Fan-application-manual-180629054956.pdf

Fan and Air System Applications Handbook AMCA International

Publication 200 Air Systems

Publication 201 Fans and Systems

Publication 202 Troubleshooting

Publication 203 Field Performance Measurements of Fan Systems

Forward Publication 200

Publication 201

Publication 202

Publication 203

Air Systems

Fans and Systems

Troubleshooting

Field Performance Measurements of Fan Systems

Air Systems is intended to provide basic information needed to design effective and energy efficient air systems. Discussion is limited to systems where there is a clear separation of the fan inlet and outlet and does not cover applications in which fans are used only to circulate air in an open space.

Fans and Systems is aimed primarily at the designer of the air moving system and discusses the effect on inlet and outlet connections of the fan’s performance. System Effect Factors, which must be included in the basic design calculations, are listed for various configurations. AMCA 202 and AMCA 203 are companion documents.

Troubleshooting is intended to help identify and correct problems with the performance and operation of the air moving system after installation. AMCA 201 and AMCA 203 are companion documents.

• System Pressure Losses

• Fan Testing and Rating

• System Checklist

• Acceptance Tests

• The Fan “Laws”

• Fan Manufacturer’s Analysis

• Test Methods and Instruments

• Master Troubleshooting Appendices

• Precautions

• Fan Performance Characteristics • System Effect • System Design Tolerances

Field Performance Measurements of Fan Systems reviews the various problems of making field measurements and calculating the actual performance of the fan and system. AMCA 201 and AMCA 202 are companion documents.

• Air Systems • Fan and System Interaction

• Limitations and Expected Accuracies

• System Effect Factors • Calculations

Review Committee These members contributed to the final review of the publications contained in this Fan Application Manual. Tom Berger Rick Bursh

Pace Company, Division of York International Illinois Blower, Inc.

Patrick Chinoda

Hartzell Fan, Inc.

Narsaiah Dasa

TLT-Babcock, Inc.

Gerald P. Jolette

AMCA Staff

Robert L. Lanier

Phelps Fan Manufacturing Co., Inc.

Tung Nguyen Sutton G. Page Scott Phillips

Emerson Ventilation Products Austin Air Balancing Corp. The New York Blower Company

Neil H. Rutherford

Delhi Industries, Inc.

Jack E. Saunders

Barry Blower/Snyder General Corp.

Paul R. Saxon

AMCA International

Erling Schmidt

Novenco, Inc.

Mark Schultz William Smiley

American Fan Company The Trane Company

James L. Smith

Aerovent, Inc.

Charles R. Voss

Phelps Fan Manufacturing Co., Inc.

Robert H. Zaleski

Acme Engineering & Manufacturing Corp.

Disclaimer This manual has been prepared by the Air Movement and Control Association, Inc. The information contained in this manual has been derived from many sources and is believed to e accurate. Please note that the recommendations contained herein do not necessarily represent the only methods or procedures appropriate for the situation discussed, but rather are inteded to present consensus opinions and practices of the air movement and control industry which may be helpful, or of interest to those who design, test, install, operate or maintain fan-duct systems. Thus, AMCA disclaimes any and all warranties, expressed or implied, regarding the accuracy of the information contained in this maual and further disclaims any liability for the use or misuse of this information. AMCA does not guarantee, certify or assure the performance of any fan-duct system designed, tested, installed, operated or maintained on the basis of the information provided in this manual. Air Movement and Control Association International, Inc. will consider and decide all written complaints regarding its standards, certification programs, or interpretations thereof. For information on procedures for submitting and handling complaints, write to: Air Movement and Control Association International 30 West University Drive Arlington Heights, IL 60004-1893 U.S.A. or AMCA International, Incorporated c/o Federation of Environmental Trade Associations 2 Waltham Court, Milley Lane, Hare Hatch Reading, Berkshire RG10 9TH United Kingdom

Table of Contents Publication 200 Air Systems 1

Introduction

3

Symbols and Subscripts

4

Properties of Air

5

Airflow

13

The Flow System

34

System Design and Tolerances

40

Annex A — SI / I-P Conversion Table

41

Annex B — Standard Atmospheric Data Versus Altitude Charts

43

Annex C — Psychrometric Density Tables

47

Annex D — Friction Charts

49

Annex E — Air Density Correction Factor Charts

Publication 201 Fans and Systems 51

Introduction

51

Symbols and Subscripts

51

Fan Testing

54

Fan Ratings

63

Catalog Performance Tables

66

Air Systems

74

System Effect Factor (SEF)

79

Outlet System Effect Factors

88

Inlet System Effect Factors

99

Effects of Factory Supplied Accessories

49

102

Annex A. — SI / I-P Conversion Table (Informative)

103

Annex B. — Dual Fan Systems - Series and Parallel

106

Annex C. — Definitions and Terminology

113

Annex D. — Examples of the Convertibility of Energy from Velocity Pressure to Static Pressure

120

Annex E. — References

Table of Contents (continued) Publication 202 Troubleshooting 122

Introduction

122

Procedure for Troubleshooting

122

Safety Precautions

122

System Checklist

128

Fan Manufacturer’s Analysis

130

Conclusion

131

Annex A. Noise

134

Annex B. Insufficient Airflow

136

Annex C. Airflow High

137

Annex D. Static Pressure Wrong

139

Annex E — Power High

140

Annex F — Fan Does Not Operate

141

Annex G — Premature Failure

142

Annex H — Vibration

Publication 203 Field Performance Measurement of Fan Systems 145

Introduction

145

Scope

145

Types of Field Tests

146

Alternatives to Conducting Field Tests

146

System Effect Factors

146

Fan Performance

146

Referenced Planes

147

Symbols and Subscripts

147

Fan Flow Rate

152

Fan Static Pressure

156

Fan Power Input

158

Fan Speed

158

Densities

Table of Contents (continued) 159

Conversion Calculations

160

Test Preparation

161

Precautions

161

Typical Fan-System Installations

165

Annex A — Field Test Examples

241

Annex B — Pitot-Static Tubes

242

Annex C — Double Reverse Tube

243

Annex D — Pitot-Static Tube Holder

244

Annex E — Static Pressure Tap

245

Annex F — Pitot-Static Tube Connections

246

Annex G — Manometer Data

248

Annex H — Distribution of Traverse Points

250

Annex J — Instrumentation Characteristics

252

Annex K — Phase Current Method for Estimating the Power Output of Three Phase Fan Motors

254

Annex L — Estimated Belt Drive Loss

256

Annex M — Density Determinations

260

Annex N — Density Charts and Tables

269

Annex P — Diffusion at Fan Outlets

270

Annex R — Diffusion at Fan Outlets

274

Annex S — Typical Format for Field Test Data

275

Annex T — Uncertainties Analysis

Air Systems 1. Introduction An air system is any assembly of ducts, filters, conditioning devices, dampers, louvers, fans, etc., the main purpose of which is to move air from one place to another in a controlled fashion. Most air systems draw air from one space and discharge it into another. Air systems are often required to operate satisfactorily in a wide range of environmental conditions. The conditions which will be encountered must be considered in the design of the ducts, pipes, etc., which will contain the airflow and constitute the boundary of the system.

1.1 Air system components A typical air system may contain one or more of the following (see Figure 1): a) System inlet b) Distribution system c) Fan d) Control device e) Conditioning device f) System outlet 1.1.1 System inlet. An air system usually includes devices such as louvers, filters, screens, guards, grilles, etc., where the air enters the system. These are used for safety reasons as well as to inhibit the entry of rain, dust, and other unwanted matter. Their appearance may be important as they are usually visible on the exterior of a structure. 1.1.2 Distribution system. Most air systems are made up of ducts specially designed and constructed to convey air from the system inlet(s) to the system outlet(s). In some cases, enclosed spaces in the structure such as plenums above ceilings or holes in walls may be used to confine and direct the flow. 1.1.3 Fan. Understanding the design and opera-tion of air systems begins with an understanding of the various types of fans, their performance characteristics, and their applications.

200 A fan is required in order to produce the pressure differential which results in the flow of air through a system. The fan must be carefully selected to meet the specified airflow and pressure for proper system operation. Different fan designs produce different pressure-volume and fan power relationships, which are critical to air system operation. Refer to Figure 4.2, AMCA Publication 201-90. 1.1.4 Control devices. In many air systems it is necessary to regulate and control the flow through the system in response to some monitoring signal, usually temperature or pressure. It may be also necessary to regulate the flow in the individual branches of the system. Control devices such as dampers function by controlling the amount of airflow. In some cases, the output of the fan can be varied by other methods (variable speed motor, variable inlet vanes, variable pitch impeller, etc.) 1.1.5 Conditioning device. Most air systems are designed to take air from the inlet and change its condition before discharging it at the outlet. Changes may include the temperature, humidity, pressure, contaminant level and cleanliness, etc., of the air. Many conditioning devices require outside energy sources, for example, heating and cooling coils; other components such as filters are passive devices and have no external energy connection. All conditioning devices increase the pressure drop across the system and this effect must be considered in the selection of the fan. 1.1.6 System outlet. An air system usually includes a special component at the termination of the system or at the end of each of the system's branches, such as a simple screen or louver. In many cases the distribution of the air at the outlet to the receiving space is very important, e.g., in an occupied air conditioned room. These systems require carefully selected outlets and diffusing devices to achieve desirable air motion and temperature conditions in the conditioned space. Typical devices are ceiling diffusers and grilles. In some cases these may incorporate control devices such as dampers and mixing boxes.

FAN

MAIN DISTRIBUTION SYSTEM (DUCT)

SYSTEM INLET

BRANCH DUCT

COIL FILTER LOUVER

DAMPER DIFFUSER SYSTEM OUTLET

Figure 1 - Typical Air System

2 | Air Systems

SYSTEM OUTLET

SYSTEM OUTLET

2. Symbols and Subscripts 2.1 Symbols and subscripted symbols Symbol

Description

SI

(I-P)

A Ae Ao ah C Cd Cn c D E ε f g γ K L μ ∆P P Ps Psx Pt Ptx Pv p Q Qx R Re rh ρ ρx SEF SR sh t td tw V v Y Z ~

Area Area-Orifice Equivalent to System Area-Nozzle with no loss Absolute Humidity Dynamic Loss Coefficient Coefficient of Discharge Coefficient of Nozzle Discharge Speed of Sound Duct Diameter and Equivalent Diameter System Resistance Curve Absolute Surface Roughness Height Friction Coefficient Gravity Ratio of Specific Heats System Effect Factor (System) Length Air Viscosity, Absolute Pressure Differential Pressure Static Pressure Static Pressure at Plane x Total Pressure Total Pressure at Plane x Velocity Pressure Atmospheric Pressure Airflow Rate Airflow Rate at Plane x Gas Constant Reynolds Number Relative Humidity Air Density Air Density at Plane x System Effect Factor (Fan) System Resistance Factor Specific Humidity (_/_ dry air) Temperature Dry-Bulb Temperature Wet-Bulb Temperature Average Velocity Velocity - At any Point Expansion Factor Altitude Is Proportional to

m2 (ft2) m2 (ft2) m2 (ft2) kg/m3 (lb/ft3) Dimensionless Dimensionless Dimensionless m/s (ft/s) m (ft) Dimensionless m (ft) Dimensionless m/s2 (ft/s2) Dimensionless Dimensionless m (ft) N-s/m2 (lbm/ft-s) Pa (in. wg) Pa (in. wg) Pa (in. wg) Pa (in. wg) Pa (in. wg) Pa (in. wg) Pa (in. wg) Pa (in. Hg) m3/s (cfm) m3/s (cfm) J/kg-K (ft-lb/lbm-°R) Dimensionless % (%) kg/m3 (lbm/ft3) kg/m3 (lbm/ft3) Pa (in. wg) m-4 (ft-4) kg/kg dry air (lb/lb) dry air °C (°F) °C (°F) °C (°F) m/s (ft/min) m/s (ft/min) Dimensionless m (ft) Dimensionless

2.2 Subscripts Subscript

Definition

Subscript

Definition

a b c d E F

Element a Element b Element c - Combined Discharge Plane of System Entry Fan

n O x x,x' 1 2

Reference to Nozzle Plane of System Outlet Plane 0, 1, 2,...as appropriate Between Planes x and x' Plane of Fan Inlet Plane of Fan Discharge

3. Properties of Air Atmospheric air is a mixture of several gases, water vapor, and impurities. The relative amounts of the important constituents for dry, sea level air are given in Table 3.1. This table may be considered representative of air at any altitude. Table 3.1 - Dry Air Composition, Fraction Component

Volume

Weight

Nitrogen

0.7809

0.7552

Oxygen

0.2095

0.2315

Argon

0.0093

0.0128

Carbon Dioxide

0.0003

0.0004

Also slight traces of neon, hydrogen, helium, krypton, ozone and others Although the gas composition of air can be considered essentially constant, the amount of water vapor contained in the air can vary greatly. The properties of moist air are dependent upon the relative amount of water vapor and dry air, therefore, in defining the properties of moist air, this relative amount must be defined (see Section 3.1.5 Humidity). The impurities in the air are of various forms, but basically can be divided into two categories: a) particulates which can be either solid or liquid, and b) mixtures, which can be either gas or vapor. The distribution of these impurities is not uniform on an atmospheric scale, but can be considered uniform for the purposes of air system design. Since air is a mixture of several gases, the behavior of air under varying conditions can be best understood by first reviewing the behavior of pure gases.

3.1 Properties of gases A gas may be defined as a compressible substance which has no free surfaces and occupies all portions of its container. The important properties of an ideal gas are listed below. 3.1.1 Density. The density of a gas is defined as the total mass of the molecules in a unit volume. In the SI system density is given in kilograms per cubic meter (kg/m3); in the I-P system, density is given in pounds per cubic foot (lbm/ft3). For purposes of uniformity, standard air has been defined as air with a density of 1.2 kg/m3 (0.075 lbm/ft3) and an absolute viscosity of 18.19 × 10-6 4 | Air Systems

N-s/m2 (1.222 × 10-5 lbm/ft-s). This is substantially equivalent to air at a temperature of 20°C (68°F), 50% relative humidity, and a barometric pressure of 101 kPa (29.92 inches mercury) at sea level. The ratio of specific heats, (γ), is taken to be 1.4, which is the expected value for a perfect diatomic gas. The temperature and barometric pressure of atmospheric air vary widely with weather conditions and geographical location, most noticeably altitude. In order to simplify design, standard atmospheric conditions have been defined which give the variation of atmospheric pressure, temperature, and, therefore, density with altitude. Annex B lists these variations. 3.1.2 Pressure In an air system, pressure is the force exerted by the air molecules on the surfaces which make up the system. Since air molecules are always in motion, they continuously collide with other air molecules or a solid surface. All these collisions are considered to be perfectly elastic and, in the case when a molecule strikes a surface, the surface experiences a force equal and opposite to the time rate of change of momentum of the rebounding molecule. This force causes the gas to exert an overall pressure on an immersed body and this force per unit area is referred to as the pressure. In air system work, the units of pressure are given in terms of force per unit area. The unit of measure for pressure in the SI system is the Pascal (Pa); in the I-P system the units are inches of water gauge (in. wg). 3.1.3 Temperature 3.1.3.1 Thermal relationships The kinetic energy of gas molecules increases with increasing temperature. The important effects of this fact are stated in Boyle's Law and Charles' Law, which state that the volume of a perfect gas varies inversely with absolute pressure and directly with absolute temperature, respectively. The total effect is more properly stated by the equation of state: PV = mRT

Eq. 3.1-1

or P = ρRT Where: P V m R T ρ

= Pressure = Volume = mass = Gas Constant = Absolute Temperature = m/V = density

Eq. 3.1-2

In the design of most air systems, it is acceptable to assume that the gas is incompressible, therefore, the air density may be considered constant, and therefore, the absolute pressure and absolute temperature are directly proportional. 3.1.3.2 Dry-bulb, wet-bulb and dew point temperature. Unless otherwise specified, the temperature of an air-water vapor mixture is that temperature which is indicated by an ordinary or drybulb thermometer. This dry-bulb temperature is the temperature of both the air and the water vapor in the mixture. The wet-bulb temperature may be determined by exposing a wetted bulb in a moving air-water vapor mixture until equilibrium is obtained. The wet-bulb temperature will be lower than the drybulb temperature as long as evaporation continues. If no evaporation is possible, the mixture is saturated and the wet and dry-bulb temperatures for this condition will be identical. The dew point temperature of an air-water vapor mixture is the saturation temperature corresponding to the absolute humidity of the mixture. The dew point temperature may also be considered as that temperature at which condensation begins when the mixture is gradually cooled. 3.1.4 Viscosity. A non-perfect gas, such as air, is capable of exerting a force parallel to the surface of a body which is moving with respect to the gas. The magnitude of the force parallel to the surface is used to define an important property of non-perfect gases - viscosity. The effects of viscosity on the behavior of real gases cause resistance to flow; the resistance is proportional to the velocity gradients which exist in the gas. The absolute viscosity (µ) is defined as the shearing stress for a unit rate of change of velocity. The absolute viscosity has units of newton-sec per meter squared (N-s/m2) in the SI system and pound mass per foot-second (lbm/ft-s) using I-P units. 3.1.5 Humidity. The density of atmospheric air is also a function of the humidity. Although the change in density due to a change in humidity is not large, it is often significant and air system designers should be aware of these changes. Remember that increasing humidity lowers the density since water vapor is lighter than dry air. The density of saturated air for various barometric and hygrometric conditions is shown in Annex C. Partially saturated air contains vapor that is superheated, that is, the temperature of the mixture and, therefore, that of the vapor is higher than the saturation temperature for the existing vapor pressure. The relative humidity (rh) of an air-water vapor

mixture is defined as the ratio of the vapor pressure existing compared to the vapor pressure at saturation for the same dry-bulb temperature. This is also equal to the ratio of the mole fractions under the same condition. Relative humidity is always expressed as a percent. Specific humidity (sh) is the actual mass (weight) of the water vapor existing per unit mass (unit weight) of dry air or gas. Absolute humidity (ah) may be expressed in kilograms (pounds) of water vapor per cubic meter (cubic foot) of mixture. The humidity of an air-water vapor mixture is often expressed by giving either relative humidity or a wet-bulb depression.

4. Airflow The flow of any fluid between two points is caused by the existence of a pressure differential between the two points. It is the purpose of this section to explain the parameters that may affect the flow of a gas between two points.

4.1 Flow conditions Most air systems are designed in the incompressible range. Where compressibility is a factor, Mach number and Reynolds number must be considered. The magnitude of these parameters gives an indication of the effects which can be expected from the deviations in the non-perfect gas behavior from that of a perfect gas. 4.1.1 Mach number. Mach number, for our purposes here, is the ratio of the velocity of an airstream to the speed of sound in that airstream. Mach number = V/c Where: V = velocity of air, m/s (ft/s) c = speed of sound in air, m/s (ft/s) The speed of sound is a function of temperature and is the speed at which very small pressure disturbances are propagated throughout the gas. The speed of sound is proportional to the square root of the absolute temperature, and for standard air is approximately 345 m/s (1130 ft/s). If the Mach number is small and no large static pressure changes are introduced by mechanical means, the flow may be considered incompressible, that is, the density is everywhere constant. Air can be considered incompressible if the fan total pressure rise is less than 2980 Pa (12 in. wg). Air Systems | 5

4.1.2 Reynolds number. The ratio of the inertia force to the viscous force caused by changes in velocity within the fluid element is known as the Reynolds number. ⎛ρ⎞ Re = DV ⎜ ⎟ ⎝μ⎠

Eq. 4.1-1A SI

⎛ ρ ⎞ Re = DV ⎜ ⎟ ⎝ 60 μ ⎠

Eq. 4.1-1A I-P

friction drag, and, for streamlined bodies closely aligned with the flow, represents the entire drag force. For blunt bodies, which may be streamlined bodies at large angles to the flow, profile drag exists. Profile drag is caused by the inability of the flow, due to its viscous effects, to follow the body shape. The inability to follow the body shape creates a wake of very turbulent flow which in effect creates the profile drag force. These wake effects are the predominant cause of flow losses in systems.

⎛ DV ⎞ =⎜ ⎟ ⎝ γ ⎠ For standard air: Re = 65970.3DV Re = 102.3DV

Eq. 4.1-1B SI

Figure 4A - Skin Friction Drag

Eq. 4.1-1B I-P

Where: D V μ γ ρ

= Any convenient reference dimension, m (ft) = Velocity, m/s (ft/min) = Absolute viscosity, N-s/m2 (lbm/ft-s) = Kinematic viscosity, m2/s (ft2/s) = Density, kg/m3 (lbm/ft3)

For flow about immersed bodies, D is normally taken as the length of body in the direction of flow. In ducted flow, D is normally taken as the diameter of the duct; in unducted flow, D is normally taken as the diameter of the opening through which the flow passes. For a fan, D is equal to the impeller tip diameter and is only proportional to conventional Reynolds numbers. The Reynolds number provides a convenient non-dimensional means of comparing two flows.

4.2 Flow about immersed bodies If a solid body is immersed in a flowing stream of a gas, the direction of flow of the gas will be parallel to the surface of the solid body. The changes in the direction of the molecules close to the body exert forces on the body which when taken over the entire body, are perpendicular to the direction of the gas flow. A non-perfect gas will also exert a force parallel to the direction of the velocity, due to the viscosity of the gas. This force, usually called drag, is due to two effects. The first is the shearing force set up within the molecules of the gas resulting from the molecules decelerating from the gas velocity to zero velocity when in contact with the body. This is called skin 6 | Air Systems

Figure 4B - Profile Drag Figures 4A and 4B illustrate skin friction drag and profile drag.

4.3 Ducted flow When air flows through a duct of constant crosssection, the average velocity remains constant and is parallel to the center line of the duct. Due to friction, the velocity at the duct wall is zero and the average velocity profile can be defined as either of two conditions: a) Laminar Flow: Flow in which the air velocity vectors are parallel to the duct wall. This type of flow is described as smooth. b) Turbulent Flow: Flow in which air velocity vectors at various points across the duct are at various angles, up to and including reverse flow. Except for extremely low air velocities, laminar flow does not exist and all duct flow involving air can be considered to be in the transition region between laminar and fully turbulent flow. The transfer of energy from the high velocity section in the center of

the duct to the low velocity section near the duct wall causes a marked resistance to the flow. This resistance varies linearly with the length of the duct and approximately with the square of the average velocity in the duct. The resistance is also a function of the Reynolds number of the flow, which is calculated using the average velocity in the duct, the duct diameter, and the surface roughness of the duct wall. The velocity profiles in a duct system for fully developed flow will vary depending on whether the flow is laminar or turbulent and the degree of duct roughness. Velocity profiles of various flow conditions are shown in Figure 4C. The absolute velocity of the air stream will vary substantially over the cross-sectional duct area, but for duct systems the velocity used for determining the velocity pressure is always the average velocity given by: V average = Q/A

Eq. 4.3-1

Where: V = Velocity, m/s (ft/min) Q = Flow rate, m3/s (cfm) A = Area of the cross-section where the flow occurs, m2 (ft2) The duct velocity profiles shown in Figure 4C are uniform along the length of the duct and symmetrical around the center line. Where there are disturbances in the ducts, such as turns, expansion or contraction, etc., the velocity profile across the duct can become very asymmetrical as shown in Figure 4D. The flow will return to a normal velocity profile after a disturbance if there is sufficient length of straight duct to allow the velocity distribution to regain uniformity. A minimum of 2½ equivalent duct diameters of straight duct is required to attain a normal velocity profile for velocities of 12.7 m/s (2500 ft/min) or less. Add one duct diameter for each additional 5.08 m/s (1000 ft/min). See AMCA Publication 201-90, Fans and Systems.

