes, typically plumbing, and required locations, typically sprinklers and lighting. With the preliminary HVAC system duct layout completed, accurate duct sizing must be undertaken either with a computer program or manually. Sizing is more straightforward if all ducts are round, as any rectangular ducts must be converted to equivalent round size for calculating the resistances. Once the ducts are sized, the final calculation of system pressure drop and location of all outlets, control items, and fire and smoke dampers can be fixed. Duct design is somewhat of an art. There is a choice of design methods; technical design must be balanced with cost and ease of installation and balancing. Slightly different pressure losses are obtained using different design methods and source data, and these are often changed somewhat as the installation contractor coordinates with other trades. With the equal-friction method, a fixed pressure drop in Pa/m is chosen and used to perform the duct sizing. The method is simple and decreases the velocity towards outlets, which provides quiet systems. Care must be taken to avoid very unequal branch resistances, which can cause significant energy waste and noise due to damper pressure drops. With the static regain method, the velocity pressure is systematically reduced to offset the prior duct run pressure drop. The method is not easy to employ manually and may need to be modified in cases where branch ducts are very different in length. The T-method is an optimization procedure, ideally run in a computer program, that designs on the basis of finding the most economic design based on initial costs and operating costs. An extended plenum is a trunk duct maintained at full size to provide a relatively equal supply pressure to each branch. A variation—the semiextended plenum—keeps the duct size up for a greater length than is necessary, often reducing the cost of numerous size reductions. An experienced designer can learn to size based on constant pressure drop modified by velocity. This is known as the velocity reduction method. For the experienced designer, a constant velocity can be chosen for sizing, especially where noise is not an issue or where all outlets include sound attenuation. Industrial exhaust systems often use constant velocity sizing to ensure particulate movement along with the exhaust airstream.

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References and Bibliography ASHRAE. 2007. ANSI/ASHRAE/IESNA Standard 90.1-2007, Energy Efficient Design of New Buildings Except Low-Rise Residential Buildings. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. ASHRAE. 2008a. ASHRAE Duct Fitting Database, v5.00.00. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. ASHRAE. 2008b. ASHRAE Handbook—Systems and Equipment. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. ASHRAE. 2009. ASHRAE Handbook—Fundamentals. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. ASTM. 2008. ASTM C680-08, Standard Practice for Estimate of the Heat Gain or Loss and the Surface Temperatures of Insulated Flat, Cylindrical, and Spherical Systems by Use of Computer Programs. West Conshohocken, PA: American Society for Testing and Materials. SMACNA. 1985. HVAC Air Duct Leakage Test Manual. Chantilly, VA: Sheet Metal and Air Conditioning Contractors’ National Association Inc. SMACNA. 2003. Fibrous Glass Duct Construction Standards. Chantilly, VA: Sheet Metal and Air Conditioning Contractors’ National Association Inc. SMACNA. 2004. Rectangular Industrial Duct Construction Standards. Chantilly, VA: Sheet Metal and Air Conditioning Contractors’ National Association Inc. SMACNA. 2005. HVAC Duct Construction Standards—Metal and Flexible, 3rd edition. Chantilly, VA: Sheet Metal and Air Conditioning Contractors’ National Association Inc.

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Skill Development Exercises for Chapter 7 Complete these questions by writing your answers on the worksheets at the back of this book. 7-1

The system depicted in the figure below consists of a fan, ductwork, and outlets. The duct is the same size as the fan outlet, so no system effect factor needs to be added. The outlets are VAV boxes with a 180 Pa pressure drop at 1200 L/s.The fan speed is adjusted to deliver 2400 L/s. The Co value of the elbow is 0.2, and the transition is style SR4-1 as in Example 2. The total pressure drop in the system is _______________.

a) about 726 Pa b) about 387 Pa c) about 539 Pa d) none of the above 7-2

Air duct system design must consider _______________. a) noise b) duct leakage, heat gains, and heat losses c) fire and smoke control d) all of the above

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7-3

Duct sizing and construction specifications are generally stated in terms of the use of _______________. a) galvanized steel b) aluminum c) fiberglass reinforced plastic d) none of the above

7-4

Generally, the most efficient and economical ducts for air systems are a) rectangular b) oval c) round d) all of the above

7-5

Duct systems of rectangular fibrous glass are generally limited to a) 12 m/s and ±0.5 kPa b) 20 m/s and ±1 kPa c) 10 m/s and ±200 Pa d) none of the above

