Fundamentals of Air System Design I-P, 2nd Ed. [PDF]

Zitiervorschau

Your Source for HVAC&R Professional Development

Fundamentals of Air System Design Second Edition (I-P) A Fundamentals of HVAC&R Series Self Directed Learning Course

1791 Tullie Circle NE • Atlanta, GA 30329 • www.ashrae.org

Air System Design I-P.indd 1

8/11/2014 3:48:40 PM

ASHRAE Learning Institute

Fundamentals of Air System Design Second Edition (I-P)

Prepared by

Robert McDowall, P.Eng. Consulting Engineer Winnipeg, Manitoba, Canada

ASHRAE 1791 Tullie Circle NE  Atlanta, GA 30329

© 2008 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.

Other Courses Available in the Fundamentals of HVAC&R Series

Print Fundamentals of: –Psychrometrics –Steam System Design –Heating and Cooling Loads –Heating Systems –Thermodynamics –Water System Design

eLearning Fundamentals of: –HVAC Control Systems –HVAC Systems –Refrigeration –Standard 62.1-2007, Ventilation for Acceptable Indoor Air Quality –Standard 90.1-2004, Energy Efficiency in New Buildings Contact ASHRAE Learning Institute for more information

Fundamentals of Air System Design, Second Edition, I-P A Self-Directed Learning Course ISBN: 978-1-933742-45-8 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 form 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: ADVANCING HVAC&R TO SERVE HUMANITY AND PROMOTE A SUSTAINABLE WORLD. ASHRAE has compiled this publication with care, but ASHRAE has not investigated, and ASHRAE expressly disclaims any duty to investigate, any product, service, process, procedure, design or the like that may be described herein. The appearance of any technical data or editorial material in this publication does not constitute endorsement, warranty or guaranty by ASHRAE of any product, service, process, procedure, design or the like. ASHRAE does not warrant that the information in this publication is free of errors. The entire risk of the use of any information in this publication is assumed by the user. ASHRAE Learning Institute: Ericka L. Reid, Manager of Professional Development Martin Kraft, Managing Editor Vickie Warren, Secretary/Administrative Assistant For course information or to order additional materials, please contact: ASHRAE Learning Institute 1791 Tullie Circle, NE Atlanta, GA 30329 Telephone: 404/636-8400 Fax: 404/321-5478 Web: http://www.ashrae.org/ali Email: [email protected] Comments, criticism and suggestions regarding the subject matter are invited. Any errors or omissions in the data should be brought to the attention of Martin Kraft, Managing Editor or emailed to SDLcorrections @ashrae.org. Updates/errata for this publication will be posted on the ASHRAE Web site at www.ashrae.org/publicationupdates. Errata noted in the list dated 12/8/10 have been corrected.

1791 Tullie Circle, NE • Atlanta, GA 30329-2305 USA • Tel 404.636.8400 • Fax 404.321.5478 • www.ashrae.org

Karen M. Murray

Email: [email protected]

Manager of Professional Development

Dear Student, Welcome to the ASHRAE Learning Institute (ALI) Fundamentals of HVAC&R Series of self-directed or group learning courses. We look forward to working with you to help you achieve maximum results from this course. You may take this course on a self-testing basis (no continuing education credits awarded) or on an ALI-monitored basis with credits (PDHs, CEUs, or LUs) awarded. ALI staff will provide support and you will have access to technical experts who can answer inquiries about the course material. For questions or technical assistance, contact us at 404-636-8400 or [email protected]. Skill Development Exercises at the end of each chapter will test your comprehension of the course material. These exercises allow you to apply the principles you have learned and develop a deeper mastery of the subject matter. If you take this course for credit, please complete the exercises in the workbook and send copies from each chapter to [email protected] (preferred method) or ASHRAE Learning Institute, 1791 Tullie Circle, Atlanta, GA 30329-2305. Please include your student ID number with each set of exercises submitted. Your student ID is composed of the last five digits of your Social Security number or other unique five-digit number you create. We will return answer sheets to the Skill Development Exercises and maintain records of your progress. Please keep copies of your completed exercises for your own records. When you finish all exercises, please submit the course evaluation, which is located at the back of your course book. Once we receive all chapter exercises and the evaluation, we will send you a Certificate of Completion indicating 35 PDHs/LUs or 3.5 CEUs of continuing education credit. Please note: The ALI does not award partial credit for self-directed learning courses. 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

Table of Contents Chapter 1: Fundamentals of Air Flow . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1-1 1.1 1.2 1.3 1.4

Static and Dynamic Compressible Fluid (Air) Laws Friction Effects The Friction Chart Density and Altitude Effects

Chapter 2: Air Distribution System Components . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2-1 2.1 2.2 2.3 2.4 2.5 2.6

Air Distribution System Overview Air Handling Units Ducts Controls Air Distribution Devices Sound Absorbers

Chapter 3: Human Comfort and Air Distribution. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3-1 3.1 Principles of Human Comfort 3.2 Principles of Space Air Distribution 3.3 Types of Air Distribution Devices

Chapter 4: Relationship of Air Systems to Load and Occupancy Demands . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 4-1 4.1 4.2 4.3 4.4 4.5

Operating System Selection Criteria System Types by Heating/Cooling Equipment Type System Type by Duct Configuration Economizers Outdoor Air Intake

Chapter 5: Exhaust and Ventilation Systems . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5-1 5.1 Design Considerations 5.2 Ventilation and Exhaust Systems 5.3 Energy Recovery

Chapter 6: Air Movers and Fan Technology . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6-1 6.1 6.2 6.3 6.4 6.5 6.6

Fan Principles Fan Drives Fan Selection Fan Installation Design Fan Controls Effect of Variable Resistance Devices

Chapter 7: Duct System Design . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 7-1 7.1 7.2 7.3 7.4 7.5

Duct System Design Overview Duct Materials Duct Construction Duct Design and Sizing Sample Systems

Chapter 8: Codes and Standards . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8-1 8.1 8.2 8.3 8.4 8.5

Building Code Requirements ASHRAE Standard 90.1-2007 ASHRAE Standard 62.1-2007 Other Codes and Standards Sources of Information

Table of Contents Chapter 9: Air System Auxiliary Components . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9-1 9.1 9.2 9.3 9.4 9.5

Dampers Air Filters Humidifiers Duct Heaters Duct Insulation

Chapter 10: Sound and Vibration in Air Systems . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 10-1 10.1 10.2 10.3 10.4

Fundamentals of Sound Sound and Vibration Sources Sound Attenuation Vibration Control

Chapter 11: Air System Start-Up and Diagnostics . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 11-1 11.1 11.2 11.3 11.4 11.5

Introduction Design Considerations Air Volumetric Measurement Methods Balancing Procedures for Air Distribution Systems Noise and Vibration Diagnostics

Chapter 12: An Actual Duct Design Problem . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 12-1 12.1 12.2 12.3 12.4

Introduction Duct Design Procedure The Building and System Working Through The Problem

Skill Development Exercises Answer and Work Sheets. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Answer Sheet 1 Evaluation Form . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Final Page

Chapter 1

Fundamentals of Air Flow Contents of Chapter 1 • • • • • • •

1.1 Static and Dynamic Compressible Fluid (Air) Laws 1.2 Friction Effects 1.3 The Friction Chart 1.4 Density and Altitude Effects Summary Bibliography Skill Development Exercises for Chapter 1

Fundamentals of Air Flow

Instructions Read the material of Chapter 1. At the end of the chapter, complete the skill development exercises without consulting the text.

Study Objectives of Chapter 1 After completing this chapter, you should be able to: • 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. • Explain duct frictional losses.

1.1 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 the applied system design concepts that follow in later chapters. The concepts are presented in the context of HVAC applications, defining terms 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

THE 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 and landing on the moon. However, their weight changed dramatically.

1–2

Fundamentals of Air System Design

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 32.2 ft/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 used 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 in the American HVAC industry, the value of gc is 32.2 lbmft/lbf sec2. The reason for this choice is that the “weight” of a poundmass, lbm, is numerically equal to a pound-force, lbf , at sea level. Two important rules result from this: • Mass and weight are inherently different but related, and pound-mass (lbm) is completely different from pound-force (lbf ). •

The symbol gc is a units conversion factor.

As a thermodynamic property, density is the ratio of mass to volume. It has the unit of lbm/ft3 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:  lb m  ft    ------ g  ---- lb f  ft 3   s 2 ------------------------------- = -----3 ft  lb m  ft  g c  -------  ---2-  lb f   s  Also, numerically g/gc = 1, because both have the value 32.2. Specific volume is the reciprocal of density, and is defined as the volume of a unit mass of material. It is expressed in cubic feet per pound-mass: v = ft3/lbm.

1–3

Fundamentals of Air Flow

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: • The atmospheric air pressure on the surface of the water; and • 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 = hA. The force required to hold the fluid column 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 = pa A + W = pa A + h A Therefore, p = pa + h. The pressure difference between A and B is:  p = yh

(1-1)

Note that the density is the weight density,  = (g/gc ).

Figure 1-1

1–4

Fluid Static System

Fundamentals of Air System Design

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.

Figure 1-2

Continuity Equation Example

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 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, . Therefore, the mass flow rate is given by AV. 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:  AV  1 =  AV  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:  AV  1 =  AV  2

(1-2)

1–5

Fundamentals of Air Flow

In the HVAC industry, flow rate almost invariably means volume flow rate, not mass flow rate. Also, the most common units are cubic feet per minute (cfm), where the velocity has been specified in feet per minute (fpm), and the areas are in square feet (ft2). The practice is so common that practitioners use cfm as a word. For example, “How many cfm do you supply to the room?” Air behaves as a perfect gas and the change in density is inversely proportional to the absolute temperature. Absolute temperature is the temperature above absolute zero, which is 460°F. To convert from our normal °F to absolute °F, we add 460°F. Thus if outside air is heated from 35°F to 75°F as it comes in through the air-conditioning system, the density, , will change from y 35 to 35 + 460 y 35  --------------------- = y 35  0.925 75 + 460 a 7.5% decrease in density. Similarly, on a hot day, cooling 100°F outside air down to 55°F will increase the density by 100 + 460 ------------------------ = 1.09 55 + 460 or 9%. The difference is even more significant in a cold climate. For example, suppose it is January and the outside temperature is 30°F. The outside air is brought in over a heating coil, and supplied at 75°F. The drop in density as the air is heated is from y –30 to – 30 + 460 y –30  ------------------------- = y –30  0.804 75 + 460 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 1,000 cfm at 30°F when heated to 75°F becomes 1,200 cfm at 75°F. 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 air streams. Suppose we have a situation where 15,000 cfm of air at 75°F from a space is being mixed with 5,000 cfm of out-

1–6

Fundamentals of Air System Design

side air at 40°F. 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 cfm is equivalent to mass and use Equation 1-3:  cfm 1  T 1  +  cfm 2  T 2  =   cfm 1 + cfm 2   T 3 

(1-3)

For example, based on Equation 1-3, the resulting volume will be: 15 000 + 5 000 = 20 000cfm  15 000  75  +  5 000  40  =   15  000 + 5 000   t m  tm = 66.25°F, which is very close to the correct answer of 65.9°F. For wide temperature differences, the inaccuracy can be almost eliminated by adjusting the incoming cfm values to be the cfm at the initially calculated mixed temperature.

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. However, 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. Each author seems 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: ftlbf . Other units of energy are the British thermal unit (Btu), kilowatt hour (kWh) and horsepower hour (hph). It is common to write energy conversion in terms of a unit of mass flowing; for example, ftlbf /lbm . Whatever the units used, it is imperative to have all values 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. For example, an experimentally derived relationship is that 1 Btu is equivalent to 778 ftlbf . 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, done by the fluid to an external machine such as an electric generator. Fluid friction,

1–7

Fundamentals of Air Flow

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.

1–8

•

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 pound 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 totally different concepts. Internal energy is associated with molecular motion, molecular bonding and other forms of molecular activity such as spinning or rotation of the molecules. Internal energy can have units of ftlbf /lbm or Btu/lbm . In an ideal gas and a liquid, u is directly related to temperature. For example, a 1 Btu increase in internal energy is represented by a 1°F rise in temperature for a pound 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 lbm book is elevated 1 ft above a desk, work in the amount of 1 ftlbf 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 have been 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 55 mph has kinetic energy, as does a baseball thrown at 90 mph. The kinetic energy is derived from a steady force applied through a distance required to accelerate the body from rest to a velocity, V: KE = mV 2/2

Fundamentals of Air System Design

Here is an excellent example where units need to be converted. As it stands, KE has units of lbmft2/sec2. The units conversion factor can be used to convert to conventional energy units by writing: 2

ft -------2 2 mV sec KE = ---------- = lb m --------------------= ft  lb f 2g c lb m ft -------  --------lb f sec 2 The Energy Equation is simply a balance of these forms of energy. 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: 2 g V W m – W f + Q + m u + pv + ---- z +------g c 2g c

in

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

(1-4) out

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 a 10 hp motor driving a fan. The Energy Equation can be turned into a Rate Equation by considering: · • Heat as a rate, Btu/h, kW or hp; Q · • Work as a rate, Btu/h, kW or hp; W · • Mass as a mass flow rate, or lbm /h; m A dot over the symbol is commonly used to indicate a rate. 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

2

in

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

(1-5a) out

1–9

Fundamentals of Air Flow

· 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

(1-5b) out

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

Figure 1-3

Energy Equation Applied to a Flow System

STATIC PRESSURE, VELOCITY PRESSURE AND TOTAL PRESSURE Now let's discuss a run of ductwork where the following conditions exist: • No machines, so all work terms are zero; • No heat transfer because the duct air is the same temperature as the room air; • No significant changes in elevation, so z is constant; and • The internal energy, u, is essentially constant. In this case, we have the simpler form of the Energy Equation: 2

2

Vpv + -----2g c

1–10

in

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

(1-6a) out

Fundamentals of Air System Design

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/, multiply through by gc /g, and replace (g/gc) by the “weight” density, . Equation 1-6b then becomes: 2

p V -- + ------ 2g

duct

p = -

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: p V2 -- + ----- 2g

Figure 1-4

= constant duct

(1-6d)

Static and Total Pressure

1–11

Fundamentals of Air Flow

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 feet. These are pressure equivalents to the weight of a column of the fluid on a unit area. Thus the units are feet of air, or feet 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

(1-6e)

Examining the units of the velocity pressure term, we find that:    lb   2 2 lb f V ft  sec -   ------ =  ------f   -----------------= ---- 2  3  2 2g  ft  ft  ft  sec   2

The unit lbf /ft2 can be converted to pounds per square inch (psi), or inches water gauge (in. wg.). 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 the following: pv V = 1096.7 ------- air

V p v = -----------4005

1–12

(1-7)

2

(1-8)

Fundamentals of Air System Design

where the numbers 1096.7 and 4005 contain the conversion factors appropriate for pv in inches of water; density, , in lbf /ft3 (standard air density is 0.075 lbm /ft3); and velocity, V, in fpm. Standard Air, for the HVAC industry, is dry air at 70°F and 14.969 psia with a mass density of 0.075 lbm/ft3. Sea level pressure is 14.969 psia, so Standard Air can be considered as typical dry air at sea level. For this reason, most airflow tables and charts are based on Standard Air. Note that defining an air flow in terms of Standard Air also defines the weight and mass flow. Thus, 1,000 ft3/min of Standard Air is also 1,000  0.075 = 75 lbm/min. As elevation increases, air density decreases and above 3,000 feet, density corrections should be considered. Because most projects are located at altitudes from sea level to 3,000 feet, most designs can use Standard Air without correction. Air expands as it is heated and the density drops. For many air-conditioning systems, this can be ignored, but be careful. In a cold climate, outside air at –30°F has a density about 20% lower at 75°F. 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 the Psychrometric Chart. The main axes on the chart are temperature along the bottom x-axis and moisture weight compared to dry air weight, lb/lb, 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 of Figure 1-5. Shown are: • Vertical temperature lines, °F • Horizontal moisture content (humidity ratio) lines, lb of moisture/lb of dry air •

•

Sloping down left to right specific volume lines, ft3/lb. For example, air at 90°F and 25% relative humidity has a specific volume of 14.0 ft3/lb. At this specific volume, 1 lb of air occupies 14 ft3. Curved relative humidity 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 capacity to hold at that temperature, we call it 50% relative humidity, or 50% rh. The chart shows the 25%, 50%, 75% and 100% relative humidity lines. The saturation line is 100% and 0% is the horizontal line along the x-axis. Note that on the chart, the relative humidity lines are not linearly related. Thus at a particular temperature, the 50% relative humidity curve is not at half the height of the saturation, 100% humidity, line.

1–13

Fundamentals of Air Flow

Figure 1-5

Psychrometric Chart

Psychrometric charts are based on Standard Air, and humidity ratio may be labeled lb, lbm or lbw. Because lbm and lbw are numerically the same at the same pressure, all the charts are graphically the same. For most above-ground terrestrial systems, the lbm and lbw issue can be ignored. But be careful with units when dealing with substantial pressure changes as occur in mines, submarines, planes and space vehicles. We will return to the Psychrometric Chart in future chapters.

1–14

Fundamentals of Air System Design

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 will be one velocity at A1 and another at A2. This process can be analyzed using the Continuity Equation. The Continuity Equation says that for a given mass flow, and by the law of conservation of matter or mass, whatever air we put in on the left side has to come out on the right side because we can neither destroy nor create air in the duct between the two points. The Continuity Equation says that the cfm in is equal to the cfm out, ignoring any kind of compressibility or temperature change. In other words, the quantity of air in (cfm1 ) is equal to the quantity of air out (cfm2 ), giving cfm1 = cfm2. Because cfm = 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:  A1  V 2 = V 1  ------   A2 

(1-9)

So as the cross-sectional area is reduced, the velocity is increased as predicted by the Continuity Equation.

Figure 1-6

Conversion of Static Pressure to Velocity Pressure

1–15

Fundamentals of Air Flow

Let’s return to the Energy Equation and the relationship that the total pressure is equal to the velocity pressure plus the static pressure (pt = pv + ps ). If the velocity increased, the velocity pressure had to increase, because velocity pressure is pv = V 2/2g. 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 did 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 flow work must decrease in direct proportion. Consequently, 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 (V 2/2g), pV2 is four times pV1 and the 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 A1. Figure 1-6 shows an accelerator, where air velocity is increased by making the duct area smaller. Suppose Figure 1-6 is reversed, as in Figure 1-7 which shows a decelerator. The air comes in at a higher velocity than it goes out. 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 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 outlets. After each outlet, the velocity is reduced and the duct size is reduced so that the static pressure at the next outlet will be the same.