4.4 System losses The losses in total pressure for flow through a system are caused by two factors: friction losses due to viscosity as the air flows along the surface of ducts and other system elements, and dynamic losses due to the turbulent wake caused by changes in direction and separation of the flow around obstructions.

In addition to the losses in total pressure in a system caused by friction losses and dynamic losses, there are losses due to System Effects. System Effects occur because of the differences between the fan inlet and outlet connections to the installed system and the standardized connections used in laboratory tests to obtain fan performance ratings. AMCA Publication 201, Fans and Systems, gives specific details on System Effects related to fans. System Effects related to series system elements are covered further in Section 4.5 of this publication. 4.4.1 Duct friction losses. In the normal range of air systems for HVAC and industrial applications, the flow falls into the transition region between laminar flow and complete turbulent flow. In this region the losses due to friction are a function of Reynolds number and the relative roughness of the duct wall. The pressure loss in the transition region will vary at slightly less than the square of the velocity. The pressure loss due to friction for flow in ducts may be calculated from the Darcy-Weisbach equation: ∆Pt = f(L/D) Pv

Eq. 4.4-1

Where: ∆Pt f L D Pv

= = = = =

Total pressure loss due to friction, Pa (in. wg) Friction factor, dimensionless Length of duct, m (ft) Diameter of pipe, m (ft) Velocity pressure, Pa (in. wg)

In the transition flow range, the value of the friction factor cannot be calculated directly. It can be obtained from the Moody diagram or by iterative solution of the Colebrook equation. See the ASHRAE Handbook: Fundamentals, chapter on Duct Design, for a more complete discussion of duct friction losses. The Moody diagram, Figure 4E, shows the relationship of the friction factor, Reynolds number and duct roughness (ε) in meters (feet). Most applications are in the transition region between laminar and full turbulent flow conditions. Using duct friction charts (see Annex D) is the most common method of determining friction losses. These charts are based on ducts having average roughness and standard air density. Correction factors must be applied for ducts having different roughness, and for variations in air density and viscosity.

Air Systems | 7

r

LAMINAR TURBULENT SMOOTH Re = 107 SMOOTH Re = 105

0

0.5

ε = 0.03D ROUGH ε = 0.008D ROUGH 1.0

1.5

v V D ε Re v V r

= Duct Diameter = Duct Roughness = Reynolds Number = Velocity at any Point = Average Velocity = Radius Figure 4C - Velocity Profiles in a Round Duct for Various Reynolds Numbers and Duct Roughness

Figure 4D - Changing Velocity Profiles 8 | Air Systems

2.0

0.10 0.09 0.08 0.05

0.07

0.04 0.06

0.03

0.05

0.02

0.010 0.008 0.006

RO

0.03

UG

0.004

H N PE DE ) Re 9a H (2 IT Eq.

W

FRICTION FACTOR, f

0.04

0.002 NC

DE

RELATIVE ROUGHNESS, ε/D

0.015 FULLY ROUGH (EQ 18) Eq. (29a)

E

0.02

0.0010 0.0008 0.0006

Eq. (27)

0.0004

SMOOTH PIPE Eqs. (28a) and (28b)

0.015

0.0002 LAMINAR

TRANSITION REGION

TURBULENT

0.00010 0.00005

0.010 0.009 0.008 103

2

3

5

104

2

3

5

105

2

3

5

106

2

3

5

107

2

3

5

0.00001 108

REYNOLDS NUMBER, Re

Figure 4E - Moody Diagram Reprinted by permission of the American Society of Heating, Refrigerating and Air-Conditioning Engineers, Atlanta, Georgia, from the 1993 ASHRAE Handbook-Fundamentals. (Moody 1944). Values on the chart are the same for both the SI and I-P systems. Equation numbers refer to equations in the source document.

Air Systems | 9

Loeffler1 has developed simplified equations for the friction factor in the normal range of flow conditions found in industrial and HVAC air systems. The equations provide for direct calculation of duct friction total pressure losses. These equations yield results that are accurate within ±5% and are conservative over most of the range of flow. For aluminum ducts, medium smooth: ε = 0.0000457 m (0.00015 ft)

Correction factors for density and viscosity variations need to be applied for conditions other than standard air. See Annexes B and E. 4.4.2 Dynamic losses. Dynamic losses occur where there are changes in velocity or direction in the air system and are due primarily to the profile drag. Dynamic losses are proportional to the square of velocity, and therefore, are proportional to the velocity pressure. ∆Pt ~ V

⎛ LQ ⎞ Pt loss = a ⎜ 4.93 ⎟ ⎝ D ⎠

2

~ Pv

1.863

Where: a = 1.764 × 10-2

Eq. 4.4-2 SI

a = 4.816 × 10-9

Eq. 4.4-2 I-P

Dynamic pressure loss data are given in a number of forms such as pressure loss for given volume or velocity, equivalent length of duct, or velocity pressure multiplier, and is available from manufacturers' data and handbooks such as: ASHRAE, the Industrial Ventilation Guide and SMACNA. Except at duct exits, dynamic losses occur along some length and cannot be separated from friction losses. For practical purposes the dynamic losses are assumed to be concentrated at one point and the friction losses are included as part of the duct friction. Dynamic loss coefficients for duct fittings are based on zero length. For friction loss calculations, the centerline length of the duct fitting is taken as the length of the fitting.

For galvanized steel ducts, average: ε = 0.0001524 m (0.0005 ft) ⎛ LQ1.921 ⎞ Pt loss = a ⎜ 5.066 ⎟ ⎝ D ⎠ Where: a = 1.717 × 10-2

Eq. 4.4-3 SI

a = 3.534 × 10-9

Eq. 4.4-3 I-P

For fiberglass ducts or lined ducts, fabric and wire flexible ducts (wire covered with fabric), medium rough, ε = 0.00091443 m (0.003 ft) ⎛ LQ1.965 ⎞ Pt loss = a ⎜ 5.208 ⎟ ⎝ D ⎠ Where: a = 2.093 × 10-2

Eq. 4.4-4 SI

a = 3.64 × 10-9

Eq. 4.4-4 I-P

4.4.3 Dynamic loss coefficient. There are two common methods of expressing dynamic losses. These are: 1) The equivalent length of duct method, and 2) the loss coefficient method. The equivalent length of duct method replaces the dynamic loss of fittings (elbows, tees, branches, etc.) with a length of duct that will have an equivalent loss. The equivalent length of duct for all of the dynamic losses are added to the straight duct length. A friction chart showing the loss is then used to determine the total loss in the system. The dynamic loss coefficient method is based on the fact that all losses in a system are functions of the velocity pressure and can be calculated by a corresponding dynamic loss factor multiplied by the velocity pressure. These losses are added to the straight duct friction loss to determine the total loss in the system.

Where: Pt loss = Total pressure loss, Pa (in. wg) Q = Flow rate, m3/s (cfm) D = Duct diameter, m (ft) (or equivalent diameter of rectangular ducts) D equivalent = (4ab/ )0.5 where a and b are the sides in m (ft) L = Duct length in m (ft) 10 | Air Systems

The dynamic loss coefficient method is preferred because it is usually quicker and offers the advantage of faster recalculation when other branch duct sizes are tried. Dynamic losses are proportional to the velocity pressure occurring in the system element and,

therefore, the pressure loss in the fitting can be related to the velocity pressure by use of a dynamic loss coefficient. The dynamic loss coefficient Co is defined as: Co = ∆Pt/Pvo

Eq. 4.4-5

Where: Co = Dynamic loss coefficient, reference to section o, dimensionless ∆Pt = Dynamic pressure loss, Pa (in. wg) Pvo = Velocity pressure at section o, Pa (in. wg) The coefficient relates the pressure loss in the element to the velocity pressure at a given crosssectioned area of the element. The pressure loss of duct system elements with known dynamic loss coeffients can be calculated by: ∆Pt = Co Pvo

Eq. 4.4-6

Where there are changes in area or divided flow in the fitting, the designer must be careful to use the proper area as noted in the loss tables for the determination of the velocity pressure to be used with the dynamic loss coefficient. The ASHRAE Handbook: Fundamentals, Chapter on Duct Design, provides a detailed discussion of the dynamic loss coefficient and tables for coefficients of many common duct elements.

4.5 System Effects Additional losses can occur in air systems because of the physical relationship of various elements in the system. These System Effect losses occur because of the difference between the way the performance of the element was determined by testing and the way the element is actually installed in the system. 4.5.1 Fan System Effects. Fan System Effects occur because of the difference in inlet and outlet conditions under laboratory test conditions and the inlet and outlet conditions as the fan is installed in the system. Detailed information on Fan System Effects is contained in AMCA Publication 201, Fans and Systems. The System Effect is accounted for as a pressure loss which must be included with the other system losses. The sum of the pressure losses is then used as the basis for selecting the fan.

4.5.2 Element System Effects in series. System Effects for other air system elements occur when two or more elements are in close proximity to one another. Loss coefficients for duct fittings, coils, filters, dampers, etc., are determined with a sufficient length of straight duct (normally 10 diameters) ahead of the element to allow for a normally distributed velocity profile entering the element, and a sufficient length of straight duct (normally 10 diameters) downstream from the element to allow a normally distributed velocity profile to be re-established. When two elements, such as elbows, or an elbow and a damper, are placed close together the air entering the second element will be highly turbulent and asymmetrical in profile, causing a higher loss than expected. In addition, any static regain occurring downstream of the first element would also be lost when sufficient length of straight duct is not present. To illustrate System Effects for duct elements, the loss coefficients for a single 90° elbow and two elbows in series are shown in Figure 4F. In the case of the two elbows in series, the difference between twice the loss of the single elbow and the actual combined loss is the System Effect. The System Effect varies substantially depending on how close the two elbows are to each other. Similar effects can be expected when any system elements are in close proximity. The amount of the System Effect will vary over a rather wide range depending upon the physical characteristics of each element and their relationship to each other in the system. Very little actual data is available on System Effects of various combinations of system elements, and the system designer must, of necessity, estimate the System Effects. The following tables for estimating System Effects, and Equation 4.5-1, are given as a guide to the designer. Actual data should be used whenever it is available. Cc = (Ca + Cb) K

Eq. 4.5-1

Where: Cc = Loss coefficient of combined elements, dimensionless Ca = Loss coefficient of element a, dimensionless Cb = Loss coefficient of element b, dimensionless K = System Effect Factor, dimensionless

1. Loeffler, J. J., Simplified Equations for HVAC Duct Friction Factors, ASHRAE Journal, January, 1980

Air Systems | 11

ONE ELBOW C1 = 1.15*

FLOW

Figure 4G should be used for elements in series where the flow is straight through, while Figure 4H should be used when a turning element (elbow, etc.) is involved. These tables have been developed on the basis of limited data and are intended only as a guide. Actual System Effects may vary from the values shown. See sample calculation in Section 5.8

TWO ELBOWS IN SERIES

a

C2 = VARIES

FLOW

b

RESISTANCE ELEMENTS

D

L L

D

L/D

1.0

2.0

4.0

10.0

C2 Loss coefficient for two elbows in series**

2.63

4.18

3.08

2.45

2 C1 Two times the loss coefficient of a single elbow System Effect (Difference)

L/D

.50

1

2

3

4

5

7.5

10

K

1.5

1.4

1.3

1.2

1.15

1.1

1.05

1.0

From SMACNA Duct Design Manual Figure 4G - Element System Effects for Straight Through Flow

2.30

+0.33 14.3%

+81.7% 1.88

+33.98% 0.78

a

b

+6.5% 0.15

L

D

* From I. E. Idelchik, Handbook of Hydraulic Resistance, 3rd Edition, 1993, p.365, Hemisphere Publishing Company. ** ibid, p.375

Figure 4F - System Effect of Duct Elements L/D

.50

1

2

3

4

5

7.5

10

K

1.0

1.08

1.75

1.5

1.30

1.2

1.1

1.0

Figure 4H - Element System Effect for Turning Elements

12 | Air Systems

5. The Flow System 5.1 Concepts of pressure The flow of air between two zones (or spaces) is due to a pressure difference between the two zones. This pressure difference forces the air to flow from the high pressure zone to the low pressure zone. Ductwork is used in most air systems to convey the air from one zone to another. The quantity of air Q in m3/s (cfm) that will flow is equal to the cross-sectional area A of the duct in m2(ft2) times the air velocity V in m/s (ft/min). Q = AV

Eq. 5.1-1

ρ is equal to the air density in kg/m3 (lbm/ft3), and standard air density equals 1.2 kg/m3 (0.075 lbm/ft3). Pv is always positive and this pressure is always exerted in the direction of airflow. Air confined within a duct or a tank, whether in motion or not, creates another kind of pressure which exerts itself in all directions at once, including perpendicular to the walls of the enclosure. This is known as static pressure Ps. Static pressure is negative when it is below atmospheric pressure, and positive when above atmospheric pressure. Total pressure Pt at any point in an air system is equal to the algebraic sum of the static pressure Ps and velocity pressure Pv. Pt can be either positive or negative, depending on its components.

The air traveling at a given velocity V in m/s (ft/min) will create a velocity pressure Pv in Pa (in. wg). The velocity pressure in these terms is:

Pt = Ps + Pv

Pv = 0.5ρV2

Of prime concern in air system design is the relationship of these pressures internal to the air system. The significance of these pressures can be demonstrated on the next several pages in Figures 5A, 5B, and 5C.

Pv = ρ(V/1096)2

Eq. 5.1-2A SI Eq. 5.1-2A I-P

And for standard air: Pv = 0.6V2 Pv = (V/4005)2

Eq. 5.1-3

5.2 Examples of pressures in duct systems

Eq. 5.1-2B SI Eq. 5.1-2B I-P

Air Systems | 13

In Figure 5A, the sealed length of duct has a static pressure of 345 Pa (1.39 in. wg) above atmospheric pressure. Since there is no airflow, velocity pressure is equal to zero. The total pressure Pt can then be calculated according to equation 5.1-3. Pt

SI = Ps + Pv = 345 + 0 = 345 Pa

Pt

TOTAL PRESSURE 345 Pa (1.39 in. wg)

I-P = Ps + Pv = 1.39 + 0 = 1.39 in. wg

STATIC PRESSURE 345 Pa (1.39 in. wg)

Figure 5A - Sealed System

14 | Air Systems

=

VELOCITY PRESSURE 0 Pa (0 in. wg)

In Figure 5B, with the duct open and a fan placed at one end blowing air through the duct, we find both static pressure and velocity pressure as illustrated by the water gauge. The total pressure is the sum of velocity pressure and static pressure.

TOTAL PRESSURE 345 Pa (1.39 in. wg)

STATIC PRESSURE 97 Pa (0.39 in. wg)

=

AIRFLOW

VELOCITY PRESSURE 248 Pa (1.00 in. wg)

20.3 m/s (4005 FPM)

In this illustration, the static pressure will be above atmospheric pressure and the total pressure is numerically greater than either static or velocity pressure.

Figure 5B - Positive Pressure System

Air Systems | 15

In Figure 5C, a fan is placed at the end of the duct and draws air through the duct. In this case, the static pressure is below atmospheric pressure. In both Figures 5B and 5C, the total pressure rise across the fan is 345 Pa (1.39 in. wg); therefore, the energy used in both systems is equal.

TOTAL PRESSURE STATIC PRESSURE -97 Pa (-0.39 in. wg) -345 Pa (-1.39 in. wg)

=

VELOCITY PRESSURE 248 Pa (1.00 in. wg)

20.3 m/s (4005 FPM)

AIRFLOW

Figure 5C - Negative Pressure System

16 | Air Systems

5.3 Conservation of energy

Therefore, to simplify Bernoulli's Theorem:

The well known Bernoulli Theorem, frequently used in the flow of fluids, states the law of conservation of energy for fluid systems. In its simpler form for airflow:

Ps1 + Pv1 = Ps2 + Pv2 + losses1,2

(V12/2g) + (P1/ρg) + Z1 = (V22/2g) + (P2/ρg) + Z2 + losses1,2 Where: 2/2g

= Kinetic energy or velocity pressure (Pv) Vx Px/ρg = Potential energy or static pressure (Ps) Zx = The elevation (normally ignored in fan systems with minimal changes in elevation) may need to be evaluated when the system elevation change causes a noticeable change in air density.

Pt

PS

PT. #1

PV

PRESSURE Pa (INCHES W.G.)

Simply stated, the sum of static pressure and velocity pressure at any point in a flow system is equal to the sum of static pressure and velocity pressure at any other point in the system, plus any losses in pressure occurring between the two points. Although it is very important to recognize that the sum of static and velocity pressure remains constant at any point in the system when disregarding losses, it is the losses that are of real importance in the design and function of air moving systems. To illustrate Bernoulli's Theorem, Figure 5D shows a typical venturi system commonly used to measure fluid flow.

Pt

PS

PV

PT. #2 Pt PS PV

Eq. 5.3-1

Pt

PT. #3

PS

PV

AIRFLOW

Pt PS PV

Figure 5D - Venturi System

Air Systems | 17

For the purposes of illustration, assume this is a perfect device with no losses. The velocity pressure, static pressure, and total pressure at each point in the system are shown on the gauges as well as on the lower portion of the illustration. Part of the static at point 1 is converted to velocity pressure as the fluid is accelerated in the contracted flow area in the neck of the venturi. At point 2, in the neck of the venturi, the velocity pressure reaches its maximum and the static pressure is reduced, but the total pressure remains constant. As the flow area is gradually increased to point 3, the velocity is slowed down to the original velocity and the velocity pressure is reduced to its original value. The static pressure increases back to its original value also, while the total pressure remains constant. In this illustration, part of the static pressure at point 1 is converted to increased velocity pressure at point 2 and the velocity pressure is then converted back into static pressure at point 3. The conversion of the velocity pressure into static pressure by reducing the velocity is known as static regain and it is very important to understand this phenomenon in the design of flow systems. In air systems, changes in velocity often occur as the air flows through the various elements comprising the system. Some total pressure loss will occur any time the velocity in the system is increased or decreased. The magnitude of these losses is dependent on the physical characteristics of the system element in which the velocity change takes place. Decreases in velocity occurring at abrupt enlargements in area result in total pressure losses approaching one velocity pressure. Figure 5E illustrates air flowing from a large plenum through a long radius flow nozzle and discharging to atmosphere.

If the plenum is very large relative to the nozzle diameter, the velocity in the plenum will approach zero, so the total pressure at point 1 (Pt1) will be equal to the static pressure (Ps1). Assuming a perfect fluid and no losses occurring in the nozzle, the static pressure in the plenum (point 1) is totally converted to velocity pressure at the discharge of the nozzle (point 2). At point 2, just beyond the discharge end of the nozzle, the static pressure (Ps2) will be zero (at atmospheric pressure) and the total pressure (Pt2) will be equal to the velocity pressure of the air stream. Applying Bernoulli's Equation for points 1 and 2 gives: Ps1 + Pv1 = Ps2 + Pv2 Since Pv1 is equal to 0 and Ps2 is equal to 0, the equation reduces to: Ps1 = Pv2 In this case, the energy of the static pressure in the plenum is totally converted to the velocity energy at the discharge of the nozzle. It should also be noted that the velocity pressure is totally lost as the airstream discharges to atmosphere. Using Equation 5.1-2A (as reduced above) the relationship between static pressure and the throat velocity can be established in the above example. Using this in Equation 5.1-2A gives: V2 = (2Ps1/ρ2)0.5 V2 = 1096(Ps1/ρ2)0.5

Eq. 5.3-2A SI Eq. 5.3-2A I-P

Or where: Ps1 = Pv2 Then for standard air, the equations would be: V2 = 1.29(Ps1)0.5

Eq. 5.3-2B SI

V2 = 4005(Ps1)0.5

Eq. 5.3-2B I-P

The throat velocity pressure would be: POINT 1

POINT 2

Figure 5E - Airflow Through a Nozzle

Pv2 = 0.5ρ2V22 Pv2 = (V2/1096)2ρ2 Where: Ps1 = Pv2

18 | Air Systems

Eq. 5.3-3A SI Eq. 5.3-3A I-P

Then, for standard air, the equation would be: Pv2 = 0.6V

2

Eq. 5.3-3B SI

Pv2 = (V2/4005)2

Eq. 5.3-3B I-P

In the above example, the quantity of flow would be dependent on the area of the nozzle discharge and the velocity in the throat as given by the equation: Q2 = V2A2 Where:

In practical system design both Cn and Y approach unity and can normally be neglected. However, these factors are important when measuring fan performance in accordance with ANSI/AMCA Standard 210. In Figure 5F, the nozzle is replaced by a sharp edged orifice. The flow through the orifice tends to neck down to a flow area smaller than the orifice diameter. The point at which the flow area reaches its minimum is called the vena contracta. The flow through the orifice is given by the equation: Q = Cd(2Ps/ρ)0.5A

m3/s

(cfm) Q = Airflow rate, V = Velocity, m/s (ft/min) A = Area, m2 (ft2)

Q = Cd1096(Ps/ρ)0.5A

Eq. 5.3-6 SI Eq. 5.3-6 I-P

Where:

Substituting for V2 from Equation 5.3-2A and 5.3-2B gives: Q = (2Ps1/ρ2)0.5A2

A = Area of orifice, m2 (ft2) Cd = Coefficient of discharge, dimensionless

Eq. 5.3-4A SI VENA CONTRACTA

Q = 1096(Ps1/ρ2

)0.5A

2

Eq. 5.3-4A I-P

And for standard air: Q = 1.29(Ps1)0.5A2

Eq. 5.3-4B SI

Q = 4005(Ps1)0.5A2

Eq. 5.3-4B I-P

The flow nozzle in the above illustration is the basis for airflow measurement in the ANSI/AMCA Standard 210 duct nozzle and chamber nozzle methods of airflow measurement. The ANSI/AMCA Standard 210 nozzle very closely approaches perfect flow conditions with almost zero losses. There are, however, some losses which vary with Reynolds number as well as the effect of the compressibility of the gas which must be accounted for. The Equation 5.3-4A must be modified by adding factors for the losses and compressibility. The equation of flow becomes: Q = CnY(2Ps/ρ)0.5A Q = CnY1096(Ps/ρ)0.5A Where: Q Cn Y A

= Airflow, m3/s (cfm) = Nozzle coefficient, dimensionless = Expansion factor, dimensionless = Area of nozzle, m2 (ft2)

Eq. 5.3-5 SI 5.3-5 I-P

Figure 5F - Airflow Through an Orifice In this case, the coefficient of discharge (Cd) accounts for the reduction in flow area of the vena contracta and losses occurring in the flow system. The value of the coefficient of discharge is dependent on the Reynolds number of the flow system. Where sharp edged orifices are used in the duct system, the coefficient of discharge is also dependent on the ratio of orifice diameter to duct diameter. As in the previous example, all of the velocity pressure is lost as the airstream discharges to atmosphere. Where ducts or plenums are used on the discharge of nozzles or orifices, a portion of the velocity pressure at the nozzle discharge is regained as static pressure as the velocity returns to a normal distribution.

Air Systems | 19

If we plot the flow through these elements versus static pressure as shown in Figure 5G, we obtain a squared curve which is typical of flow versus pressure for any constant system.

Resistance Curve (see Figure 5H) the point of operation is at the intersection of the fan performance curve and the System Resistance curve.