7-6

Compression of flexible ducts significantly decreases airflow resistance. a) True b) False

7-7

Sealant systems have been developed that can substitute for mechanical joining of ductwork. a) True b) False

Codes and Standards

Study Objectives After completing this chapter, you should be able to list the following principle codes and standards affecting air system design and briefly state what they cover and why they are important: T

T T T T T

ANSI/ASHRAE/IESNA Standard 90.1-2007, Energy Efficient Design of New Buildings Except Low-Rise Residential Buildings, ANSI/ASHRAE/ IESNA 90.1 User’s Manual (ASHRAE 2007a, 2007b) and other ASHRAE energy conservation standards ANSI/ASHRAE Standard 62.1-2010, Ventilation for Acceptable Indoor Air Quality (ASHRAE 2010) NFPA 90A–Installation of Air Conditioning and Ventilating Systems (NFPA 2002) NFPA 90B—Installation of Warm Air Heating and Air-Conditioning Systems (NFPA 2006) NFPA 96–Standard Ventilation Control and Fire Protection of Commercial Cooking Operations (NFPA 2004) HVAC Duct Construction Standards—Metal and Flexible (SMACNA 2005)

Instructions Read the material in Chapter 8. Verify the examples presented in the chapter with your own calculations. At the end of the chapter, complete the skill development exercises. Review those sections of the chapter as needed to complete the exercises.

Building Code Requirements In the private sector, each new construction or renovation project is normally governed by national, state, or local laws that require compliance with specific health, safety, property protection, and energy conservation regulations. Figure 8-1 depicts relationships between laws, ordinances, codes, and standards that can affect the design and construction of private-sector HVAC duct systems in the

194

Chapter 8 Codes and Standards

Figure 8-1

Hierarchy of building codes and standards.

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195

United States. However, Figure 8-1 may not list all applicable regulations and standards for a specific locality. It is the designer’s responsibility to establish what laws are applicable to a project. Generally, code officials will provide help if asked before the design is submitted. Code changes require long cycles for consensus approval. Because the development of safety codes, energy codes, and standards proceed independently, the most recent edition of a code or standard may not be adopted by a local jurisdiction. HVAC designers must know which code compliance obligations affect their designs. If a provision conflicts with the design intent, the designer should resolve the issue with local building officials. New or different construction methods can be accommodated by the provisions for equivalency that are incorporated into codes. Staff engineers from the model code agencies are available to help resolve conflicts, ambiguities, and equivalencies.

ASHRAE/IES Standard 90.1 Codes and standards have become much more important. With substantial increase in energy demands, many codes now also incorporate a minimum energy performance requirement. Specifically, many model codes in the United States reference ASHRAE/IES Standard 90.1. Originally drafted in 1975, ASHRAE Standard 90 was revised and reissued in 1980, 1989, 1999, 2004, and 2007. The original standard dealt with all buildings and was split into 90.1, for all but low-rise residential buildings, and 90.2 for low-rise residential buildings. ASHRAE/IES Standard 90.1 has been revised into code language to make it code enforceable. The standard is on an ANSI continuous maintenance schedule; addenda are issued for review when ready and approved when they have passed the public review process. To assist with code enforcement, the standard is reprinted with all addenda incorporated every three years, with the latest printed revision in 2007. The original ASHRAE Standard 90 was very important because it was one of the first documents that truly addressed what can be done in the design of buildings to conserve energy. It went through an extensive review and was commented on by thousands of engineers across the country. Much of the information in ASHRAE/IES Standard 90.1 has been adopted by model building codes. The standard deals with all design aspects of building energy use including Section 5, “Building Envelope”; Section 7, “Service Water Heating”; Section 8, “Power”; and Section 9, “Lighting.” The most relevant section for this course is Section 6, “Heating, Ventilating, and Air Conditioning,” although choices made in the other sections will affect the air system choice, sizing, and zoning. HVAC systems are one of the most significant energy users in the buildings covered by Standard 90.1. However, the designer has significant latitude in the

196

Chapter 8 Codes and Standards energy costs and consumption of HVAC systems; a poorly designed system easily can have twice the annual energy costs of an energy-conserving design. Analyzing the energy use and cost of an HVAC system is complicated by system interactions. An efficient system is not merely characterized as one that uses efficient equipment. System-level efficiency must account for installation, control, maintenance, system losses, and component interactions (such as reheat or heat recovery). As a conceptual model, overall HVAC system efficiency may be defined as the ratio of loads the system must handle (space heating and cooling as well as water heating) to the energy the system consumes. An efficient system minimizes energy use by minimizing system losses, maximizing equipment efficiencies, using free heating/cooling, and recovering heat where possible. Section 6 of ASHRAE/IES Standard 90.1 approaches the regulations of HVAC system design by addressing the following fundamental factors of system efficiency: • • • • • • • • • • •