Figure 1-7

1–16

A Decelerator

Fundamentals of Air System Design

1.2 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 p f = f ----  -----D 2g

in lbf /ft2

(1-10)

Or in terms of head as feet of fluid flowing, the Darcy-Weisbach Equation can be written as: 2 p f L V -------- = f ---- ------D 2g 

in feet of fluid flowing

(1-11)

For Standard Air, the Darcy-Weisbach Equation can also be written as: L V p f = f ---- -----------D 4005

2

in in. wg

(1-12)

This is a purely empirical formula which states that the frictional pressure drop is proportional to length, L, inversely proportional to diameter, D, and proportional to velocity pressure or velocity head. One would hope that the proportionality constant, f (a dimensionless constant called the friction factor), would be truly constant, and that turns out to be partially true. When the flow is fast, f 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 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 flowing 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: Re = VD -----------

(1-13)

where μ = absolute viscosity, lbm/ftsec; V = velocity, fps; D = diameter, ft; and  = density, lbm /ft3 .

1–17

Fundamentals of Air Flow

The Reynolds number is the ratio of the momentum of the flow (V) 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 air flow 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 manifested in the behavior of the friction factor as shown on the Moody Chart (Figure 1-8). This chart shows the friction factor as a function of the Reynolds number. Note that both axes have logarithmic scales.

Figure 1-8

1–18

Moody Chart

Fundamentals of Air System Design

Several interesting features are present on the Moody Chart: •

The laminar flow region is shown for Reynolds numbers smaller than about 2,000. The dashed extension of the solid line indicates that, under some circumstances, the relationship can be extended up to 4,000. 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: f = 64  Re

•

If this value is substituted into Equation 1-12: L V p f = 64v ------2 -------------2- in in.wg D 4005

•

•

(1-14)

(1-15)

where , the kinematic viscosity, is / with units of ft2/sec. Note that the frictional pressure drop varies with the first power of velocity. 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, . The relative roughness values are shown. For example, the roughness of commercial steel pipe is 0.00015 in. Relative to a 4 in. pipe, /D is about 0.00045. 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.

Test work performed by ASHRAE and its predecessor organization ASHVE (American Society of Heating and Ventilating Engineers) indicated prior to Moody’s work1 that  for galvanized sheet metal ductwork was about 0.0005 ft. This is based on transverse joints spaced at 30 in. intervals. When joints are spaced at 46 in. intervals, the value is reduced to 0.0003 ft.

1–19

Fundamentals of Air Flow

THE SYSTEM CONSTANT FORM OF THE DARCY-WEISBACH EQUATION HVAC air and piping systems usually use a simplified form of the Darcy-Weisbach 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). So we lump all of the constants together and write two forms that are essentially the same – the second being an inversion of the first – with two constants, K and Cs : p f = K  cfm 

2

(1-16)

or cfm = C s p f

(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 cfm and pf is required to determine K or Cs .

Figure 1-9

1–20

Typical System Curve

Fundamentals of Air System Design

Consider a complex air handling system where we want to move 10,000 cfm through the system. The pressure drop in the system is calculated to be 4 in. wg. The system constant form of the Darcy-Weisbach Equation can be used to find the system constant: 10 000 C s = ------------------ = 5 000 4

Similarly, we find that K = 4  10-8. Values of pf and cfm can be plotted on a graph. Other values can be determined by using: p f = K  cfm 

2

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

1.3 The Friction Chart In 1945, D.K. Wright published “A New Friction Chart For Round Ductwork” in the ASHVE Transactions.2 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 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  value for galvanized sheet metal ductwork of 0.0005 ft . Since that time, ASHRAE and the Sheet Metal and Air Conditioning National Contractors' Association (SMACNA) have 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. The friction factor chart (see Figure 1-10) was revised based on standard galvanized sheet metal ductwork with an absolute  roughness of 0.0003 ft instead of 0.0005 ft. Other factors, including the shape of the duct, the roughness of the material of construction, and fittings used must be taken into consideration. These will be discussed later in Chapter 7.

1–21

Fundamentals of Air Flow

Figure 1-10

1–22

Friction Factor Chart

Fundamentals of Air System Design

1.4 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 the 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 Engineering Guide for Altitude Effects.3 Table 1-1 Typical Altitude Correction Factors4 Item Compressors Condensers, air-cooled Condensers, evaporative Chillers Induction room terminals (chilled water) Fan-coil units Total capacity (*SHF = .40-.95) Sensible capacity (SHF = .40-.95) Total capacity (SHF = .95-1.00) Packaged air-conditioning units, air-cooled condenser Total capacity (*SHF = .40-.95) Sensible capacity (SHF = .40-.95) Total capacity (SHF = .95-1.00)

Altitude (ft above sea level) 2500 1.00 0.95 1.00 1.00 0.93

5000 1.00 0.90 1.01 1.00 0.86

7000 1.00 0.85 1.02 1.00 0.80

10,000 1.00 0.80 1.03 1.00 0.74

0.97 0.92 0.93

0.95 0.85 0.86

0.93 0.78 0.79

0.91 0.71 0.73

0.98 0.92 0.96

0.96 0.85 0.82

0.94 0.78 0.88

0.92 0.71 0.84

*SHF = Sensible Heat Factor

The Next Step This chapter has introduced the theory needed and included some discussion about air flowing in ducts. Chapter 2 will introduce the other common components of air systems that condition air and deliver it to the occupied space. Included will be their function and main operating characteristics. More detailed issues of choosing components and their detailed operation will be explained in later chapters.

1–23

Fundamentals of Air Flow

Summary The chapter began by explaining the difference between mass and weight. Mass is a property of matter that is invariant with location, but weight changes depending on the local gravitation. Conveniently, for most building designers, gravity is constant, with lbm and lbw being numerically the same. Hydrostatic pressure, commonly referred to as static pressure, is the pressure exerted by a fluid at rest. The pressure is the same in all directions at any point. In a duct with the air flowing under pressure, the static pressure around the duct will be 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. The useful, but not absolutely correct, formula for calculating the result of mixing airstreams was introduced: (cfm1  T1) + (cfm2  T2) = [(cfm1 + cfm2)  T3] 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. Bernoulli’s Equation states that in a system without energy losses or gains, the sum of static pressure and velocity pressure are constant: 2 p --- + V ------ 2g

1–24

= constant duct

Fundamentals of Air System Design

The valuable concept in Bernoulli’s Equation is that if the velocity is reduced due to a wider duct, the drop in velocity pressure (V 2 reduces) is exactly matched by an increase in static pressure, friction ignored. Velocity pressure in inches water gage for Standard Air equals: 2

V----------4005

in. wg

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, Moody Chart and Darcy-Weisbach Equation. The critical point to remember is that in a fixed system, the pressure drop through the system will be about pro2 portional to the square of the flow: p f = K  cfm  Thus, doubling the flow will create 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 the ASHRAE chart shown in Figure 1-9 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.

Bibliography 1. Moody, L. 1944. "Friction factors for pipe flow." ASME Transactions. New York, NY: American Society of Mechanical Engineers. 2. Wright, D. 1945. "A new friction chart for round ductwork." ASHVE Transactions. Atlanta, GA: ASHRAE. 3. Carrier Corp. Engineering Guide for Altitude Effects. ASHRAE produces four Handbooks: Fundamentals; HVAC Systems and Equipment; HVAC Applications; and Refrigeration. Each Handbook is updated and reissued on a fouryear cycle. Handbook sections that relate to the material are listed in the bibliography for each chapter. For this chapter, see HandbookFundamentals for general theory fluid flow and duct sizing.

1–25

Fundamentals of Air Flow

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

1–26

1-1.

In the figure below, Area A1 = 2 ft2, Area A2 = 1.25 ft2, and velocity V1 = 1,000 fpm. Calculate V2 (fpm). a) 1,600 fpm b) 625 fpm c) 1,406 fpm d) 2,569 fpm

1-2.

The total pressure at a certain point in a system is determined to be 5 in. wg, and the static pressure at that point is determined to be 2 in. wg. What is the velocity pressure (in. wg) at that point? a) 21 in. wg b) 7 in. wg c) 3 in. wg d) 2 in. wg

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 6 in. pressure drop through the system at a flow of 8,000 cfm. What is the system constant? a) 1.5 b) 1,333 c) 3,265 d) 4,000

1-5.

The product of fluid pressure and specific volume is ______? a) Internal energy b) Reynolds number c) Kinetic energy d) Flow work e) Viscosity

1-6.

What does a water manometer measure? a) Velocity b) Pressure c) Temperature d) All of the above

Fundamentals of Air System Design

1-7.

Fan pressures are typically indicated in what units? a) in. wg b) in. Hg c) cfm d) None of the above

1-8.

If the cross-sectional area of a duct decreases in size, the velocity of an airstream passing through the duct will increase. a) True b) Cannot tell c) False

1-9.

Air is passing 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 0.85 in. wg to 0.60 in. wg along the suspect section of duct. Approximately what percentage of air is being lost through the leak? a) 50% b) 31% c) 16% d) 11%

1-10.

In an air-conditioning system, 3000 cfm of outside air at 34°F is drawn in over a heater and delivered into the building at 74°F. What volume of air is delivered? a) 6,529 cfm b) 1,378 cfm c) 2,775 cfm d) 3,243 cfm

1-11.

In an air-conditioning system, 30,000 cfm of return air at 78°F is mixing with 4,600 cfm of outside air at 95°F. What is the approximate resulting volume and temperature? a) 36,400 cfm, 79.2°F b) 36,400 cfm, 80.3°F c) 34,600 cfm, 80.3°F d) 34,600 cfm, 79.2°F

1–27

Fundamentals of Air System Design

Chapter 2

Air Distribution System Components Contents of Chapter 2 • • • • • • • • •

2.1 Air Distribution System Overview 2.2 Air Handling Units 2.3 Ducts 2.4 Controls 2.5 Air Distribution Devices 2.6 Sound Absorbers Summary Bibliography Skill Development Exercises for Chapter 2

2–1

Air Distribution System Components

Instructions Read the material of Chapter 2. At the end of the chapter, complete the skill development exercises without consulting the text.

Study Objectives of Chapter 2 The goal of this chapter is to give you an overview of air distribution system components and their schematic symbols, to 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. • Identify the schematic symbols of air distribution system components.

2.1 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 will provide 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.

2.2 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 that moves air. It may also be used to clean, heat, cool, humidify, dehumidify and mix the air. Figure 2-1 shows a large typical central air-handling unit.Types of air-handling units include: • • • • •

2–2

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.

Fundamentals of Air System Design

Figure 2-1

Air Handling Unit

FANS, MOTORS AND FAN DRIVES A fan is an air pump that creates a pressure difference and causes air flow. The fan impeller does work on the air, imparting to it both static and kinetic energy, varying in proportion depending on the fan type. Fans are generally classified as centrifugal fans or axial flow fans according to the direction of air flow 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.

2–3

Air Distribution System Components

2–4

Figure 2-2

Centrifugal Fan Configuration and Symbol

Figure 2-3

Axial Fan Configuration and Symbol

Fundamentals of Air System Design

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 adjustments 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.

COILS A coil is a cooling or heating element made of pipe or tube. Coils are usually finned, and are found in a number of shapes (serpentine, helical, etc.). Some coils commonly encountered in air systems include: • 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 inlet end 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 will 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: • A disposable filter has elements that are discarded after use. Efficiencies range from very low to relatively high depending on the construction.

2–5

Air Distribution System Components

•

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 air flow 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 a 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 micrometers (microns) in size. A particular type, the High Efficiency Particulate Air (HEPA) filter, is tested and rated to an ASTM standard to remove at least 99.97% of particles 0.3 microns in diameter.

•

An electrostatic filter (active) has the airstream passing through a high-voltage ionizing field to impart a positive electrical charge to the particles, which are then collected on electrically negative plates.

•

An electrostatic filter (static) consists of plastic media that generate an electrostatic attraction by the air flow over the plastic.

•

A carbon filter (adsorber) uses a mass of granulated activated carbon to chemically 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.

2.3 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.

2–6

Fundamentals of Air System Design

Figure 2-4

Duct Symbols

2–7

Air Distribution System Components

2.4 Controls A control is a device that regulates a variable such as temperature, velocity or pressure. Controls may be manual or automatic. For example, an air handling unit might have a manually set minimum flow of outside air and an automatic control to increase the outside air as the building becomes more densely occupied. If automatic, the implication is that the control is responsive to a measured 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, permit the passage of control air when the temperature of the controlled air is within the limits at which the thermostat is set, and other temperature control purposes. Schematic symbols for these controls are shown in Figure 2-5.

Figure 2-5

2–8

Controls Symbols

Fundamentals of Air System Design

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 fully discussed in the next chapter .

Figure 2-6

Air Distribution Devices

2–9

Air Distribution System Components

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.

Figure 2-7

Louvers

2.6 Sound Absorbers A proper acoustical environment is as important for human comfort as other environmental factors controlled by air-conditioning systems. The objective of sound control is to achieve an appropriate sound level for all activities 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 soundabsorbing units, acoustical linings and other means. The schematic symbol for a sound absorber in ductwork is shown in Figure 2-8.

2–10

Figure 2-8

Sound Absorber

Fundamentals of Air System Design

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. This is the subject of Chapter 3.

Summary This chapter has 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 the fans provide the necessary static pressure and velocity to the air flow. A fan is an air pump. The fan creates a pressure difference (static pressure) and causes air flow (kinetic energy). The first main fan type is the centrifugal fan where the air enters the center of the drum-shaped impeller 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 impeller directly on an extended motor shaft. This arrangement 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 impeller 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, thereby dehumidifying 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. This will be 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.

2–11

Air Distribution System Components

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 goes into more detail.

Bibliography ASHRAE Handbooks: HandbookFundamentals contains information on air contaminants and odors HandbookHVAC Systems and Equipment contains detailed information on HVAC components

2–12

Fundamentals of Air System Design

Skill Development Exercises for Chapter 2 Complete these questions by marking your answers on the worksheets at the back of this book. 2-1.

This is the symbol for:

a) Centrifugal fan b) Axial fan c) Diffuser d) None of the above

2-2.

This ductwork is _____________, and the dimension of the side shown is _________.

a) Dropping, 20 b) Dropping, 12 c) Rising, 12 d) None of the above

2-3.

This is the symbol for a flexible duct:

a) True b) False c) Cannot be determined from the information given.

2–13

Air Distribution System Components

2-4.

This symbol shows

a) A blanked-off duct, with a top dimension of 12 b) A return air duct, with a side dimension of 18 c) A supply air duct, with a side dimension of 18 d) None of the above

2-5.

The shown dimension of this duct is 24:

a) True b) False c) Cannot be determined from the information given.

2-6.

A filter that uses a liquid as an adhesive is a: a) Carbon filter b) Electrostatic filter c) Viscous filter d) All of the above e) None of the above

2-7.

An air handling unit may be used to: a) Move air b) Mix air c) Heat air d) All of the above e) None of the above

2–14

Fundamentals of Air System Design

2-8.

This is the symbol for:

a) Manually operated damper b) Electrically controlled damper c) Manual damper d) All of the above e) None of the above

2-9.

This is the symbol for:

a) Pneumatically operated damper b) Inline psychrometric observation device c) Fire damper d) All of the above e) None of the above

2–15

Air Distribution System Components

2-10.

This is the symbol for:

a) Temperature relay b) Test station c) Remote bulb thermostat d) All of the above e) None of the above

2–16

Fundamentals of Air System Design

Chapter 3

Human Comfort and Air Distribution Contents of Chapter 3 • • • • • •

3.1 Principles of Human Comfort 3.2 Principles of Space Air Distribution 3.3 Types of Air Distribution Devices Summary Bibliography Skill Development Exercises for Chapter 3

3–1

Human Comfort and Air Distribution

Instructions Read the material of Chapter 3. At the end of the chapter, complete the skill development exercises.

Study Objectives of Chapter 3 After completing this chapter, you should be able to list and explain: • • • •

The main issues of thermal comfort; The principles of air distribution as they relate to human comfort; The principles of space air distribution; and The functions of the different types of air distribution devices.

3.1 Principles of Human Comfort Human comfort depends on a variety of factors. There are factors relating to the space, the individual and the individual’s current activity level. The space temperature, humidity, air quality and acoustics are controlled, or influenced, by the air conditioning. However, other space factors such as lighting are not controlled by the 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 levels of the individuals and their clothing will influence their comfort. We will start with thermal comfort and then go 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 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 total energy plant that operates the same way as any other 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.

3–2

Fundamentals of Air System Design

As shown in Figure 3-1, the body uses three 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. The 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.

Figure 3-1

Body Heat Mechanisms

3–3

Human Comfort and Air Distribution

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: • • • • • • •

Dry-bulb temperature, °F Relative humidity, rh Thermal radiation Air movement, fpm 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/full length skirt, underwear, shirt and jacket. Typical light summer clothing, including shorts/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 18.4 Btu/h/ft2. A resting adult typically produces 1 met; light office work produces 1 to 1.3 met; and walking at 2 mph produces 2 met. Figure 3-2, which is adapted from ASHRAE Standard 55-2004, Thermal Environmental Conditions for Human Occupancy, 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 40 fpm

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, the cooler the space needs to be to provide the same thermal comfort for the occupants.

3–4

Fundamentals of Air System Design

Figure 3-2

Acceptable Range of Temperature and Humidity

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 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 permit mold growth can make the building uninhabitable until very expensive remedial work has been completed. The comfort chart indicates that relative humidity 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 lbmoisture/lbdry air is far higher than acceptable in a building in a moist climate. Because mold can grow in relative humidities above 60%, it is prudent to maintain buildings in hot humid climates with a humidity ratio significantly lower, at about 0.010

3–5

Human Comfort and Air Distribution

lbmoisture/lbdry air. In addition, relative humidity 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: • • •

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 in the air-conditioning system.

Where winter humidification is provided for comfort, a minimum relative humidity of 20% is generally acceptable in cold climates. If a higher humidity level 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 78°F dry-bulb temperature, enter the annual wet-bulb degree-hours above 66°F in the occupied space at the bottom left. Next, intersect this value with the indoor relative humidity chosen, and then draw a vertical line to the weekly hours of cooling system operation and read the energy used (in million Btu per 1,000 cfm) on the upper-left scale.