PRESSURE

PRESSURE

SYSTEM RESISTANCE

P = KQ2

POINT OF OPERATION

FAN PRESSURE CURVE FLOW

FLOW

Figure 5G - Typical System Resistance Curve These curves are called System Resistance curves, and define the relationship of flow versus pressure for any system with constant resistance. System Resistance curves are defined by the following equations: Q = (2Ps/ρ)0.5Ae

Eq. 5.3-7A SI

Q = 1096(Ps/ρ)0.5Ae

Eq. 5.3-7A I-P

Then for standard air: )0.5A

Q = 1.29(Ps

e

Q = 4005(Ps)0.5Ae

Eq. 5.3-7B SI Eq. 5.3-7B I-P

Where: Q Ps ρ Ae

= = = =

Airflow, m3/s (cfm) Pressure, Pa (in. wg) Gas density, kg/m3 (lbm/ft3) Area of an orifice having resistance equivalent to the system resistance, m2 (ft2) (equivalent orifice)

The System Resistance concept is very useful in understanding flow in complete systems or elements of flow systems. If we add a fan curve to the System

Figure 5H - Typical Point of Operation

5.4 Fan total and static pressure The flow of a gas through a system of ducts and various system elements requires energy: a) To accelerate the air from ambient conditions at the entry to the system b) To overcome the losses in the system from friction and system element shock losses c) For the loss of energy dissipated as velocity at the system outlet d) To overcome any static pressure at the entry or outlet of the system The fan provides this energy by the increase in total pressure from the fan's inlet to the fan's outlet. The inlet plane of a fan is referred to as Plane 1 and the outlet plane as Plane 2. The total pressure provided by the fan is made up of static pressure and velocity pressure components. The total pressure of a fan is defined as PtF = Pt2 - Pt1 Or: PtF = Ps2 + Pv2 - Ps1 - Pv1

20 | Air Systems

Eq. 5.4-1

Eq. 5.4-2

Eq. 5.4-4

Both static pressure and total pressure curves of the fan and system resistance are shown. Either set of curves can be used for the flow rate because the point of operation is the same. The difference in pressure at the point of operation between the total pressure curve and the static pressure curve is the velocity pressure at the fan discharge (plane 2).

Fan static pressure, as defined, is a term that is peculiar to fans and is not consistent with the normal meaning of static pressure rise. Fan static pressure is derived from the method of testing fans where the fan static pressure is equal to gauge static pressure at the fan discharge when the fan draws air from surrounding atmosphere through a well shaped inlet.

Most fans are rated in terms of static pressure and flow, however, fans having high discharge velocities such as vaneaxial fans are quite often rated in terms of total pressure. Be aware of these different methods of rating and be certain whether fan static pressure or fan total pressure was used to determine the fan selection.

Special care must be used when using fan static pressure for purposes of matching the required fan performance against system total pressure losses. The relationship of these pressures is covered in detail for various systems further on in this section.

5.4.2 Fan system pressure relationships. Figure 5K shows the relationship of total pressure, static pressure and velocity pressure for a fan with free inlet conditions and discharging through a duct against some system. In this example the fan is shown as being equipped with a short inlet duct and an inlet bell. For simplicity's sake it is assumed that there are no losses at the inlet to the fan.

The velocity pressure of a fan is defined as: PvF = Pv2

Eq. 5.4-3

The static pressure of a fan is defined as: PsF = PtF - PvF

The static pressure of a fan can also be stated in several other forms. Substituting Equation 5.4-2 for the total pressure of the fan in Equation 5.4-4 gives: PsF = Ps2 + Pv2 - Ps1 - Pv1 - Pv2

Eq. 5.4-5

Simplifying: PsF = Ps2 - Ps1 - Pv1

Eq. 5.4-6

Since: Ps1 + Pv1 = Pt1 The equation can be restated as: PsF = Ps2 - Pt1

Eq. 5.4-7

5.4.1 Fan performance specification. The system designer should be aware of the effect of the velocity pressure at the outlet of the system and the velocity pressure of the fan discharge (plane 2) on the determination of fan total or static pressure for the system. The net result of the fan total or static pressure requirements at a given flow rate for the system is the fan performance specification, which is normally stated as flow at a specific static, or total pressure. This statement of required fan performance is, in reality, a statement of one point on a system resistance or equivalent orifice curve, which then defines the flow and pressure relationship of the system being designed. The actual point of operation of the combined fan and system will be the intersection of the fan performance curve and the system curve as shown in Figure 5J.

The total pressure prior to the entry of the fan in Figure 5K is zero and since it was assumed that there are no entry losses, the total pressure remains zero until the flow is acted upon by the fan. As air enters the fan, its velocity and the pressure due to that velocity (velocity pressure) increases while static pressure decreases in direct proportion. (In actual conditions there will be some entry losses which will be accounted for in the fan performance rating.) Referring to Figure 5K, the fan total pressure is equal to the total pressure at plane 2 minus the total pressure at plane 1. PtF = Pt2 - Pt1

Eq. 5.4-8

The fan static pressure is equal to the total pressure at plane 2 minus the velocity pressure at plane 2. PsF = Pt2 - Pv2

Eq. 5.4-9

The static pressure of the fan can also be stated as the static pressure at plane 2 minus the static pressure at plane 1 minus the velocity pressure at plane 1. PsF = Ps2 - Ps1 - Pv1

Eq. 5.4-10

Or, as the static pressure at plane 2 minus the total pressure at plane 1. PsF = Ps2 - Pt1

Eq. 5.4-11

Air Systems | 21

The actual static pressure rise across the fan from plane 1 to plane 2 will be greater than the fan static pressure by the amount of the velocity pressure at the fan inlet, plane 1. The difference in the actual static pressure rise across the fan and the fan static pressure represents the energy required to accelerate the air to its entry velocity. This kinetic energy is retained by the moving air stream until there is a change in velocity in the system or it is dissipated at the point of discharge, and as such does not represent a loss in total pressure until it is discharged.

5.5 The total system

The equipment arrangement shown in Figure 5K is typical of the test conditions for fans in ANSI/AMCA Standard 210, and is the basis for fan performance ratings. When fans are tested in accordance with ANSI/AMCA Standard 210, the inlet and discharge conditions are rigidly specified for each test method.

PtE = PtO + Pt loss E,O - PtF

When a fan is installed in an air system where the actual entry and exit conditions are different than the test conditions, the performance of the fan may be altered and System Effect factors must be used to account for the altered performance. AMCA Publication 201, Fans and Systems describes various System Effects and provides quantitative data for calculating System Effect losses.

A fan provides the total pressure to move the air through a system and the flow rate will reach a point of equilibrium (point of operation) when the total pressure provided by the fan equals the total pressure losses in the system at that flow rate. A typical system with inlet and discharge resistance is illustrated in Figure 5L. Applying Bernoulli's Equation at the point of entry and the outlet: Eq. 5.5-1

Rearranging: PtF = Pt loss, E,O + PtO - PtE

Eq. 5.5-2

Restating in terms of Ps and Pv: PtF = Pt loss E,O + PsO + PvO - PsE - PvE PtF = Pt loss E,O + PvO + (PsO - PsE - PvE)

Eq. 5.5-3

Substituting from Equation 5.4-4 for PtF :

PRESSURE

AIRFOIL FAN - SWSI

PvF SYSTEM RESISTANCE (TOTAL PRESSURE)

PtF

PsF SYSTEM RESISTANCE (STATIC PRESSURE)

VOLUME FLOW RATE Figure 5J - Constant Speed Performance Curve with System Resistance 22 | Air Systems

PsF = [Pt loss E,O + PvO + (PsO - PsE - PvE)] - PvF

is included in the total pressure loss of the discharge element.

Eq. 5.5-4

System total pressure loss = total pressure loss internal to the system, plus the velocity pressure loss at the outlet(s) of the system.

Equations 5.5-3 and 5.5-4 are the general statements of fan total pressure or fan static pressure required for flow through a system. The Pt loss term is the loss internal to the system from friction and shock losses. The PvO term represents the energy loss to the system contained in the velocity at the outlet of the system.

The term (PsO - PsE - PvE) represents the change in fan total pressure or fan static pressure requirements because of the static pressure conditions existing at the system entry or outlet and any velocity pressure present at the system entry. Velocity and velocity pressure generated by external sources, such as wind, at the system entry are seldom encountered, so the PvE term can generally be disregarded.

In Equation 5.5-4, the items enclosed in large brackets represent the fan total pressure. In the normal method of calculating system total pressure losses, the velocity pressure at the outlet(s)

PLANE 2

PLANE 1 E

ENTRY

PtF = Pt2 - Pt1 PsF = PtF - Pv2

RESISTANCE ELEMENT

OUTLET

FLOW

0

FAN

Pt

Pv1

Pt2 = PtF

Pv

+

Pv2

PtE= 0

Ps2 = PsF 0

Ps1 _

Ps

ABSOLUTE PRESSURE

Volume Flow Rate m3/sec (cfm)

Figure 5K - Fan with Discharge Resistance (AMCA Installation Type B) Air Systems | 23

RESISTANCE ELEMENT

PLANE 1

PLANE 2

RESISTANCE ELEMENT

E

ENTRY

FLOW

O

OUTLET

FAN

Pv2

PtF ATMOSPHERIC PRESSURE

Pv

Pt2

PvO = PtO

Ps2

+ PtE = 0 _

0 PsF

Pt Ps

ABSOLUTE PRESSURE

Ps1 Pvt

PtE = PtO + Pt LOSS - PtF PtF = Pt LOSS + PvO + (PsO - PsE - PvE) PsF = [Pt LOSS + PvO + (PsO - PsE - PvE)] - Pv2

Figure 5L - Fan with Inlet and Discharge Resistance (AMCA Installation Type D)

PLANE 1

PLANE 2 E

O

OUTLET

ENTRY FLOW

FAN Ptf = PvF + 0

PtF = PvF + PsF since PsF = 0 PtF = PvF = Pv2

_

Figure 5M - Fan with No Resistance at Either the Inlet or the Discharge (AMCA Installation Type A) 24 | Air Systems

RESISTANCE ELEMENT

E

PLANE 1

ENTRY

FLOW

PLANE 2

FAN

OUTLET 0

ATMOSPHERIC PRESSURE

+ PtE = 0 _

Pt FAN

Pv

PtO = Pv2

Pt Ps

Pt1

Ps FAN Ps1

Pv

ABSOLUTE PRESSURE

PtF = Pt LOSS + PvO Where:

PsF = Pt LOSS + PvO - Pv2 since PvO = Pv2 PsF = Pt LOSS

Pv1

(Pressure Loss Internal to System)

Figure 5N - Fan with Inlet Resistance (AMCA Installation Type C)

Air Systems | 25

5.6 Types of fan system

PsF = Pt loss + PvO - Pv2

There are four basic system Installation Types:

In the special case where the velocity pressure at the outlet is equal to the velocity pressure at the fan discharge, the fan static pressure will equal the total pressure loss. If these velocity pressures are different the fan total pressure and fan static pressure must be increased or decreased by the amount of the difference in these velocity pressures.

AMCA INSTALLATION TYPE A: Free Inlet, Free Outlet

5.6.3 AMCA Installation Type C: Fan system with ducted inlet and free outlet. Figure 5N illustrates a system with all system losses on the inlet side of the fan. Since the velocity pressure at the outlet equals the velocity pressure of the fan discharge and is also equal to the total pressure at the fan discharge, the fan static pressure will be equal to the total pressure losses of the system.

AMCA INSTALLATION TYPE B: Free Inlet, Ducted Outlet

PsF = Pt loss AMCA INSTALLATION TYPE C: Ducted Inlet, Free Outlet

Fans designed for use at the end of a system, such as power roof ventilators, include the loss at the discharge in the fan ratings and no system effect loss is needed.

Figure 5P - Installation Types 5.6.1 AMCA Installation Type A: Fan system with free inlet and free outlet. An AMCA Type A installation covers equipment such as window fans, panel fans and power roof ventilators. This type of installation, in which there is no resistance at the fan inlet or outlet, is shown in Figure 5M. The fan provides the total pressure necessary to move air to the velocity at the fan outlet. The total pressure of the fan for this special case is equal to the fan velocity pressure Pv2. 5.6.2 AMCA Installation Type B: Fan system with free inlet and ducted outlet. Figure 5K shows a fan system with discharge resistance. In this system:

Where: PsO = Ps1 + Pv1 = 0 And: 26 | Air Systems

Eq. 5.6-3

The fan in Figure 5N has a short discharge duct which is the way ducted fans are normally tested. If a fan is used at the end of a system and is not equipped with a 2 to 3 diameter length of duct, the system effect loss at the discharge must be included when determining the total pressure loss.

AMCA INSTALLATION TYPE D: Ducted Inlet, Ducted Outlet

PtF = Pt loss + PvO

Eq. 5.6-2

Eq. 5.6-1

5.6.4 AMCA Installation Type D: Fan system with ducted inlet and ducted outlet. Figure 5L illustrates a system with system resistance on both inlet and discharge sides of the fan. In this case the general equations 5.5-3 and 5.5-4 apply. In the special case where the velocity pressure at the outlet is equal to the velocity pressure of the fan discharge, the fan static pressure is equal to the total pressure loss internal to the system. The system shown in Figure 5L has higher velocity pressure, relative to the static pressure, than would normally be expected in a system. This is done to emphasize the velocity pressure effects on system total pressure losses for purposes of illustration. To illustrate the effect of the outlet velocity on the total pressure requirements, Figure 5Q shows the same system as in Figure 5L with the addition of an evasé outlet (diffuser) on the system. The evasé outlet greatly reduces the system outlet velocity pressure. The reduction in fan total pressure and fan static pressure is clearly evident when compared to the system in Figure 5L and is equal to the reduction in the velocity pressure at the duct outlet minus the total

pressure loss in the evasé section. Losses will be quite small for a long evasé outlet of good design (optimum included angle is about 10 degrees). The change in fan static pressure, because of the change in the outlet velocity, is accounted for in the terms "PvO - PvF" in Equation 5.5-4. When system velocities exceed 15 m/s (3000 ft/min), consideration should be given to the use of an evasé outlet to reduce the system pressure requirements. See ASHRAE Handbook, Fundamentals, chapter on Duct Design for more information on the subject.

5.7 System resistance factors The flow through any system is proportional to the square root of the pressure causing the flow. This relationship, which defines the flow versus pressure characteristics of a particular system, is very useful in fan and system designs. The flow rate in any system was given earlier in equation 5.3-4B and can be restated here as: Q = 1.29(Ps)0.5Ao

Eq. 5.7-1 SI

Q = 4005(Ps)0.5Ao

Eq. 5.7-1 I-P

For standard air, where:

SR =

CoPvo (Q / 1.29)2

Eq. 5.7-4 SI

SR =

CoPvo (Q / 4005)2

Eq. 5.7-4 I-P

5.7.1 System resistance factors in series. The particular value of using the system resistance factor SR is that for resistances in series, the SR factors of each element can be added to determine the system resistance factor of the total system. As an example, the three Figure 5R resistance factors in series can be added to obtain the system resistance factor of the complete system. In this case, the resistance factors (given for SI and (I-P) units, respectively) of 34.49 (0.3), 22.99 (0.2), and 17.25 (0.15), total 74.73 (0.65). The pressure loss of this system would be defined by the equation: Pt loss = (Q/1.29)2SR

Eq. 5.7-5 SI

Pt loss = (Q/4005)2SR

Eq. 5.7-5 I-P

For an airflow rate of 2.83 m3/s (6000 cfm): SI:

Ao = area of flow nozzle with no loss

Pt loss

This can also be stated as: Pt = (Q/1.29)2SR

Eq. 5.7-2 SI

Pt = (Q/4005)2SR

Eq. 5.7-2 I-P

Where:

The system resistance factor can be calculated from known pressure loss information: SR =

ΔPt (Q / 1.29)2

Eq. 5.7-3 SI

SR =

ΔPt (Q / 4005)2

Eq. 5.7-3 I-P

for standard air. They can also be calculated from the dynamic loss coeffient:

(Q/1.29)2SR (2.83/1.29)2 (34.49 + 22.99 + 17.25) (2.195)2 (74.73) (4.818)(74.73) 360 Pa

= = = = =

(Q/4005)2SR (6000/4005)2 (0.3 + 0.2 + 0.15) (1.498)2 (0.65) (2.244)(0.65) 1.45 in. wg

I-P: Pt loss

SR = System resistance factor, m-4 (ft-4) = 1/Ao2

= = = = =

5.7.2 System resistance factors in parallel. Similar relationships can be established for flow through parallel systems. The total pressure loss through each branch of a parallel system must be equal to establish equilibrium. In Figure 5S the system resistance factor of each branch is given as 0.3 and 0.2; and since the pressure loss will be equal in both branches we can equate these losses.

Air Systems | 27

RESISTANCE PLANE 1 ELEMENT

PLANE 2

RESISTANCE ELEMENT

E

0

ENTRY

FLOW

ATMOSPHERIC PRESSURE Pv

FAN

OUTLET

Pv2

PtF Pt2

Pv0 = Pt0

Ps2

+ PtE = 0 _

0

Ps

PsF

Pt

ABSOLUTE PRESSURE

Ps1 Pvt

Figure 5Q - Fan with Inlet and Discharge Resistance - Evasé Outlet Added

RESISTANCE 1

RESISTANCE 2

RESISTANCE 3

SR1 = 34.49 m-4 (0.3 ft.-4)

SR2 = 22.99 m-4 (0.2 ft.-4)

SR2 = 17.25 m-4 (0.15 ft.-4)

Figure 5R - Resistance in Series

RESISTANCE 1

Q1

SR1 = 34.49 m-4 (0.3 ft.-4) QT

RESISTANCE 2 SR2 = 22.99 m-4 (0.2 ft.-4) Figure 5S - Resistance in Parallel 28 | Air Systems

Q2

2

2

⎡ Q1 ⎤ ⎡ Q2 ⎤ ⎢ C ⎥ SR1 = ⎢ C ⎥ SR2 ⎣ ⎦ ⎣ ⎦

⎡ 0.4472 ⎤ = QT ⎢ ⎥ ⎣ 0.9949 ⎦

Where:

= QT (0.4495)

C = 1.29 (for SI units) = 4005 (for I-P units)

If:

And:

QT = 0.472 m3/s or (1000 cfm)

SR2 Q1 = Q2 SR1

And: Q1 = QT (0.4495)

⎡ S ⎤ Q1 = Q2 ⎢ R2 ⎥ ⎢⎣ SR1 ⎥⎦

Eq. 5.7-6

Q1 = (0.472 m3/s)(0.4495) or (1000 cfm)(0.4495) = 0.212 m3/s or (449.5 cfm)

Substituting: Q2 = QT - Q1 ⎡ SR2 Q1 = QT ⎢ ⎢⎣ SR1 + SR2

Then:

And: ⎤ ⎥ ⎥⎦

For the example shown in Figure 5S:

Eq. 5.7-7

Q2 = QT - Q1 = 0.472 m3/s - 0.212 m3/s or (1000 cfm - 449.5 cfm) = 0.260 m3/s or (550.5 cfm) It can further be demonstrated that the relationship of the system resistance factors for parallel systems is:

SI: ⎡ SR2 Q1 = QT ⎢ ⎢⎣ SR1 + SR2

⎤ ⎥ ⎥⎦

⎡ ⎤ 22.99 = QT ⎢ ⎥ ⎢⎣ 34.49 + 22.99 ⎥⎦ 4.795 ⎡ ⎤ = QT ⎢ ⎥ ⎣ 5.873 + 4.795 ⎦ ⎡ 4.795 ⎤ = QT ⎢ ⎥ ⎣10.667 ⎦ = QT ( 0.4495 ) I-P: ⎡ ⎤ SR2 Q1 = QT ⎢ ⎥ ⎢⎣ SR1 + SR2 ⎥⎦ ⎡ ⎤ 0 .2 = QT ⎢ ⎥ ⎢⎣ 0.3 + 0.2 ⎥⎦ ⎡ ⎤ 0.4472 = QT ⎢ ⎥ + ( 0 . 5477 0 . 4472 ) ⎣ ⎦

1 1 1 1 = + + ... + SRT SR1 SR2 SRn

System resistance factors can be quite useful in many system design and analysis problems.

5.8 System design and loss calculations There are a number of design methods for sizing duct work such as Equal Friction, Static Regain, etc., which are commonly used. The specific details of these various methods will not be covered in this publication, and the reader is referred to the ASHRAE Guide and the Industrial Ventilation Guide on this subject. The procedure for calculating the total pressure loss of the system is included as it is vital to the selection of the fan. Figure 5T, shown previously, will be used as an example of the method of calculating the system total pressure loss and fan static pressure required. It is necessary to include in the loss calculation all factors that contribute pressure loss, including System Effects. The general method is to determine the loss of each element of the system as they occur Air Systems | 29

and total the losses. In systems that include parallel branches it is customary to determine the loss of the branch path that has the highest resistance to establish the fan requirements. Other branch paths that have lower losses must have resistance added to them to balance the system since the total pressure loss of each path must be equal. 5.8.1 Example: System loss calculations for branch 1 (dynamic loss coefficient method) ITEM K - DISCHARGE DIFFUSER 0.3048 m (12 in.) diameter with a dynamic loss coefficient of 0.50. (1000 cfm) Airflow = 0.47 m3/s Pv = 25.08 Pa (0.101 in. wg) Pt Loss = 0.50 × 25.08 (0.50 × 0.101) = 12.54 Pa (0.051 in. wg) NOTE: Loss coefficient includes velocity lost at discharge and static regained by diffusion. ITEM J - BALANCING DAMPER, 0.3048 m (12 in.) diameter with a dynamic loss coefficient of 0.52. (1000 cfm) Airflow = 0.47 m3/s Pv = 25.08 Pa (0.101 in. wg) Pt Loss = 0.52 × 25.08 (0.52 × 0.101) = 13.04 Pa (0.053 in. wg) ITEM DUCT, 6.096 m (20 ft.) of 0.3048 m (12 in.) diameter. Airflow = 0.47 m3/s (1000 cfm) = 10.18 Pa (0.041 in. wg) Pt Loss (from Annex D) ITEM A - ELBOW, 0.3048 m (12 in.) diameter coefficient of 0.22. Airflow = 0.47 m3/s Pv = 25.08 Pa Pt Loss = 0.22 × 25.08 = 5.518 Pa

with dynamic loss (1000 cfm) (0.101 in. wg) (0.22 × 0.101) (0.022 in. wg)

ITEM DUCT, 6.096 m (20 ft.) of 0.3048 m (12 in.) diameter. Airflow = 0.47 m3/s (1000 cfm) = 10.18 Pa (0.041 in. wg) Pt Loss ITEM C - DIVIDED FLOW FITTING, 0.4064 m (16 in.) diameter to two 0.3048 m (12 in.) diameter with a main branch dynamic loss coefficient of 0.12. Airflow = 0.94 m3/s (2000 cfm) Pv = 32.29 Pa (0.13 in. wg) Pt Loss = 0.12 × 32.29 (0.12 × 0.13) = 3.88 Pa (0.016 in. wg) 30 | Air Systems

ITEM DUCT, 6.096 m (20 ft.) of 0.4064 m (16 in.) diameter. (2000 cfm) Airflow = 0.94 m3/s = 8.94 Pa (0.036 in. wg) Pt Loss ITEM D - DIVIDED FLOW FITTING, 0.4572 m (18 in.) to 0.4064 m (16 in.) and 0.3048 m (12 in.) diameter, with a main branch dynamic loss coefficient of 0. (3000 cfm) Airflow = 1.416 m3/s Pv = 44.70 Pa (0.18 in. wg) Pt Loss = 44.70 × 0 (0.18 × 0) = 0 Pa (0 in. wg) NOTE: The net loss in the main branch of this fitting is zero (0), since there is static regain to offset the loss. ITEM DUCT, 12.192 m (40 ft.), 0.4572 m (18 in.) diameter. (3000 cfm) Airflow = 1.42 m3/s Pt Loss = 21.36 Pa (0.086 in. wg) E through H -These losses may be included in the manufacturer's rating data. ITEM E - ENTRANCE, from plenum 0.4572 m (18 in.) diameter, dynamic loss coefficient of 0.5. (3000 cfm) Airflow = 1.42 m3/s Pv = 44.70 Pa (0.18 in. wg) Pt Loss = 44.70 × 0.5 (0.18 × 0.5) = 22.35 Pa (0.09 in. wg) ITEM F - COIL Airflow = 1.416 m3/s Pt Loss = 74.51 Pa (from manufacturer's data)