Specifying minimum equipment efficiencies Reducing system losses from ductwork through sealing and insulation Reducing system losses from piping through insulation Reducing system operation through the use of automatic time controls and zone isolation Reducing system inefficiencies by minimizing simultaneous heating and cooling Reducing system inefficiencies by shutting OFF outdoor ventilation during setback and warm up Reducing system operation through requirements for zone controls Reducing system inefficiencies by limiting equipment oversizing Reducing distribution losses, limiting HVAC fan energy demand, and requiring efficient balancing practices Requiring systems to take advantage of cool weather to provide free cooling Requiring energy recovery on systems of over 2500 L/s and 70% outside air

The format of the standard is intended to be general and flexible so it may be applied to many different building types, HVAC system types, and climates. Although compliance with Section 6 assures a minimum level of HVAC system performance, designers are encouraged to view the requirements as a starting point and investigate designs that exceed these requirements. Careful design and application of heat recovery, solar energy, or high-efficiency equipment can create systems that are more efficient than the standard requires and offer excellent returns on investment. The process of life-cycle costing is used to determine that the proposed alternates have an economic payback.

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Compliance Methods There are three primary subsections in Section 6. The first is a simplified approach for smaller buildings with simple HVAC systems. Second, there are mandatory requirements in Section 6.4 that must be met for either compliance path. Last of all, the prescriptive requirements in Section 6.5 include measures that must be met to show compliance via the prescriptive method. In this method, the designer must choose equipment with required performance and obey a number of design requirements. These prescriptive requirements do not have to be met with the energy cost budget method detailed in Section 11 of the standard. In the energy cost budget method, the building designers must show that their design would have no greater energy cost than a building designed under the prescriptive route. Many of the Section 6 requirements apply to larger, multiple-zone systems. The breadth of the section may seem overwhelming to designers of simpler, single-zone HVAC systems that are typically used in one- or two-story buildings under 2500 m2.

ASHRAE Standard 62.1 ASHRAE Standard 62.1 was introduced briefly in Chapter 3 as the standard that sets minimum outdoor-air ventilation rates and requirements for exhaust. The standard also sets requirements to provide acceptable indoor air quality during the building’s lifetime, and it requires documentation of the design assumptions and that they are available for the system’s operation. The standard includes requirements in the system planning that address the following questions: • • •

• • •

How much outside air is required in each space? How will the differing requirements for each space be achieved? When variable-air-volume systems are used, how will the required ventilation air volume be maintained when the supply volume to a space is reduced? How effectively is the ventilation air distributed to the occupants in the space? What quality of air can be recirculated from one space to another space? What ventilation is required when occupancy varies over time? The standard includes specific construction requirements for

• • • •

outdoor-air intakes to minimize moisture problems due to rain and snow, filtration requirements to prevent wet coils from excessive dirt collection, drain pans’ slope and drainage arrangement to ensure that condensation drains away, access for maintenance and cleaning of coils,

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Chapter 8 Codes and Standards • •

duct construction, and system startup and balancing.

The standard has requirements for the system’s ongoing operation and maintenance after construction, including inspection and measurement of outdoor airflow.

Other Codes and Standards Several organizations other than ASHRAE produce codes and standards tat relate to HVAC duct design. Included among these are the National Fire Protection Association (NFPA), the Sheet Metal and Air Conditioning Contractors’ National Association (SMACNA), and the American Conference of Governmental Industrial Hygienists (ACGIH).

National Fire Protection Association The NFPA issues a wide range of standards. Three of interest to HVAC designers are NFPA 90A—Installation of Air Conditioning and Ventilating Systems (2002), NFPA 90B—Installation of Warm Air Heating and Air-Conditioning Systems (2006), and NFPA 96—Installation of Equipment for the Removal of Smoke and Grease-Laden Vapors from Commercial Cooking Equipment (2004). NFPA 90A applies to systems for air movement in • • •

structures over 1000 m3 in volume; buildings of Type III, IV, and V construction over three stories in height, regardless of volume; and buildings, spaces, occupants or processes not covered by other NFPA standards.