Figure 3-3

3–6

Summer Dehumidification Energy Requirements

Repeating this procedure for a different value of relative humidity yields the energy savings obtainable by raising relative humidity. However, be cautious about choosing excessively high humidities. Computer rooms, particularly computer printers and drafting rooms, are two applications for which relative humidity in excess of 50% to 55% is undesirable or unacceptable due to effects of moisture on the paper products.

Fundamentals of Air System Design

INDOOR AIR QUALITY Air contaminants. Indoor air contaminants can be solid or liquid particles, gases or vapors. Some can be irritants or odiferous, thus affecting occupant comfort. The same 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. 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 meter (μg/m3). They include all particle sizes or the total suspended particulate (TSP) 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 (RSP) 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) • 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) • 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. In addition, 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

3–7

Human Comfort and Air Distribution

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 relative humidity 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. ASHRAE Standard 62.1-2007, Ventilation for Acceptable Indoor Air Quality tabulates enforceable and guideline maximum concentration levels of common indoor contaminants. It also includes the US National Primary Ambient-Air Quality Standards for Outdoor Air 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. Standard 62.1-2007 provides designers with a means of determining ventilation rates needed 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 5 cfm per sedentary occupant plus a minimum of 0.06 cfm/ft2 to deal with pollutants from the space. These minimums are increased for more active occupants, an example being 20 cfm/person in an exercise room. Similarly, the space ventilation rate is increased where there are anticipated contaminants; an example being 0.12 cfm/ft2 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 them 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, inactivated 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. Standard 62.1-2007 specifies the exhaust rate for many spaces

3–8

Fundamentals of Air System Design

in terms of cfm/ft2. Examples are 0.5 cfm/ft2 for barber shops, arenas, locker rooms and copy/print rooms. Twice the exhaust, 1.0 cfm/ft2, is required for darkrooms, janitor, trash, recycling and science class rooms due to the higher anticipated pollution to be removed. Air movement effect. 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 experience of many commercial building operators has shown air motion is a significant benefit to comfort in mechanically ventilated spaces. The standard further prescribes a maximum rate of air movement of 40 fpm to avoid drafts. Higher air speeds (up to 160 fpm) 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 air flow 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 air flow of 0.6 to 0.8 cfm/ft2 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 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 air flow, 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 40 fpm 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 ceilinged room) is the division of air into a series of temperature layers. If conditioned air is introduced at about the 10 ft level or below, the space close to the floor is conditioned. The cooling requirements of the elements above the 10 ft level may be reduced.

3–9

Human Comfort and Air Distribution

3.2 Principles of Space Air Distribution Room air distribution systems include mixing, under-floor, 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 creates 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 (see next section). Mixing air outlets have been 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 can handle relatively large quantities of air at large temperature differentials when cooling.

Figure 3-4

3–10

Air Motion Characteristics of Group A Outlets

Fundamentals of Air System Design

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 non-spreading jet (see Figure 3-5). This figure 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 15 ft to 20 ft between the drop region and the exposed wall is a conservative design value.

Figure 3-5

Air Motion Characteristics of Group B Outlets

A comparison of Figures 3-4 and 3-5 for heating shows that the stagnant region is smaller for Group B than 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 would occur with Group A outlets. Cooling effectiveness of Group B is inferior to Group A for the same reasons.

3–11

Human Comfort and Air Distribution

•

Group C outlets are mounted in or near the floor and discharge air in a vertical spreading jet (see Figure 3-6). Although outlets of this group are related to Group B, they have 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. 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.

Figure 3-6

•

3–12

Air Motion Characteristics of Group C Outlets

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 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.

Fundamentals of Air System Design

•

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 to and follows the floor, producing a stagnant region near the ceiling. During heating, the conditioned air flow reaches the floor and folds back toward the ceiling. If projected air does not reach the floor, a stagnant zone results.

Figure 3-7

Air Motion Characteristics of Group D Outlets

Figure 3-8

Air Motion Characteristics of Group E Outlets

3–13

Human Comfort and Air Distribution

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

The primary air from the outlet down to a velocity of about 150 fpm can be treated analytically. The heating or cooling load has a strong effect on the characteristics of the primary air.

•

The conditioned air (which is shown by the 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 150 fpm), with air temperatures generally within 1°F of 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 a stagnant zone 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 20 fpm), 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.

•

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 mode. For cooling mode, a standard method has been developed for rating diffusers 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, , is between –3°F and +2F and the air velocity is below 70 fpm.

3–14

Fundamentals of Air System Design

The draft temperature is a measure of perceived difference in temperature at a location as compared to a location at the average temperature and air velocity of 30 fpm. The draft temperature is calculated as: Draft temperature =  = (tactual – taverage) – 0.07(local air velocity – 30) For example, a location temperature 2F cooler than average and with an air velocity of 20 fpm will have a draft temperature of  = (–2) – 0.07(20–30) = –2 + 0.7 = –1.3F. To calculate the ADPI, a test room with air supplied 20F 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 ASHRAE Standard 113-2005, Method of Testing for Room Air Diffusion. Outlet performance selection data from numerous tests are shown for high sidewall outlets in Table 3-1 below. Now consider a 20 ft2 office, with a cooling load of 22 Btu/hft2 being cooled by a sidewall diffuser. The length of throw by which the air velocity has dropped to 50 fpm, T50, is used as the criteria. For a room load of 20 Btu/hft2, the maximum ADPI of 85 is obtained with a T50/L of 1.5 and, for over 80%, the ADPI range is 1.01.9. Aiming for the maximum, choose a grille with a 50 fpm terminal velocity throw of 1.5 times the room length; T50/20 = 1.5, so T50 = 30 ft. Table 3-1 Outlet Performance Selection Data Terminal Device High sidewall grilles

Room Load T50/L for (Btu/hft2) Max. ADPI 80 60 40 20

1.8 1.8 1.6 1.5

Maximum For ADPI ADPI Greater Than 68 72 78 85

 70 70 80

Range of T50/L  1.5  2.2 1.2  2.3 1.0  1.9

3–15

Human Comfort and Air Distribution

UNDER-FLOOR SYSTEMS Under-Floor Air Distribution (UFAD) is supplied from a raised floor through numerous small floor grilles. The floor typically consists of 24 in.2 metal plates, or tiles, supported by a 1018 in.-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 re-location, addition or removal a simple task, as shown in 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 58°64°F, 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 50 fpm within 4 feet of the floor. 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 air column takes contaminants with it up and out of the breathing zone. This sweep-away action is considered more effective than mix-and-dilute. As a result, the ventilation requirements of ASHRAE Standard 62.1 can be satisfied with 10% less outside air.

FAN COIL Figure 3-9

3–16

Under-Floor Air Distribution

Fundamentals of Air System Design

There are numerous outlets, because the individual outlet volume is typically limited to 100 cfm. 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 will work 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 is rising 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 somewhat acts as a passive thermal storage unit. This storage can be used to reduce peak loads, but it means 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 1015 feet to the perimeter fan coils to maintain an adequately low supply temperature. In a similar way, conference rooms that have highly variable loads can use a thermostatically controlled fan to boost the flow into the room when it is occupied. A modification of the under-floor 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 grille and has yet to gain popularity. The under-floor air delivery system has the following advantages: • Changing the layouts of partitions, 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-horsepower savings. • Less ventilation outside air can potentially be used. Disadvantages include: • A significant cost per square foot for the floor system supply, installation and maintenance.

3–17

Human Comfort and Air Distribution

• •

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. 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 100 fpm or less. The outlets are located 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 because 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.

Figure 3-10

3–18

Schematic of Displacement Ventilation

Fundamentals of Air System Design

Unidirectional air flow systems. Here, 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 air flow across the entire room. This type of system is primarily used for ventilating cleanrooms, or for 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 air flow 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 towards 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 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 open- Figure 3-11 Localized Ventilation ing. 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, then to a common

3–19

Human Comfort and Air Distribution

return or exhaust because smoke would accumulate in the main egress pathway in case of fire. Most building codes restrict this application. Room system balancing. Room system balancing is adjusting the air flows 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 air supply 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. Balancing an air system will be discussed in Chapter 11.

3.3 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, and ceiling diffuser outlets. 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. 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 is similar to the adjustable single-deflection 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 double-deflection 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.

3–20

Fundamentals of Air System Design

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 air-conditioning installations in interior zones and may cause drafts in perimeter applications. Accessories available for grille outlets include: • 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 air flow, 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.

Figure 3-12

Grille and Register Outlet Accessory Controls

3–21

Human Comfort and Air Distribution

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 3 in. generally meet the performance criteria for slot diffusers. Slot diffusers are generally equipped with accessory devices for uniform supply air discharge along the entire length of the slot. While accessory devices help correct the air flow pattern, proper approach conditions for the airstream are also important for satisfactory performance. When the plenum supplying a slot diffuser is being designed, the transverse velocity in the plenum should be less than the discharge velocity of the jet, as recommended by the manufacturer and also as shown by experience. If tapered ducts are used to introduce supply air into the diffuser, they should be sized to maintain a velocity of about 500 fpm, and tapered to maintain constant static pressure. Slot diffusers having 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 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: • Multi-passage 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.

3–22

•

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 in laboratories, hospital operating rooms and other spaces having high air change rates to provide laminar flow. Designers are cautioned to thoroughly investigate the air flow and induction characteristics under both cooling and heating conditions for this type of diffuser, particularly in applications with varying air flows such as VAV systems.

Fundamentals of Air System Design

•

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 4 to 12 in. beyond the outer edge of the diffuser. Their purpose is to minimize ceiling smudging.

Dampers 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 (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.

Figure 3-13

Ceiling Diffuser Outlet Accessory Controls

3–23

Human Comfort and Air Distribution

•

Blankoff baffles. These baffles are used for minor adjustments of the air flow 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 for balancing. Procedure for outlet selection. The following procedure is generally used in selecting outlet locations and types: •

Determine the amount of air to be supplied to each room based on system design and heating/cooling load calculations.

•

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. Table 3-2 Outlet Usage Guide

Outlet Type Grille Slot Perforated panel Ceiling diffuser Perforated ceiling

3–24

Floor space air loading (cfm/ft2)

Approx. max. ACH for 10-ft ceiling

0.6  1.2 0.8  2.0 0.9  3.0 0.8  5.0 1.0  10.0

    

Fundamentals of Air System Design

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 lowpressure 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 will occur when using ceiling and slot diffusers. Dirt particles held in suspension in the room air are subjected to turbulence at the outlet face. This turbulence is primarily responsible for smudging. The cleanliness of the room will affect when the smudging becomes visible. An outlet’s sound level is a function of the damper arrangement, discharge velocity and transmission of systemic noise, which is a function of 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 will be covered in greater detail in Chapter 10. Suggested duct velocities where takeoffs to grilles or diffusers are close to the outlet are: • Acceptable high noise levels: 1,500 fpm maximum • General office of classroom: 1,000 fpm maximum • Noise sensitive areas: 800 fpm 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 air flow. 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 air flow 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 inlet’s performance. Usually no other accessory devices are required.

3–25

Human Comfort and Air Distribution

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

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 55–2004 details requirements for thermal comfort. ASHRAE Standard 62.1-2007 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, under-floor, 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 it mixes with room air, 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 significantly influenced by the location of the return air outlet. The performance of various types of mixing outlets has been analyzed for their relative ability to maintain comfort conditions throughout a space. This data is presented as ADPI and can be used to make choices about outlets. Under-Floor Air Distribution (UFAD) uses a plenum created above the structural floor using 2 ft2 metal panels on support columns. The air, at 58°64°F, is supplied up through diffusers distributed among the floor panels. The system uses the vertical supply and convection to lift the air towards ceiling outlets. As the air flows across the structure, the structure acts as a thermal buffer and the system is slow to change.

3–26

Fundamentals of Air System Design

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-tofloor height, and a need for very 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, with convection lifting the contaminated air to high level return outlets. The system minimizes mixing. A wide range of grille outlets with fixed or/and adjustable vanes provide 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 adjustable by using the grille blades or adjustable damper and turning vane accessories. A slot diffuser is a grille with one long dimension and is often designed to be installed endto-end for long continuous air supply. Due to their length, the air supply must be carefully designed to obtain consistent performance. Their length makes them aesthetically suitable to install alongside fluorescent lighting fixtures. 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 wellmixed circulating air. Having established the quantity of supply air, a preliminary choice of outlet style and layout can be made. Using manufacturers’ data on air flow and sound generation, the final airflows and layout are determined. The diffuser choice will often be 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.

3–27

Human Comfort and Air Distribution

Bibliography ANSI/ASHRAE Standard 55-2004, Thermal Environmental Conditions for Human Occupancy. Atlanta, GA: ASHRAE. ANSI/ASHRAE Standard 62.1-2007, Ventilation for Acceptable Indoor Air Quality. Atlanta, GA: ASHRAE. ASHRAE Standard 113-2005, Method of Testing for Room Air Diffusion. Atlanta, GA: ASHRAE. ASHRAE. 1993. Air-Conditioning Systems Design Manual. Atlanta, GA: ASHRAE. ASHRAE Handbook-Fundamentals: thermal comfort, indoor environmental health, odor, space air diffusion, heating and cooling load calculations ASHRAE Handbook-HVAC Applications: control of gaseous indoor air contaminants ASHRAE Handbook-Systems and Equipment: air-diffusing equipment

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

3–28

3-1.

The human body uses which of the following heat transfer mechanisms: a) Radiation b) Convection c) Evaporation d) All of the above

3-2.

The perception of comfort relates to: a) Individual physical condition b) Body heat exchange with the surroundings c) Physiological characteristics d) All of the above e) None of the above

3-3.

Which of the following would be within the acceptable range of temperature and humidity for human comfort when wearing light summer clothing? a) 72°F, 20% rh b) 70°F, 65% rh c) 80°F, 30% rh d) All of the above

3-4.

In a system with 8,000 annual wet-bulb degree hours above 66°F, with a 60% indoor relative humidity desired, and 56 hours of cooling system operation per week, the energy used will be _______ Btu  106 per year per 1,000 cfm. a) 51 b) 25 c) 36 d) None of the above

3-5.

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 b) Ventilation Rate c) Contaminant Mitigation d) All of the above e) None of the above

Fundamentals of Air System Design

3-6.

Many designers have adopted a minimum total supply air flow of _____ for office applications. a) 0.1 to 0.3 cfm/ft2 b) 0.6 to 0.8 cfm/ft2 c) 0.2 to 2.0 cfm/ft2 d) All of the above e) None of the above

3-7.

The airstream velocity at the end of the throw is called: a) Terminal velocity b) Primary velocity c) Airstream velocity d) All of the above e) None of the above

3-8.

_________________ air distribution systems create relatively uniform air conditions in the occupied zone. a) Unidirectional b) Local c) Mixing d) All of the above e) None of the above

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 b) The same as c) Smaller than d) All of the above

3-10.

In displacement systems, the outlets are frequently located: a) At or near the floor level b) In the walls c) In the ceiling d) A and B e) None of the above

3-11.

Smudging is most likely to occur from dirt particles held in suspension in: a) The room air b) The supply air c) The return air d) All of the above e) None of the above

3-12.

The fan horsepower for under-floor 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 air flow in the plenum c) The insulating value of the floor and carpet

3-13.

The under-floor supply systems work well for large open areas and the most effective control is a thermostat in the return duct. True or false? a) True b) False

3–29

Fundamentals of Air System Design

Chapter 4

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

4.1 Operating System Selection Criteria 4.2 System Types by Heating/Cooling Equipment Type 4.3 System Type by Duct Configuration 4.4 Economizers 4.5 Outdoor Air Intake Summary Bibliography Skill Development Exercises for Chapter 4

4–1

Relationship of Air Systems to Load and Occupancy Demands

Study Objectives of Chapter 4 After completing this chapter, you should be able to describe: • • • •

Operating system criteria; Air systems by heating/cooling equipment type; Air systems by duct configurations; and Considerations for outdoor air intake.

4.1 Operating System Selection Criteria To select an operating system, detailed building design and use information plus 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:

4–2

•

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 as to 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.

•

Thermal zones. The thermal zones within the building should be identified. For example, external offices with windows will have different thermal characteristics than windowless rooms in the building’s interior. Additionally, some areas of the building may have to be kept at different temperatures than others.

•

Room pressures. 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 will prevent 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.

•

Building uses. The uses to which the building will be put will 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 dry-bulb temperatures, daily range, heating and cooling degree days, elevation, etc.) and select outdoor design conditions from local weather stations. The ASHRAE

Fundamentals of Air System Design

Handbook–Fundamentals also lists outdoor design conditions for many 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 ASHRAE Standard 55-2004 has wide limits on moisture content in the space as it is dealing 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 example, polluted areas such as toilets.

•

System design and sizing. The proper design and sizing of central heating and air-conditioning systems require 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 that system performance be analyzed as a series of psychrometric processes. The ASHRAE Handbook–HVAC Systems and Equipment and the ASHRAE Handbook–Fundamentals 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 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, in buildings having a large amount of glass located at 32°N latitude, the peak load times are shown in Table 4-1.

4–3

Relationship of Air Systems to Load and Occupancy Demands

Table 4-1 Perimeter Zone East West South North Northeast & Southeast Northeast & Southwest

Peak Load Times

Peak Load Time

Month

8:00 AM 4:00 PM 12:00 NOON 12:00 NOON 10:00 AM 2:00 PM

August August December June March & October March & October

Interior zones peak at times of peak occupancy.

•

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 will be present whether or not the system is used at all (taxes, insurance, etc.). Operating costs are what it costs to run the system including energy and maintenance. The ASHRAE Handbook–HVAC Applications provides a detailed discussion of this subject.

4.2 System Types by Heating/Cooling Equipment Type UNITARY EQUIPMENT SYSTEMS Unitary equipment systems are systems that are factory-assembled into an integrated package including fans, filters, heating coil, cooling coil, refrigerant compressors, refrigerantside controls, airside 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 airconditioning 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. They are installed in buildings requiring many temperature control zones (such as motels, apartments and dormitories). These systems are applicable for renovation of existing 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.

4–4

Fundamentals of Air System Design

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 that 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, 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 carefully evaluate the use of unitary equipment in cases of more than 25% outside air. Many unitary systems will not remove sufficient moisture at high outside air quantities. For adequate part-load performance at high outside 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 ASHRAE Handbook–HVAC Systems and Equipment.