(3000 cfm) (0.3 in. wg)

ITEM G - FILTER Airflow = 1.416 m3/s Pt Loss = 86.93 Pa (from manufacturer's data)

(3000 cfm) (0.35 in. wg)

ITEM FAN, Bulkhead Discharge (SEF due to lack of fan discharge ductwork). Airflow = 1.416 m3/s (3000 cfm) Fan vel. = 6.53 m/s (1285 fpm) SEF = 39.74 Pa (0.16 in. wg) (from AMCA Publication 201) ITEM FAN, Fan enclosed in a cabinet; SEF due to plenum wall being too close to fan inlet). (3000 cfm) Airflow = 1.42 m3/s Inlet vel. = 5.31 ms (1045 fpm) SEF = 7.45 Pa (0.03 in. wg) (from AMCA Publication 201)

ITEM H - INTAKE LOUVER Airflow = 1.42 m3/s Pt Loss = 19.87 Pa (from manufacturer's data)

(3000 cfm) (0.08 in. wg)

SYSTEM TOTAL PRESSURE LOSS BRANCH 1 = 336.4 Pa (1.355 in. wg). 5.8.2 Example: System loss calculations for branch 2: ITEM K - DIFFUSER Pt Loss = 12.42 Pa (0.050 in. wg) NOTE: Loss coefficient includes velocity lost at discharge and static regained by diffusion. ITEM J - DAMPER Pt Loss = 12.91 Pa

(0.052 in. wg)

ITEM DUCT, 6.096 m (20 ft.) of 0.3048 m (12 in.) diameter. = 10.18 Pa (0.041 in. wg) Pt Loss ITEM B - 45° ELBOW, 0.3048 m (12 in.) diameter, dynamic loss coefficient of 0.13. Airflow = 0.47 m3/s (1000 cfm) Pv = 25.08 Pa (0.101 in. wg) Since the elbow is located one duct diameter from the divided flow fitting there will be a System Effect loss that will need to be included. K factor from Figure 4H = 1.08 Pt Loss = 25.08 × 0.13 × 1.08 (0.101 × 0.13 × 1.08) = 3.52 Pa (0.014 in. wg) ITEM C - DIVIDED FLOW FITTING, 0.4064 m (16 in.) diameter to 0.3048 m (12 in.) diameter with a branch dynamic loss coefficient of 0.46. Airflow = 0.94 m3/s (2000 cfm) Pv = 32.29 Pa (0.13 in. wg) K factor from Figure 4H = 1.08 = 0.46 × 32.29 × 1.08 (0.46 × 0.13 × 1.08) Pt Loss = 16.04 Pa (0.065 in. wg) Losses for the balance of the system are the same as those previously calculated. These total 281.1 Pa (1.132 in. wg). SYSTEM TOTAL PRESSURE LOSS BRANCH 2 = 336.2 Pa (1.354 in. wg). The loss through branch 2 is equal to the loss through branch 1, and the branches are in balance. Similar calculations for branch 3 show that it has a lower resistance than branches 1 and 2, and some

resistance will need to be added by adjusting its damper to balance the system. 5.8.3 Fan selection for the example system. The total system pressure loss of the branch with the highest resistance must be used to determine the fan pressure requirements. The fan will need to provide a total pressure of 336.2 Pa (1.354 in. wg) at 1.42 m3/s (3000 cfm). The fan static pressure PsF is equal to the fan total pressure PtF minus the calculated fan discharge velocity pressure Pv2. PsF = PtF - Pv2 = 336.2 - 25.58 = 310.62 Pa or (1.354 - 0.103 = 1.251 in. wg) This procedure applies to any system regardless of its complexity. The important point is that all losses, including System Effect losses, need to be included in the calculations. In this, as in many systems, the System Effect losses are a significant portion of the total pressure loss. AMCA Publication 201 should be used to determine the System Effect losses for various fan inlet and outlet conditions. Figure 5W shows the point of operation of this system, where the system resistance curve intersects the fan performance curve.

5.9 Density effects in air systems Since the density of the air varies with temperature, pressure (altitude), and humidity, it is necessary to evaluate the effect of density on the system design and fan performance. Because of the variations in density encountered in all air systems, a standard density was established, and is used as the basis for determining fan performance and system pressure losses. Standard air density is defined as air with a density of 1.2 kg/m3 (0.075 lbm/ft3). Fan performance ratings and system element pressure loss ratings are based on handling air at standard density. The system designer must evaluate the actual air density that will be handled by the system in order to properly determine the volume of flow required and the actual pressure losses in the system. Since fans are essentially constant volume machines, the volume of air handled by the fan will remain constant regardless of the density, but the total pressure developed by the fan and the power required by the fan will vary in direct proportion to the density. Similarly, the pressure losses in the system due to friction and shock losses will also vary directly with density. In many applications it is actually the mass flow rate that is important, and, therefore, the volume of air required should be determined from the mass flow Air Systems | 31

12.2m

6.1m

6.1m

FAN D

1.42m3/s

0.46m DIA. E COIL F FILTER G LOUVER H

C

A

0.94m3/s

0.472m3/s

0.41m DIA.

0.30m DIA.

B

B

0.30m DIA. 0.472m3/s

0.30m DIA. 0.472m3/s

J

DAMPER

J

DIFFUSER K

K

J

K

3

2

1

SI

40 ft.

20 ft.

20 ft.

FAN D

3000 CFM

C 2000 CFM

18” DIA. E COIL F FILTER G LOUVER H

16” DIA.

12” DIA.

B

B

12” DIA. 1000 CFM

12” DIA. 1000 CFM

J

J

DIFFUSER K

2

I-P

Figure 5T - Typical Air System

DAMPER

J

K

K 3

32 | Air Systems

A 1000 CFM

1

0.3810m DWDI FAN 0.2165M2 OUTLET AREA 1316 RPM

600

PRESSURE, Pa

DESIGN SYSTEM

300

PT vs FLOW

0

0.5

1.0

1.5

2.0

3

m /sec

15” DWDI FAN 2.33 FT2 OUTLET AREA 1316 RPM

PRESSURE, H2O in. wg

2

DESIGN SYSTEM

1

PT vs FLOW

0

1

2

3

4

CFM × 1000

Figure 5W - Fan Performance Versus System Air Systems | 33

rate required at the design conditions. As an example, in a system requiring a mass flow rate of 226.8 kg/min (500 lbm/min) at 121.1°C (250°F) and at an altitude of 914.4 m (3000 ft), the air density from Annex E is 0.800 kg/m3 (.05 lbm/ft3). The required volume flow rate can then be determined by dividing the mass flow rate by the design density:

conditions (Annex E) has a density of 1.072 kg/m3 (0.067 lbm/ft3). The power at these conditions would be the power for standard air multiplied by the density ratio:

SI (Volume flow rate) 226.8 kg/min ÷ (60 s/min × 0.800 kg/m3) = 4.725 m3/s

I-P (Power required at 68°F and 3000 ft) 14.5 BHP × (0.067 lbm/ft3 ÷ 0.075 lbm/ft3) = 12.95 BHP

I-P (Volume flow rate) 500 lbm/min ÷ 0.5 lbm/ft3 = 10000 cfm

In normal HVAC applications the effects of density changes other than for operation at higher altitudes, are quite often ignored and the system design is based on handling standard air. The system designer should, however, be aware of the effects of density change and take them into consideration when making field measurements of system performance or balancing the system.

The pressure loss in the system would be calculated based on 4.725 m3/s (10000 cfm) at standard air density of 1.2 kg/m3 (.075 lbm/ft3). If in this example the pressure loss at standard air is 1490.16 Pa (6 in. wg) the pressure loss at actual conditions would be this value multiplied by the density ratio: SI (Pressure loss - actual conditions) 1490.15 Pa × (0.8 kg/m3 ÷ 1.2 kg/m3) = 993.44 Pa I-P (Pressure loss - actual conditions) 6 in. wg × (.05 lbm/ft3 ÷ .075 lbm/ft3) = 4 in. wg A fan for this system must be selected based on its performance at standard conditions. For this example select the fan for a performance of 4.725 m3/s (10000 cfm) at 1490.16 Pa (6 in. wg) static pressure. Determine the fan power required at standard density from the fan performance data. The power required at actual conditions would then be calculated by multiplying the catalog fan power by the density ratio. In this example the fan power required at standard conditions is 10.82 kW (14.5 BHP). At the actual operating conditions of 121.1°C (250°F) air at 914.4 m (3000 ft) altitude, the fan power required would be the power for standard air multiplied by the density ratio: SI (Power at actual conditions) 10.82 kW × (0.8 kg/m3 ÷ 1.2 kg/m3) = 7.21 kW I-P (Power at actual conditions) 14.5 BHP × (0.05 lbm/ft3) ÷ .075 lbm/ft3) = 9.67 BHP Consider other density effects when selecting a fan for elevated temperature operations. If in this example the fan would be required to start and run some period of time at normal temperature as the system warms up, the motor should be selected on the basis of the cold air density. Since the fan will be at 914.4 m (3000 ft) altitude, the air density at 20°C (68°F) and 914.4 m (3000 ft) should be used to determine the required fan power. Air at these 34 | Air Systems

SI (Power required at 20°C and 914.4 m) 10.82 kW × (1.072 kg/m3 ÷ 1.2 kg/m3) = 9.67 kW

Use Annex C to determine the density of air over a range of barometric pressures, temperatures, and relative humidities. Annex E gives the density ratios for a wide range of temperatures and altitudes. The air density at the various conditions is obtained by multiplying standard air density of 1.2 kg/m3 (0.075 lbm/ft3) by the factors shown in the table.

6. System Design and Tolerances Before making a final determination of the fan selection there are several factors in the design of a system and the selection of a fan that need to be understood and evaluated: a) The effect of variation in the resistance of the actual installed system versus the resistance of the designed system, i.e.; point of operation b) The fan performance characteristics system/performance tolerances

and

c) The effect of changes in the system, either intentionally or unintentionally, on the point of operation d) The upper and lower system resistance design points in systems that have variable resistance characteristics (constant volume systems), or that have variable fan performance characteristics (variable air volume systems).

6.1 Point of operation The system resistance of the actual installed system can vary substantially from that calculated for the system design, because of a number of factors:

a) The installed system is different from the designed system, such as the addition of elbows and offsets to meet field conditions, failure to provide turning vanes in elbows, or the change in position of various system elements with respect to each other;

These curves are only shown to make the system designer aware of the effect of the slope of the fan curve on the expected system performance with variations in system resistance. Many other factors enter into the determination of the best fan type and size for a given application.

b) Excessive leakage or increased resistance due to poor quality workmanship at the installation;

6.2.2 Fan performance tolerance. The fan performance also has a tolerance which must be considered. The AMCA Check Test Tolerances are described in AMCA Publication 211, Certified Ratings Program -Air Performance, Product Rating Requirement Subsection B. The AMCA Check Test Tolerance is shown on Figure 6B. This tolerance is to be applied along a parabolic system line. The power required by the AMCA Check Test Tolerances shall not exceed the rated data at the measured volume by more than 5% or 37 watts, whichever is greater. The fan curve in Figure 6B has dashed lines indicating the tolerance range of fan performance, and when combined with the system resistance tolerance curves, an area of probable system performance is indicated by the tolerance limits shown for the system resistance and the tolerance limits shown for the fan flow-pressure. As can be seen, the probable flow range could be from 3.15 m3/s (6674 cfm) to 3.39 m3/s (7185 cfm) which is -4.7 to +2.6%. An installed system tolerance range approaching ± 5.0% of flow could be expected.

c) Loss coefficients of the various system elements such as coils, filters, dampers, diffusers, elbows, etc., improperly accounted for; d) System Effects: • not properly accounted for • ignored in the original system design • not accounted for because of on-site installation changes. The degree to which all of the various tolerances and the field changes affect the actual system resistance varies quite widely. Experience indicates that the difference to be expected between the calculated and actual system resistance can be as much as ±10%. In extreme cases, greater system resistance differentials have been experienced. Not accounting for system effects in the design will result in a higher system resistance and reduced flow.

6.2 Fan performance There is a wide variety of basic fan designs in axial, centrifugal and mixed flow variations. Curves of several typical basic fans, all selected for a point of operation of 3.30 m3/s (7000 cfm) at a static pressure of 1490.16 Pa (6 in. wg), are shown on Figure 6A. All of these fans pass through the design point of operation, but with different slopes. The point of highest efficiency will typically occur somewhat to the right of the peak pressure point. 6.2.1 System resistance effect on performance. The system resistance curves for the point of operation and curves for ±10% and ±25% of the design pressure are also shown in Figure 6A. The intersection of these system lines with the various fan curves show what range of volume performance can be expected for each fan over this range of system resistance. The amount of variation in flow rate with changes in the system resistance will be dependent on the slope of the fan performance curve in the range of operation.

6.2.3 Performance safety factor. Evaluate the fan performance tolerance and system resistance tolerances to determine if the lower or upper limits of the probable flow in the system are acceptable. The combination of these tolerances should also be evaluated to ensure that the high side system resistance curve does not fall into the unstable portion of the fan curve. With a few exceptions, all fans have an unstable range of performance. Operation in this area of the curve should be avoided and precautions taken to ensure operation outside of the unstable area at the highest expected system resistance. 6.2.3.1 Static pressure safety factor. It has been common practice among system designers to apply a performance safety factor to the calculated system requirements. This is often accomplished by adding a nominal percentage of pressure to the system pressure requirements. Some system designers will size the system for a higher flow rate than is required. The use of safety factors is discouraged when all system components and system effects are properly accounted for. The use of safety factors is not required when system effect factors and all known losses are accounted for.

Air Systems | 35

6.3 Effects of system changes Some air systems are designed to operate at more than one system condition, such as an exhaust system serving multiple inlets where some of the inlets can be closed off, or supply systems where some of the outlets can be closed off or dampered for reduced flow. The effect of these changes in the system need to be evaluated in the system design and the selection of a fan for this service. The main concern would be that the fan is not forced to operate in the unstable range. Also, the fan performance should be such that the system performance is acceptable over the range of operating conditions desired. The motor must be selected to cover this range of operating conditions. The system resistance and system performance, for the example used in Section 5.8, were calculated for the design condition and also for two other system conditions. [Figure 6C illustrates the effect of system changes.] The system resistance and flow were calculated for the condition where all the dampers were open and for the condition where one of the dampers was closed. Each of these conditions has a different system resistance curve resulting in a different operating point on the fan curve and a different total flow for the system. The flow in each branch of the system will also change. This leads to an important conclusion: IN A FIXED SYSTEM, A CHANGE IN RESISTANCE IN ANY ELEMENT WILL CHANGE THE TOTAL SYSTEM RESISTANCE, AND AS A RESULT, CHANGE THE POINT OF OPERATION ON THE FAN CURVE AND THE FLOW RATE THROUGH ALL OTHER ELEMENTS OF THE SYSTEM. It is because of this interaction of the total system with changes in any part of the system that the job of balancing a system is very difficult.

6.4 Variable systems Where systems are designed to be variable over some range of operation, or where both the fan and system are variable, the point of operation needs to be evaluated at the upper and lower limits of operation, relative to the tolerance of the fan and system.

36 | Air Systems

Figure 6D shows a typical fan curve with system resistance curves for a variable system where the system pressure is allowed to vary as the system demand for airflow changes. The system may be varied by volume control, dampers, or other control devices to provide a varying flow rate as demanded by the system. The tolerance ranges are shown for both the fan and the system resistance. The most critical point in the design of this type of system will be at the low flowhigh pressure condition. The fan selection and system limits should be such that the fan will operate in the stable portion of the fan curve at the maximum resistance condition. Make sure the fan power requirements over the tolerance range can be met by the motor selected. Figure 6E shows a typical set of fan curves for a centrifugal fan with inlet vane control with system resistance curves for a variable resistance system. In this system, both the pressure and flow characteristics of the fan are varied by changing the inlet vanes position to meet the flow rate demand of the system. Similar systems employ axial fans with variable pitch control, or fans with variable speed capability. The critical area of fan selection is near the peak of the pressure curve. Almost all fans exhibit some degree of instability to the left of the peak pressure point. It is wise to avoid operation in this range without the expressed approval of the fan manufacturer. There are many system variations to meet various design criteria that the designer may encounter. Not all of the possibilities can be covered in the scope of this publication. If fan users apply the principles outlined in this publication to the specific system, they can expect to design a good, functional system and avoid many of the pitfalls often encountered in air systems.

PRESSURE

1490.16 Pa (6.0 in. wg)

+25% +10% -10%

BACKWARD INCLINED FAN

-25%

RADIAL BLADE FAN

VANEAXIAL FAN

3.3m/s (7000 cfm) VOLUME FLOWRATE

Figure 6A - Performance of Example Fans with System Varations

+10% DESIGN POINT OF OPERATION PRESSURE

-10% FAN CURVE

POTENTIAL SYSTEM FLOW RANGE 3.15 m3/s (6701 cfm)

3.39 m3/s (7210 cfm)

3.3 m3/s (7000 cfm)

AMCA CERTIFIED RATING TOLERANCE

VOLUME FLOWRATE Figure 6B - Air Performance with Certified Ratings Tolerance

Air Systems | 37

DWDI FAN

TOTAL PRESSURE

DESIGN SYSTEM

ONE DAMPER CLOSED ALL DAMPERS CLOSED

Pt vs FLOW

VOLUME FLOWRATE Figure 6C - Fan and System Curves - System Changes

+10% HIGH PRESSURE DESIGN POINT

-10%

FAN CURVE

PRESSURE

+10%

LOW PRESSURE DESIGN POINT

AMCA CERTIFIED RATING TOLERANCE

VOLUME FLOWRATE Figure 6D - Fan and System Curves - Variable Resistance System 38 | Air Systems

-10%

PRESSURE

+10%

-10%

MAXIMUM FLOW DESIGN POINT

VARIABLE VOLUME SYSTEM RESISTANCE

IVC SETTING 1

IVC SETTING 2

IVC SETTING 3

IVC SETTING 4

STATIC PRESSURE CONTROL POINT

VOLUME FLOWRATE

Figure 6E - Fan and System Curves - Variable Volume System

Air Systems | 39

Annex A. SI / I-P Conversion Table Conversion factors between SI and I-P systems:

Quantity

I-P to SI

SI to I-P

Length

(ft) 0.3048 = m

(m) 3.2808 = ft

Mass (weight)

(lbs) 0.4536 = kg

(kg) 2.2046 = lbs.

Time

The unit of time is the second in both systems

Velocity

(ft-s) 0.3048 = ms (ft/min) 0.00508 = ms

(ms) 3.2808 = fts (ms) 196.85 = ft/min

Acceleration

(in./s2) 0.0254 = m/s2

(m/s2) 39.370 = in.s/2

Area

(ft2) 0.09290 = m2

(m2) 10.764 = ft2

Volume Flow Rate

(cfm) 0.000471948 = m3/s

(m3/s) 2118.88 = cfm

Density

(lb/ft3) 16.01846 = kg/m3

(kg/m3) 0.06243 = lb/ft3

Pressure

(in. wg) 248.36 = Pa (in. wg) 0.24836 = kPa (in. wg) 3.3864 = kPa

(Pa) 0.004026 = in. wg (kPa) 4.0264 = in. wg (kPa) 0.2953 = in. Hg

Viscosity: Absolute Kinematic

(lbm/ft-s) 1.4882 = Pa s (ft2/s) 0.0929 = m2/s

(Pa s) 0.6719 = (lbm/ft-s) (m2/s) 10.7639 = ft2/s

Gas Constant

(ft lb/lbm-°R) 5.3803 = J-kg/K

(j-kg/K) 0.1858 = (ft lb/lbm-°R)

Temperature

(°F - 32°)/1.8 = °C

(1.8 × °C) + 32° = °F

Power

(BHP) 746 = W (BHP) 0.746 = kW

(W)/746 = BHP (kW)/0.746 = BHP

40 | Air Systems

Annex B. Standard Atmospheric Data Versus Altitude Charts

Chart B.1 - SI Standard Atmospheric Data Versus Altitude Z Altitude

t Temperature

p Atmospheric Pressure

ρ Gas Density

μ Absolute Viscosity

v Kinematic Viscosity

c Speed of Sound

m

°C

kPa

kg/m3

Pa•s

m2/s

m/s

0 100 200 300

15.00 14.35 13.70 13.05

101.32 100.13 98.94 97.77

1.230 1.215 1.201 1.189

1.793x10-5 1.790x10-5 1.786x10-5 1.784x10-5

1.456×10-5 1.473×10-5 1.487×10-5 1.500×10-5

340.43 340.05 339.66 339.28

400 500 600 700

12.40 11.76 11.11 10.46

96.61 95.46 94.32 93.20

1.177 1.166 1.155 1.145

1.780x10-5 1.777x10-5 1.774x10-5 1.771x10-5

1.512×10-5 1.524×10-5 1.536×10-5 1.546×10-5

338.89 338.51 338.19 337.73

800 900 1000 1100

9.81 9.16 8.51 7.86

92.08 90.98 89.88 88.80

1.134 1.123 1.112 1.102

1.768x10-5 1.765x10-5 1.761x10-5 1.758x10-5

1.559×10-5 1.571×10-5 1.584×10-5 1.595×10-5

337.34 336.95 336.57 336.18

1200 1300 1400 1500

7.21 6.56 5.90 5.25

87.72 86.66 85.61 84.56

1.091 1.080 1.069 1.058

1.755x10-5 1.751x10-5 1.748x10-5 1.745x10-5

1.609×10-5 1.621×10-5 1.635×10-5 1.649×10-5

335.79 335.40 335.01 334.62

1600 1700 1800 1900

4.60 3.95 3.30 2.65

83.53 82.50 81.49 80.49

1.047 1.037 1.026 1.016

1.741x10-5 1.738x10-5 1.735x10-5 1.732x10-5

1.663×10-5 1.676×10-5 1.691×10-5 1.705×10-5

334.22 333.83 333.44 333.05

2000 2100 2200 2300

2.00 1.35 0.70 0.53

79.49 78.51 77.54 76.57

1.006 0.996 0.986 0.976

1.728x10-5 1.725x10-5 1.722x10-5 1.718x10-5

1.718×10-5 1.732×10-5 1.746×10-5 1.760×10-5

332.66 332.26 331.87 331.48

2400 2500 2600 2700

-0.60 -1.25 -1.90 -2.55

75.62 74.68 73.74 72.82

0.967 0.957 0.948 0.938

1.715x10-5 1.712x10-5 1.708x10-5 1.705x10-5

1.774×10-5 1.789×10-5 1.802×10-5 1.818×10-5

331.08 330.69 330.29 329.90

2800 2900 3000 3100

-3.20 -3.85 -4.50 -5.15

71.91 71.00 70.11 69.23

0.929 0.919 0.909 0.900

1.702x10-5 1.699x10-5 1.695x10-5 1.692x10-5

1.832×10-5 1.845×10-5 1.865×10-5 1.880×10-5

329.50 329.10 328.71 328.31

3200 3300 3400 3500

-5.80 -6.46 -7.11 -7.76

68.35 67.48 66.62 65.77

0.890 0.880 0.871 0.862

1.689x10-5 1.685x10-5 1.682x10-5 1.679x10-5

1.898×10-5 1.914×10-5 1.931×10-5 1.948×10-5

327.51 327.11 326.70 326.70

Air Systems | 41

Chart B.2 - I-P Standard Atmospheric Data Versus Altitude Z Altitude

t Temperature

p Atmospheric Pressure

ρ Air Density

μ Dynamic Viscosity

ν Kinematic Viscosity

c Speed of Sound

ft

°F

in. Hg

lbm/ft3

lbm/ft-s

ft2/s

ft/s

-1000 -500

62.6 60.8

31.02 30.47

.0787 .0776

1.212×10-5 1.208×10-5

1.538×10-4 1.556×10-4

1120.7 1118.8

0

59.0

29.92

.0765

1.205×10-5

1.576×10-4

1116.9

500 1000 1500 2000

57.2 55.4 53.7 51.9

29.38 28.86 28.33 27.82

.0754 .0743 .0732 .0721

1.202×10-5 1.198×10-5 1.195×10-5 1.192×10-5

1.595×10-4 1.614×10-4 1.633×10-4 1.653×10-4

1115.0 1113.1 1111.1 1109.2

2500 3000 3500 4000

50.1 48.3 46.5 44.7

27.32 26.82 26.33 25.84

.0710 .0700 .0689 .0679

1.189×10-5 1.185×10-5 1.182×10-5 1.179×10-5

1.673×10-4 1.694×10-4 1.714×10-4 1.735×10-4

1107.3 1105.3 1103.4 1101.4

4500 5000 5500 6000

43.0 41.2 39.4 37.6

25.37 24.90 24.43 23.98

.0669 .0659 .0649 .0639

1.175×10-5 1.172×10-5 1.169×10-5 1.165×10-5

1.757×10-4 1.778×10-4 1.800×10-4 1.823×10-4

1099.5 1097.5 1095.6 1093.6

6500 7000 7500 8000

35.8 34.0 32.3 30.5

23.53 23.09 22.65 22.22

.0630 .0620 .0610 .0601

1.162×10-5 1.158×10-5 1.155×10-5 1.152×10-5

1.846×10-4 1.869×10-4 1.892×10-4 1.916×10-4

1091.7 1089.7 1087.7 1085.7

8500 9000 9500 10000

28.7 26.9 25.1 23.3

21.80 21.39 20.98 20.58

.0592 .0583 .0574 .0565

1.148×10-5 1.145×10-5 1.142×10-5 1.138×10-5

1.904×10-4 1.965×10-4 1.990×10-4 2.015×10-4

1083.8 1081.8 1079.8 1077.8

42 | Air Systems

Annex C. Psychrometric Density Tables

Chart C.1 - Psychrometric Density Table (SI) Density of Saturated Air for Various Barometric Conditions - kg/m3 Dry-Bulb Temp. °C