As stated in the standard, the purpose of NFPA 90A is “to prescribe minimum requirements for safety to life and property from fire.” The requirements of NFPA 90A are intended to • • •

• •

restrict the spread of smoke through air duct systems in a building or into a building from the outside; restrict the spread of fire through air duct systems from the area of fire origin, whether it be within the building or from outside; maintain the fire-resistant integrity of building components and elements (such as floors, partitions, roofs, walls, and floor/roof-ceiling assemblies) affected by the installation of air duct systems; minimize ignition sources and combustibility of the elements of the air duct systems; and permit the air duct systems in a building to be used for the additional purpose of emergency smoke control.

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NFPA 90A provides requirements for HVAC systems (equipment and air distribution), integrating HVAC systems with building construction, controls, and acceptance testing. Of particular interest with respect to duct design, Figure 8-2 shows required treatments of penetrations of walls or partitions and location of fire and smoke dampers. The requirement is that fire dampers be shown on the drawings. The building’s architectural design determines its fire separations and the requirements for duct fire and smoke dampers. Fire dampers are a significant cost, and access doors must be provided for checking and servicing them. When laying out the ductwork, choices can often be made to reduce the number of fire dampers and to position the access doors to minimize costs. NFPA 90B applies to all warm-air heating and air-conditioning systems that serve one- or two-family dwellings, and spaces not exceeding 1000 m3 in volume in any occupancy (for example, light commercial). Other systems are covered by NFPA 90A. NFPA 90B addresses system components, fire integrity of building construction, equipment, wiring, and controls. NFPA 96 covers basic requirements for the design, installation, and use of exhaust system components, including hoods; grease removal devices; exhaust ducts; dampers; air-moving devices; auxiliary equipment; and fire extinguishing equipment for the exhaust system and the cooking equipment used in commercial, industrial, institutional, and similar cooking applications. Other topics discussed in NFPA 96 include duct systems, air movement, procedures for use and maintenance of equipment, and minimum safety requirements for cooking equipment. This standard does not apply to installations for normal residential family use.

Sheet Metal and Air Conditioning Contractors’ National Association The SMACNA HVAC Duct Construction Standards—Metal and Flexible (2005) cover basic duct construction; fittings and other construction; round, oval, and flexible ducts; hangers and supports; exterior components; casings; functional criteria for demonstrating equivalency; and duct sealing classifications. Also included are highly valuable appendices, useful in duct construction, and fibrous glass duct construction standards.

American Conference of Governmental Industrial Hygienists The American Conference of Governmental Industrial Hygienists (ACGIH) publishes and regularly updates Industrial Ventilation: A Manual of Recommended Practice for Design (ACGIH 2010). This book includes general information on ventilation and numerous examples of industrial ventilation and the removal of contaminants from specific industrial processes.

Figure 8-2

Wall and partition penetrations and smoke dampers.

200 Chapter 8 Codes and Standards

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Sources of Information Many sources of information are available to HVAC designers. ASHRAE produces an extensive range of publications, including Standards, Handbook, and Advanced Energy Design Guides, which can be located online at the ASHRAE Web site (www.ashrae.org). Other resources include the following: Air Movement and Control Association Inc. (AMCA) 30 West University Drive Arlington Heights, IL 60004-1893 Phone: (708) 394-0150 Fax: (708) 253-0088 www.amca.org American Conference of Governmental Industrial Hygienists (ACGIH) Kemper Woods Center 1330 Kemper Meadow Dr. Cincinnati, OH 45240 Phone: (513) 742-2020 Fax: (513) 742-3355 www.acgih.org National Fire Protection Association (NFPA) 1 Batterymarch Park Quincy, MA 02269-9101 Phone: (617) 770-3000 Fax: (617) 770-0700 www.nfpa.org Sheet Metal and Air Conditioning Contractors’ National Association Inc. (SMACNA) 4201 Lafayette Center Drive Chantilly, VA 22021-1209 Phone: (703) 803-2980 Fax: (703) 803-3732 www.smacna.org In addition, each chapter of each ASHRAE Handbook contains a detailed bibliography. An extensive list of HVAC codes and standards is included in the 2009 ASHRAE Handbook–Fundamentals (ASHRAE 2009).

The Next Step Chapter 9 addresses some air system components, including dampers, air filters, humidifiers, duct heaters, and duct insulation.