Figure 4-1

Multiple Packaged Units

4–5

Relationship of Air Systems to Load and Occupancy Demands

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, 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, research and development facilities), as well as many industrial/manufacturing facilities. All-air systems have the following advantages:

4–6

•

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.

•

These 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, including simultaneous heating and cooling, even during off-season periods.

•

Air-to-air and other heat recovery systems may be readily incorporated.

•

Good design flexibility is permitted for optimum air distribution, draft control and adaptability to varying load requirements.

•

These systems are well suited to applications requiring unusual exhaust or makeup air quantities (negative or positive pressurization, etc.).

•

All-air 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.

Fundamentals of Air System Design

•

By increasing the air change rate, these systems are able to maintain operating conditions of ±1.0°F dry-bulb and ±5% relative humidity fairly simply. Some systems 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 on 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 air flow, temperatures, relative humidity, loads and psychrometrics are covered in the ASHRAE Handbook–Fundamentals. It is important that the designer understands the operation of the various system components, 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 dual-duct. 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

Constant volume single-duct. 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 remote from the space it serves, and it may operate with or without distributing ductwork.

4–7

Relationship of Air Systems to Load and Occupancy Demands

In Figure 4-2, heat flows and air flows are indicated by arrows, and temperatures are indicated by t. The subscripts in the sequence of the air flow are: R = room rp = return plenum o = outside air m = mixed air r = return 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 outside 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 the operation of adjacent areas. They are energy efficient, easy to control, and easily adaptable to economizers. Their disadvantage 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, 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

4–8

Fundamentals of Air System Design

Figure 4-2

Single Zone Schematic and Psychrometric Chart

4–9

Relationship of Air Systems to Load and Occupancy Demands

relation to the outside air inlet pressure and ambient. 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, ASHRAE Standard 90.1 severely restricts the application of reheat, limiting this option to special cases because of the high energy consumption of the system.1 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-andbypass 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. 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. 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 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.

4–10

Fundamentals of Air System Design

Figure 4-3

VAV System Schematic and Psychrometric Chart

4–11

Relationship of Air Systems to Load and Occupancy Demands

With the current concern for indoor air quality, care should be exercised to provide minimum ventilation in any occupied space and required outside air quantities under all operating conditions. The pressure relationships of the system change when the supply fan is throttled. Means such as outside 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: •

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 air flow, usually at least 50% of the full air flow. 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 not desirable. Fan speed control is preferred. Three VAV box arrangements are shown in Figure 4-4. The first is the simplest. 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. The second diagram 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 air flow. When the primary airflow is reduced, the fan draws more air from the ceiling plenum, maintaining the 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.

4–12

Fundamentals of Air System Design

Figure 4-4

VAV Box Arrangements

4–13

Relationship of Air Systems to Load and Occupancy Demands

4–14

•

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 year-round, 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 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.

•

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 will still be delivered to the occupied zone when the primary air is throttled. A distributed outside 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 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 outside 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.1 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

Fundamentals of Air System Design

other where cooling is required throughout the year. The typical terminal device (box) will have two inlets, one for cold air and one for hot air or bypass air. Each inlet will have a throttling damper and actuator. Typically, the cold damper will be 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 air flow. Under these conditions, the induction effect of the diffuser is kept high, and cold air mixes in the space.

One important, and difficult, issue with VAV systems is providing enough ventilation air to each space all the time. Consider a simple example 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.2 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 that ventilation effectiveness is maintained even at times of low air flow. 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). The two types of constant volume dual-duct systems are: •

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).

4–15

Relationship of Air Systems to Load and Occupancy Demands

Figure 4-5

Dual Duct System

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 from two static pressure controllers: one located in the hot deck, and the other in the cold deck. The duct requiring the highest pressure governs the fan air flow

•

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 multizone system has a heating coil in each mixed air zone, which is energized only when the cooling damper is closed.

4–16

Fundamentals of Air System Design

Figure 4-6

Dual Duct, Dual Fan System

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

4–17

Relationship of Air Systems to Load and Occupancy Demands

Figure 4-7

Air-Water Induction Terminal

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. 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.

4–18

Fundamentals of Air System Design

Figure 4-8

Fan-Coil Unit

Figure 4-9

Ceiling Panel Example

.

4–19

Relationship of Air Systems to Load and Occupancy Demands

EVAPORATIVE COOLING SYSTEMS Evaporative coolers exchange sensible heat for latent heat. Evaporative air cooling evaporates water into an airstream. Figure 4-10 illustrates the 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 or the extent to which complete saturation is approached is expressed as a percentage evaporative cooling or saturation effectiveness and is defined:  t1 – t2  e c = ------------------ t 1 – t'  where: ec = evaporative cooling or saturation effectiveness, percent 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).

4–20

Figure 4-10

Thermodynamic Interaction of Water and Air

Fundamentals of Air System Design

4.3 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. 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 will be discussed in Chapter 7.

4.4 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 so-called airside 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 airside 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, airmixing chambers and louvers necessary to dissipate the air pressure caused by supplying 100% outside air.

Figure 4-11

Airside Economizer

4–21

Relationship of Air Systems to Load and Occupancy Demands

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; the dampers are positioned to bring in the minimum outdoor air required for ventilation.

As a simple rule-of-thumb, airside 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.

Figure 4-12

Airside Temperature Economizer Cycle

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 airside free cooling, depending on climate. Another form of free cooling involves purging the conditioned areas with cool night air prior to occupancy the following morning. This can avoid the cooling energy necessary to

4–22

Fundamentals of Air System Design

overcome the heat buildup from lights, office equipment and heat flowing back to the conditioned space from concrete floor slabs absorbing heat during the day. This purging cycle is highly effective in dry climates with low nighttime temperatures, such as in the southwestern United States. Do not use it in hot humid climates because of the potential moisture buildup.

Figure 4-13

Enthalpy Economizer Cycle

4.5 Outdoor Air Intake Outdoor air is air outside a building, or 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; and The energy associated with heating or cooling this outdoor air is a significant space-conditioning load.

In large buildings, the effects of infiltration and ventilation on distribution and interzone air flow patterns, which include smoke circulation patterns in the event of fire, should be determined, see “Fire and Smoke Management” in the ASHRAE Handbook–HVAC Applications. Outdoor 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 equip-

4–23

Relationship of Air Systems to Load and Occupancy Demands

ment. 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 2.5 ft 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). When outdoor air must be drawn in through the roof, a gooseneck outdoor air intake like the one 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: “Use rain hoods sized for no more than 500 fpm (2.5 m/s) face velocity with a downward-facing intake such that all intake air passes upward through a horizontal plane that intersects the solid surfaces of the hood before entering the system” to minimize rain entrainment.2

Figure 4-14

4–24

Outdoor Air Louver

Fundamentals of Air System Design

Figure 4-15

Gooseneck Outdoor Air Intake3

The Next Step We have been considering supply air systems in this chapter. In Chapter 5, we will consider exhaust systems to remove excess air and contaminants from the building.

Summary System selection depends on many factors including: • • • • • •

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 costs

4–25

Relationship of Air Systems to Load and Occupancy Demands

Once the requirements are known, the most appropriate system can be selected. Complete, factory-assembled units range from the small window air-conditioner serving a single room to large packages serving a whole building. In the larger sizes, the unit may be supplied as a set of bolt-together parts. These units 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 advantages are that the equipment is 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 also provide air for processes with high exhaust needs as well as flexibility in zoning and control performance. All-air systems have disadvantages including 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 either 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 them: • • • •

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, one with cold air run through the building. At each zone, air from each duct is connected to a dual-duct box that chooses the proportion of warm and cool air to deliver to the zone to maintain temperature control.

Air-and-water systems provide ventilation and humidity control by supplying air to each zone while most of the cooling and heating loads are handled by water coils in the zone. The ventilation air may be used as the power source for inducing room air over the coil as in induction systems, 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 will be discussed in Chapter 7.

4–26

Fundamentals of Air System Design

Mechanical cooling can be minimized by using outside air whenever the outside 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 also should be located to minimize drawing in pollutants.

Bibliography 1. ASHRAE. 2004. ASHRAE/IESNA Standard 90.1-2004, Energy Efficient Design of New Buildings Except Low-Rise Residential Buildings. Atlanta, GA: ASHRAE. 2. ASHRAE. 2007. ANSI/ASHRAE Standard 62.1-2007, Ventilation for Acceptable Indoor Air Quality. Atlanta, GA: ASHRAE. 3. Carrier Corp. 1965. Handbook of Air Conditioning System Design. New York, NY: McGrawHill. ASHRAE HandbookFundamentals: load calculations, psychrometrics; HandbookHVAC Systems and Equipment: HVAC system analysis and selection, system types and equipment, heat recovery; HandbookHVAC Applications: energy use, owning and operating costs, building intake and exhaust design, evaporative cooling

4–27

Relationship of Air Systems to Load and Occupancy Demands

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

4–28

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 e) None 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) All of the above e) 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 b) False

4-5.

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 e) None 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 outside air louvers where light snowfall is expected: a)1.0 ft b) 2.5 ft c) 4.0 ft d) All of the above e) None of the above

Fundamentals of Air System Design

Chapter 5

Exhaust and Ventilation Systems Contents of Chapter 5 • • • • • •

5.1 Design Considerations 5.2 Ventilation and Exhaust Systems 5.3 Energy Recovery Summary Bibliography Skill Development Exercises for Chapter 5

5–1

Exhaust and Ventilation Systems

Instructions Read the material of Chapter 5. At the end of the chapter, complete the skill development exercises without consulting the text.

Study Objectives of Chapter 5 After completing this chapter, you should be able to describe design considerations for exhaust and ventilation systems and some energy recovery systems.

5.1 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:

5–2

•

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.

•

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.

Fundamentals of Air System Design

•

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, ASHRAE Standard 62.1, Ventilation for Acceptable Indoor Air Quality has requirements for ventilation air and exhaust, as was described in Chapter 2.1 For industrial applications, minimum values are prescribed for local exhaust systems in Industrial Ventilation, A Manual of Recommended Practice2 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, 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 due to 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, the longer 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 continuous operation. Exhaust ducts from high-moisture areas (such as shower rooms) must have drains and watertight bottoms. Corrosion-resistant material should be considered.

•

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

5–3

Exhaust and Ventilation Systems

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: •

To replace air exhausted through combustion processes and local and general exhaust systems.

•

To eliminate uncomfortable cross-drafts by proper arrangement of supply air, and prevent infiltration through doors, windows and similar openings that may make exhaust hoods unsafe or ineffective, defeat environmental control, bring in or stir up dust, or adversely affect processes.

•

To obtain clean air. Supply air can be filtered, infiltration air cannot. Also, supply air can be preheated to prevent spot freeze-up, infiltration air cannot.

•

To control building pressure and air flow from space to space. Such control is necessary: • To avoid positive or negative pressures that will 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 control temperature, humidity and air movement positively. • To recover heat and conserve energy.

Table 5-1 Negative Pressure (in.wg) 0.01  0.02 0.01  0.05 0.02  0.05 0.03  0.10 0.05  0.10 0.10  0.25

5–4

Negative Pressures That May Cause Unsatisfactory Building Conditions Adverse Conditions Worker Draft Complaints: High velocity drafts through doors and windows Natural Draft Stacks Ineffective: Ventilation through roof exhaust ventilators, flow through stacks with natural draft greatly reduced Carbon Monoxide Hazard: Back-drafting will occur in hot water heaters, unit heaters, furnaces and other combustion equipment not provided with induced draft General Mechanical Ventilation Reduced: Air flows reduced in propeller fans and low pressure supply and exhaust systems Doors Difficult to Open: Serious injury may result from nonchecked, slamming doors Local Exhaust Ventilation Impaired: Centrifugal fan exhaust reduced

Fundamentals of Air System Design

STACK EFFECT Temperature differences between indoors and outdoors cause density differences and, therefore, pressure differences that drive infiltration. During the heating season, the 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 stack effect is reduced and pressures reversed, because the indoor-outdoor temperature differences are smaller and 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 (NPL). Above this point (during the heating season), the interior pressure is greater than the exterior; below this point, the greater exterior pressure causes air flow into the building.

Figure 5-1

Pressure Differences Due to Stack Effect (Heating Season)

5–5

Exhaust and Ventilation Systems

The pressure difference due to the stack effect at height h is: p s = C 2   o –  i g  h – h NPL  = C 2  i g  h – h NPL   T i – T o T o

(5-1)

where: ps = pressure difference due to stack effect, in. wg

 = air density, lbm /ft3 (about 0.075 for indoor conditions) g = gravitational constant, 32.2 ft/s2 h = height of observation, ft hNPL = height of neutral pressure level, ft T = average absolute temperature, °R C2 = unit conversion factor, 0.00598 Regarding the unit conversion factor, water weighs 62.4 lbm per ft3. So 1 inch of water is 62.4/12 lbm per inch depth. For ps to be in in. wg, ps must be multiplied by (62.4/12)  32.2 = 167.17. Moving 167.17 to the right hand side of the equation changes it to 1/167.17 = 0.00598. Subscripts: i = inside; o = outside

5.2 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 warmmoist 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

5–6

Fundamentals of Air System Design

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 grilles, ceiling diffusers or supply plenums to distribute the air adequately throughout the area. Pressure relationships must be carefully considered to prevent air flow from locker rooms, toilets and shower spaces to other occupied spaces. ASHRAE Standard 62.1 includes general exhaust requirements including those shown in Table 5-2. Note that where the lockers are being used for laboring employees with wet, sweaty clothes, the rate should be increased to the higher of 1 cfm/ft2 or 7 cfm exhausted from each locker. Where heavy labor is involved and the clothes may be wet and have picked up odors, the rate should be increased to the higher of 3 cfm/ft 2 or 10 cfm exhausted from each locker.

Table 5-2 Ventilation for Locker Rooms, Ancillary and Toilet Spaces Space Locker rooms Locker/dressing rooms Janitor, trash, recycling Toilets - public Toilets - private

CFM per Unit

CFM/FT2 0.50 0.25 1.00

50/70* 25/50*

*The toilet rate is per water closet and/or urinal. Provide the higher rate where periods of heavy use are expected to occur; for example, toilets in theaters, schools and sports facilities. The lower rate may be used where use is intermittent.

5–7

Exhaust and Ventilation 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 space, with additional factors including: • • • •

Extremely variable loads with high peaks, in many cases occurring twice daily High sensible and latent heat gains because of gas, steam and electric appliances, people and food Heavy infiltration of outdoor air through doors during rush hours in commercial establishments 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.

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 smokefilled, 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:

5–8

•

Stack effect. As discussed earlier, when the air outside a building is colder than the building air, 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.

Fundamentals of Air System Design

•

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

(5-2)

where: Qout = volumetric flow rate of smoke out of the fire compartment, cfm Qin = volumetric flow rate of smoke into the fire compartment, cfm Tout = absolute temperature of smoke leaving the fire compartment, °R Tin = absolute temperature of smoke entering the fire compartment, °R •

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 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 ASHRAE Handbook–HVAC Applications.

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 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, air flow through the door should be sufficient to

5–9

Exhaust and Ventilation Systems

prevent smoke backflow. Designing such a system is difficult because of the many combinations of open stairwell doors and weather conditions affecting air flow. 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 over-pressurization, 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 atmosphere. Alternatively, pressure sensors measuring the pressure between a floor and the stairwell can control a damper on a short-circuit 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 0.20.3 in. wg so that it can be opened. The minimum pressure to hold back smoke is about 0.08 in. wg, 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 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 air supply 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 45 ft apart (see Figures 5-2a and 5-2b). Compartmentation of a stairwell is illustrated in Figure 5-3. Additional information on stair pressurization can be found in the ASHRAE Handbook– HVAC Applications.

5–10

Fundamentals of Air System Design

Figure 5-2a

Stairwell Pressurization, Ground-Level Fan

Figure 5-3

Figure 5-2b

Stairwell Pressurization, Roof-Mounted Fan

Compartmentation of Pressurized Stairwell

5–11

Exhaust and Ventilation Systems

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 The need for different temperature and humidity requirements for various areas 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 ASHRAE Handbook–HVAC Applications.

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

ENERGY RECOVERY COILS Run-Around Coils: One way to achieve energy recovery is with run-around energy recovery coils. A typical run-around 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 pre-cool 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 mixture of very cold fluid from the supply air coil and

5–12

Fundamentals of Air System Design

diverted fluid mix to a temperature that is high enough to avoid causing frost on the exhaust air coil. The maximum 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 run-around 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 outside air intake low in the building and the fume hood exhaust on the roof, with the run-around pipes connecting the two coils.

•

The run-around 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.

Figure 5-4

Run-Around Energy Recovery Coils

5–13

Exhaust and Ventilation Systems

HEAT PIPES A heat pipe is a length of pipe with an interior wick that contains a refrigerant charge, as shown in Figure 5-5. The type and quantity of refrigerant that is 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. 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 end 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, giving up its latent heat, which heats the air. The re-condensed 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 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 units is very low cross-contamination.

5–14

Fundamentals of Air System Design

Figure 5-5

Cutaway Section of a Heat Pipe

Figure 5-6

Heat Pipe Assembly in Exhaust and Outside Air Entry Pipe

5–15

Exhaust and Ventilation Systems

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 through the other half. Suppose it is a hot summer day, so the exhaust is cooler and drier than the supply of outside air. The chemical coating in the section of the coil 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, a section is moving 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. The use of heat recovery is required in many energy codes, particularly for larger systems and systems with a high proportion of outside air. ASHRAE Standard 90.1-2007Energy Standard for Buildings Except Low-Rise Residential Buildings has several mandatory requirements for the use of heat recovery equipment.3

Figure 5-7

5–16

Desiccant Wheel Matrix and Operation Pipe

Fundamentals of Air System Design

The Next Step This chapter has covered ventilation and exhaust. The next chapter will cover fans and the movement of air through systems.