-2.0 -1.5 -1.0 -0.5 0.0 0.5 1.0 1.5 2.0 2.5 3.0 3.5 4.0 4.5 5.0 5.5 6.0 6.5 7.0 7.5 8.0 8.5 9.0 9.5 10.0 10.5 11.0 11.5 12.0 12.5 13.0 13.5 14.0 14.5 15.0 15.5 16.0 16.5 17.0 17.5 18.0

Barometric Pressure kPa 97 1.244981 1.242122 1.239396 1.236782 1.234260 1.231812 1.229423 1.227079 1.224768 1.222480 1.220207 1.217942 1.215680 1.213416 1.211147 1.208871 1.206587 1.204295 1.201994 1.199687 1.197375 1.195060 1.192743 1.190428 1.188116 1.185810 1.183512 1.181224 1.178948 1.176683 1.174432 1.172192 1.169963 1.167742 1.165527 1.163312 1.161092 1.158860 1.156606 1.154320 1.151991

98.5 1.263273 1.260977 1.258667 1.256345 1.254012 1.251672 1.249325 1.246973 1.244618 1.242261 1.239902 1.237545 1.235188 1.232834 1.230483 1.228135 1.225792 1.223453 1.221119 1.218791 1.216468 1.214150 1.211838 1.209530 1.207227 1.204927 1.202631 1.200338 1.198047 1.195757 1.193466 1.191174 1.188879 1.186581 1.184277 1.181965 1.179644 1.177313 1.174968 1.172609 1.170232

100 1.282390 1.280094 1.277753 1.275377 1.272975 1.270553 1.268119 1.265679 1.263236 1.260796 1.258360 1.255931 1.253510 1.251098 1.248697 1.246304 1.243921 1.241546 1.239179 1.236817 1.234459 1.232105 1.229752 1.227399 1.225045 1.222689 1.220330 1.217968 1.215603 1.213236 1.210866 1.208497 1.206131 1.203771 1.201420 1.199084 1.196770 1.194483 1.192231 1.190025 1.187875

101.5 1.302927 1.300086 1.297353 1.294710 1.292141 1.289629 1.287163 1.284731 1.282324 1.279934 1.277553 1.275177 1.272800 1.270421 1.268037 1.265645 1.263247 1.260842 1.258431 1.256015 1.253595 1.251173 1.248752 1.246334 1.243920 1.241512 1.239113 1.236723 1.234343 1.231974 1.229616 1.227266 1.224925 1.222588 1.220251 1.217911 1.215560 1.213191 1.210795 1.208361 1.205877

103 1.324194 1.322000 1.319731 1.317400 1.315018 1.312595 1.310140 1.307661 1.305166 1.302659 1.300147 1.297634 1.295123 1.292618 1.290121 1.287634 1.285157 1.282692 1.280239 1.277798 1.275367 1.272946 1.270533 1.268128 1.265728 1.263332 1.260938 1.258544 1.256148 1.253747 1.251342 1.248928 1.246506 1.244075 1.241632 1.239178 1.236712 1.234235 1.231747 1.229250 1.226746

104.5 1.340401 1.337965 1.335505 1.333027 1.330532 1.328024 1.325506 1.322979 1.320447 1.317912 1.315376 1.312841 1.310307 1.307778 1.305254 1.302735 1.300224 1.297720 1.295225 1.292738 1.290260 1.287790 1.285328 1.282875 1.280428 1.277988 1.275553 1.273122 1.270693 1.268266 1.265837 1.263406 1.260970 1.258527 1.256073 1.253607 1.251125 1.248624 1.246101 1.243553 1.240975

Air Systems | 43

Chart C.1 - Psychrometric Density Table (SI) (Continued) Density of Saturated Air for Various Barometric Conditions - kg/m3 Dry-Bulb Temp. °C

18.5 19.0 19.5 20.0 20.5 21.0 21.5 22.0 22.5 23.0 23.5 24.0 24.5 25.0 25.5 26.0 26.5 27.0 27.5 28.0 28.5 29.0 29.5 30.0 30.5 31.0 31.5 32.0 32.5 33.0 33.5 34.0 34.5 35.0 35.5 36.0 36.5 37.0 37.5 38.0

44 | Air Systems

Barometric Pressure kPa 97 1.148567 1.146325 1.144073 1.141813 1.139548 1.137279 1.135008 1.132735 1.130461 1.128188 1.125917 1.123646 1.121378 1.119111 1.116846 1.114582 1.112318 1.110055 1.107790 1.105523 1.103253 1.100978 1.098695 1.096404 1.094102 1.091787 1.089456 1.087106 1.084735 1.082339 1.079915 1.077460 1.074970 1.072440 1.069867 1.067247 1.064575 1.061846 1.059056 1.056198

98.5 1.167391 1.164887 1.162437 1.160033 1.157668 1.155335 1.153029 1.150742 1.148470 1.146207 1.143949 1.141691 1.139431 1.137164 1.134888 1.132601 1.130299 1.127983 1.125650 1.123300 1.120932 1.118548 1.116147 1.113730 1.111299 1.108856 1.106402 1.103942 1.101478 1.099014 1.096553 1.094100 1.091661 1.089240 1.086844 1.084478 1.082149 1.079865 1.077632 1.075460

100 1.185062 1.182780 1.180492 1.178197 1.175897 1.173591 1.171279 1.168962 1.166639 1.164311 1.161977 1.159639 1.157295 1.154946 1.152592 1.150234 1.147871 1.145503 1.143131 1.140754 1.138373 1.135988 1.133599 1.131206 1.128809 1.126408 1.124004 1.121596 1.119184 1.116769 1.114351 1.111930 1.109506 1.107079 1.104649 1.102216 1.099780 1.097342 1.094902 1.092459

101.5 1.203323 1.200987 1.198647 1.196304 1.193957 1.191607 1.189254 1.186898 1.184537 1.182174 1.179806 1.177435 1.175060 1.172681 1.170298 1.167912 1.165521 1.163126 1.160726 1.158323 1.155915 1.153503 1.151086 1.148664 1.146239 1.143808 1.141372 1.138932 1.136487 1.134037 1.131582 1.129122 1.126656 1.124186 1.121710 1.119229 1.116742 1.114250 1.111753 1.109249

103 1.225071 1.222584 1.220116 1.217665 1.215229 1.212804 1.210388 1.207980 1.205577 1.203177 1.200778 1.198380 1.195979 1.193576 1.191169 1.188756 1.186338 1.183912 1.181480 1.179039 1.176591 1.174134 1.171669 1.169195 1.166714 1.164226 1.161731 1.159230 1.156724 1.154213 1.151700 1.149185 1.146669 1.144155 1.141644 1.139139 1.136640 1.134151 1.131673 1.129210

104.5 1.240138 1.237641 1.235154 1.232675 1.230205 1.227740 1.225283 1.222830 1.220383 1.217939 1.215499 1.213061 1.210625 1.208190 1.205755 1.203320 1.200883 1.198445 1.196003 1.193559 1.191110 1.188656 1.186196 1.183730 1.181257 1.178775 1.176286 1.173786 1.171277 1.168756 1.166224 1.163679 1.161121 1.158549 1.155963 1.153361 1.150743 1.148108 1.145455 1.142784

Chart C.2 - Psychrometric Density Table (I-P) Density of Saturated Air for Various Barometric Conditions - lbm/ft3 Dry-Bulb Temp. °F

30 31 32 33 34 35 36 37 38 39 40 41 42 43 44 45 46 47 48 49 50 51 52 53 54 55 56 57 58 59 60 61 62 63 64

Barometric Pressure in. Hg 28.5 .07703 .07687 .07671 .07654 .07638 .07621 .07605 .07589 .07573 .07557 .07541 .07525 .07509 .07493 .07477 .07461 .07445 .07429 .07413 .07397 .07381 .07366 .07350 .07334 .07318 .07302 .07287 .07271 .07255 .07240 .07224 .07208 .07193 .07177 .07161

29.0 .07839 .07822 .07806 .07789 .07772 .07756 .07739 .07723 .07706 .07690 .07674 .07657 .07641 .07625 .07609 .07592 .07576 .07560 .07544 .07528 .07512 .07496 .07479 .07464 .07447 .07431 .07415 .07399 .07383 .07367 .07352 .07336 .07320 .07304 .07288

29.5 .07974 .07957 .07940 .07924 .07907 .07890 .07873 .07856 .07840 .07823 .07806 .07790 .07773 .07757 .07740 .07724 .07707 .07691 .07674 .07658 .07642 .07625 .07609 .07593 .07576 .07560 .07544 .07528 .07512 .07495 .07479 .07463 .07447 .07430 .07414

30.0 .08111 .08093 .08075 .08058 .08041 .08024 .07807 .07990 .07973 .07956 .07939 .07922 .09705 .07889 .07872 .07855 .07838 .07822 .07805 .07788 .07772 .07755 .07739 .07722 .07706 .07689 .07673 .07656 .07640 .07623 .07607 .07590 .07574 .07557 .07541

30.5 .08245 .08228 .08210 .08193 .08175 .08158 .08141 .08123 .08106 .08089 .08072 .08055 .08038 .08021 .08004 .07986 .07970 .07953 .07936 .07919 .07902 .07885 .07868 .07852 .07835 .07818 .07801 .07784 .07768 .07751 .07734 .07718 .07701 .07684 .07668

31.0 .08380 .08363 .08345 .08327 .08310 .08292 .08274 .08257 .08239 .08222 .08205 .08187 .08170 .08153 .08135 .08118 .08101 .08084 .08066 .08049 .08032 .08015 .07998 .07981 .07964 .07947 .07930 .07913 .07896 .07879 .07862 .07845 .07828 .07811 .07794

Air Systems | 45

Chart C.2 - Psychrometric Density Table (I-P) Density of Saturated Air for Various Barometric Conditions - lbm/ft3 Dry-Bulb Temp. °F

65 66 67 68 69 70 71 72 73 74 75 76 77 78 79 80 81 82 83 84 85 86 87 88 89 90 91 92 93 94 95 96 97 98 99 100

46 | Air Systems

Barometric Pressure in. Hg 28.5 .07145 .07130 .07114 .07098 .07083 .07067 .07051 .07035 .07020 .07004 .06988 .06972 .06956 .06940 .06925 .06909 .06893 .06877 .06861 .06845 .06829 .06812 .06796 .06780 .06764 .06748 .06731 .06715 .06698 .06682 .06665 .06648 .06632 .06615 .06598 .06581

29.0 .07272 .07256 .07240 .07224 .07208 .07192 .07176 .07160 .07144 .07128 .07112 .07096 .07080 .07064 .07048 .07032 .07015 .07000 .06983 .06967 .06950 .06934 .06917 .06901 .06885 .06868 .06852 .06835 .06818 .06801 .06785 .06768 .06751 .06734 .06717 .06700

29.5 .07398 .07382 .07366 .07350 .07333 .07317 .07301 .07285 .07268 .07252 .07236 .07220 .07203 .07187 .07171 .07155 .07138 .07122 .07105 .07089 .07072 .07056 .07039 .07022 .07005 .06989 .06972 .06955 .06938 .06921 .06904 .06887 .06870 .06853 .06835 .06818

30.0 .07525 .07508 .07492 .07475 .07459 .07442 .07426 .07410 .07393 .07377 .07360 .07343 .07327 .07310 .07294 .07277 .07261 .07244 .07227 .07211 .07194 .07177 .07160 .07143 .07126 .07109 .07092 .07075 .07058 .07041 .07024 .07006 .06989 .06972 .06954 .06937

30.5 .07651 .07634 .07618 .07601 .07584 .07568 .07551 .07534 .07517 .07501 .07484 .07467 .07451 .07434 .07417 .07400 .07383 .07366 .07349 .07333 .07316 .07299 .07281 .07264 .07247 .07230 .07213 .07195 .07178 .07161 .07143 .07126 .07108 .01091 .07073 .07055

31.0 .07770 .07760 .07744 .07727 .07710 .07693 .07676 .07659 .07642 .07625 .07603 .07591 .07574 .07557 .07540 .07523 .07506 .07489 .07472 .07454 .07437 .07420 .07403 .07385 .07368 .07351 .07333 .07316 .07298 .07280 .07263 .07245 .07227 .07209 .07191 .07174

50

0.1 20

0.2

0.3

0.5 0.4

0.7

1

2

3

4

5

7

10

63

1.2

1.4

1.6

50

1.8

2.0

80 2.5

3

10 100

4

0 3.5

12 5

5 6

7

200

8

0 10

16 9

12

14

0

5 31 1,000

50

0 2,000

0 5,000

80

63

0

40

0 AIR QUANTITY, L/s at 1.20 kg/m3 (ε = 0.09 mm)

500

18 16

0 20

20

25

20

12

10

00 10,000

30

35

50 45 40

00 20,000

16

30

25 70

80

50,000

60

00

40

20

50

30

00 25

70

s

50

LO CI TY m/ 90

200,000

9

20

0.2

0.3

0.5 0.4

0.7

1

2

3

4

5

7

10

20

30

40

50

70

100

400,000

mm R,

00

VE

IAM CT D

31 DU

ET E

40

100

Annex D. Friction Charts Chart D.1 - SI

FRICTION LOSS, Pa/m

Reprinted by permission of the American Society of Heating, Refrigerating and Air-Conditioning Engineers, Atlanta, Georgia, from the 1993 ASHRAE Handbook-Fundamentals.

Air Systems | 47

50

0.01 50

0.02

0.03

0.04

0.05

0.08

0.1

0.2

0.3

0.4

0.5

0.8

100

200

0

30

5

0

40

6

0

50

0

500

0

70 0

60

7

80

0

00

10

8

90

9 00

12

1,000

00

14

10

20 0 18 0 00 16 00

00

24

12

00

28

2,000

14

1

16

2

5,000

36

30 32

10,000

00

18

45

20,000

00

40

50,000

70

40

AIR QUANTITY, cfm at 0.075 lb/ft3 (ε = 0.0003 ft)

18

3

20 22 24 26

4

VE LO

90

5

80

00 70 65 00 0 60 0 0 55 0 0 50 0 00 45 00 40 00 36 00 32 00

100,000

0

200,000

00

80

00

0

00

12 00

10 90

m

TY , fp

CI

10 0 IAM

7

50 55 60

80

CT D DU

in. R,

48 | Air Systems ET E

10

400,000

0.01

0.02

0.03

0.04

0.05

0.08

0.1

0.2

0.3

0.4

0.5

0.8

1

2

3

4

5

7

10

Chart D.2 - I-P

FRICTION LOSS, in. of water per 100 ft of duct

Reprinted by permission of the American Society of Heating, Refrigerating and Air-Conditioning Engineers, Atlanta, Georgia, from the 1993 ASHRAE Handbook-Fundamentals. 4

3

Annex E. Air Density Correction Factor Charts Chart E.1 - Air Density Correction Factor (SI) (Multiply Standard Air Density, 0.075 lbm/ft3 × the Factor to obtain Density at Condition B.) Altitude, m Sea Level 300 600 900 1200 Barometer, mm Hg kPa

759.97 101.32

733.47 97.79

707.46 94.32

682.43 90.98

657.90 87.71

Air Temp. °C -20 0 20 40 60 80 100 120 140 160 180 200 220 240 260 280 300 320 340

1.22 1.12 1.03 .95 .93 .85 .79 .77 .73 .69 .65 .62 .57 .57 .56 .54 .52 .50 .49

1.12 1.04 .97 .91 .86 .80 .76 .72 .69 .65 .61 .57 .56 .55 .53 .52 .50 .48 .46

1.08 1.01 .94 .88 .83 .78 .73 .70 .67 .63 .59 .55 .54 .53 .51 .50 .48 .49 .48

1.04 .97 .90 .85 .80 .75 .70 .67 .64 .61 .57 .53 .52 .51 .49 .48 .46 .45 .43

1.00 .94 .87 .81 .77 .72 .68 .65 .62 .59 .55 .51 .50 .49 .47 .46 .44 .43 .41

Altitude, m

1500

1800

2100

2400

2700

Barometer, mm Hg kPa

634.34 84.57

611.3 81.50

588.98 78.52

567.17 75.62

546.30 72.83

.97 .90 .84 .78 .74 .69 .65 .62 .60 .57 .53 .49 .48 .47 .45 .44 .43 .41 .40

.93 .87 .81 .75 .71 .67 .63 .61 .58 .55 .51 .47 .46 .46 .44 .43 .41 .40 .39

.91 .84 .78 .73 .69 .65 .61 .59 .56 .53 .50 .47 .45 .46 .43 .42 .40 .39 .37

.87 .81 .75 .70 .67 .63 .59 .57 .54 .51 .49 .45 .43 .43 .41 .40 .38 .37 .36

.84 .80 .72 .68 .64 .60 .58 .58 .53 .49 .47 .43 .42 .41 .39 .38 .37 .35 .34

Air Temp. °C -20 0 20 40 60 80 100 120 140 160 180 200 220 240 260 280 300 320 340

Air Systems | 49

Chart E.2 - Air Density Correction Factor (I-P) (Multiply Standard Air Density, 0.075 lbm/ft3 × the Factor to obtain Density at Condition B.) Altitude, ft. -1000 Sea Level 1000 2000 3000

4000

Barometer, in. Hg in wg.

31.02 422.2

29.92 407.5

28.86 392.81

27.82 378.6

26.82 365.0

25.84 351.7

Air Temp. °F -40 0 40 70 100 150 200 250 300 350 400 450 500 550 600 700 800 900 1000

1.31 1.19 1.10 1.04 0.98 0.90 0.83 0.77 0.72 0.68 0.64 0.60 0.57 0.54 0.52 0.47 0.44 0.40 0.37

1.26 1.15 1.06 1.00 0.95 0.87 0.80 0.75 0.70 0.65 0.62 0.58 0.55 0.53 0.50 0.46 0.42 0.39 0.36

1.22 1.11 1.02 0.96 0.92 0.84 0.77 0.72 0.67 0.62 0.60 0.56 0.53 0.51 0.48 0.44 0.40 0.37 0.35

1.17 1.07 0.99 0.93 0.88 0.81 0.74 0.70 0.65 0.60 0.57 0.54 0.51 0.49 0.46 0.43 0.39 0.36 0.33

1.13 1.03 0.95 0.89 0.85 0.78 0.71 0.67 0.62 0.58 0.55 0.52 0.49 0.47 0.45 0.41 0.37 0.35 0.32

1.09 0.99 0.92 0.86 0.81 0.75 0.69 0.64 0.60 0.56 0.53 0.50 0.47 0.45 0.43 0.39 0.36 0.33 0.31

Altitude, ft.

5000

6000

7000

8000

9000

10,000

Barometer, in. Hg in wg.

24.90 338.9

23.98 326.4

23.09 314.3

22.22 302.1

21.39 291.1

20.58 280.1

Air Temp. °F -40 0 40 70 100 150 200 250 300 350 400 450 500 550 600 700 800 900 1000

1.05 0.95 0.88 0.83 0.78 0.72 0.66 0.62 0.58 0.54 0.51 0.48 0.45 0.44 0.41 0.38 0.35 0.32 0.30

1.01 0.91 0.85 0.80 0.75 0.69 0.64 0.60 0.56 0.52 0.49 0.46 0.44 0.42 0.40 0.37 0.33 0.31 0.29

0.97 0.89 0.82 0.77 0.73 0.67 0.62 0.58 0.54 0.51 0.48 0.45 0.43 0.41 0.39 0.35 0.32 0.30 0.27

0.93 0.85 0.79 0.74 0.70 0.65 0.60 0.56 0.52 0.49 0.46 0.43 0.41 0.39 0.37 0.34 0.31 0.29 0.26

0.90 0.82 0.76 0.71 0.68 0.62 0.57 0.58 0.50 0.47 0.44 0.42 0.39 0.38 0.35 0.33 0.30 0.28 0.26

0.87 0.79 0.73 0.69 0.65 0.60 0.55 0.51 0.48 0.45 0.42 0.40 0.38 0.36 0.34 0.32 0.29 0.27 0.25

50 | Air Systems

Fans and Systems ANSI/AMCA 210 Laboratory Methods of Testing Fans For Aerodynamic Performance Rating, offers the system design engineer guidance as to how the fan was tested and rated. AMCA Publication 201 Fans and Systems, helps provide guidance as to what effect the system and its connections to the fan have on fan performance. Recognizing and accounting for losses that affect the fan’s performance, in the design stage, will allow the designer to predict with reasonable accuracy, the installed performance of the fan.

1.1 Purpose This part of the AMCA Fan Application Manual includes general information about how fans are tested in the laboratory, and how their performance ratings are calculated and published. It also reviews some of the more important reasons for the "loss" of fan performance that may occur when the fan is installed in an actual system. Allowances, called System Effect Factors (SEF), are also given in this part of the manual. SEF must be taken into account by the system design engineer if a reasonable estimate of fan/system performance is to be determined.

1.2 Some limitations It must be appreciated that the System Effect Factors given in this manual are intended as guidelines and are, in general, approximations. Some have been obtained from research studies, others have been published previously by individual fan manufacturers, and many represent the consensus of engineers with considerable experience in the application of fans. Fans of different types and even fans of the same type, but supplied by different manufacturers, will not necessarily react with the system in exactly the same way. It will be necessary, therefore, to apply judgment based on actual experience in applying the SEF. The SEF represented in this manual assume that the fan application is generally consistent with the method of testing and rating by the manufacturer. Inappropriate application of the fan will result in SEF values inconsistent with the values presented. Mechanical design of the fan is not within the scope of this publication.