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Chapter 8 Codes and Standards

Summary In the private sector, each new construction or renovation project is normally governed by state laws and/or local ordinances that require compliance with specific health, safety, property protection, and energy conservation regulations. These requirements are based on existing design methods, and negotiation may be needed with the authorities to use new design methods. ASHRAE/IES Standard 90.1 has become the legal requirement in the United States. The standard is a consensus document with public review. It is adopted by the American National Standards Institute (ANSI) and undergoes continuous improvement through new addenda. Standard 90.1 covers the building fabric and all permanent energy-using plants and equipment in the building. Section 6, “Heating, Ventilating, and Air Conditioning,” includes two methods of achieving compliance. The prescriptive method follows a specific set of requirements, including minimum requirements for plant efficiency, guidelines for when economizers and heat recovery must be included, and insulation and control strategies to minimize wasting energy. Simple rules are included for some small buildings. With the energy cost budget method, the building is designed to have no greater energy cost than a system designed under the prescriptive approach. ASHRAE Standard 62 sets out requirements for ventilation with outside air and exhaust from polluted spaces, design of the systems to facilitate correct operation through the life of the building, operations and maintenance requirements, and documentation requirements. The NFPA offers three standards applicable to HVAC systems. NFPA 90A applies to systems for air movement in larger buildings, with an emphasis on life safety to reducing the risk of fire and smoke and their effect when they do occur; NFPA 90B applies to smaller buildings; and NFPA 96 covers kitchen exhausts and fire suppression. SMACNA publication, HVAC Duct Construction Standards, covers the design, construction, and installation of galvanized ductwork in detail and other materials more generally. ACGIH publication, Industrial Ventilation, A Manual of Recommended Practice for Design, includes general information on ventilation and numerous examples of industrial ventilation and the removal of contaminants from specific industrial processes.

References and Bibliography ACGIH. 2010. Industrial Ventilation, A Manual of Recommended Practice for Design, 25th Edition. Cincinnati, OH: American Conference of Governmental and Industrial Hygienists ASHRAE. 2007a. ANSI/ASHRAE/IESNA Standard 90.1-2007, Energy Efficient Design of New Buildings Except Low-Rise Residential Buildings. Atlanta:

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American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. ASHRAE. 2007b. ANSI/ASHRAE/IESNA 90.1 User’s Manual. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. ASHRAE. 2010. ANSI/ASHRAE Standard 62.1-2010, Ventilation for Acceptable Indoor Air Quality. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. ASHRAE. 2007. ASHRAE Handbook—HVAC Applications. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. ASHRAE. 2008. ASHRAE Handbook—Systems and Equipment. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. ASHRAE. 2009. ASHRAE Handbook—Fundamentals. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. ASHRAE. 2010. ASHRAE Handbook—Refrigeration. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. NFPA. 2002. NFPA 90A–Installation of Air Conditioning and Ventilating Systems. Quincy, MA: National Fire Protection Association. NFPA. 2004. NFPA 96–Standard Ventilation Control and Fire Protection of Commercial Cooking Operations. Quincy, MA: National Fire Protection Association. NFPA. 2006. NFPA 90B—Installation of Warm Air Heating and Air-Conditioning Systems. Quincy, MA: National Fire Protection Association. SMACNA. 2005. HVAC Duct Construction Standards—Metal and Flexible. Chantilly, VA: Sheet Metal and Air Conditioning Contractor’ National Association Inc.

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Chapter 8 Codes and Standards

Skill Development Exercises for Chapter 8 Complete these questions by writing your answers on the worksheets at the back of this book. 8-1

8-2

Combustibility and toxicity ratings are normally based on tests of _______________. a) new materials

b) old work

c) fibrous materials

d) all of the above

In the private sector, new construction is normally governed by ____________. a) state laws b) local ordinances c) codes

8-3

Zone temperature controls are required for all systems, with special requirements for perimeter heating systems. a) True

8-4

8-5

8-6

a) NFPA 90A

b) NFPA 90B

c) NFPA 96

d) all of the above

SMACNA HVAC Duct Construction Standards covers _______________. a) basic duct construction

b) hangers and supports

c) duct sealing classifications

d) all of the above

ASHRAE/IES Standard 90.1 has a somewhat easier compliance route for many small air-conditioned buildings.

b) False

HVAC systems are one of the most significant energy users in the types of buildings covered by ASHRAE/IES Standard 90.1. a) True

8-10

b) False

HVAC designers must know which code compliance obligations affect their designs. a) True

8-9

b) False

Compliance with ASHRAE/IES Standard 90.1, Section 6, assures a minimum level of HVAC system performance. a) True

8-8

b) False

Which of the following standards applies to structures not exceeding 700 m3 in volume?

a) True 8-7

d) all of the above

b) False

A very efficient HVAC system could have an overall efficiency greater than one. a) True

b) False

Air System Auxiliary Components Study Objectives After completing this chapter, you should understand the function, selection, and sizing of dampers, air filters, humidifiers, duct heaters, and duct insulation.