Summary Ventilation and exhaust systems control heat, odors and contaminants. Exhausts can be: •

Local: removing the contaminant before it mixes with the air in the space

•

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 there is a balance. Failure to provide adequate supply air makeup can create problems of pressure difference. Therefore, exhausts must be designed with the supply system. For human comfort, the supply of outside air was the criteria. But for many commercial and industrial processes, the exhaust is the criteria. The process often determines the volume and the construction of the exhaust system to deal with corrosion and erosion. 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 0.1 in. wg, 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

5–17

Exhaust and Ventilation Systems

•

Expansion of the air due to temperature around the fire

•

Wind blowing past the building, creating a higher pressure on the windward side and a lower pressure on the leeward side

•

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 as the pressure must be maintained even with doors open but limited to prevent doors being held shut by the pressure. Barometric dampers and short-circuit ducts 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. 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: Transfer heat using the boiling and condensation of refrigerant in sealed lengths of pipe to transfer heat from one end of the tube to the other. Capacity control is by tilting the pipes. Some cross-contamination may occur. Desiccant wheels: 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 1. ANSI/ASHRAE Standard 62.1-2007, Ventilation for Acceptable Indoor Air Quality. Atlanta, GA: ASHRAE. 2. ACGIH. 2004. Industrial Ventilation, A Manual of Recommended Practice. Cincinnati, OH: American Conference of Governmental Industrial Hygienists. 3. ASHRAE Standard 90.1-2007, Energy Standard for Buildings Except Low-Rise Residential Buildings. Atlanta, GA: ASHRAE. ASHRAE HandbookFundamentals: ventilation, stack effect; HandbookHVAC Applications: ventilation and exhaust for specific applications, energy recovery, fire and smoke management; HandbookHVAC Systems and Equipment: air-to-air energy recovery

5–18

Fundamentals of Air System Design

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 will produce a significant stack effect: a) True b) False

5-2.

Care must be taken in exhaust systems to minimize: a) Corrosion b) Dissolution c) Melting d) All of the above e) None of the above

5-3.

All other things being equal, ductwork is least subject to condensation corrosion when the runs are: a) Long and horizontal b) Short and vertical c) Direct to the terminal discharge d) All of the above e) None of the above

5-4.

Kitchen air pressure should be kept ______________ relative to other areas. a) Positive b) Neutral c) Negative d) All of the above e) None of the above

5-5.

Smoke movement is driven by: a) Stack effect b) Buoyancy c) Expansion d) All of the above e) None of the above

5-6.

To prevent smoke infiltration on a fire floor, a pressurized stairwell must maintain a _________________ pressure difference across a closed stairwell door. a) Positive b) Neutral c) Negative d) All of the above e) None of the above

5-7.

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 e) None of the above

5–19

Fundamentals of Air System Design

Chapter 6

Air Movers and Fan Technology Contents of Chapter 6 • • • • • • • • •

6.1 Fan Principles 6.2 Fan Drives 6.3 Fan Selection 6.4 Fan Installation Design 6.5 Fan Controls 6.6 Effect of Variable Resistance Devices Summary Bibliography Skill Development Exercises for Chapter 6

6–1

Air Movers and Fan Technology

Study Objectives of Chapter 6 After completing this chapter, you should be able to: • • • •

List and explain fan principles; List and describe the main types of HVAC fans, fan drives and fan controls; Explain factors to be considered when selecting an appropriate fan for a given set of conditions; and Explain factors to be considered when installing a fan, once it has been selected.

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

V Wo

= fan outlet area, ft2 = fan size or impeller diameter = rotational speed, rpm (sometimes revolutions per second) = volume flow rate moved by fan at fan inlet conditions, cfm = fan total pressure rise; fan total pressure at outlet minus fan total pressure at inlet, in. wg = fan velocity pressure; pressure corresponding to average velocity determined from the volume flow rate and fan outlet area, in. wg = fan static pressure rise; fan total pressure rise diminished by fan velocity pressure, in. wg. The fan inlet velocity head is assumed equal to zero, because the inlet is not connected to ductwork and unobstructed for fan rating purposes. = fan inlet or outlet velocity, fpm = power output of fan; based on fan volume flow rate and fan total pressure, hp

Wi

= power input to fan; measured by power delivered to fan shaft, hp

ht

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

hs

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



= gas density, lb/ft3

A D N Q ptf pvf psf

6–2

Fundamentals of Air System Design

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. Centrifugal fan impellers produce pressure from: •

The centrifugal force created by rotating the air column enclosed between the blades

•

The kinetic energy imparted to the air by virtue of its velocity leaving the impeller

Axial flow fans produce pressure from the change in velocity passing through the impeller, with none being produced by centrifugal force. The basic fan types can be further subdivided 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 illustrates most of these fans, 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 centre 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) when the only components subjected to the high temperature are the impeller and drive shaft.

6–3

Air Movers and Fan Technology

Figure 6-1 6–4

Types of Fans

Fundamentals of Air System Design

Figure 6-1 Types of Fans (cont.)

6–5

Air Movers and Fan Technology

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 of changing size, speed or density on volume flow, pressure and power. Fan Law 2 shows the effect of changing size, pressure or density on volume flow rate, speed and power. Fan Law 3 shows the effect of changing size, volume flow or density on speed, pressure and power. 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 1/2

1c

W1 = W2



(D1/D2)5 (N1/N2)3 1/2

2a

Q1 = Q2



(D1/D2)2 (p1/p2)1/2 ( 2 /1)1/2

2b

N1 = N 2



(D2/D1) (p1/p2)1/2 (2 /1)1/2

2c

W1 = W2



(D1/D2)2 (p1/p2)3/2 (2 /1)1/2

3a

N1 = N 2



(D2/D1)3 (Q1/Q2)

3b

p1 = p2



(D2/D1)4 (Q1/Q2)2 1/2

3c

W1 = W2



(D2/D1)4 (Q1/Q2)3 1/2

Fan Law 2

Fan Law 3

Notes: 1. Subscript 1denotes 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

The fan laws simplify analyzing a given fan because there is no change in fan size (D) or density (). 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.

6–6

Fundamentals of Air System Design

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 = 6,000 cfm and ptf2 = 1.13 in. wg Using Fan Law 1a at Point E, Q1 = 6,000  650/600 = 6,500 cfm Using Fan Law 1b, ptf1 = 1.13 (650/600)2 = 1.33 in. wg 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.

Figure 6-2

Sample Application of the Fan Laws

6–7

Air Movers and Fan Technology

If equivalent points of rating are joined (as shown by the dotted lines in Figure 6-2), these points will form parabolas that are defined by the relationship expressed in the following equation: Q  p --------2-  =  ------2   p 1   Q1 

2

(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 test. Finally, the horsepower required by a fan is related to both the volume flow rate, and the pressure. The relationship can be expressed in several ways: hp~p  Q or hp~Q 3 or hp~N 3 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 cfm. For example, if the air flow rate is reduced by 20% to 80% of the previous, Q2 /Q1 = 0.80. Therefore, Q2 3 hp 2 3 -------- =  ------  = 0.8  Q1  hp 1

0.512 =

(6-2)

The fan power is reduced to 51.2% of the original amount. Another way to express this is that a 20% reduction in air flow results in a 48.8% 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 AMCA Standard 210 and ASHRAE Standard 51, and the efficiency is reflected in the rating.1,2

6–8

Fundamentals of Air System Design

Fan total pressure ptf is a true indication of the energy imparted to the airstream by the fan. System pressure loss (p) 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 lowvelocity systems. (ASHRAE Handbook-Fundamentals has further information.)3 A very important relationship is: V 2 p v =  ------------  in. wg. 4005

(6-3)

To specify the pressure performance of a fan, the relationship of ptf , psf and pvf must be understood, especially when negative pressures are involved. Most importantly, psf is a defined term 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 illustrate the relationships among these various pressures. Note that, as defined, p tf = p t2 – p t 1 . Figure 6-3 illustrates a fan with an outlet system but no connected inlet system. In this particular case, the fan static pressure p sf equals the static pressure rise across the fan. Figure 6-4 shows a fan with an inlet system but Figure 6-3 Pressure Relationships of Fan With Outlet no outlet system. Figure System Only 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.

6–9

Air Movers and Fan Technology

Figure 6-4

Figure 6-5

6–10

Pressure Relationships of Fan With Inlet System Only

Pressure Relationships of Fan With Equal-Sized Inlet and Outlet Systems

Fundamentals of Air System Design

All of the systems illustrated 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 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 the figure, 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 this rather normal fan situation, 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.

Figure 6-6

Pressure Relationships of Fan With Diverging Cone Outlet

6–11

Air Movers and Fan Technology

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 used to determine the air flow 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.

Figure 6-7

6–12

Method of Obtaining Fan Performance Curves

Fundamentals of Air System Design

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 of the fan and provides stable, uniform flow conditions at the plane of measurement. This allows the centrifugal fan with a cutoff, and the vaneaxial or propeller fan discharge velocities to equalize along the duct, and the difference in velocity pressure is converted to available 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 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 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.)

6.2 Fan Drives A proper motor and drive selection aids in long life and minimum service requirements. Direct drive fans are normally used on 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 brake horsepower requirements of the fan plus drive losses. The motor must be large enough to operate within its rated horsepower capacity including drive losses and reductions in line voltages and shortterm conditions. Normal torque motors are generally used for fan duty.

6–13

Air Movers and Fan Technology

6.3 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, 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 illustrated in Figure 6-8. If the fan is to operate in zone A, a larger fan will be more efficient. Conversely, if the fan is to operate in zone B, a smaller fan will be 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 will also dictate 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.

Figure 6-8

6–14

Optimum Fan Selection Zone

Fundamentals of Air System Design

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 36.5 in. diameter impeller or wheel. The value of a chart like Figure 6-9 is that, once it is prepared for a given type of fan, you can enter the X-axis with the cfm 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 cfm/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 36.5 in. fan. It is quite likely that you can go down to the 33 in. fan. While it will be less efficient, it will be down only a few points, it will not be 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 horsepower goes up as the cube of the speed ratio (hp = cfm3). 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 = 1.331), and few systems can easily tolerate that big an increase. Happily, this works in reverse. That is, if the cfm can be reduced, the whole process is reversed. Instead of being worried about overloading the system and requiring new equipment, you are cutting the power bill appreciably because of the cube ratio.

DENSITY, TEMPERATURE AND ALTITUDE Unless otherwise identified, fan performance data are based on dry air at standard conditions: 14.696 psi and 70°F, with a density of 0.075 lb/ft3. In actual applications, the fan may be required to handle air or gas at some other density. The change in density may be because of temperature, composition of the gas, or altitude. As indicated by the fan laws,

6–15

Air Movers and Fan Technology

Figure 6-9

6–16

Maximum Efficiency Lines for Various Fan Sizes

Fundamentals of Air System Design

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 10,000 cfm at sea level, it will produce 10,000 cfm at 5,000 ft above sea level, but not at the same pressure. The flow rate will remain the same no matter what the density. The change is strictly in pressure. Happily, a reduction in horsepower also occurs, which comes down as the density reduces. The point to remember is that catalog information is developed at standard density, and 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 air flow for a given energy delivered may need to be increased, which results in higher pressures and possibly higher horsepower requirements.

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 1 in. wg, and assume a fan that delivers 4,000 cfm across an outlet area of 1 ft2, giving a velocity of 4,000 fpm. From the equation for velocity pressure: V 2 4000 2 p v =  ------------  =  ------------  = 1in.  4005   4005 

(6-4)

If we arbitrarily establish a static pressure of 1 in. at the fan outlet, the total pressure becomes 1 in. + 1 in. = 2 in. Now consider another fan. Here is the same 4,000 cfm, but this fan has 2 ft2 of output area. Therefore, the velocity is 2,000 fpm. Again using Equation 6-1, the velocity pressure equals (2000/4005)2, or 0.25 in. To make things equal, we want 2 in. total pressure just as before. Subtracting the 0.25 in. velocity pressure from the total pressure leaves 1.75 in. of static pressure. The trouble starts when we consider efficiency. The equation for efficiency, , is:

6–17

Air Movers and Fan Technology

 cfm   p   = ---------------------------- 6362   hp  where,

(6-5)

 = efficiency p = pressure, in. wg

If we use static pressure in this equation, we get static efficiency. If we use total pressure, we get total efficiency. Some computations will demonstrate the effect of output area on static efficiency. Assume that from tests it has been determined that hp = 1.57.  cfm   p t   t = -----------------------------  100%  6362   hp 

(6-6)

Total efficiency in both cases:  4000   2   t = ----------------------------------  100% = 80%  6362   1.57  Static efficiency in case 1:  4000   1.00  s = -------------------------------- 6362   1.57   100% = 40%

(6-7)

Static efficiency in case 2:  4000   1.75  s = -------------------------------- 6362   1.57   100% = 70%

(6-8)

So here are two fans with the same horsepower (1.57), the same total pressure (2 in.), the same cfm (4,000), and the same total efficiency (80%). However, by doubling the outlet area, the static efficiency has been increased from 40% to 70%. 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 easily 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.

6–18

Fundamentals of Air System Design

However, particularly in systems with higher velocities (>1,500 fpm), it is important to deal with total pressure, not static pressure. A fan introducing unheated outside air will discharge a larger cfm of air after the air is heated. The fan motor should be selected for this added horsepower. The density of air varies as the absolute temperature difference where standard air in degrees Rankine (°R) is (70° + 460°) = 530°R. Therefore, a 4,000 cfm rated fan whose discharge air is heated to 170°F (630°R) introduces (4000  630°/530°) = 4,750 cfm 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 system 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 Figure 6-10 Operating Points and pressure; in the case shown, 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 set-

6–19

Air Movers and Fan Technology

ting, 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 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 flow 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, non-uniform 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 Figure 6-11 Deficient Duct System Performance3 must be added to the calculated system pressure losses to determine the actual system curve. The system effect 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 will, therefore, vary across the range of flow volumes of the fan (see Figure 6-12).

6–20

Fundamentals of Air System Design

In Figure 6-11, the point of intersection between the fan performance curve and the actual system curve is Point 4. The actual flow volume will, therefore, 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 this section. 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 effect factor is given in in. wg, and must be added to the total system pressure losses, as shown in Figure 6-11. The velocity figure used in entering the chart will be either the inlet or the outlet velocity of the fan. This will be dependent 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 the AMCA publication Fans and Systems.3

6–21

Air Movers and Fan Technology

Figure 6-12

6–22

System Effect Curves3

Fundamentals of Air System Design

6.4 Fan Installation Design COMPUTING THE EFFECT OF FAN OUTLET CONDITIONS Imagine an ideal uniform flow downstream from the fan. However, the reality is quite different. Figure 6-13 shows the flow patterns of a centrifugal fan and an axial fan. In either case, the flow is non-uniform 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 2,000 fpm. The slope of transition elements should not be greater than 15° for the converging elements, nor greater than 7° for the diverging elements.

Figure 6-13

Blast Areas for Centrifugal and Axial Fans3

6–23

Air Movers and Fan Technology

There will be a system effect for most fans at effective duct lengths of less than 100% of straight duct. Closer than that, and there will be an effect such as the one illustrated in Table 6-2, and the losses at other duct components (elbows, tees, etc.) will be higher than listed in the ASHRAE or SMACNA handbooks. Table 6-2 Blast Area Ratios for Various Fan Types Fan Type

Blast Area Ratio

Centrifugal Airfoil Backward-curved Backward-inclined Modified radial Radial Forward-curved

0.70 0.70 0.70 0.60 0.80 0.50

Propeller

0.90

Axial Hub ratio: 0.3 0.4 0.5 0.6 0.7

0.90 0.85 0.75 0.65 0.50

Note: Use actual manufacturer’s data when available.

One way to calculate effective duct length for round duct is as follows: •

If the duct velocity is greater than 2,500 fpm:  Vo Ao  L e = -------------------10 600

•

If the duct velocity is less than 2,500 fpm: A L e = ---------o 4.3

6–24

(6-9)

Fundamentals of Air System Design

where: Vo = duct velocity, fpm Le = effective duct length, ft Ao = duct area, in.2 If the duct is rectangular, the equivalent duct diameter is calculated by: 4HW D h = -------------------------2H + W 

(6-10)

where: Dh = equivalent duct diameter, in. H = rectangular duct height, in. W = rectangular duct width, in. In those cases where you use a shorter discharge length than one effective duct length, an additional pressure loss will result. 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:  V  p = K 1   ------------   1096 

2

(6-11)

where: p = pressure loss, in. wg V = velocity at outlet plane, fpm K1 = factor from appropriate tables  = density, lb/ft3 The blast area ratio is calculated by: Blast Area Ratio = Blast Area/Outlet Area 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.

6–25

Air Movers and Fan Technology

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 will result, 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 6-3. The additional pressure loss may also be determined by using Equation 6-11.

Figure 6-14

6–26

Outlet Duct Elbows3

Fundamentals of Air System Design

Table 6-3 System Effect Curves for Outlet Elbows3 Blast Area Outlet Elbow No Outlet 12% Effective 25% Effective 50% Effective 100% Effective Outlet Area Position Duct Duct Duct Duct Duct 0.4

A B C D

N M LM LM

O MN M M

PQ O N N

S R Q Q

0.5

A B C D

P NO MN MN

Q OP NO NO

R PQ OP OP

T S RS RS

0.6

A B C D

Q B NO O

QR Q OP P

RS R PQ QR

U T S ST

0.7

A B C D

ST RS QR R

T S R RS

U T S ST

W V UV UV

0.8

A B C D

S R Q QR

ST RS QR R

TU ST RS S

VW UV U UV

0.9

A B C D

S-T R-S R RS

T S RS S

U T ST T

W V UV V

1.0

A B C D

RS ST RS RS

S T S S

T U T T

V W V V

No System Effect Factor

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

6–27

Air Movers and Fan Technology

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 such obstructions are unavoidable, the resulting pressure losses can be estimated by using Equation 6-11. The K Factors for inlet area obstructions are shown in Table 6-4. Table 6-4 K Factor for Inlet Area Obstructions % Inlet Area Obstructed K Factor

5

10

15

25

50

75

0.22

0.40

0.53

0.80

1.20

1.60

When you estimate the percentage of inlet area remaining obstructed, use that part of the projected area of the obstruction perpendicular to the air flow 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 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 horsepower goes down. If you are getting lower horsepower 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,

6–28

Fundamentals of Air System Design

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

Figure 6-15a

Figure 6-15b

Inlet Duct Connections Causing Inlet Spin3

Corrections for Inlet Spin3

6–29

Air Movers and Fan Technology

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 6-5. The additional pressure losses may also be calculated using Equation 6-11. Table 6-5 K Factor for Enclosure Restrictions3 Length (L) 0.75  inlet diameter 0.50  inlet diameter 0.40  inlet diameter 0.30  inlet diameter

System Effect Curves* VW U T S

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

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 below for 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 of the fans 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-8 shape is a plot of all possible combinations of volume flow at each pressure value for the individual fans. All

6–30

Fundamentals of Air System Design

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 air flow than is shown on the p 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 p E–E, which requires a pressure of 1.0 in. wg and a volume of 5,000 cfm. The requirements of this system can be satisfied with each fan delivering 2,500 cfm at 1.0 in. wg, at Point CE. The system can also be satisfied at Point CE by one fan operating at 1,400 cfm at 0.9 in. wg, while the second fan delivers 3,400 cfm at the same 0.90 in. wg. 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.