201

2. Symbols and Subscripts For symbols and subscripted symbols, see Table 2.1. For subscripts, see Table 2.2.

3. Fan Testing Fans are tested in setups that simulate installations. The four standard installation types are as shown in Figure 3.1. AMCA INSTALLATION TYPE A: Free Inlet, Free Outlet

AMCA INSTALLATION TYPE B: Free Inlet, Ducted Outlet

AMCA INSTALLATION TYPE C: Ducted Inlet, Free Outlet

AMCA INSTALLATION TYPE D: Ducted Inlet, Ducted Outlet

Figure 3.1 - Standard Fan Installation Types

3.1 ANSI/AMCA Standard 210 Most fan manufacturers rate the performance of their products from tests made in accordance with ANSI/AMCA 210 Laboratory Methods of Testing Fans for Aerodynamic Performance Rating. The purpose of ANSI/AMCA 210 is to establish uniform methods for laboratory testing of fans and other air moving devices to determine performance in terms of airflow,

Table 2.1 - Symbols and Subscripted Symbols

SYMBOL

DESCRIPTION

UNITS OF MEASURE SI I-P

A

Area of cross section

m2

ft2

D

Diameter, impeller

mm

in.

D

Diameter, Duct

m

ft

H

Fan Power Input

kw

hp

H/T

Hub-to-Tip Ratio

Dimensionless

Kp

Compressibility Coefficient

Dimensionless

Cp

Loss Coefficient

Dimensionless

N

Speed of Rotation

rpm

rpm

Ps

Fan Static Pressure

Pa

in. wg

Pt

Fan Total Pressure

Pa

in. wg

Pv

Fan Velocity Pressure

Pa

in. wg

pb

Corrected Barometric Pressure

kPa

in. Hg

PL

Plane of Measurement

---

---

Q

Airflow

m3/s

ft3/min

Re

Fan Reynolds Number

SEF

System Effect Factor

Pa

in. wg

td

Dry-Bulb Temperature

°C

°F

tw

Wet-Bulb Temperature

°C

°F

μ

Air Viscosity

Pa•s

lbm/ft•s

V

Velocity

m/s

fpm

W

Power Input to Motor

watts

watts

ηs

Fan Static Efficiency

%

%

ηt

Fan Total Efficiency

%

%

ρ

Air Density

kg/m3

lbm/ft3

Dimensionless

Table 2.2 - Subscripts SUBSCRIPT

DESCRIPTION

a c x 1 2 3 5 6 8

Atmospheric conditions Converted Value Plane 0, 1, 2, ...as appropriate Fan Inlet Plane Fan Outlet Plane Pitot Traverse Plane Plane 5 (nozzle inlet station in chamber) Plane 6 (nozzle discharge station in chamber) Plane 8 (inlet chamber measurement station)

52 | Fans and Systems

pressure, power, air density, speed of rotation and efficiency, for rating or guarantee purposes. Two methods of measuring airflow are included: the Pitot tube and the long radius flow nozzle. These are incorporated into a number of "setups" or "figures". In general, a fan is tested on the setup that most closely resembles the way in which it will be installed in an air system. Centrifugal and axial fans are usually tested with an outlet duct. Propeller fans are normally tested in the wall of a chamber or plenum. Power roof ventilators (PRV) are tested mounted on a curb exhausting from the test chamber. It is very important to realize that each setup in ANSI/AMCA 210 is a standardized arrangement that is not intended to reproduce exactly any installation likely to be found in the field. The infinite variety of possible arrangements of actual air systems makes it impractical to duplicate every configuration in the fan test laboratory.

3.2 Ducted outlet fan tests Figure 3.2 is a reproduction of a test setup from ANSI/AMCA 210. Note that this particular setup includes a long straight duct connected to the outlet of the fan. A straightener is located upstream of the Pitot traverse to remove swirl and rotational components from the airflow and to ensure that airflow at the plane of measurement is as nearly uniform as possible.

the fan outlet is limited to ensure that uniform airflow will be maintained. A steep transition, or abrupt change of cross section would cause turbulence and eddies. The effect of this type of airflow disturbance at the fan outlet is discussed later. Uniform airflow conditions ensure consistency and reproducibility of test results and permit the fan to develop its maximum performance. In any installation where uniform airflow conditions do not exist, the fan's performance will be measurably reduced. As illustrated in Figure 3.3 Plane 2, the velocity profile at the outlet of a fan is not uniform. The section of straight duct attached to the fan outlet controls the diffusion of the outlet airflow and establishes a more uniform velocity as shown in Figure 3.3 Plane X. The energy loss when a gas, such as air, passes through a sudden enlargement is related to the square of the velocity. Thus the ducted outlet with its more uniform velocity significantly reduces the loss at the point of discharge to the atmosphere. A manufacturer may test a fan with or without an inlet duct or outlet duct. For products licensed to use the AMCA Certified Ratings Seal, catalog ratings will state whether ducts were used during the rating tests. If the fans are not to be applied with the same duct(s) as in the test setup, an allowance should be made for the difference in performance that may result.

The angle of the transition between the test duct and

1

Transition Piece

2

Straightener

FOR FAN INSTALLATION TYPES: B: Free Inlet, Ducted Outlet D: Ducted Inlet, Ducted Outlet Figure 3.2 - Pitot Traverse in Outlet Duct Fans and Systems | 53

3.3 Free inlet, free outlet fan tests Figure 3.4 illustrates a typical multi-nozzle chamber test setup from ANSI/AMCA 210. This simulates the conditions under which most exhaust fans are tested and rated. Fan performance based on this type of test may require adjustment when additional accessories are used with the fan. Fans designed for use without duct systems are usually rated over a lower range of pressures. They are commonly cataloged and sold as a complete unit with suitable drive and motor.

3.4 Obstructed inlets and outlets The test setups in ANSI/AMCA 210 result in unobstructed airflow conditions at both the inlet and the outlet of the fan. Appurtenances or obstructions located close to the inlet and/or outlet will affect fan performance. Shafts, bearings, bearing supports and other appurtenances normally used with a fan should be in place when a fan is tested for rating. Variations in construction which may affect fan performance include changes in sizes and types of sheaves and pulleys, bearing supports, bearings and shafts, belt guards, inlet and outlet dampers, inlet vanes, inlet elbows, inlet and outlet cones, and cabinets or housings. Since changes in performance will be different for various product designs, it will be necessary to make suitable allowances based on data obtained from the applicable fan catalog or directly from the manufacturer. Most single width centrifugal fans are tested using Arrangement 1 fans. Some allowance for the effect of bearings and bearing supports in the inlet may be necessary when using Arrangement 3 or Arrangement 7. The various AMCA standard arrangements are shown on Figures 3.5, 3.6, and 3.7.

4. Fan Ratings 4.1 The Fan Laws It is not practical to test a fan at every speed at which it may be applied. Nor is it possible to simulate every inlet density that may be encountered. Fortunately, by use of a series of equations commonly referred to as the Fan Laws, it is possible to predict with good accuracy the performance of a fan at other speeds and densities than those of the original rating test. The performance of a complete series of geometrically similar (homologous) fans can also be 54 | Fans and Systems

calculated from the performance of smaller fans in the series using the appropriate equations. Because of the relationship between the airflow, pressure and power for any given fan, each set of equations for changes in speed, size or density, applies only to the same Point of Rating, and all the equations in the set must be used to define the converted condition. A Point of Rating is the specified fan operating point on its characteristic curve. The Fan Law equations are shown below as ratios. The un-subscripted variable is used to designate the initial or test fan values for the variable and the subscript c is used to designate the converted, dependent or desired variable. Qc = Q × (Dc/D)3 × (Nc/N) × (Kp/Kpc) Ptc = Pt × (Dc/D)2 × (Nc/N)2 × (ρc/ρ) × (Kp/Kpc) Pvc = Pv × (Dc/D)2 × (Nc/N)2 × (ρc/ρ) Psc = Ptc - Pvc Hc = H × (Dc/D)5 × (Nc/N)3 × (ρc/ρ) × (Kp/Kpc)

ηtc = (Qc × Ptc × Kp) / Hc ηtc = (Qc × Ptc × Kp) / (6362 • Hc)

(SI) (I-P)

ηsc = ηtc × (Psc/Ptc) These equations have their origin in the classical theories of fluid mechanics, and the accuracy of the results obtained is sufficient for most applications. Better accuracy would require consideration of Reynolds number, Mach number, kinematic viscosity, dynamic viscosity, surface roughness, impeller blade thickness and relative clearances, etc.

4.2 Limitations Under certain conditions the properties of gases change and there are, therefore, limitations to the use of the Fan Laws. Accurate results will be obtained when the following limitations are observed: a. Fan Reynolds Number (Re). The term Reynolds number is associated with the ratio of inertia to viscous forces. When related to fans, investigations of both axial and centrifugal fans show that performance losses are more significant at low Reynolds number ranges and are effectively negligible above certain threshold Reynolds numbers. In an effort to simplify the comparison of the Reynolds numbers of two fans, the fan industry

PL X

BLAST AREA

PL 2

DISCHARGE DUCT OUTLET AREA

CUTOFF

CENTRIFUGAL FAN

PL 2

PL X

AXIAL FAN Figure 3.3 - Controlled Diffusion and Establishment of a Uniform Velocity Profile in a Straight Length of Outlet Duct

38mm ±6mm (1.5in. ±0.25 in.) PL.5 PL.6

PL.8

PL.1

PL.2 0.5 M MIN.

0.2M MIN.

0.5M MIN.

0.2 M MIN. 0.3 M MIN.

t d2 AIRFLOW

M

FAN

VARIABLE SUPPLY SYSTEM

t d3 0.1 M MIN. SETTLING MEANS

SETTLING MEANS (See note 4)

Ps5

ΔP

Pt8

Figure 3.4 - Inlet Chamber Setup - Multiple Nozzles in Chamber (ANSI/AMCA 210-99, Figure 15) Fans and Systems | 55

ANSI/AMCA Standard 99-2404-03

Page 1 of 2

Drive Arrangements for Centrifugal Fans An American National Standard - Approved by ANSI on April 17, 2003

AMCA Drive Arrangement

1 SWSI

ISO 13349 Drive Arrangement

1 or 12 (Arr. 1 with sub-base)

Description

Fan Configuration

Alternative Fan Configuration

For belt or direct drive. Impeller overhung on shaft, two bearings mounted on pedestal base. Alternative: Bearings mounted on independant pedestals, with or without inlet box.

2 SWSI

For belt or direct drive.

2

Impeller overhung on shaft, bearings mounted in bracket supported by the fan casing. Alternative: With inlet box.

3 SWSI

3 or 11 (Arr. 3 with sub-base)

For belt or direct drive. Impeller mounted on shaft between bearings supported by the fan casing. Alternative: Bearings mounted on independent pedestals, with or without inlet box.

3 DWDI

6 or 18 (Arr. 6 with sub-base)

For belt or direct drive. Impeller mounted on shaft between bearings supported by the fan casing. Alternative: Bearings mounted on independent pedestals, with or without inlet boxes.

4 SWSI

4

For direct drive. Impeller overhung on motor shaft. No bearings on fan. Motor mounted on base. Alternative: With inlet box.

5 SWSI

5

For direct drive. Impeller overhung on motor shaft. No bearings on fan. Motor flange mounted to casing. Alternative: With inlet box.

AMCA International, Inc. | 30 W. University Dr. | Arlington Heights, IL, 60004-1893 | U.S.A

Figure 3.5 - AMCA Standard 99-2404 / Page 1 56 | Fans and Systems

AMCA 201-02 Page 2 of 2

ANSI/AMCA Standard 99-2404-03

AMCA Drive Arrangement

ISO 13349 Drive Arrangement

7 SWSI

7

Description

Fan Configuration

Alternative Fan Configuration

For coupling drive. Generally the same as Arr. 3, with base for the prime mover. Alternative: Bearings mounted on independent pedestals with or without inlet box.

7DWDI

17 (Arr. 6 with base for motor)

For coupling drive. Generally the same as Arr. 3 with base for the prime mover. Alternative: Bearings mounted on independent pedestals with or without inlet box.

8 SWSI

8

For direct drive. Generally the same as Arr. 1 with base for the prime mover. Alternative: Bearings mounted on independent pedestals with or without inlet box.

9 SWSI

9

For belt drive. Impeller overhung on shaft, two bearings mounted on pedestal base. Motor mounted on the outside of the bearing base. Alternative: With inlet box.

10 SWSI

10

For belt drive. Generally the same as Arr. 9 with motor mounted inside of the bearing pedestal. Alternative: With inlet box.

AMCA International, Inc. | 30 W. University Dr. | Arlington Heights, IL, 60004-1893 | U.S.A

Figure 3.6 - AMCA Standard 99-2404 / Page 2 Fans and Systems | 57

ANSI/AMCA Standard 99-3404-03

Page 1 of 1

Drive Arrangements for Axial Fans An American National Standard - Approved by ANSI on June 10, 2003 Note: All fan orientations may be horizontal or vertical

AMCA Drive Arrangement 1

ISO 13349 Drive Arrangement 1 12 (Arr. 1 with sub-base)

Description

Alternative Fan Configuration

Fan Configuration

For belt or direct drive. Impeller overhung on shaft, two bearings mounted either upstream or downstream of the impeller. Alternative: Single stage or two stage fans can be supplied with inlet box and/or discharge evasé.

3

3 11 (Arr. 3 with sub-base)

For belt or direct drive. Impeller mounted on shaft between bearings on internal supports. Alternative: Fan can be supplied with inlet box, and/or discharge evasé.

4

4

For direct drive. Impeller overhung on motor shaft. No bearings on fan. Motor mounted on base or integrally mounted.

M

M

M

M

Alternative: With inlet box and/or with discharge evasé.

7

7

For direct drive. Generally the same as Arr. 3 with base for the prime mover.

M

M

Alternative: With inlet box and/or discharge evasé.

8

8

For direct drive. Generally the same as Arr. 1 with base for the prime mover.

M

M

Alternative: Single stage or two stage fans can be supplied with inlet box and/or discharge evasé.

9

9

For belt drive. Generally same as Arr. 1 with motor mounted on fan casing, and/or an integral base. Alternative: With inlet box and/or discharge evasé

M

AMCA International, Inc. | 30 W. University Dr. | Arlington Heights, IL, 60004-1893 | U.S.A

Figure 3.7 - AMCA Standard 99-3404 / Page 1 58 | Fans and Systems

calculated using the proper specific heat ratio for the gases being handled.

has adopted the term Fan Reynolds Number. Re = (πND2ρ) / (60μ) where: N = impeller rotational speed, rpm D = impeller diameter, m(ft) ρ = air density, kg/m3 (lbm/ft3) μ = absolute viscosity, 1.8185 × 10-5 Pa•s (5°C to 38°C) (1.22 × 10-5 lbm/ft•s (40°F to 100°F))

(SI) (I-P)

The threshold fan Reynolds number for centrifugal and axial fans is about 3.0 × 106. That is, there is a negligible change in performance between the two fans due to differences in Reynolds number if both fans are operating above this threshold value. When the Reynolds number of a model fan is below 3.0 × 106, there may be a gain in efficiency (size effect) for a full size fan operating above the threshold compared to one operating below the threshold. This occurs only when both fans are operating near peak efficiency. Therefore, when a model test is being conducted to verify the rating of a full size fan, the Reynolds number should be above 3.0 ×106 to avoid any uncertainty relating to Reynolds number effects. b. Point of Rating. To predict the performance of a fan from a smaller model using the Fan Laws, both fans must be geometrically similar (homologous), and both fans must operate at the same corresponding rating points on their characteristic curves. Two or more fans are said to be operating at corresponding “points of rating” if the positions of the operating points, relative to the pressure at shutoff and the airflow at free delivery, are the same. c. Compressibility. Compressibility is the characteristic of a gas to change its volume as a function of pressure, temperature and composition. The compressibility coefficient (Kp) expresses the ratio of the fan total pressure developed with an incompressible fluid to the fan total pressure developed with a compressible fluid (See ANSI/AMCA 210). Differences in the compressibility coefficient between two similar fans must be

d. Specific Heat Ratio (Cp). Model fan tests are usually based on air with a specific heat ratio of 1.4. Induced draft fans may handle flue gas with a specific heat ratio of 1.35. Even though these differences may normally be considered small, they make a noticeable difference in the calculation of the compressibility coefficient. Refer to AMCA Publication 802, Annex A, for calculation procedures. e. Tip Speed Mach Parameter (Mt). Tip speed Mach parameter is an expression relating the tip speed of the impeller to the speed of sound at the fan inlet condition. When airflow velocity at a point approaches the speed of sound, some blocking or choking effects occur that reduce the fan performance.

4.3 Fan performance curves A fan performance curve is a graphic presentation of the performance of a fan. Usually it covers the entire range from free delivery (no obstruction to airflow) to no delivery (an air tight system with no air flowing). One, or more, of the following characteristics may be plotted against volume airflow (Q). Fan Static Pressure Fan Total Pressure Fan Power Fan Static Efficiency Fan Total Efficiency

Ps Pt H ηs ηt

Air density (ρ), fan size (D), and fan rotational speed (N) are usually constant for the entire curve and must be stated. A typical fan performance curve is shown in Figure 4.1. Figure 4.2 illustrates examples of performance curves for a variety of fan types.

Fans and Systems | 59

SIZE 30 FAN AT N RPM

Pt

100

Ps

ηt

ηs

70 60 50 40

H 30 20 OPERATION AT STANDARD DENSITY

10 0

AIRFLOW, Q

Figure 4.1 - Fan Performance Curve at N RPM

60 | Fans and Systems

EFFICIENCY, η PERCENT

80

POWER, H

PRESSURE, P

90

TYPE

BACKWARDINCLINED BACKWARDCURVED

HOUSING DESIGN

• Highest efficiency of all centrifugal fan designs. • Ten to 16 blades of airfoil contour curved away from direction of rotation. Deep blades allow for efficient expansion within blade passages • Air leaves impeller at velocity less than tip speed. • For given duty, has highest speed of centrifugal fan designs

• Scroll-type design for efficient conversion of velocity pressure to static pressure. • Maximum efficiency requires close clearance and alignment between wheel and inlet

• Efficiency only slightly less than airfoil fan. • Ten to 16 single-thickness blades curved or inclined away from direction of rotation • Efficient for same reasons as airfoil fan.

• Uses same housing configuration as airfoil design.

• Higher pressure characteristics than airfoil, backward-curved, and backward-inclined fans. • Curve may have a break to left of peak pressure R and fan should not be operated in this area. • Power rises continually to free delivery.

RADIAL

CENTRIFUGAL FANS

AIRFOIL

IMPELLER DESIGN

R

• Scroll. Usually narrowest of all centrifugal designs. • Because wheel design is less efficient, housing dimensions are not as critical as for airfoil and backward-inclined fans.

M

CENTRIFUGAL AXIAL

CENTRIFUGAL

A

POWER ROOF VENTILATORS

SPECIAL DESIGNS

TUBULAR

VANEAXIAL

TUBEAXIAL

AXIAL FANS

PROPELLER

FORWARDCURVED

M

B

• Flatter pressure curve and lower efficiency than the airfoil, backward-curved, and backward-inclined. • Do not rate fan in the pressure curve dip to the left of peak pressure. • Power rises continually toward free delivery. Motor selection must take this into account.

• Scroll similar to and often identical to other centrifugal fan designs. • Fit between wheel and inlet not as critical as for airfoil and backward-inclined fans.

• Low efficiency. • Limited to low-pressure applications. • Usually low cost impellers have two or more blades of single thickness attached to relatively small hub. • Primary energy transfer by velocity pressure.

• Simple circular ring, orifice plate, or venturi. • Optimum design is close to blade tips and forms smooth airfoil into wheel.

• Somewhat more efficient and capable of developing more useful static pressure than propeller fan. • Usually has 4 to 8 blades with airfoil or singlethickness cross section. • Hub usually less than transfer by velocity pressure.

• Cylindrical tube with close clearance to blade tips.

• Good blade design gives medium- to high-pressure capability at good efficiency. • Most efficient of these fans have airfoil blades. • Blades may have fixed, adjustable, or controllable pitch. • Hub is usually greater than half fan tip diameter.

• Cylindrical tube with close clearance to blade tips. • Guide vanes upstream or downstream from impeller increase pressure capability and efficiency.

• Performance similar to backward-curved fan except capacity and pressure are lower. • Lower efficiency than backward-curved fan. • Performance curve may have a dip to the left of peak pressure.

• Cylindrical tube similar to vaneaxial fan, except clearance to wheel is not as close. • Air discharges radially from wheel and turns 90° to flow through guide vanes.

• Low-pressure exhaust systems such as general factory, kitchen, warehouse, and some commercial installations. • Provides positive exhaust ventilation, which is an advantage over gravity-type exhaust units. • Centrifugal units are slightly quieter than axial units.

• Normal housing not used, since air discharges from impeller in full circle. • Usually does not include configuration to recover velocity pressure component.

• Low-pressure exhaust systems such as general factory, kitchen, warehouse, and some commercial installations. • Provides positive exhaust ventilation, which is an advantage over gravity-type exhaust units.

• Essentially a propeller fan mounted in a supporting structure • Hood protects fan from weather and acts as safety guard. • Air discharges from annular space at bottom of weather hood.

Figure 4.2 - Types of Fans Adapted with permission from 1996 ASHRAE Systems and Equipment Handbook (SI) Fans and Systems | 61

PERFORMANCE CURVES

a

PERFORMANCE CHARACTERISTICS

Pt

8

Ps

10

6 ηt 4

8 6

ηs wo

2

4 2

VOLUME FLOW RATE, Q

0 0

2

EFFICIENCY

PRESSURE-POWER

10

4

6

8

6

8 6

4

4 2

VOLUME FLOW RATE, Q 0

2

4

6

8

10

6

8 6

4

4 2

2

VOLUME FLOW RATE, Q 0

2

4

6

8

EFFICIENCY

PRESSURE-POWER

8

0

10

6

8 6

4

4 2

2

VOLUME FLOW RATE, Q 0

2

4

6

8

EFFICIENCY

PRESSURE-POWER

8

0

0 10

8

10

6

8 6

4

4 2

2

VOLUME FLOW RATE, Q

0 0

2

4

6

8

EFFICIENCY

PRESSURE-POWER

10

8

10

6

8 6

4

4 2

2

VOLUME FLOW RATE, Q

0 0

2

4

6

8

EFFICIENCY

PRESSURE-POWER

• Primarily for materials handling in industrial plants. Also for some high-pressure industrial requirements. • Rugged wheel is simple to repair in the field. Wheel sometimes coated with special material. • Not common for HVAC applications.

• Pressure curve less steep than that of backward-curved fans. Curve dips to left of peak pressure. • Highest efficiency to right of peak pressure at 40 to 50% of wide open volume. • Rate fan to right of peak pressure. • Account for power curve, which rises continually toward free delivery, when selecting motor.

• Primarily for low-pressure HVAC applications, such as residential furnaces, central station units, and packaged air conditioners.

• High flow rate, but very low-pressure capabilities. • Maximum efficiency reached near free delivery. • Discharge pattern circular and airstream swirls.

• For low-pressure, high-volume air moving applications, such as air circulation in a space or ventilation through a wall without ductwork. • Used for makeup air applications.

• High flow rate, medium-pressure capabilities. • Performance curve dips to left of peak pressure. Avoid operating fan in this region. • Discharge pattern circular and airstream rotates or swirls.

• Low- and medium-pressure ducted HVAC applications where air distribution downstream is not critical. • Used in some industrial applications, such as drying ovens, paint spray booths, and fume exhausts.

• High-pressure characteristics with medium-volume flow capabilities. • Performance curve dips to left of peak pressure due to aerodynamic stall. Avoid operating fan in this region. • Guide vanes correct circular motion imprated by wheel and improve pressure characteristics and efficiency of fan.

• General HVAC systems in low-, medium-, and high-pressure applications where straight-through flow and compact installation are required. • Has good downstream air distribution • Used in industrial applications in place of tubeaxial fans. • More compact than centrifugal fans for same duty.