Instructions Read the material in Chapter 9. Verify the examples presented in the chapter with your own calculations. At the end of the chapter, complete the skill development exercises. Review those sections of the chapter as needed to complete the exercises.

Dampers Types of Dampers Two damper arrangements are used for air-handling system flow control: parallel blade and opposed blade (see Figure 9-1). The linkages shown in the figure are attached to the blades. Moving the linkage upwards on the parallel-blade damper opens the damper, and lowering the linkage closes the damper. Note that the ends of the damper blades have opposed grooves. This is so that the grooves interlock when the damper is closed to improve the seal and provide rigidity to the damper blade. Having the linkage in the airstream increases the damper resistance and, at higher air speeds, can produce air noise. The preferable alternative, although a little more costly, is for the linkage to be external and connected to the damper shafts. The sheet metal blade section shown in Figure 9-1 is made by forming three grooves (one at each edge and a central one around the shaft) and is called a triple-V blade. Blades are also made in an airfoil section, providing a lower resistance to airflow and lower noise generation. The power to drive the linkage comes from an actuator. Optimum control of airflow is obtained with a linear relationship between airflow and the degree to which the damper is open. Conventional wisdom was that parallel-blade dampers were useful for open-closed control and adequate for modulating control if they accounted for the major pressure drop and directional airflow was not a problem. Opposed-blade dampers were considered

206

Chapter 9 Air System Auxiliary Components

Figure 9-1

Parallel- and opposed-blade dampers.

preferable because they were thought to provide better control. The reality, however, is not that simple. The characteristic curves of the two damper types are shown in Figures 9-2 and 9-3. In both figures, the parameter a is the ratio of the pressure drop across the fully open damper at design flow to the total subsystem pressure drop, including fully open control damper pressure drop. These idealized curves are correct in concept but not realized in practice. Recent research, particularly RP-1157, “Flow Resistance and Modulating Characteristics of Control Dampers” (ASHRAE 2004), shows that the performance of dampers is highly dependent on the following: •

Construction: This differs from manufacturer to manufacturer for the same style, triple-V or airfoil. In Figure 9-4, triple-V dampers from two manufacturers have very different performance curves in both arrangements.

•

Relative size of the damper to the duct or plenum and the arrangement. A simple example is the situation where the damper is the same size as the duct, so the airflow is relatively straight into the damper. In contrast, a small damper in a large wall has air coming from all directions into the

Fundamentals of Air System Design SI

Figure 9-2

Installed parallel-blade dampers.

Figure 9-3

Installed opposed-blade dampers.

207

208

Chapter 9 Air System Auxiliary Components

Figure 9-4

Two parallel-blade triple-V dampers from different manufacturers.

damper, creating a different flow characteristic. In Figure 9-4, the performance characteristic is modified for an intake louver and damper to a damper and relief louver. •

Location relative to other components, including changes in duct direction. Figure 9-5 shows an example where the opposed-blade damper characteristic is degraded by being placed inside an inlet louver.

Therefore, actual performance data on their specific dampers must be obtained from the manufacturers, and the situational conditions that influence performance must be considered. The one situation where parallel-blade dampers consistently provide more linear control is in the mixing box, typically mixing outdoor air and return air to provide supply air. The combination of three parallel-blade dampers working in unison provides a more linear control characteristic than using opposedblade dampers. Damper leakage is important, particularly where tight shutoff is required. For example, an outdoor-air damper must close tightly to prevent coils and pipes from freezing. Low-leakage dampers are more costly and require larger operators because of the friction of the seals in the closed position. Therefore, they should be used only when necessary, including in any location where the

Fundamentals of Air System Design SI

Figure 9-5

209

Effect of inlet louver on an opposed-blade damper characteristic.

tight-closing damper reduces energy consumption significantly. Literature from manufacturers expresses leakage rates based on specific pressure differentials across the closed damper.