Figure 6-16

Two Forward Curve Centrifugal Fans in Parallel Operation

6–31

Air Movers and Fan Technology

6.5 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 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, increases 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 Figure 6-17 System Total Pressure Loss Curves can be used even in cases where essentially continuous control is needed. However, a system effect loss is 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 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).

6–32

Fundamentals of Air System Design

Figure 6-18

Effect of Inlet Vane Control on Backward Curve Centrifugal Fab Performance

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 6-19

Effect of Blade Pitch on Controllable Pitch Vaneaxial Fan Performance

6–33

Air Movers and Fan Technology

6.6 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 will also change. 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 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 will push the system curve to the left, while dry coils will push it to the right. If a coil is not dehumidifying and becomes dry, there will be less of a pressure drop, the system curve will slide to the right, and more air will be delivered. As the coil begins to dehumidify, or take moisture out, the pressure drop will be greater and the system curve will slide back up to its original range. Figure 6-21 shows two system curves for a variable Figure 6-20 Fan Curve and System Curve volume system. In this case, the volume will be 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.

6–34

Fundamentals of Air System Design

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 will reduce the effective pressure available at the fan, and will keep 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.

Figure 6-21

Figure 6-22

Variable Volume System at 3/4 Flow

Variable Volume System at 1/2 Flow

6–35

Air Movers and Fan Technology

The Next Step Having learned how fans work and produce airstreams with static and dynamic pressure, the next chapter covers ducts that distribute the air around the facility.

Summary A fan is an air pump with rotating blades that creates an increase in static and velocity pressure. The two main types of fan are the centrifugal where the air enters the eye of a barrel and is thrown radially out into the spiral scroll, and the axial fan. The fan laws enable one to calculate fan performance with changes in rpm, fan size, and air density. For a specific fan connected to a system, the volume rises with the rpm1, static pressure with rpm2, and power input with rpm3. A fan creates a velocity pressure and 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 on a low velocity system, but probably will not on a higher velocity system. The reason 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 the risk of error, designs should be based on total fan pressure and not just on fan static pressure. Direct drives are used on smaller systems where oversizing the fan is easier than matching the fan to the load. On 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 horsepower 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 horsepower and operating costs.

6–36

Fundamentals of Air System Design

Fan performance data are normally given for Standard Air. At constant speed, the power and pressure vary with gas density, proportional to absolute temperature, R. At altitudes significantly above sea level, the air density drops. A fan will provide the same volume flow, but the same volume will transfer less thermal energy due to the lower air density. The design volume will typically be 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. Due to velocity changes at inlet and outlet connections the use of static pressure alone can cause errors. Particularly with system velocities over 1,500 fpm, it is important to work with total pressure not just static pressure. Remember, velocity pressure equals (V/4005)2 in. wg. When installed, a fan may not provide the expected flow. This may be due to the system pressure losses being different from design calculations or the fan is being affected by the way the inlet and outlet are configured. A difference in system pressure loss will result in the fan riding its curve until the system curve and fan curve meet. Correction may be possible by fan speed adjustment. 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 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 even 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 will be 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 will cause an additional static pressure loss, which 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 will be 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.

6–37

Air Movers and Fan Technology

Bibliography 1. AMCA. 1985. Standard 210, Laboratory Methods of Testing Fans for Rating. Arlington Heights, IL: Air Movement and Control Association Inc. 2. ASHRAE. 1999. ASHRAE Standard 51, Laboratory Methods of Testing Fans for Rating. Atlanta, GA: ASHRAE. 3. AMCA. 1990. Fans and Systems. Arlington Heights, IL: Air Movement and Control Association Inc. Publication 201-90. ASHRAE HandbookFundamentals and HandbookSystems and Equipment

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

A fan is delivering 6,000 cfm at a pressure of 1.5 in. wg 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) 4,800 cfm, 1.2 in. wg b) 4,800 cfm, 0.96 in. wg c) 3,840 cfm, 0.96 in. wg d) 3,840 cfm, 1.2 in. wg. e) None of the above

6-2.

Given a fan operating at 4,000 cfm, 3 in. wg total pressure, and 2.5 hp, what is the fan total efficiency? a) 85% b) 80% c) 75% d) All of the above e) None of the above

6-3.

Given a fan operating at 4,000 cfm, using 1.5 hp, what is the fan total efficiency? a) 85% b) 75% c) 65% d) None of the above

6-4.

What is one effective duct length for a duct with a duct velocity of 4,000 fpm and an area of 200 in.2? a) 80 ft b) 3.3 ft c) 5.66 ft d) None of the above e) Cannot be determined from the information given

6–38

Fundamentals of Air System Design

6-5.

What is one effective duct length for a duct with a duct velocity of 2,000 fpm and an area of 225 in.2? a) 3.5 ft b) 3.0 ft c) 52.3 ft d) None of the above e) Cannot be determined from the information given

6-6.

A rectangular duct is 10 in. high and 20 in. wide. What is the equivalent duct diameter of this duct? a) 200 in.2 b) 254 in. c) 16 in. d) None of the above e) Cannot be determined from the information given

6-7.

For any given system, the system effect factor is constant across the range of flow volumes of the fan. a) True b) False c) Cannot be determined from the information given

6-8.

A fixed fan system is drawing 3 hp to deliver 10,000 cfm. If the air flow requirement can be reduced to 7,000 cfm by decreasing the fan speed, the horsepower requirement will be reduced to: a) 2.1 hp b) 1.0 hp c) 0.44 hp d) All of the above e) None of the above f) Cannot be determined from the information given

6-9.

The ___________________ is the highest efficiency centrifugal fan design. a) Radial b) Forward-curved c) Backward-inclined, backward-curved d) All of the above e) None of the above

6-10.

Power roof ventilators ___________ : a) Usually operate without discharge ductwork b) Operate at low pressure c) Operate at high volume d) All of the above e) None of the above

6–39

Fundamentals of Air System Design

Chapter 7

Duct System Design Contents of Chapter 7 • • • • • • • •

7.1 Duct System Design Overview 7.2 Duct Materials 7.3 Duct Construction 7.4 Duct Design and Sizing 7.5 Sample Systems Summary Bibliography Skill Development Exercises for Chapter 7

7–1

Duct System Design

Study Objectives of Chapter 7 After completing this chapter, you should be able to: •

Layout and size a simple duct system that will transport the required quantity of air from the fan to the conditioned space using appropriate methods and materials; and

•

Calculate the pressure losses in a duct system.

7.1 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 presents duct construction, system design considerations and calculating a system’s frictional and dynamic resistance to air flow.

7.2 Duct Materials A variety of materials are used in the construction of ducts. Selection of duct materials should receive the same careful consideration as the 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 considered. Materials used for ducts include: galvanized steel, carbon (black) steel, aluminum, stainless steel, 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. Duct sizing and construction specifications are generally stated in terms of use of galvanized steel, and correction factors for other materials must be used. Unless otherwise noted, this chapter will consider 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 liner, pressure sensitive tapes, sealants, reinforcements and hangers. 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.

7–2

Fundamentals of Air System Design

Table 7-1 Duct Materials1 Material

Applications

Galvanized Steel Widely used for most air handling applications. Not recommended for corrosive product handling or temperatures above 400°F. Carbon Steel Breechings, flues, stacks, hoods, other (Black Iron) high temperature duct systems, kitchen exhausystems, ducts requiring paint or special coatings Aluminum Duct systems for moisture-laden air, louvers, special exhaust systems, ornamental duct systems. Often substituted for galvanized steel in HVAC duct systems. Stainless Steel Duct systems for kitchen exhaust, moisture-laden air fume exhaust

Copper

Advantages

Limitations

High strength, rigidity, durability, rust resistance in ordinary conditions, availability, nonporous, workability High strength, rigidity, durability, availability, paintability, weldability, non-porous

Weldability, paintability, weight, corrosion resistance Corrosion resistance, weight

Weight, resistance to some forms Low strength, material of corrosion, availability cost, weldability, thermal expansion

High resistance to many common forms of corrosion (but care is definitely required in alloy selection) Duct systems for exposure to outside ele- Accepts solder readily, durable, ments and moisture-laden air resists corrosion, non-magnetic

Fiberglass Reinforced Plastic (FRP)

Chemical exhaust, scrubbers, underground duct systems

Corrosion resistant, ease of modification

Polyvinyl Chloride (PVC) Polyvinyl Steel (PVS)

Exhaust systems for chemical fumes and hospitals, underground duct systems Underground duct systems, moistureladen air, corrosive air systems

Corrosion resistance, weight, weldability, ease of modification Corrosion resistance, weight, workability fabrication, rigidity

Concrete

Underground ducts, air shafts

Rigid Fibrous Glass

Interior HVAC low-pressure duct systems

Compressive strength, corrosion resistance (steel reinforcement in concrete must be properly treated) Weight, thermal insulation and vapor barrier, acoustical qualities, ease of modification, inexpensive tooling for fabrication

Gypsum Board

Ceiling plenums, corridor ducts, airshafts

Cost, availability

Material cost, workability, availability

Cost, electrolytic action if in contact with galvanized steel, thermal expansion, stains Cost, weight, range of chemical and physical properties, brittleness, fabrication, code acceptance Cost, fabrication, code acceptance Susceptible to coating damage, temperature limitations (250°F max.), weldability, code acceptance Cost, weight, porous, fabrication (requires forming processes) Cost, susceptible to damage, system pressure, code acceptance, questionable cleanability Weight, code acceptance, leakage, deterioration when damp

7–3

Duct System Design

7.3 Duct Construction DUCT TYPES Rectangular metal ducts. Table 7-2 lists construction requirements for rectangular steel ducts. It shows that construction requirements are determined by duct thickness, duct dimension and duct pressure. The table indicates what kind of reinforcing (if any) is required for any given combination of these factors. Combinations of factors that are not allowed are also indicated. For 4 in. wg and higher systems, this is for positive pressure only. For negative pressure systems with internal duct wall supports, consult the SMACNA Industrial Duct Construction standards.1 Reinforcing is indicated by a letter and number (for example, D-10). The letter indicates the rigidity class of the required reinforcing, and the number indicates the spacing of the reinforcements, in feet. Specifications for rigidity class and transverse joint reinforcement are shown in Table 7-3. Tables 7-2 and 7-3 are samples of the many tables detailing duct construction requirements. For additional detail, consult the ASHRAE Handbook–HVAC Systems and Equipment.2 The SMACNA publication HVAC Duct Construction Standards–Metal and Flexible gives the functional criteria on which Tables 7-2 and 7-3 are based.3 Transverse joints and, when necessary, intermediate structural members are designed to reinforce the duct system. Ducts larger than 96 in. 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. Where fitting curvature or internal member attachments provide equivalent rigidity, such features may be credited as reinforcement.

7–4

Fundamentals of Air System Design

Table 7-2

Rectangular Ferrous Metal Ducts for Commercial Systems

Table Notes a  Table 7-2 is reproduced from the 1992 ASHRAE Handbook-Systems and Equipment, and is based on Tables 1-3 through 1-9 in HVAC Duct Contruction StandardsMetal and Flexible (SMACNA 1985). For tie rod details, refer to this standard. b  For a given duct thickness,

numbers indicate maximum spacing (feet) between duct reinforcement; letters indicate type (rigidity class) of duct reinforcement. Transverse joint spacing is unrestricted on unreinforced ducts. To qualify joints on reinforced ducts, select transverse joints from Table 7-3. Tables are based on steel construction. Designers should specify galvanized, uncoated or painted steel joint and intermediate reinforcement. Use the same metal duct thickness on all duct sides. Evaluate duct reinforcement on each duct side separately. When required on four sides, for +4, +6 and +10 in. wg systems, corners must be tied. When required on two sides, corners must be tied with rods or angles at the ends for +4, +6 and +10 in. wg systems. Duct sides over 18 in. width with less than 0.0356 in. thickness, which have more than 10 ft2 of unbraced panel area, must be cross-broken or beaded, unless they are lined or insulated externally. Lined or externally insulated ducts are not required to have crossbreaking or beading. continued on next page

7–5

Duct System Design

Table 7-2 (cont.) Table Notes, cont. c The reinforcement tables are

based on galvanized steel of the indicated thickness. They apply to galvanized, painted, uncoated and stainless steel whenever the base metal thickness is not less than 0.0015 in. below that indicated for galvanized steel.

d  Blank spaces indicate that no reinforcement is required. e  See SMACNA’s publication for

alternative reinforcements using tie rods or tie straps for positive pressure.

f  Sheet metal 0.0466 in. thick is acceptable.

g  Tie rods with a minimum diameter of 0.375 in. (or 0.25 in. if the maximum length is 36 in.) must be used on these constructions. The rods for positive pressure ducts are spaced a maximum of 60 in. apart along joints and reinforcements.

7–6

Fundamentals of Air System Design

Table 7-3

Transverse Joint Reinforcement

7–7

Duct System Design

Pressure classification in relation to the fan curve must be considered, especially with VAV systems where the dampers may throttle the air flow, raising the duct pressure. 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 7-4 shows the SMACNA Duct Pressure Classification scheme. Table 7-4 SMACNA Duct Pressure Classifications3 Static Pressure Pressure Class Operating Pressure 0.5 in. wg 1 in. wg 2 in. wg 3 in. wg 4 in. wg 6 in. wg 10 in. wg

Up to 0.5 in. wg Over 0.5 in. wg to 1 in. wg Over 1 in. wg to 2 in. wg Over 2 in. wg to 3 in. wg Over 3 in. wg to 4 in. wg Over 4 in. wg to 6 in. wg Over 6 in. wg to 10 in. wg

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 factoryapplied 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 2,400 fpm and ±2 in. wg. 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 and compression of these ducts significantly increases air flow resistance, they should be kept as short as possible, and fully extended. For further information on fibrous glass ducts, consult the SMACNA publication Fibrous Glass Duct Construction Standards4, or manufacturers’ construction standards.

7–8

Fundamentals of Air System Design

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. It should be realized that 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.

7.4 Duct Design and Sizing HVAC system duct design follows after the room loads and desired air quantities have been determined. 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 will affect the type of air handling apparatus selected.

AIR DISTRIBUTION First, locate the supply air outlets, and then select the size and type required for proper air distribution in each conditioned space (refer to Chapter 3 of this course). 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’s occupied zone (floor to 6 ft 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 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 air flow throw, drop, air pattern, terminal velocities, acoustics, ceiling heights, etc.

7–9

Duct System Design

Supply outlets on the same branch should be chosen with approximately the same pressure loss (no more than 0.05 in. wg 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 ASHRAE Handbook–HVAC Systems and Equipment and to air distribution equipment manufacturers’ application engineering data. However, some of the basic procedures used in the selection of air distribution equipment are: •

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 when used with systems such as a variable air volume (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 will be required for both perimeter zones and interior zones. In general, the exterior zone will be divided into zones that will be 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 would 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.

PRELIMINARY LAYOUT The next step is to draw a preliminary schematic diagram for the ductwork that will convey the design air quantity to the selected zones and outlets by the most efficient and economical path. This layout should be made on a reproducible tracing of the architectural floor-

7–10

Fundamentals of Air System Design

plans. By doing this, the designer will have a better feel for the final relationship of air terminals, branch ducts, main ducts, risers and apparatus. This procedure will help 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 air flows throughout the system. If a constant volume system is chosen, it will be the arithmetic sum of the cfm of each terminal (including branches) working back from the end of the longest run to the fan. However, if a VAV system is chosen, the designer must apply the proper diversity factors to allow a summarization of the peak design air flows 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. This is to size the ductwork properly, and to enable the designer to evaluate the effect of the total HVAC system design, balancing the proper proportions of supply air to return air, exhaust air and outside makeup air. Pressure losses due to fittings and transitions must also be included in the calculation.

DUCT SIZING Having completed the preliminary HVAC system duct layout, the designer will then proceed to use one of the methods for sizing the duct system discussed later in this chapter. Generally, these methods will give the equivalent round duct sizes and the pressure losses for the various elements of the duct system. The designer will then incorporate this information into the preliminary duct layout. If round ductwork is to be used throughout, the duct sizing efforts are completed, providing the ductwork will physically fit 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 will fit 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. A further consideration in the sizing and routing of a ductwork system is the space and access requirements for air terminals, mixing boxes, VAV boxes, fire and smoke dampers, balancing dampers, reheat coils and other accessories.

7–11

Duct System Design

DESIGN METHODS No single design method will automatically provide the most economical duct system for all conditions. A careful evaluation of all cost variables entering into a duct system should be made with each design method or combination of methods. The 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. The careful use of these methods will allow the designer to efficiently size HVAC duct systems for larger residences, 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 foot 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 air flow, thus reducing the possibility of generating noise (against the air flow 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 7-1) has the pressure drop in in. wg per 100 ft, with the shaded area indicating the suggested design limits. Many designers use 0.1 in. wg per 100 ft for ductwork with no acoustic treatment. For systems with VAV boxes, which provide a measure of sound attenuation, 0.2 in. wg per 100 ft might be used from the supply from fan to VAV boxes and dropping to 01 in. wg per 100 ft 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. The beginning of

7–12

Fundamentals of Air System Design

Figure 7-1

Friction Chart

7–13

Duct System Design

such a table is shown in Table 7-5. The reason for including velocity and velocity pressure will become obvious when calculating the pressure drop through fittings in the ductwork. Table 7-5 Sample Data for Duct Sizing* Flow (cfm)

Diameter (in.)

Velocity (fpm)

Velocity Pressure (in. wg)

50 100 200

5 6 7

480 580 630

0.35 0.38 0.40

* At 0.1 in. wg per 100 ft

The Friction Chart is for round duct, but often rectangular duct must be used. For equal flow and pressure drop, the equivalent rectangular duct can be read from a table such as Table 7-6. Note that the velocity will be lower in the equivalent rectangular duct.