• Performance similar to backward-curved fan, except capacity and pressure is lower. • Lower efficiency than backward-curved fan because air turns 90°. • Performance curve of some designs is similar to axial flow fan and dips to left of peak pressure.

• Primarily for low-pressure, return air systems in HVAC applications. • Has straight-through flow.

• Usually operated without ductwork; therefore, operates at very low pressure and high volume. • Only static pressure and static efficiency are shown for this fan.

• Low-pressure exhaust systems, such as general factory, kitchen, warehouse, and some commercial installations. • Low first cost and low operating cost give an advantage over gravity flow exhaust systems. • Centrifugal units are somewhat quieter than axial flow units.

• Usually operated without ductwork; therefore, operates at very low pressure and high volume. • Only static pressure and static efficiency are shown for this fan.

• Low-pressure exhaust systems, such as general factory, kitchen, warehouse, and some commercial installations. • Low first cost and low operating cost give an advantage over gravity flow exhaust systems.

0 10

10

0 10

10 8

10

6

8 6

4

4 2

2

VOLUME FLOW RATE, Q

0 0

2

4

6

8

EFFICIENCY

PRESSURE-POWER

• Higher pressure characteristics than airfoil and backwardcurved fans. • Pressure may drop suddenly at left of peak pressure, but this usually causes no problems. • Power rises continually to free delivery.

0 10

10

0 10

10 8

10

6

8 6

4

4 2

2

VOLUME FLOW RATE, Q

0 0

2

4

6

8

EFFICIENCY

PRESSURE-POWER

• Same heating, ventilating, and air-conditioning applications as airfoil fan. • Used in some industrial applications where airfoil blade may corrode or erode due to environment.

0 10

10

0 10

10 8

10

6

8 6

4

4 2

2 VOLUME FLOW RATE, Q

0 0

2

4

6

8

EFFICIENCY

PRESSURE-POWER

• Similar to airfoil fan, except peak efficiency slightly lower. 10 EFFICIENCY

PRESSURE-POWER

8

0

0 10

10 8

10

6

8 6

4

4 2

2

0 0

VOLUME FLOW RATE, Q 2 4 6

8

EFFICIENCY

PRESSURE-POWER

• General heating, ventilating, and air-conditioning applications. • Usually only applied to large systems, which may be low-, medium-, or high-pressure applications. • Applied to large, clean-air industrial operations for significant energy savings.

0 10

10

2

APPLICATIONS

• Highest efficiencies occur at 50 to 60% of wide open volume. This volume also has good pressure characteristics. • Power reaches maximum near peak efficiency and becomes lower, or self-limiting, toward free delivery.

0 10

a: These performance curves reflect general characteristics of various fans as commonly applied. They are not intended to provide complete selection criteria, since other parameters, such as diameter and speed, are not defined.

Figure 4.2 - Types of Fans Adapted with permission from 1996 ASHRAE Systems and Equipment Handbook (SI) 62 | Fans and Systems

5.1 Type A: Free inlet, free outlet fans

1) Type B: Free inlet, ducted outlet 2) Type C: Ducted inlet, free outlet 3) Type D: Ducted inlet, ducted outlet

Fans designed for use other than with duct systems are usually rated over a lower range of pressures. They are commonly cataloged and sold as a complete unit with suitable drive and motor.

The performance of fans intended for use with duct systems is usually published in the form of a "multirating" table. A typical multi-rating table, as illustrated in Figure 5.2 shows:

Typical fans in this group are propeller fans and power roof ventilators. They are usually available in direct or belt-drive arrangements and performance ratings are published in a modified form of the multirating table. Figure 5.1 illustrates such a table for part of a line of belt-drive propeller fans.

a) the speed (N) in rpm b) the power (H) in kw (hp) c) the fan static pressure (Ps) in Pa (in. wg) d) the outlet velocity (V) in m/s, (fpm) e) the airflow (Q) in m3/s (cfm)

5. Catalog Performance Tables

5.2 Ducted fans There are three types of ducted fans, as described in Section 3: SIZE No. of Motor Peak rpm (cm) Blades kW kW 0.19 862 0.13 0.19 960 0.20 61 3 0.25 1071 0.27 0.37 1220 0.40 0.19 806 0.20 0.25 883 0.27 69 3 0.37 1035 0.43 0.56 1165 0.62 0.37 825 0.42 0.56 945 0.62 0.75 1045 0.82 84 3 1.12 1190 1.19 1.49 1306 1.64 TYPICAL RATING TABLE

Figure 5.3 shows constant speed characteristic curves superimposed on a section of the multi-rating table for the same fan. A brief study of this figure will assist in understanding the relationship between curves and the multi-rating tables.

AIRFLOW (m3/s) @ STATIC PRESSURE (Pa) 0 31 62 93 124 155 186 217 2.02 1.58 0.58 2.25 1.87 0.97 2.51 2.18 1.76 0.76 2.86 2.57 2.24 1.70 0.81 2.89 2.36 1.05 3.17 2.68 1.94 0.76 3.71 3.30 2.85 1.56 0.95 4.18 3.83 3.44 3.01 1.60 1.10 4.36 3.76 3.04 1.27 4.99 4.48 3.92 2.38 1.42 5.23 5.08 4.57 4.01 2.31 1.52 6.29 5.90 5.47 5.01 4.48 2.79 1.94 6.91 6.53 6.15 5.75 5.32 4.81 3.05 2.24 FOR A SERIES OF BELT-DRIVEN PROPELLER FANS

248

1.84

SIZE No. of Motor Peak AIRFLOW (ft3/min) @ STATIC PRESSURE (in. wg) rpm (in.) Blades hp bhp 0 1/8 1/4 3/8 1/2 5/8 3/4 7/8 1 1/4 862 0.18 4,283 3,350 1,230 1/4 960 0.27 4,770 3,960 2,050 24 3 1/3 1071 0.36 5,321 4,620 3,730 1,600 1/2 1220 0.54 6,062 5,450 4,750 3,600 1,710 1/4 806 0.27 6,123 4,990 2,230 1/3 883 0.36 6,708 5,675 4,100 1,620 27 3 1/2 1035 0.57 7,862 7,000 6,035 3,315 2,020 3/4 1165 0.83 8,850 8,110 7,290 6,385 3,400 2,330 1/2 825 0.56 9,240 7,970 6,430 2,700 3/4 945 0.83 10,580 9,500 8,300 5,040 3,010 1 1045 1.1 11,710 10,755 9,685 8,490 4,890 3,215 33 3 1½ 1190 1.6 13,335 12,490 11,580 10,610 9,500 5,905 4,100 2 1306 2.2 14,630 13,845 13,030 12,185 11,280 10,200 6,470 4,740 3,900 TYPICAL RATING TABLE FOR A SERIES OF BELT-DRIVEN PROPELLER FANS Figure 5.1 - Propeller Fan Performance Table Fans and Systems | 63

IMPELLER DIAMETER: TIP SPEED IN m/s: Volume m3/s 1.81 2.17 2.53 2.89 3.25 3.61 3.97 4.33 4.69 5.06 5.42 5.78 6.14 6.50 6.86 7.22 7.94 8.67 9.39 10.11 10.83 11.55 12.28 13.00 13.72 14.44

Outlet Vel. (m/s) 2.55 3.06 3.56 4.07 4.58 5.08 5.59 6.10 6.61 7.13 7.63 8.14 8.65 9.15 9.66 10.17 11.18 12.21 13.23 14.24 15.25 16.27 17.30 18.31 19.32 20.34

62 Pa rpm

kW

222 236 253 272 292 314 338 361 385 409 434 458 483 508

0.14 0.17 0.22 0.27 0.34 0.42 0.51 0.62 0.74 0.88 1.03 1.21 1.41 1.63

927 mm .0485 × RPM 93 Pa rpm

kW

270 284 300 317 337 358 379 402 426 449 473 498 522 547 571 621

0.25 0.30 0.36 0.43 0.52 0.62 0.74 0.87 1.01 1.18 1.37 1.58 1.81 2.06 2.34 2.99

OUTLET AREA: MAXIMUM kW:

124 Pa rpm

kW

313 327 343 360 378 398 419 441 464 488 511 535 559 585 633 682

0.39 0.45 0.53 0.63 0.73 0.86 1.00 1.16 1.33 1.53 1.75 1.99 2.25 2.54 3.20 3.98

155 Pa rpm

kW

352 366 382 399 417 437 457 479 501 525 538 571 595 644 693 742 791

0.55 0.64 0.74 0.85 0.98 1.13 1.30 1.49 1.69 1.92 2.16 2.44 2.74 3.41 4.20 5.13 6.20

.71 SQ METERS 13.65 × (RPM/1000)3

186 Pa

217 Pa

rpm

kW

rpm

kW

389 403 419 436 454 473 494 515 537 560 584 607 654 703 752 801 850

0.75 0.86 0.98 1.11 1.26 1.44 1.63 1.86 2.09 2.35 2.62 2.93 3.63 4.44 5.38 6.47 7.70

411 424 438 455 472 489 509 529 550 572 595 616 665 712 761 810 859 908

0.87 0.98 1.10 1.25 1.41 1.58 1.79 2.01 2.26 2.54 2.82 3.14 3.85 4.68 5.64 6.73 7.99 9.40

246 Pa rpm

kW

443 458 472 489 506 524 543 564 585 606 629 675 721 769 818 867 916 965 1015

1.10 1.19 1.39 1.56 1.74 1.95 2.18 2.43 2.71 3.01 3.34 4.07 4.93 5.90 7.01 8.27 9.70 11.30 13.06

310 Pa

373 Pa

rpm

kW

rpm

kW

494 507 522 538 555 572 590 610 630 651 695 741 788 834 883 932 981 1030 1072 1129

1.52 1.68 1.86 2.06 2.28 2.53 2.78 3.07 3.39 3.74 4.52 5.40 6.41 7.57 8.87 10.32 11.95 13.77 15.78 17.98

540 554 568 584 600 617 635 654 674 715 759 805 852 898 946 995 1044 1093 1142

1.99 2.18 2.39 2.62 2.89 3.16 3.45 3.78 4.15 4.96 5.89 6.94 8.11 9.47 10.96 12.62 14.46 16.50 18.76

TYPICAL MULTISPEED RATING TABLE FOR A SINGLE WIDTH, SINGLE INLET CENTRIFUGAL FAN

IMPELLER DIAMETER: TIP SPEED IN FPM: Volume CFM 3825 4590 5355 6120 6885 7650 8415 9180 9945 10710 11475 12240 13005 13770 14535 15300 16830 18360 19890 21420 22950 24480 26010 27540 29070 30600

Outlet Vel. (fpm) 500 600 700 800 900 1000 1100 1200 1300 1400 1500 1600 1700 1800 1900 2000 2200 2400 2600 2800 3000 3200 3400 3600 3800 4000

36.5 IN 9.56 × RPM

OUTLET AREA: MAXIMUM BHP:

7.65 SQ FT 18.3 × (RPM/1000)3

1/4 in. wg

3/8 in. wg

1/2 in. wg

5/8 in. wg

3/4 in. wg

7/8 in. wg

rpm

bhp

rpm

bhp

rpm

bhp

rpm

bhp

rpm

rpm

222 236 253 272 292 314 338 361 385 409 434 458 483 508

0.185 0.233 0.292 0.365 0.450 0.560 0.682 0.826 0.989 1.175 1.387 1.626 1.895 2.191

270 284 300 317 337 358 379 402 425 449 473 498 522 547 571 621

0.334 0.400 0.483 0.579 0.695 0.832 0.988 1.163 1.360 1.587 1.837 2.115 2.424 2.767 3.144 4.003

313 327 343 360 378 398 419 441 464 488 511 535 559 585 633 682

0.519 0.608 0.716 0.840 0.981 1.149 1.340 1.553 1.780 2.048 2.346 2.665 3.017 3.403 4.289 5.335

352 366 383 399 417 437 457 479 501 525 538 571 595 644 693 742 791

0.743 0.856 0.992 1.144 1.314 1.514 1.741 1.993 2.269 2.570 2.901 3.275 3.672 4.577 5.632 6.885 8.308

bhp

389 1.01 403 1.15 419 1.31 436 1.49 454 1.69 473 1.93 494 2.19 515 2.49 537 2.80 560 3.15 584 3.52 607 3.93 654 4.87 703 5.96 752 7.22 801 8.67 850 10.32

bhp

1 in. wg rpm

411 1.17 424 1.31 443 438 1.48 458 455 1.68 472 472 1.89 489 489 2.12 506 509 2.40 524 529 2.70 543 550 3.03 564 572 3.40 585 595 3.78 606 618 4.21 629 665 5.16 675 712 6.28 721 761 7.56 769 810 9.03 818 859 10.71 867 908 12.50 916 965 1015

bhp

1.48 1.60 1.86 2.09 2.34 2.61 2.92 3.26 3.64 4.04 4.48 5.46 6.61 7.91 9.40 11.09 13.01 15.16 17.52

1-1/4 in. wg 1-1/2 in. wg rpm

bhp

rpm

bhp

494 507 522 538 555 572 590 610 630 651 695 741 788 834 883 932 981 1030 1072 1129

2.04 2.25 2.49 2.76 3.06 3.39 3.73 4.12 4.55 5.02 6.06 7.24 8.60 10.15 11.89 13.84 16.03 18.47 21.16 24.11

540 554 568 584 600 617 635 654 674 715 759 805 852 898 946 995 1044 1093 1142

2.67 2.92 3.20 3.52 3.87 4.24 4.63 5.07 5.56 6.65 7.90 9.30 10.88 12.70 14.70 16.92 19.39 22.13 25.16

TYPICAL MULTISPEED RATING TABLE FOR A SINGLE WIDTH, SINGLE INLET CENTRIFUGAL FAN Figure 5.2 - Centrifugal Fan Performance Tables 64 | Fans and Systems

OUTLET VELOCITY 500 600 700 800 900 1000 1100 1200 1300 1400 1500 1600 1700 1800 1900 2000 2200 2400 2600 2800 3000 3200 3400 3600 3800 4000

VOLUME CFM

3825 4590 5355 6120 6885

7650 8415 9180 9945 10710

11475 12240 13005 13770 14535

15300 16830 18360 19890 21420

22950 24480 26010 27540 29070 30600

.185 .233 .292 .365 .450

BHP

434 1.387 456 1.626 482 508 2.19

.334 .400 .483 .579

BHP

571 3.744 629 4.003

449 1.587 473 1.837 493 2.115 522 2.424 547 2.767

BHP

584 3.403 633 4.289 682 5.335

464 1.78 488 2.048 511 2.346 535 2.665 559 3.017

360 .840 378 .981 398 1.149 419 1.340 441 1.553

313 .519 327 .608 343 .716

RPM

1/2” SP

BHP

596 644 4.577 693 5.632 742 6.885 791 8.308

479 1.995 501 2.269 525 2.570 538 2.901 571 3.276

332 .992 399 1.144 417 1.314 437 1.514 457 1.741

352 .743 366 .856

RPM

5/8” SP

RPM

3.93 4.87 5.76 7.22 8.67

2.19 2.49 2.80 3.15 3.52

1.15 1.31 1.49 1.69 1.93

850 10.32

607 654 703 752 801

494 515 537 560 584

403 419 436 454 473

389 1.01

BHP

3/4” SP

390 RPM

337 .695 358 .822 379 .988 482 1.163 426 1.360

270 284 300 317

RPM

3/8” SP

BHP

4.21 5.16 6.28 7.56 9.03

2.40 2.70 3.03 3.40

1.31 1.48 1.58 1.89 2.12

859 10.71 908 12.60

618 665 712 761 810

509 529 550 572 595

424 438 455 472 489

411 1.17

RPM

7/8” SP

BHP

4.48 5.46 6.81 7.91 8.48

2.61 2.92 3.26 3.84 4.04

1.48 1.60 1.86 2.09 2.34

3.06 3.49 3.73 4.12 4.55

2.04 2.25 2.49 2.76

BHP

651 5.02 695 6.06 741 7.24 788 8.60 834 10.15

555 572 590 610 630

494 507 522 538

RPM

1-1/4” SP

BHP

3.52 3.87 4.24 4.63 5.07 674 5.56 715 6.65 759 7.90 9.30 852 10.88

584 600 617 635 654

540 2.67 554 2.92 568 3.28

RPM

1-1/2” SP

BHP

3.99 4.36 4.76 5.18 5.63 696 6.11 736 7.24 778 822 10.02 867 11.65

612 627 643 661 678

584 3.37 598 3.66

RPM

1-3/4” SP

867 11.09 883 11.89 898 12.70 914 13.48 916 13.01 932 13.84 946 14.70 960 15.56 965 15.16 981 16.03 995 16.92 1009 17.83 1015 17.52 1030 18.47 1044 19.39 1057 20.35 1079 21.16 1093 22.13 1106 23.12 1129 24.11 1142 25.16 1155 26.18

629 675 721 769 818

524 543 564 585 606

443 458 472 489 506

RPM

1” SP

490 RPM

314 .560 338 .682 361 .826 335 .988 409 1.175

222 236 253 272 292

RPM

1/4” SP

PRESSURE IN IN. WG BRAKE HORSEPOWER

585 RPM 810 RPM

RECOMMENDED SELECTION RANGE

CFM

Figure 5.3 - Typical Fan Performance Table Showing Relationship to a Family of Constant Speed Performance Curves

Fans and Systems | 65

Most performance tables do not cover the complete range from no delivery to free delivery but cover only the typical operating range. Figure 5.4 illustrates the recommended performance range of a centrifugal fan. Comparison of Figure 5.4 with Figure 5.3 will show that the published performance table also covers only the recommended performance range of the fan.

6. Air Systems 6.1 The system An air system may consist simply of a fan with ducting connected to either the inlet or outlet or to both. A more complicated system may include a fan, ductwork, air control dampers, cooling coils, heating coils, filters, diffusers, sound attenuation, turning vanes, etc. See AMCA Publication 200 Air Systems, for more information.

It should be remembered that fans are generally tested without obstructions in the inlet and outlet and without any optional airstream accessories in place. Catalog ratings will, therefore, usually apply only to the bare fan with unobstructed inlet and outlet.

6.2 Component losses

Fan performance adjustment factors for airstream accessories are normally available from either the fan catalog or the fan manufacturer.

Every system has a combined resistance to airflow that is usually different from every other system and is dependent upon the individual components in the system.

Fans are usually tested in arrangement 1, or similar (see Figure 3.5). Rating tables will, therefore, also apply only to the tested arrangement. Allowances for the effect of bearing supports used in other arrangements should be obtained from the manufacturer if not shown in the catalog.

The determination of the "pressure loss" or "resistance to airflow," for the individual components can be obtained from the component manufacturers. The determination of pressure losses for ductwork design is well documented in standard handbooks such as the ASHRAE Handbook of Fundamentals.

SELECTION NOT USUALLY RECOMMENDED IN THIS RANGE

RECOMMENDED SELECTION RANGE

SY

ST

EM

CU RV E

PRESSURE

RE

SU

ES

PR

DU CT

E RV

EM

SELECTION NOT USUALLY RECOMMENDED IN THIS RANGE

CU

ST

T UC

SY

D

AIRFLOW Figure 5.4 - Recommended Performance Range of a Typical Centrifugal Fan 66 | Fans and Systems

In a later section, the effects of some system components and fan accessories on fan performance are discussed. The System Effects presented will assist the system designer to determine fan selection.

The system curve of a "fixed system" plots as a parabola in accordance with the above relationship. Typical plots of the resistance to flow versus volume airflow for three different and arbitrary fixed systems, (A, B, and C) are illustrated in Figure 6.1. For a fixed system an increase or decrease in airflow results in an increase or decrease in the system resistance along the given system curve only. Also, as the components in a system change, the system curve changes.

6.3 The system curve At a fixed airflow through a given air system a corresponding pressure loss, or resistance to this airflow, will exist. If the airflow is changed, the resulting pressure loss, or resistance to airflow, will also change. The relationship between airflow pressure and loss can vary as a function of type of duct components, their interaction and the local velocity magnitude. In many cases, typical duct systems operate in the turbulent flow regime and the pressure loss can be approximated as a function of velocity (or airflow) squared. The simplifying relationship used in this publication governing the change in pressure loss as a function of airflow for a fixed system is:

Refer to Figure 6.1, Duct System A. With a system at the design airflow (Q) and at a design system resistance (P), an increase in airflow to 120% of Q will result in an increase in system resistance P of 144% since system resistance varies with the square of the airflow. Likewise, a decrease in airflow Q to 50% would result in a decrease in system resistance P to 25% of the design system resistance. In Figure 6.1, System Curve B is representative of a system that has more component pressure loss than System Curve A, and System Curve C has less component pressure loss than System Curve A.

Pc/P = (Qc/Q)2

Notice that on a percentage basis, the same relationships also hold for System Curves B and C. These relationships are characteristic of typical fixed systems.

A more through discussion of duct system pressure losses can be found in AMCA Publication 200 Air Systems.

200

160 140 SY

E ST

M

C

120 100

EM

B

80 ST

S SY

60

SY

PERCENT OF SYSTEM RESISTANCE

180

TE

M

A

SYSTEM DESIGN POINT

40 20 0 0

20

40

60

80

100

120

140

160

180

200

PERCENT OF SYSTEM AIRFLOW Figure 6.1 - System Curves Fans and Systems | 67

is now at Point 3 (the intersection of the fan curve and the new System C), with the airflow at approximately 120% of Q.

6.4 Interaction of system curve and fan performance curve If the system characteristic curve, composed of the resistance to system airflow and the appropriate SEF have been accurately determined, then the fan will deliver the designated airflow when installed in the system.

6.5 Effect of changes in speed Increases or decreases in fan rotational speed will alter the airflow through a system. According to the Fan Laws (see below), the % increase in airflow is directly proportional to the fan rotational speed ratio, and the fan static pressure is proportional to the square of the fan rotational speed ratio. Thus, a 10% increase in fan rotational speed will result in a new fan curve with a 10% increase in Q, as illustrated in Figure 6.3. Since the system components did not change, System Curve A remains the same. With airflow increasing by 10% over the original Q, the system resistance increases along System Curve A to Point 2, at the intersection with the new fan curve.

The point of intersection of the system curve and the fan performance curve determines the actual airflow. System Curve A in Figure 6.2 has been plotted with a fan performance curve that intersects the system design point. The airflow through the system in a given installation may be varied by changing the system resistance. This is usually accomplished by using fan dampers, duct dampers, mixing boxes, terminal units, etc.

The greater airflow moved by the fan against the resulting higher system resistance to airflow is a measure of the increased work done. In the same system, the fan efficiency remains the same at all points on the same system curve. This is due to the fact that airflow, system resistance, and required power are varied by the appropriate ratio of the fan rotational speed.

A

200

ST

EM

180 160

SY

PERCENT OF SYSTEM RESISTANCE

Figure 6.2 shows the airflow may be reduced from design Q by increasing the resistance to airflow, i.e., changing the system curve from System A to System B. The new operating point is now at Point 2 (the intersection of the fan curve and the new System B) with the airflow at approximately 80% of Q. Similarly, the airflow can be increased by decreasing the resistance to airflow, i.e., changing the system curve from System A to System C. The new operating point

140

SY

E ST

M

C

FAN CURVE

2

120

1

100 80

SYSTEM DESIGN POINT

3

60

EM

B

T YS

40

S

20 0 0

20

40

60

80

100

120

140

160

PERCENT OF SYSTEM AIRFLOW

Figure 6.2 - Interaction of System Curves and Fan Curve 68 | Fans and Systems

180

200

air density of 1.2 kg/m3 (0.075 lbm/ft3) is standard in the fan industry throughout the world. Figure 6.4 illustrates the effect on the fan performance of a density variation from the standard value.