Damper Operators Damper operators are available using either electricity or compressed air as a power source: •

•

Electric damper operators can be either unidirectional spring return or reversible. This type of operator is available with many options for rotational shaft travel (expressed in degrees of rotation) and timing (expressed in the number of seconds to move through the rotational range). Pneumatic damper operators use air pressure to produce a linear motion, which moves the crank arm to rotate the damper open or closed. Normally open or normally closed refers to the position of the dampers when no air pressure is applied to the operator; this is also known as the failed position. Positive positioners are important for accurate positioning due to the varying pressures on the damper.

Damper Functions Dampers have a wide variety of functions: •

Shutoff dampers are used to regulate the flow of air through a duct. When fully closed, they shut off the flow aside from any leakage that may occur.

210

Chapter 9 Air System Auxiliary Components •

•

Balancing dampers are used to make final adjustments in the flow of air through a duct when the system is first being commissioned. In smaller ducts, balancing dampers are often a flat metal plate as they are just a variable resistance to be set up by the balancing contractor. The balancing damper may be used to adjust the total flow in a single duct system or to adjust the ratio of flows in ducts with multiple ducts. In fire and smoke control, openings for ducts in walls and floors with fire resistance ratings should be protected by fire dampers and ceiling dampers, as required by local codes. Note that fire dampers are manufactured in two styles: with the damper in the duct section and with the damper outside the duct section. Having the damper in the duct section may be required where space is very tight. The significant resistance must be allowed for, particularly in small ducts. Air transfer openings should also be protected. A smoke damper can be used for either traditional smoke management (smoke containment) or for smoke control. In smoke management, a smoke damper inhibits the passage of smoke under the forces of buoyancy, stack effect, and wind. Generally, for smoke containment, smoke dampers should have low leakage characteristics at elevated temperatures. However, smoke dampers are only one of many elements (partitions, floors, doors, etc.) intended to inhibit smoke flow. In smoke management applications, the leakage characteristics of smoke dampers should be selected to be appropriate with the leakage of the other system elements. In a smoke-control system, a smoke damper inhibits the passage of air that may or may not contain smoke. Low leakage characteristics of a damper are not necessary when outdoor air is on the high-pressure side of the damper, as is the case for dampers that shut off supply air from a smoke zone or that shut off exhaust air from a nonsmoke zone. In these cases, moderate leakage of smoke-free air through the damper does not adversely affect the control of smoke movement. Smoke-control supply air systems should be designed so that only smoke-free air is on the high-pressure side of a smoke damper. These dampers should be classified and labeled in accordance with UL-555 Standards (UL 1999, 2006a, 2006b)

Air Filters The purpose of a filter is to remove contaminants from an airstream. Contaminants may be gaseous, such as odors from an adjacent restaurant, or particulates from outside and inside the building. Gaseous filtration is costly to install and maintain. Activated carbon filters may be used for general organic vapor removal. In other situations, specific gaseous compounds can use filters containing chemicals to remove the contaminant. Gaseous filtration is a specialized field and is not covered in this course. The most common application of air filters is particulate removal. The characteristics of airstreams that most affect the performance of an air filter include particle size and shape, mass, concentration, and electrostatic properties. The most important of these is particle size.

Fundamentals of Air System Design SI

211

Particle size may be defined in numerous ways. Particles less than 2.5 µm (microns, or millionths of a metre) in diameter are generally referred to as fine, and those greater than 2.5 µm are considered coarse. From an industrial hygiene perspective, particles that are 5 µm or greater are considered the nonrespirable fraction of dust, which means that they are filtered in the nasal passages before reaching the lungs. Particles less than 5 µm are considered the respirable fraction. Particle size in this discussion refers to aerodynamic particle size (defined as the diameter of a unit-density sphere having the same gravitational settling velocity as the particle in question). Therefore, larger particles with lower densities could be found in the lungs. Also note that fibers differ from particles in that fiber shape, diameter, and density all affect where a fiber will settle in the body (NFPA 2009a). Atmospheric dust is a complex mixture of smokes, mists, fumes, dry granular particles, microorganisms, other biologically produced particles, and natural and synthetic fibers. When suspended in air, this mixture is called an aerosol. A sample of atmospheric dust usually contains soot, smoke, silica, clay, decayed animal and vegetable matter, organic materials in the form of lint and plant fibers, and metallic fragments. It may also contain living organisms, such as mold spores, bacteria, and plant pollens, which may cause disease or allergic responses. Major factors influencing filter design and selection include degree of air cleanliness required, specific particle size range or aerosols that require filtration, and aerosol concentration. Note that filters are used to protect ductwork and equipment as well as occupied spaces. Cooking facilities require a grease filter that both reduces the grease load in the duct and also acts as a fire stop between the cooking surface and the ducting. Clothes dryers require filters to reduce the buildup of fibers in the duct and on any exhaust screen. These situations are often quite specifically mandated in local building and fire codes.