7–14

Fundamentals of Air System Design

Table 7-6 Equivalent Round and Rectangular Duct Sizes Circular Duct Diameter (in.) 5 5.5 6 6.5 7 7.5 8 8.5 9 9.5 10 10.5 11 11.5 12 12.5 13 13.5 14 14.5 15 16 17 18 19 20 21 22 23 24 25 26 27 28 29 30 31 32 33 34 35 36 37 38 39

Length One Side Of Rectangular Duct (a), in. 4 5 6 8 9 11 13 15 17 20 22 25 29 32

5

5 6 7 8 10 11 13 15 17 19 21 23 26 29 32 35 38

6

6 7 8 9 10 12 13 15 16 18 20 22 24 27 29 32 35 38 45

7 8 9 10 12 14 16 18 20 Length Adjacent Side of Rectangular Duct (b), in.

7 8 9 10 11 12 14 15 17 18 20 22 24 26 28 30 36 41 47 54

8 9 10 12 13 14 15 17 18 20 22 24 25 30 34 39 44 50 57 64

9 10 11 12 13 15 16 17 19 20 22 25 29 33 38 43 48 54 60 66

10 11 12 13 14 15 17 18 19 22 25 29 33 37 41 46 51 57 63 69 76

12 13 14 15 16 18 20 23 26 29 33 36 40 44 49 54 59 64 70 76 82 89 96

14 15 17 19 22 24 27 30 33 36 40 44 48 52 56 61 66 71 76 82 88 95 101 108

16 17 19 21 23 26 28 31 34 37 40 43 47 51 55 59 64 68 73 78 83 89 95

18 19 20 23 25 27 29 32 35 38 41 44 47 51 54 58 62 67 71 76 80

20 22 24 26 28 31 33 36 39 41 44 48 51 54 58 62 66 70

22

24

22 24 26 28 30 32 35 37 40 42 45 48 51 55 58 62

24 25 27 29 31 34 36 38 41 44 46 49 52 55

7–15

Duct System Design

The whole business of reading off 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 100 ft and read off the other corresponding variables. For example, setting the pressure drop at 0.2 in.wg and volume at 4,000 fpm, you can read off 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 air flow 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, allowing 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 will stay 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. Note: The loss coefficients for duct fittings found in the ASHRAE Handbook–Fundamentals include static pressure regain or loss for the velocity condition changes that occur at divided flow or change-of-size duct fittings.7 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. 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 ASHRAE Handbook–Fundamentals chapter on

7–16

Fundamentals of Air System Design

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, variable air volume box, etc.), extended as a plenum to serve multiple outlets and/or branch ducts with essentially equal pressure. A semi-extended plenum is a trunk duct system utilizing the concept of the extended plenum incorporating a minimum number of size reductions. This modification can be used with equal friction and static regain design methods. Some of the advantages may be: lower first-costs, lower operating costs, ease of balancing, and adaptability to branch duct or outlet changes. A disadvantage is that low air flow velocities could result in additional heat gain or loss to the airstream through the duct walls. 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 can use 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 the SMACNA HVAC Duct Construction Standards-Metal and Flexible3 and the SMACNA HVAC Air Duct Leakage Test Manual.8 Leakage in ducts varies with the fabricating machinery 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.

7–17

Duct System Design

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 = Cp S

(7-1)

where: Q = leakage rate, cfm C = constant reflecting area characteristics of leakage path ps = static pressure differential from duct interior to exterior, in. wg N = exponent relating turbulence or laminar flow in leakage path Analysis of the AISI/ASHRAE/SMACNA/TIMA 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: 0.65

Q = C L p S

(7-2)

where: Q = Leakage rate per unit surface area, cfm/100 ft2 CL = Leakage class, cfm per 100 ft2 duct surface at 1 in. wg static pressure Figure 7-2 shows how duct pressure affects the leakage rate for each leakage class. ASHRAE Standard 90.1 prescribes minimum sealing requirements for supply, return, and exhaust ducts run outside, in conditioned spaces and unconditioned spaces.5 Specifying allowable leakage rates of less than CL3 should be avoided due to both cost and difficulty. Leakage class is defined as the cfm leaked per 100 ft2 of duct surface area at 1 in. wg. A selected series of leakage classes based on Equation 7-2 is shown in Figure 7-2. Table 7-7 is a summary of the leakage class attainable for good 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.

7–18

Fundamentals of Air System Design

Figure 7-2

Duct Leakage Classifications1

Table 7-7 Applicable Leakage Classes1 Duct Class Applicable Sealing Rectangular Metal Round and Oval Metal Rectangular Fibrous Glass Round Fibrous Glass

0.5, 1, 2 in.wg

3 in. wg

4, 6, 10 in. wg

N/A

Transverse Joints Only

Transverse Joints and Seams

All Joints, Seams and Wall Penetrations

48

24

12

6

30

12

6

3

N/A

6

N/A

N/A

N/A

3

N/A

N/A

7–19

Duct System Design

EXAMPLE 7.1 Question: Given the system shown below, with average pressure of 2.5 in. wg, find the leakage, in cfm, of the supply ductwork between points A and B.

Answer: From Table 7-6, it is determined that the leakage class for a 3 in. wg duct class round metal duct is 6. Using Figure 7-2, the leakage factor is determined to be 10.6 cfm/100 ft2. The 30 in. diameter duct from A to B has 785 ft2 of duct surface, and the leakage is determined using Equation 7-2: 10.6 cfm -  785 ft 2 = 83 cfm Leakage = -------------------2 100 ft

(7-3)

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 ASHRAE Handbook–Fundamentals. In this section, some of the 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 air flows are proportioned properly. The air in the ductwork still gets warmer or cooler as it passes through the conditioned space, thus decreasing the temperature difference. As a result, less air is 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

7–20

Fundamentals of Air System Design

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 Standard 90.1 requires thermal insulation of all duct systems and their components (such as ducts, plenums and enclosures) installed in or on buildings.5 To estimate duct heat transfer and entering or leaving air temperatures, use Equations 7-4, 7-5 and 7-6: UPL  t e + t 1 Q 1 = -----------  -------------- – t a 12  2 

(7-4)

t 1  y + 1  – 2t a t e = --------------------------------y – 1

(7-5)

t e  y – 1  + 2t a t 1 = -------------------------------y + 1

(7-6)

where: y = 2.4 AV/UPL for rectangular ducts = 0.6 DV/UL for round ducts A = cross-sectional area of duct, in.2 V = average velocity, fpm D = diameter of duct, in. L = duct length, ft Ql = heat loss/gain through duct walls U = overall heat transfer coefficient of duct wall, Btu/hft2F P = perimeter of bare or insulated duct, in.  = density, lbm /ft3 te = temperature of air entering duct, °F tl = temperature of air leaving duct, °F ta = temperature of air surrounding duct, °F Use Figure 7-3 to determine the U-values for insulated and uninsulated ducts. For a 2 in. thick, 0.75 lb/ft3 fibrous glass blanket compressed 50% during installation, the heat transfer rate increases approximately 20%, as shown in Figure 7-3. Pervious flexible duct liners also influence heat transfer significantly, as shown in Figure 7-3. At 2,500 fpm, the pervious liner U-value is 0.33 Btu/hft 2F. For an impervious liner, the U-value is 0.19 Btu/hft2F.

7–21

Duct System Design

Figure 7-3

7–22

Heat Transfer Coefficients

Fundamentals of Air System Design

EXAMPLE 7.2 Example 7.2 is adapted from SMACNA HVAC Systems Duct Design, p. 5.27–28. Question: A 65 ft length of 24 in.  36 in. uninsulated sheet metal duct, freely suspended, conveys heated air through a space maintained above freezing at 40°F. Based on heat loss calculations for the heated zone, 17,200 cfm of standard air at a supply air temperature of 122°F is required. The duct is connected directly to the heated zone. Determine the required air temperature entering the duct, and the duct heat loss. Answer: a. Calculate the duct velocity: V = cfm -------- = ---------------------------------------------------------------------------= 17 200 cfm 2 2 A  24 in.  36 in.    144 in.  ft 

2900 fpm

Select U = 0.73 Btu/hft2F (from Figure 7-3) Calculate P = 2(24 in. + 36 in.) = 120 in. 3

2.4   24 in.   36 in. =   2900 fpm   0.075 lb/ft  y =  2.4AV    UPL  = --------------------------------------------------------------------------------------------------------------- 0.73   120 in.   65 ft 

79.2

b. Calculate the entering air temperature: 122F  79.2 + 1  –  2  40F  t e = ------------------------------------------------------------------------ = 124.1F  79.2 – 1  c. Calculate the duct heat loss:  0.73   120 in.   65 ft  124.1F + 122F Q 1 = ------------------------------------------------------   ----------------------------------------- – 40F   12 2

= 39 200 Btu/h

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 tables such as the ones found in the ASHRAE Handbook–Fundamentals. These loss coefficients represent the ratio of total pressure loss to the dynamic

7–23

Duct System Design

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 (pt) loss of a fitting is determined using the loss coefficient in the following equation: p t = C o  p v

(7-7)

where: pt = total pressure loss (in. wg) Co = Dimensionless local loss coefficient Pv = Velocity pressure (in. wg) 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 will be the closest possible approximation of the actual system total pressure requirements for the fan

EXAMPLE 7.3 Question: To demonstrate the use of the loss coefficient tables, assume a velocity of 2,500 fpm in a 9 in. 7 Gore (segment), 90° elbow, as shown in Figure 7-4. According to the Figure 7-4 table, Co for this fitting is 0.10.

Figure 7-4

CD3-10 Elbow

CO Values for CD3-10 Elbow D (in.) Co

3 0.16

6 0.12

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

7–24

9 0.10

12 0.08

15 0.07

18 0.06

27 0.05

60 0.03

Fundamentals of Air System Design

Answer: Using Equation 1-8, we determine that the velocity pressure is 0.39 in. wg: V 2 2500 2 p v =  ------------  =  ------------  = 0.39 in. wg  4005   4500  Using Equation 7-7, we determine that the total pressure loss is 0.039 in. wg: p t = C o  p v = 0.10  0.39 =

0.039 in. wg

7.5 Sample Systems The following two simplified sample systems illustrate how to calculate pressure drop in “real” systems.

SYSTEM 1 As depicted in Figure 7-5, 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 1 in. wg pressure drop at 4,000 cfm. The fan speed will be adjusted to deliver 4,000 cfm. What is the pressure drop in the system?

Figure 7-5

System 1

7–25

Duct System Design

Solution: 1. Calculate the circular equivalent of the rectangular duct using the formula or use Table 7-6.   ab  0.625  - D e = 1.30  -------------------------  a + b  0.250

= 12.9

where: De = circular equivalent of rectangular duct for equal length, fluid resistance, and air flow, in. a = length of one side of duct, in. b = length of adjacent side of duct, in. 2. Calculate the velocity of the airstream in the duct: 4000 ------------------------ = 4114 fpm  14 ------  10 ------  12 12 2

2  De  12.9 -  --------------------------------Ae =   = 0.9076  =  4  144  4  144

Q- = ----------------------4000 V = ---= cfm2 Ae 0.9076 ft

4407 fpm

3. Calculate the fan outlet velocity pressure from Equation 1-8:  V  - p v =  ----------4005  

2

 4407 - =  ----------4005  

2

= 1.211 in. wg

4. Find the pressure drop in the duct per 100 ft. using the Friction Chart shown in Figure 7-1 (also in the ASHRAE Handbook–Fundamentals): 1.9 in. wg per 100 ft 5. Calculate the pressure drop in the 30 ft run of duct: 30 ft------------ 1.9 in. wg = 0.57 in. wg 100 ft

7–26

Fundamentals of Air System Design

6. Add the pressure drops: Pressure drop in 30 ft duct run 0.57 in.wg Fan outlet velocity pressure 1.21 in.wg Required VAV outlet pressure drop 1.00 in. wg Total pressure drop required at fan

2.78 in. wg

SYSTEM 2 Figure 7-6 is the same system as in Figure 7-5, except that the outlet at the end of the duct run was removed, and a 20° 14 10 in. to 24 10 in. rectangular transition was added. Attached to the transition outlet are two 12  10 in. elbows with r = 1.5 (Co = 0.2). Attached to each elbow is a branch duct. One branch is 17 ft long, the other branch is 12 ft long. There is a balancing damper in the 12 ft branch. At each end of the duct extension is a VAV box with a 1 in. wg pressure drop at 2,000 cfm. What is the pressure drop in the system, and how should the balancing damper be adjusted to equalize the pressure in both branch runs?

Figure 7-6

System 2

7–27

Duct System Design

Solution: 1. Calculate the pressure drop, velocity and pressure at the end of the 30 ft run. Because this is the same configuration as System 1, on the main 30 ft run, the pressure drop is 0.57, the velocity is 4,407 fpm, and the velocity pressure is 1.211 in., the 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: Inlet area = 14 in.  10 in. = 140 in.2 Outlet area = 24 in.  10 in. = 240 in.2 Outlet area- = 240 --------------------------------- = 1.7 Inlet area 140 b. Refer to the table below (from the ASHRAE Handbook–Fundamentals), which gives Co values for rectangular transitions. Because the table does not give an exact value for an outlet/inlet ratio of 1.71, by interpolation, the Co value is estimated to be 0.43. Multiply the pressure at the inlet by the Co value to calculate the pressure drop across the transition: 1.211 in. wg  0.43 = 0.52 in. wg

7–28

Fundamentals of Air System Design

SR4-1 Rectangular Transition*

Ao/A1

10

15

20

Co Values  30 45 60

0.10 0.05 0.05 0.05 0.05 0.17 0.05 0.04 0.04 0.04 0.25 0.05 0.04 0.04 0.04 0.50 0.06 0.05 0.05 0.05 1.00 0.00 0.00 0.00 0.00 2.00 0.56 0.52 0.60 0.96 4.00 2.72 3.04 3.52 6.72 10.00 24.00 26.00 36.00 53.00 16.00 66.56 69.12 102.40 143.36

0.07 0.05 0.06 0.06 0.00 1.40 9.60 69.00 181.76

0.08 0.07 0.07 0.07 0.00 1.48 10.88 82.00 220.16

90

120

150

180

0.19 0.18 0.17 0.14 0.00 1.52 11.20 93.00 256.00

0.29 0.28 0.27 0.20 0.00 1.48 11.04 93.00 253.44

0.37 0.36 0.35 0.26 0.00 1.44 10.72 92.00 250.88

0.43 0.42 0.41 0.27 1.00 1.40 10.56 91.00 250.88

*Two sides parallel, symmetrical. supply air systems

3. Calculate the pressure drop in the elbows. The calculation for each elbow is the same. The velocity of the airstream entering each elbow is: V = Q ---- = ------------------------------------------= 4000 A  2   12  10  144 

2400 fpm

a. Convert the 12 in. x 10 in. rectangular duct to circular measurement (or use the table of equivalents found in the ASHRAE Handbook–Fundamentals): 0.625

0.625

 ab   12  10  D e = 1.30 -----------------------= 1.30 -------------------------------= 0.25 0.25 a + b  12 + 10 

12.0

b. Calculate the velocity pressure in the elbows:  2400 ------------  4005

2

= 0.359 in. wg

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: 0.359 in. wg  0.2 = 0.072 in. wg 4. Calculate the pressure drop in the 17 ft branch, the longest branch.

7–29

Duct System Design

a. Find the pressure drop in the duct per 100 ft. using the Friction Chart in Figure 7-1: 0.6 in. wg per 100 ft b. Calculate the pressure drop for the 17 ft branch run: 17 ft------------ 0.6 in. wg = 0.102 in. wg 100 ft 5. The loss of the VAV box is given as 1.0 in. wg (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 balancing damper in the 12 ft branch should be throttled so that the pressure drop in the 12 ft branch is equal to the pressure drop in the 17 ft branch (0.102 in. wg). To calculate the pressure drop required across the damper: a. Calculate the pressure drop in the 12 ft branch run without a damper: 12 ft------------ 0.7 in. wg = 0.084 in. wg 100 ft b. Subtract the pressure drop in the 12 ft branch run from the pressure drop in the 17 ft branch run: 0.119 in. wg – 0.084 in. wg = 0.035 in. wg c. Therefore the balancing damper must be adjusted to obtain a 0.035 in. wg pressure drop to obtain equal flow in each branch. 7. Add the pressure requirements:

7–30

Pressure drop in 30 ft duct run Pressure drop at transition 12 in.  10 in. elbow 17 ft branch duct Required VAV outlet pressure Fan outlet velocity pressure

0.57 in. wg 0.52 in. wg 0.07 in. wg 0.10 in. wg 1.00 in. wg 1.21 in. wg

Total pressure drop

3.47 in. wg

Fundamentals of Air System Design

The Next Step The next chapter 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 plus operating costs. Many materials are used for ductwork, but the vast majority is 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 SMACNA standard specifications for size, static pressure (positive or negative), material thickness, jointing, reinforcing and supports. When choosing the pressure rating, take care to allow for probable maximum and minimum pressures on all but the smallest systems. Round and oval metal 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 access. 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 2,400 fpm and ±2 in. wg. Flexible ducts are typically manufactured from a coiled wire and fabric, and are used for connection of other ducts to diffusers. Ducts must be sufficiently airtight to ensure economical and quiet performance of the system. Leakage classifications are given in cfm/100ft2 at 1 in. wg. Actual leakage is Q = CLp0.65. Many materials, gaskets and tapes are available, but many have an unfortunately short life. ASHRAE 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 tempera-

7–31

Duct System Design

ture of the space the duct runs through, there will be some heat gain or loss which should be approximately included at this stage. For designs to meet 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. Due to variations in loads, the HVAC system will be zoned. Typically, interior and exterior spaces will be on separate zones, and the duct layout must accommodate the zoning and associated air control devices. Next the preliminary layout is drawn, ideally over the architectural layout for supply, return and exhaust ducts. Some very preliminary sizing will be done at this stage to ensure adequate space for the main duct runs. At this stage, extra space for duct joints and insulation must be included, as well as allowing for the other services. The need to accommodate all three may significantly influence the final choice of layout when crossovers and available space are evaluated. With the preliminary HVAC system duct layout done, 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, 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 will often be changed somewhat as the installation contractor deals with coordination with other trades. With the equal friction method, a fixed pressure drop per 100 feet is chosen and used to do the duct sizing. This 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 do manually and may need to be modified in cases where branch ducts are very different in length.

7–32

Fundamentals of Air System Design

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  semi-extended plenum  keeps the duct size up for a greater length than necessary often reducing the cost of numerous size reductions. In the velocity reduction method, velocity reductions are chosen by experienced designer. It can be considered as the constant pressure drop method improved by experience. 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.

Bibliography 1. SMACNA. 2004. Rectangular Industrial Duct Construction Standards. Chantilly, VA: Sheet Metal and Air Conditioning Contractors' National Association Inc. 2. SMACNA. 1990. HVAC Systems–Duct Design. Chantilly, VA: Sheet Metal and Air Conditioning Contractors' National Association Inc. 3. SMACNA. 2005. HVAC Duct Construction Standards–Metal and Flexible. Chantilly, VA: Sheet Metal and Air Conditioning Contractors National Association Inc. 4. SMACNA. 2003. Fibrous Glass Duct Construction Standards. Chantilly, VA: Sheet Metal and Air Conditioning Contractors' National Association Inc. 5. ASHRAE. 2004. ASHRAE/IESNA Standard 90.1-2004, Energy Efficient Design of New Buildings Except Low-Rise Residential Buildings. Atlanta, GA: ASHRAE. 6. SMACNA. 1985. HVAC Air Duct Leakage Test Manual. Chantilly, VA: Sheet Metal and Air Conditioning Contractors' National Association Inc. ASHRAE HandbookHVAC Systems and Equipment, duct construction; HandbookFundamentals, duct design and sizing.

7–33

Duct System Design

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

7–34

7-1.

As depicted in the figure below, this system 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 1 in. wg pressure drop at 2,500 cfm. The fan speed will be adjusted to deliver 5,000 cfm. The Co value of the elbow is 0.2. What is the total pressure drop in the system? a) 3.2 in. wg b) 1.8 in. wg c) 1.6 in. wg 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 e) None of the above

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) All of the above e) None of the above

Fundamentals of Air System Design

7-4.

Generally the most efficient and economical ducts for air systems are: a) Rectangular b) Oval c) Round d) All of the above e) None of the above

7-5.

Duct systems of rectangular fibrous glass are generally limited to: a) 2,400 fpm and ±2 in. wg b) 4,000 fpm and ±3 in. wg c) 1,000 fpm and ±3 in. wg d) None of the above

7-6.

Compression of flexible ducts significantly decreases air flow resistance. a) True b) False c) Cannot be determined from the information given

7-7.

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

7–35

Fundamentals of Air System Design

Chapter 8

Codes and Standards Contents of Chapter 8 • • • • • • • •

8.1 Building Code Requirements 8.2 ASHRAE Standard 90.1-2007 8.3 ASHRAE Standard 62.1-2007 8.4 Other Codes and Standards 8.5 Sources of Information Summary Bibliography Skill Development Exercises for Chapter 8

8–1

Codes and Standards

Study Objectives of Chapter 8 After completing this chapter, you should be able to list the principle codes and standards affecting air system design, and briefly state what they cover and why they are important: • • • • • •

ASHRAE Standard 90.1-2007 ASHRAE Standard 62–20047 NFPA 90A NFPA 90B NFPA 96 SMACNA HVAC Duct Construction Standards

8.1 Building Code Requirements In the private sector, each new construction or renovation project is normally governed by state laws or local ordinances 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 HVAC duct systems. However, Figure 8-1 may not list all applicable regulations and standards for a specific locality. Specifications for federal government construction are promulgated by such agencies as the Federal Construction Council, General Services Administration, Department of Defense, Department of Energy, and by Executive Orders. Model code changes require long cycles for approval by the consensus process. Because the development of safety codes, energy codes and standards proceed independently, the most recent edition of a code or standard may not have been 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.

8–2

Fundamentals of Air System Design

Figure 8-1

Hierarchy of Building Codes and Standards

8–3

Codes and Standards

8.2 ASHRAE Standard 90.1-2007 Codes and standards have become much more important. With the 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 Standard 90.1-2007, Energy Efficient Design of New Buildings Except Low-Rise Residential Buildings.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, but it was split into 90.1 for all but low-rise residential buildings and 90.2 for low-rise residential buildings. Standard 90.1 is referenced in the Energy Act of 1992 and has been revised into code language to make it code enforceable. The standard is on ANSI continuous maintenance; 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 every three years, with the latest printed revision in 2007. The original 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 Standard 90.1 has been adopted by model building codes. The standard’s format is intended to be general and flexible, so it may be applied to many different climates, building types and HVAC system types. The standard deals with all aspects of building energy use including the Building Envelope in Section 5, Service Water Heating in Section 7, Power in Section 8, and Lighting in Section 9. 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 energy costs and consumption of HVAC systems; a poorly designed system can easily 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.

8–4

Fundamentals of Air System Design

An efficient system will minimize energy use by minimizing system losses, maximizing equipment efficiencies, using free heating/cooling, and recovering heat where possible. A very efficient system could have an overall efficiency greater than one. Section 6 addresses 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 over 5,000 cfm and 70% outside air

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 lifecycle costing is used to determine that proposed alternates have an economic payback.

COMPLIANCE METHODS There are three primary subsections in Section 6. First, there is a simplified approach for smaller buildings with simple HVAC systems. Then there are mandatory requirements in Section 6.4 that must be met for either compliance path. Lastly, the prescriptive requirements in Section 6.5 include measures that must be met to show compliance via the prescriptive method. In this prescriptive method, the designer must choose equipment with required performance and obey a number of design requirements.

8–5

Codes and Standards

These prescriptive requirements do not have to be met with the energy cost budget method, which is detailed in Section 11. In the energy cost 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 this section may seem overwhelming to designers of simpler, single zone HVAC systems that are typically used in one- or two-story buildings under 25,000 ft2.

8.3 ASHRAE Standard 62.1-2007 ASHRAE Standard 62.1-2007, Ventilation for Acceptable Indoor Air Quality, was briefly introduced in Chapter 3 as the standard that sets minimum outside 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 deal with the following questions: •

How much outside air is required in each space?

•

How will the differing requirements for each space be achieved?

•

When VAV 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

•

Duct construction

•

System startup and balancing

After construction, the standard has requirements for the system’s ongoing operation and maintenance, including inspection and measurement of outdoor air flow.

8–6

Fundamentals of Air System Design

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

NATIONAL FIRE PROTECTION ASSOCIATION The National Fire Protection Association (NFPA) issues a wide range of standards. Three of interest to HVAC designers are NFPA 90A, NFPA 90B and NFPA 96.2-4 NFPA 90A–Installation of Air Conditioning and Ventilating Systems applies to systems for air movement in: •

Structures over 25,000 ft3 in volume

•

Buildings of Type III, IV and V construction over three stories in height regardless of volume

•

Buildings, spaces, occupants and 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-resistive 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

•

Permit the air duct systems in a building to be used for the additional purpose of emergency smoke control

NFPA 90A provides requirements for HVAC systems (equipment and air distribution), integrating HVAC systems with building construction, controls and acceptance testing.

8–7

Codes and Standards

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 will determine its fire separations and the requirements for duct fire and smoke dampers. Fire dampers are a significant cost, and accessible 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–Installation of Warm Air Heating and Air-Conditioning Systems applies to all warm air heating and air-conditioning systems that serve: one- or two-family dwellings; and spaces not exceeding 25,000 ft3 in volume in any occupancy (for example, light commercial). Other systems are covered by NFPA 90A. Standard NFPA 90B addresses: system components; fire integrity of building construction; equipment; wiring; and controls. NFPA 96–Installation of Equipment for the Removal of Smoke and Grease-Laden Vapors from Commercial Cooking Equipment 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.

8–8

Fundamentals of Air System Design

Figure 8-2

Wall and Partition Penetrations and Smoke Dampers

8–9

Codes and Standards

SHEET METAL AND AIR CONDITIONING CONTRACTORS’ NATIONAL ASSN. The SMACNA HVAC Duct Construction Standards cover: basic duct construction; fittings and other construction; round, oval and flexible duct; 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.5

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. This book includes general information on ventilation and numerous examples of industrial ventilation and the removal of contaminants from specific industrial processes. The 25th edition was published in 2004.

8.5 Sources of Information Many sources of information are available to HVAC designers: •

ASHRAE produces an extensive range of standards, Handbooks and Advanced Energy Design Guides which can be located at the website: www.ashrae.org.

•

National Fire Protection Association (NFPA), 1 Batterymarch Park, Quincy, MA 02269-9101; 617/770-3000, Fax 617/770-0700; website: www.nfpa.org.

•

Sheet Metal and Air Conditioning Contractors’ National Association Inc. (SMACNA), 4201 Lafayette Center Drive, Chantilly, VA 22021-1209; 703/803-2980, Fax 703/803-3732; website: www.smacna.org.

•

American Conference of Governmental Industrial Hygienists (ACGIH), Kemper Woods Center, 1330 Kemper Meadow Dr., Cincinnati, OH 45240; 513/742-2020, Fax 513/742-3355; website: www.acgih.org.

•

Air Movement and Control Association Inc. (AMCA), 30 West University Drive, Arlington Heights, IL 60004-1893; 708/394-0150, Fax 708/253-0088; website: www.amca.org.

In addition, each chapter of the ASHRAE Handbooks contains a detailed bibliography. An extensive list of HVAC codes and standards is included in the ASHRAE Handbook–Fundamentals.6

8–10

Fundamentals of Air System Design

The Next Step The next chapter will discuss some components in air systems including dampers, air filters, humidifiers, duct heaters and duct insulation.

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 Standard 90.1 has become the legal requirement in the USA. The standard is a consensus document with public review. It is adopted by the American National Standards Institute (ANSI), and undergoes continuous improvement through addenda. The latest printed edition was released in 2007. The standard covers the building fabric and all permanent energy-using plant and equipment in the building. Section 6HVAC includes two methods of achieving compliance: •

Prescriptive approach: Follow specific set of requirements, including: minimum requirements for plant efficiency; 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.

•

Energy cost budget method: Design the building to have no greater energy cost than a system designed under the prescriptive approach.

ASHRAE Standard 62-2007 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

•

Documentation requirements

The National Fire Protection Association (NFPA) has three standards applicable to HVAC systems: •

NFPA 90A–Installation of Air Conditioning and Ventilating Systems applies to systems for air movement in larger buildings with the emphasis on life safety.

8–11

Codes and Standards

The focus is on reducing the risk of fire and smoke and their effect when they do occur. •

NFPA 90B–Installation of Warm Air Heating and Air-Conditioning Systems applies to smaller buildings.

•

NFPA 96–Installation of Equipment for the Removal of Smoke and Grease-Laden Vapors from Commercial Cooking Equipment covers kitchen exhausts and fire suppression.

The SMACNA book HVAC Duct Construction Standards covers the design, construction and installation of galvanized ductwork in detail and other materials more generally. The ACGIH book Industrial Ventilation, A Manual of Recommended Practice includes general information on ventilation and numerous examples of industrial ventilation and the removal of contaminants from specific industrial processes.

Bibliography 1. ASHRAE. 2007. ASHRAE/IESNA Standard 90.1-2007, Energy Efficient Design of New Buildings Except Low-Rise Residential Buildings. Atlanta, GA: ASHRAE. 2. ASHRAE. 2007. ASHRAE/IESNA Standard 62.1-2007, Ventilation for Acceptable Indoor Air Quality. 3. NFPA. 2002. NFPA 90A–Installation of Air Conditioning and Ventilating Systems. Quincy, MA: National Fire Protection Association. 4. NFPA. 2006. NFPA 90B–Installation of Warm Air Heating and Air-Conditioning Systems. Quincy, MA: National Fire Protection Association. 5. NFPA. 2004. NFPA 96–Installation of Equipment for the Removal of Smoke and Grease-Laden Vapors from Commercial Cooking Equipment. Quincy, MA: National Fire Protection Association. 6. SMACNA. 2005. HVAC Duct Construction Standards. Chantilly – Metal and Flexible, VA: Sheet Metal and Air Conditioning Contractors' National Association Inc. ASHRAE HandbookFundamentals, codes and standards; HandbookHVAC Systems and Equipment, codes and standards relevant to specific systems and equipment; Handbook Refrigeration, codes and standards; HandbookApplications, codes and standards relevant to specific applications.

8–12

Fundamentals of Air System Design

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

Combustibility and toxicity ratings are normally based on tests of: a) New materials b) Old work c) Fibrous materials d) All of the above

8-2.

In the private sector, new construction is normally governed by: a) State laws b) Local ordinances c) Codes d) All of the above

8-3.

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

8-4.

Which of the following standards applies to structures not exceeding 25,000 ft3 in volume? a) NFPA 90A b) NFPA 90B c) NFPA 96 d) All of the above

8-5.

SMACNA HVAC Duct Construction Standards covers: a) Basic duct construction b) Hangers and supports c) Duct sealing classifications d) All of the above

8-6.

ASHRAE Standard 90.1 has a somewhat easier compliance route for many small air-conditioned buildings. a) True b) False

8-7.

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

8-8.

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

8-9.

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

8-10.

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

8–13

Fundamentals of Air System Design

Chapter 9

Air System Auxiliary Components Contents of Chapter 9 • • • • • • • •

9.1 Dampers 9.2 Air Filters 9.3 Humidifiers 9.4 Duct Heaters 9.5 Duct Insulation Summary Bibliography Skill Development Exercises for Chapter 9

9–1

Air System Auxiliary Components

Study Objectives of Chapter 9 After completing this chapter, you should understand the function, selection and sizing of: • • • • •

Dampers Air filters Humidifiers Duct heaters Duct insulation

9.1 Dampers TYPES OF DAMPERS Two damper arrangements are used for air-handling system flow control: parallelblade and opposed-blade (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 Figure 9-1 Parallel and Opposed Blade Dampers 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 thus called a triple-V blade. Blades

9–2

Fundamentals of Air System Design

are also made in aerofoil section, providing a lower resistance to airflow and lower noise generation. The power to drive the linkage is from an actuator. Optimum control of airflow is obtained with a linear relationship between air flow and the degree to which the damper is open. For many years, the information about damper performance has been: Parallel-blade dampers are adequate for two-position control and can be used for modulating control when they are the primary source of pressure drop and directional air flow is not a problem. Opposed-blade dampers are preferable, because they normally provide better control. The characteristic curves of installed parallel blade dampers and installed opposed blade dampers are shown in Figures 9-2 and 9-3, respectively. The parameter a in both figures 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.

Figure 9-2

Installed Parallel Blade Dampers

Figure 9-3

lnstalled Opposed Blade Dampers

9–3

Air System Auxiliary Components

These idealized curves are correct in concept but not realized in practice. Recent research, particularly ASHRAE Research Report 1157 Flow Resistance and Modulating Characteristics of Control Dampers,1 shows that the performance of dampers is highly dependent on: •

Construction. Differs from manufacturer to manufacturer for the same style, triple-V or aerofoil. See Figure 9-4 where 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 will have air coming from all directions into the 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. See Figure 9-5 for an example where the opposed blade damper characteristic is degraded by being placed inside an inlet louver.

Figure 9-4

9–4

Two Parallel Blade Triple-V Dampers From Different Manufacturers

Fundamentals of Air System Design

Therefore, actual performance data must be obtained from the manufacturers on their specific dampers 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 outside air and return air to provide supply air. The Figure 9-5 Effect of Inlet Louver on an Opposed Blade Damper Characteristic combination of three parallel bladed dampers working in unison provides a more linear control characteristic than when using opposed blade 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 any location where the tight-closing damper will reduce energy consumption significantly. Literature from manufacturers expresses leakage rates when exposed to specific pressure differentials across the closed damper. Details about sizing dampers and leakage are given in the ASHRAE course Fundamentals of HVAC Controls.

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 of the shaft through which a linkage moves the crank arm to open or close the dampers. Normally-open or normally-closed operation refers to the position of the dampers when no air pressure is applied at the operator, the failed position. Positive positioners are important for sequencing the damper with other devices.

9–5

Air System Auxiliary Components

DAMPER FUNCTIONS Dampers have a wide variety of functions: •

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

•

Balancing dampers are used to make final adjustments in the air flow through a duct when the system is first being commissioned. In smaller ducts, balancing dampers are often flat metal plates 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 systems 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, but 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 outside 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.2,3,4

9–6

Fundamentals of Air System Design

9.2 Air Filters The purpose of a filter is to remove contaminants from an airstream. Contaminants may be gaseous, such as odors from the 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 be removed using filters containing chemicals to remove the specific contaminant. Gaseous filtration is a specialized field and is not covered in this course. The most common situation 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. Particle size may be defined in numerous ways. Particles less than 2.5 μm (microns, or millionths of a meter) in diameter are generally referred to as fine, with those greater than 2.5 μm being considered as 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 out in the nasal passage 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 are different than particles in that fiber shape, diameter and density all affect where a fiber will settle in the body.5 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 diseases 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.

9–7

Air System Auxiliary Components

RATING FILTERS In addition to criteria affecting the degree of air cleanliness, factors such as cost (initial investment and maintenance), space requirements and air flow resistance have encouraged 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, air flow 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 for most filters and applications. However, because the efficiency of many drytype filters increases with dust load, in applications with low dust concentrations, the initial (clean filter) efficiency should be considered for design.

•

Air flow resistance (or resistance) is the pressure drop across the filter at a given air flow 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 specified air flow rate to some maximum resistance value or before its performance drops seriously as a result of the collected dust.

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

9–8

•

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 ASHRAE Standard 52.1 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. ASHRAE Standard 52.2 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 0.3 μm to 10 μm. Based on the results, filters are classified into 20 categories called 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 extreme 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.

Fundamentals of Air System Design

•

DOP Penetration Test. This is a U.S. Army specification test (MIL-STD-282) using a chemical, DOP, producing 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 high efficiency filters, MERV 17 to 20.

Table 9-1 describes the common range of particulate filters, typical use, and their performance as tested under ASHRAE Standards 52.1 and 52.2. Read through the table’s Application Guidelines and MERV ratings so you understand the range of filter performance and typical uses. This will provide context for the later discussion on filter selection. Table 9-1 Std. 52.2 MERV Rating

Filter Types and Performance*

Approx. Std. 52.1 Results Dust Spot Efficiency

Arrestance

20

n/a

n/a

19

n/a

n/a

18

n/a

n/a

17

n/a

n/a

16

n/a

n/a

15

>95%

n/a

14

90-95%

>98%

13

80-90%

>98%

12

70-75%

>95%

11

60-65%

>95%

10

50-55%

>95%

9

40-45%

>90%

8

30-35%

>90%

7

25-30%

>90%

6