Rating Filters In addition to criteria affecting the degree of air cleanliness, factors such as cost (initial investment and maintenance), space requirements, and airflow resistance have stimulated the development of a wide variety of filters. Accurate comparisons of different filters can be made only from data obtained by standardized test methods. The three main operating characteristics that distinguish the various types of filters are efficiency, airflow resistance, and dust-holding capacity: •

Efficiency measures the filter’s ability to remove particulate matter from an airstream. Average efficiency during the life of the filter is the most meaningful metric for most filters and applications. However, because the efficiency of many dry-type filters increases with dust load, in applications with low dust concentrations, the initial (clean filter) efficiency should be considered for design.

212

Chapter 9 Air System Auxiliary Components •

•

Airflow resistance (or resistance) is the pressure drop across the filter at a given airflow rate. The term pressure drop is used interchangeably with resistance. Dust-holding capacity defines the weight of dust that a filter can hold when it is operated at a specific airflow to some maximum resistance value, or before its performance drops significantly as a result of the collected dust.

In general, four types of tests together with certain variations determine filter efficiency: •

Arrestance. A standardized synthetic dust consisting of various particle sizes is fed into the filter, and the weight fraction of the dust removed is determined. In the old ASHRAE Standard 52.1-1992, Gravimetric and Dust-Spot Procedures for Testing Air-Cleaning Devices Used in General Ventilation for Removing Particulate Matter (ASHRAE 1992) test, this type of efficiency measurement is named synthetic dust weight arrestance to distinguish it from other efficiency values. The synthetic dust used contains fibers and is generally coarser than normally experienced dust, so the test is of limited value.

•

Dust spot efficiency. As defined in ASHRAE Standard 52.1, a standardized atmospheric dust is passed into the filter, and the discoloration effect of the cleaned air on filter paper targets is compared with that of the incoming air. This type of measurement is called atmospheric dust spot efficiency.

•

Particle size removal efficiency test. ANSI/ASHRAE Standard 52.2-2010, Method of Testing General Ventilation Air-Cleaning Devices for Removal Efficiency by Particle Size (ASHRAE 2010) details this method. An optical particle counter measures the number of particles upstream and downstream of the filter. The measurements are made for particles in the range of 0.3 to 10 µm. Based on the results, filters are classified into a scale of 20 categories known as the minimum efficiency reporting value (MERV). The MERV 1 filter is the least efficient, typically collecting long fibers and particles over 10 µm. At the other end of the scale are the MERV 17 to 20 filters used in industrial and medical facilities to remove dusts of 0.3 µm at better than 99.97% efficiency. DOP Penetration Test. This is a U.S. Army specification test based on MIL-STD-282, Military Standard, Filter Units, Protective Clothing, GasMask Components, and Related Products: Performance Test Methods (U.S. DOD 1956) using a chemical, dioctyl phthalate (DOP), that produces a particle cloud of around 0.3 µm. The rating is based on the proportion of particles penetrating through the filter. This test is used to rate very highefficiency filters, MERV 17 to 20.

•

Table 9-1 describes the performance of dry media filters as tested under ASHRAE Standards 52.1 and 52.2. Read through the table’s application guidelines and MERV ratings to understand the range of filter performance

Fundamentals of Air System Design SI Table 9-1

Filter Types and Performance*

Approx. Std. 52.1 Results

Std. 52.2 MERV Rating Dust Spot Arrestance Efficiency

Application Guidelines Typical Controlled Contaminant Larger than 0.3 μm particles: Virus, all combustion smoke, sea salt, radon progeny

20 19 18

n/a n/a n/a

n/a n/a n/a

17

n/a

n/a

16 15 14

n/a >95% 90%–95%

n/a n/a >98%

13

80%–90%

>98%

12 11 10

70%–75% 60%–65% 50%–55%

>95% >95% >95%

9

40%–45%

>90%

8 7 6

30%–35% >90% 25%–30% >90% 3.0–10.0 μm