6.5.1 Fan Laws - effect of change in speed - (fan size and air density remaining constant) For the same size fan, Dc = D and, therefore, (Dc/D) = 1. When the air density does not vary, ρc = ρ and the air density ratio (ρc/ρ) = 1. Kp is taken as equal to unity in this and following examples.

6.6.1 Fan Laws - effect of change in density - (fan size and speed remaining constant) When the speed of the fan does not change, Nc = N and, therefore, (Nc/N) = 1. The fan size is also fixed, Dc = D and therefore (Dc/D) = 1.

Qc = Q × (Nc/N) Ptc = Pt × (Nc/N)2 Psc = Ps × (Nc/N)2

Qc = Q

Pvc = Pv × (Nc/N)2

Ptc = Pt × (ρc/ρ) Psc = Ps × (ρc/ρ)

Hc = H × (Nc/N)3

Pvc = Pv × (ρc/ρ)

6.6 Effect of density on system resistance

Hc = H × (ρc/ρ)

SY CT

S (AT 1.1N)

160

PRESSURE

H (AT 1.1N) 133

140

S (AT N) PRESSURE

120

2 H (AT N)

100

1 100

80 60 50

40 20

PERCENT OF POWER

DU

PERCENT OF SYSTEM RESISTANCE

ST

EM

A

The resistance of a duct system is dependent upon the density of the air flowing through the system. An

110%

0 0

20

40

60

80

100

120

140

160

180

200

PERCENT OF SYSTEM AIRFLOW

Figure 6.3 - Effect of 10% increase in Fan Speed Fans and Systems | 69

PERCENT OF SYSTEM RESISTANCE AND FAN PRESSURE

FAN PRESSURE CURVE @ DENSITY ρ

SYSTEM A @ DENSITY ρ FAN INLET

100

SYSTEM A @ DENSITY ρ/2 FAN INLET

FAN PRESSURE CURVE @ DENSITY ρ/2

80 60 40 20 0

PERCENT OF POWER

100 POWER @ DENSITY ρ

80 60 40 POWER @ DENSITY ρ/2

20 0 0

20

40

60

80

100

120

140

PERCENT OF SYSTEM AIRFLOW

Figure 6.4 - Density Effect

70 | Fans and Systems

160

180

200

6.7 Fan and system interaction When system pressure losses have been accurately estimated and desirable fan inlet and outlet conditions have been provided, design airflow can be expected, as illustrated in Figure 6.5. Note again that the intersection of the actual system curve and the fan curve determine the actual airflow. However, when system pressure losses have not been accurately estimated as in Figure 6.6, or when undesirable fan inlet and outlet conditions exist as in Figure 6.7, design performance may not be obtained.

6.8 Effects of errors in estimating system resistance 6.8.1 Higher system resistance. In Figure 6.6, System Curve B shows a situation where a system has greater resistance to airflow than designed (Curve A). This condition is generally a result of inaccurate allowances of system resistance. All pressure losses must be considered when calculating system resistance or the actual system will be more restrictive to airflow than intended. This

condition results in an actual airflow at Point 2, which is at a higher pressure and lower airflow than was expected. If the actual duct system pressure loss is greater than design, an increase in fan speed may be necessary to achieve Point 5, the design airflow. CAUTION: Before increasing fan rotational speed, check with the fan manufacturer to determine whether the fan rotational speed can be safely increased. Also determine the expected increase in power. Since the power required increases as the cube of the fan rotational speed ratio, it is very easy to exceed the capacity of the existing motor and that of the available electrical service. 6.8.2 Lower system resistance. Curve C in Figure 6.6 shows a system that has less resistance to airflow than designed. This condition results in an actual airflow at Point 3, which is at a lower pressure and higher airflow than was expected.

CALCULATED SYSTEM CURVE PEAK FAN PRESSURE

1

DESIGN RESISTANCE

FAN PRESSURE CURVE

DESIGN AIRFLOW

Figure 6.5 - Fan/System Curve at Design Point Fans and Systems | 71

the fan speed, adjusting the variable inlet vane (VIV), if installed, or inlet dampers. The system resistance could also be increased to Point 1 on Curve A, Figure 6.6. The change in fan operating point should be evaluated carefully, since a change in fan power consumption may occur.

6.9 Safety factors It has been common practice among system designers to add safety factors to the calculated system resistance to account for the “unexpected”. In some cases, safety factors may compensate for resistance losses that were unaccounted for and the actual system will deliver the design airflow, Point 1, Figure 6.6. If the actual system resistance is lower than the design system resistance, including the safety factors, the fan will run at Point 3 and deliver more airflow. This result may not be advantageous because the fan may be operating at a less efficient point on the fan’s performance curve and may require more power than a properly designed system. Under these conditions, it may be desirable to reduce the fan performance to operate at Point 4 on Curve C, Figure 6.6. This may be accomplished by reducing

The system designer should also evaluate the fan performance tolerance and system resistance tolerance to determine if the lower or upper limits of the probable airflow in the system are acceptable. The combination of these tolerances should be evaluated to ensure that the “high-side” system resistance curve does not fall into the unstable range of performance. Operation in this area of the curve should be avoided and precautions taken to ensure operations outside of the unstable area, especially at the highest expected system resistance.

CURVE B: ACTUAL SYSTEM

ACTUAL SYSTEM RESISTANCE MORE THAN DESIGN

CURVE A: CALCULATED SYSTEM

5 CURVE C ACTUAL SYSTEM PEAK FAN PRESSURE

2 1

DESIGN RESISTANCE

3

ACTUAL SYSTEM LESS THAN DESIGN

4

FAN PRESSURE CURVE

DESIGN AIRFLOW

72 | Fans and Systems

Figure 6.6 - Fan/System Curve Not at Design Point

6.10 Deficient fan/system performance The most common causes of deficient fan/system performance are improper fan inlet duct design, fan outlet duct design, and fan installation into the duct system. Any one or a combination of these conditions that alter the aerodynamic characteristics of the air flowing through the fan such that the fan’s full airflow potential, as tested in the laboratory and cataloged, is not likely to be realized. Other major causes of deficient performance are: • The air performance characteristics of the installed system are significantly different from the system designer's intent (See Figure 6.6). This may be due to a change in the system by others or unexpected behavior of the system during operation. • The system design calculations did not include adequate allowances for the effect of accessories and appurtenances (See Section 10). • The fan selection was made without allowing for the effect of appurtenances on the fan's performance (See Section 10). • Dirty filters, dirty ducts, dirty coils, etc., will increase the system resistance, and consequently, reduce the airflow - often significantly. • The "performance" of the system has been determined by field measurement techniques that have a high degree of uncertainty. Other "on-site" problems are listed in AMCA Publication 202 Troubleshooting, which includes detailed checklists and recommendations for the correction of problems with the performance of air systems.

6.11 Precautions performance

to

prevent

deficient

• Use appropriate allowances in the design calculations when space or other factors dictate the use of less than optimum arrangement of the fan outlet and inlet connections (See Sections 8 and 9). • Design the connections between the fan and the system to provide, as nearly as possible, uniform airflow conditions at the fan outlet and inlet connections (See Sections 8 and 9).

• Include adequate allowance for the effect of all accessories and appurtenances on the performance of the system and the fan. If possible, obtain from the fan manufacturer data on the effect of installed appurtenances on the fan's performance (See Section 10). • Use field measurement techniques that can be applied effectively on the particular system. Be aware of the probable accuracy of measurement and conditions that affect this. Refer to AMCA Publication 203 Field Performance Measurement of Fan Systems; for more precise measurement see AMCA Standard 803 Industrial Process/Power Generation Fans: Site Performance Test Standard. Also, refer to AABC National Standards, Chapter 8, Volume Measurements, Associated Air Balance Council, 5th Edition, 1989.

6.12 System Effect Figure 6.7 illustrates deficient fan/system performance resulting from one or more of the undesirable airflow conditions listed in Section 6.10. It is assumed that the system pressure losses, shown in system curve A, have been accurately determined, and a suitable fan selected for operation at Point 1. However, no allowance has been made for the effect of the system connections on the fan's performance. To account for this System Effect it will be necessary to add a System Effect Factor (SEF) to the calculated system pressure losses to determine the actual system curve. The SEF for any given configuration is velocity dependent and will vary across a range of airflow. This will be discussed in more detail in Section 7. (See Figure 7.1). In Figure 6.7 the point of intersection between the fan performance curve and the actual system curve B is Point 4. The actual airflow will be deficient by the difference 1-4. To achieve design airflow, a SEF equal to the pressure difference between Point 1 and 2 should have been added to the calculated 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. The System Effect includes only the effect of the system configuration on the fan's performance.

Fans and Systems | 73

7. System Effect Factor (SEF)

7.1 System Effect Curves

A System Effect Factor is a value that accounts for the effect of conditions adversely influencing fan performance when installed in the air system.

Figure 7.1 shows a series of 19 System Effect Curves. 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 SEF for a particular configuration.

CURVE B ACTUAL SYSTEM WITH SYSTEM EFFECT

CURVE A CALCULATED SYSTEM WITH NO ALLOWANCE FOR SYSTEM EFFECT

2 SYSTEM EFFECT LOSS AT DESIGN AIRFLOW

4

DESIGN RESISTANCE

1 3

SYSTEM EFFECT AT ACTUAL AIRFLOW

FAN CATALOG PRESSURE CURVE

AIRFLOW DEFICIENCY

DESIGN AIRFLOW

Figure 6.7 - Deficient Fan/System Performance - System Effect Ignored

74 | Fans and Systems

FG H I J K L

1000

M

N

O

P

900 Q

800 700

R 600 500 S

SYSTEM EFFECT FACTOR PRESSURE, Pa

400

300 T U 200

V

100 90

W

80 70 60 X

50 40

30

20 2.5

3

4

5

6

7

8

9 10

20

30

AIR VELOCITY, (m/s) (Air Density = 1.2 kg/m3)

Figure 7.1 - System Effect Curves (SI)

Fans and Systems | 75

FG H I J K L

M

N

O

5.0 P 4.0 Q 3.0 R

SYSTEM EFFECT FACTOR - PRESSURE, in. wg

2.5 2.0

S 1.5 T 1.0 0.9

U

0.8 0.7 0.6

V

0.5 0.4

W

0.3 0.25 X

0.2

0.15

0.1

5

6

7

8 9 10

15

20

25

30

AIR VELOCITY, ft/min × 100 (Air Density = 0.075 lbm/ft3)

Figure 7.1 - System Effect Curves (I-P)

76 | Fans and Systems

40

50

60

Table 7.1 - System Effect Coefficients

Curve in Figure 7.1

Dynamic Pressure Loss Coefficient C

F G H I J K L M N O P Q R S T U V W X

16.00 14.20 12.70 11.40 9.50 7.90 6.40 4.50 3.20 2.50 1.90 1.50 1.20 0.75 0.50 0.40 0.25 0.17 0.10 2

⎛ V ⎞ SEF = C ⎜ ⎟ ρ ⎝ 1.414 ⎠

SI

2

⎛ V ⎞ SEF = C ⎜ ⎟ ρ ⎝ 1097 ⎠

I-P

Fans and Systems | 77

The SEF is given in Pascals (in. wg) and must be added to the total system pressure losses as shown on Figure 7.2. The velocity used when entering Figure 7.1 will be either the inlet or the outlet velocity of the fan. This will depend on whether the configuration in question is related to the fan inlet or the fan outlet. Most catalog ratings include outlet velocity figures but, for centrifugal fans, it may be necessary to calculate the inlet velocity (See Figure 9.14). The inlet velocity and outlet velocity of an axial fan can be approximated by using the fan impeller diameter to determine the airflow area. The necessary dimensioned drawings are usually included in the fan catalog. In Sections 8 and 9, typical inlet and outlet configurations are illustrated and the appropriate System Effect Curve is listed for each configuration. If more than one configuration is included in a system, the SEF for each must be determined separately and the total of these System Effects must be added to the total pressure losses.

The System Effect Curves are plotted for standard air at a density of 1.2 kg/m3 (0.075 lbm/ft3). Since the System Effect is directly proportional to density, values for other densities can be calculated as below: ⎛d ⎞ SEF2 = SEF1 ⎜ 2 ⎟ ⎝ d1 ⎠ Where: SEF2 = SEF at actual density SEF1 = SEF at standard density d2 = actual density d1 = standard density Alternatively, the SEF may be calculated by the method shown in Table 7.1. Determine the configuration being evaluated and use the appropriate loss coefficient, Cp, and application velocity, V. The SEF can then be calculated using the equations shown in Table 7.1.

FAN POWER

ACTUAL SYSTEM RESISTANCE

ACTUAL POWER REQUIRED

ACTUAL SYSTEM W/ SEF

SEF

CALCULATED SYSTEM W/NO ALLOWANCE FOR SEF

FAN PRESSURE

DESIGN AIRFLOW

Figure 7.2 - Effect of System on Fan Selection 78 | Fans and Systems

should examine catalog ratings carefully for statements defining whether the published ratings are based on tests made with A: free inlet, free outlet; B: free inlet, ducted outlet; C: ducted inlet, free outlet or D; ducted inlet, ducted outlet.

7.2 Power determination When all the applicable System Effect Factors (SEF) have been added to the calculated system pressure losses the power shown in the catalog for the actual point of operation, Figure 7.2 or Table 7.1 may be used without further adjustment.

8.1 Outlet ducts

ANSI/AMCA 210 specifies an outlet duct that is no greater than 105% or less than 95% of the fan outlet area. It also requires that the slope of the transition elements be no greater than 15° for converging elements or greater than 7° for diverging elements.

As previously discussed, fans intended primarily for use with duct systems are usually tested with an outlet duct in place (See Figure 3.2). In most cases it is not practical for the fan manufacturer to supply this duct as part of the fan, but rated performance will not be achieved unless a comparable duct is included in the system design. The system design engineer

Figure 8.1 shows changes in velocity profiles at various distances from centrifugal and axial flow fan outlets. By definition, 100% "effective duct length" is a minimum of two and one half (2½) equivalent duct diameters. For velocities greater than 13 m/s (2500 fpm), add 1 duct diameter for each additional 5 m/s (1000 fpm).

8. Outlet System Effect Factors

BLAST AREA DISCHARGE DUCT CUTOFF

OUTLET AREA

25% 50% 75% CENTRIFUGAL FAN 100% EFFECTIVE DUCT LENGTH AXIAL FAN

To calculate 100% duct length, assume a minimum of 2½ duct diameters for 12.7 m/s (2500 fpm) or less. Add 1 duct diameter for each additional 5.08 m/s (1000 fpm). EXAMPLE: 25.4 m/s (5000 fpm) = 5 equivalent duct diameters. If the duct is rectangular with side dimensions a and b, the equivalent duct diameter is equal to (4ab/π)0.5. Figure 8.1 - Fan Outlet Velocity Profiles Fans and Systems | 79

8.1.1 Axial flow fan - outlet ducts. Most exhaust axial flow fans are tested and/or rated with two to three equivalent duct diameters attached to the fan outlet. Often, fans are installed without an outlet duct, either because of available space or for economic reasons. Tubeaxial fans installed with no outlet ducts have System Effect Factors (SEF) approaching zero. Vaneaxial fans, however, do not perform as cataloged when they are installed with less than 50% "effective duct length." System Effect Curves for tubeaxial and vaneaxial fans with less than optimum outlet duct are shown in Figure 8.2. To determine the applicable SEF, calculate the average velocity in the outlet duct and enter the System Effect Curve (Figure 7.1) at this velocity, utilizing the appropriate System Effect Curve selected from Figure 8.2, then read over horizontally to the System Effect Factor, Pascals (in. wg) on the ordinate. 8.1.2 Centrifugal flow fan - outlet ducts. Centrifugal fans are sometimes installed with a less than optimum outlet duct. If it is not possible to use a

full-length outlet duct, then a SEF must be added to the system resistance losses. System Effect Curves for centrifugal fans with less than optimum outlet duct length are shown in Figure 8.3.

8.2 Outlet diffusers Many air systems are space-constricted and must, of necessity, use relatively small ducts having high static pressure losses. If space is not severely constricted, the use of larger ductwork and moving air at a lower velocity may be beneficial. Larger ductwork (within reason) reduces system pressure requirements. To effectively transition from a smaller duct size to a larger duct size it is necessary to use a connection piece between the duct sections that allows the airstream to expand gradually. This piece is called a diffuser, or evasé. These terms are used interchangeably in the industry. A properly designed evasé has a smooth and gradual transition between the duct sizes so that airflow is relatively undisturbed. An evasé operates on a very simple principle: air flowing from the smaller area to the larger area loses

AXIAL FAN

100% EFFECTIVE DUCT LENGTH To calculate 100% duct length, assume a minimum of 2½ duct diameters for 12.7 m/s (2500 fpm) or less. Add 1 duct diameter for each additional 5.08 m/s (1000 fpm). EXAMPLE: 25.4 m/s (5000 fpm) = 5 equivalent duct diameters

No Duct

12% Effective Duct

25% Effective Duct

50 % Effective Duct

100% Effective Duct

Tubeaxial Fan

---

---

---

---

---

Vaneaxial Fan

U

V

W

---

---

Determine SEF by using Figure 7.1 Figure 8.2 - System Effect Curves for Outlet Ducts - Axial Fans 80 | Fans and Systems

velocity as it approaches the larger area, and a portion of the change (reduction) in velocity pressure is converted into static pressure. This process is called “static regain”, and is simply defined as the conversion of velocity pressure to static pressure. The efficiency of conversion (or loss of total pressure) will depend upon the angle of expansion, the length of the evasé section, and the blast area/outlet area ratio of the fan. The fan manufacturer will, in most cases, be able to provide design information for an efficient diffuser.

See AMCA Publication 200 Air Systems, for an example showing the effect of a diffuser on a duct exit.

8.3 Outlet duct elbows Values for pressure losses through elbows, which are published in handbooks and textbooks, are based upon a uniform velocity profile at entry into the elbow. Any non-uniformity in the velocity profile ahead of the elbow will result in a pressure loss greater than the industry-accepted value.

BLAST AREA DISCHARGE DUCT OUTLET AREA

CUTOFF

100% EFFECTIVE DUCT LENGTH

CENTRIFUGAL FAN To calculate 100% duct length, assume a minimum of 2½ duct diameters for 2500 fpm or less. Add 1 duct diameter for each additional 1000 fpm. EXAMPLE: 5000 fpm = 5 equivalent duct diameters. If the duct is rectangular with side dimensions a and b, the equivalent duct diameter is equal to (4ab/π)0.5.

Pressure Recovery

No Duct

12% Effective Duct

25% Effective Duct

50% Effective Duct

100% Effective Duct

0%

50%

80%

90%

100%

W W W-X — — — —

— — — — — — —

Blast Area Outlet Area 0.4 0.5 0.6 0.7 0.8 0.9 1.0

System Effect Curve P P R-S S T-U V-W —

R-S R-S S-T U V-W W-X —

U U U-V W-X X — —

Determine SEF by using Figure 7.1 Figure 8.3 - System Effect Curves for Outlet Ducts - Centrifugal Fans Fans and Systems | 81

Since the velocity profile at the outlet of a fan is not uniform, an elbow located at or near the fan outlet will develop a pressure loss greater than the industryaccepted value.

8.3.1 Axial fans - outlet duct elbows. Tubeaxial fans with two-piece and four-piece mitered elbows at varying distances from the fan outlet have a negligible SEF (see Figure 8.4).

The amount of this loss will depend upon the location and orientation of the elbow relative to the fan outlet. In some cases, the effect of the elbow will be to further distort the outlet velocity profile of the fan. This will increase the losses and may result in such uneven airflow in the duct that branch- takeoffs near the elbow will not deliver their design airflow. (See Section 8.6)

Vaneaxial fans with two and four-piece mitered elbows at varying distances from the fan outlet resulted in System Effect Curves as shown in Figure 8.4.

Wherever possible, a length of straight duct should be installed at the fan outlet to permit the diffusion and development of a uniform airflow profile before an elbow is inserted in the duct. If an elbow must be located near the fan outlet then it should be a radius elbow having a minimum radius-to-duct-diameter ratio of 1.5.

8.3.2 Centrifugal fans - outlet duct elbows. The outlet velocity of centrifugal fans is generally higher toward one or adjacent sides of the rectangular duct. If an elbow must be located near the fan outlet it should have a minimum radius-to-duct-diameter ratio of 1.5, and it should be arranged to give the most uniform airflow possible. Figure 8.5 gives System Effect Curves that can be used to estimate the effect of an elbow at the fan outlet. It also shows the reduction in losses resulting from the use of a straight outlet duct.

TUBEAXIAL FAN SHOWN

% EFFECTIVE DUCT LENGTH

% EFFECTIVE DUCT LENGTH

VANEAXIAL FAN SHOWN

90° Elbow

No Duct

12% Effective Duct

25% Effective Duct

50 % Effective Duct

100% Effective Duct

Tubeaxial Fan

2 & 4 Pc

---

---

---

---

---

Vaneaxial Fan

2 Pc

U

U-V

V

W

---

Vaneaxial Fan

4 Pc

W

---

---

---

---

Determine SEF by using Figures 7.1 and 8.1 Figure 8.4 - System Effect Curves for Outlet Duct Elbows - Axial Fans 82 | Fans and Systems

POSITION C

POSITION D

POSITION B

E TIV TH C G FE EF LEN % CT DU

INL

ET

POSITION A

SWSI CENTRIFUGAL FAN SHOWN

Note: Fan Inlet and elbow positions must be oriented as shown for the proper application of the table on the facing page. Figure 8.5 - Outlet Elbows on SWSI Centrifugal Fans

Fans and Systems | 83

Outlet Elbow Position

No Outlet Duct

12% Effective Duct

25% Effective Duct

50% Effective Duct

0.4

A B C D

N M-N L-M L-M

O N M M

P-Q O-P N N

S R-S Q Q

0.5

A B C D

O-P N-O M-N M-N

P-Q O-P N N

R Q O-P O-P

T S-T R-S R-S

0.6

A B C D

Q P N-O N-O

Q-R Q O O

S R Q Q

U T S S

0.7

A B C D

R-S Q-R P P

S R-S Q Q

T S-T R-S R-S

V U-V T T

0.8

A B C D

S R-S Q-R Q-R

S-T S R R

T-U T S S

W V U-V U-V

0.9

A B C D

T S R R

T-U S-T S S

U-V T-U S-T S-T

W W V V

1.0

A B C D

T S-T R-S R-S

T-U T S S

U-V U T T

W W V V

SYSTEM EFFECT CURVES FOR SWSI FANS

DETERMINE SEF BY USING FIGURES 7.1 AND 8.1 For DWDI fans determine SEF using the curve for SWSI fans. Then, apply the appropriate multiplier from the tabulation below MULTIPLIERS FOR DWDI FANS ELBOW POSITION A = ∆P × 1.00 ELBOW POSITION B = ∆P × 1.25 ELBOW POSITION C = ∆P × 1.00 ELBOW POSITION D = ∆P × 0.85 Figure 8.5 - Outlet Elbows on SWSI Centrifugal Fans 84 | Fans and Systems

100% Effective Duct

NO System Effect Factor

Blast Area Outlet Area

a large plenum or to free space a parallel blade damper may be satisfactory.

8.4 Turning vanes Turning vanes will usually reduce the pressure loss through an elbow, however, where a non-uniform approach velocity profile exists, such as at a fan outlet, the vanes may serve to continue the nonuniform profile beyond the elbow. This may result in increased losses in other system components downstream of the elbow.

8.5 Volume control dampers Volume control dampers are manufactured with either "opposed" blades or "parallel" blades. When partially closed, the parallel bladed damper diverts the airstream to the side of the duct. This results in a non-uniform velocity profile beyond the damper and airflow to branch ducts close to the downstream side may be seriously affected. The use of an opposed blade damper is recommended when air volume control is required at the fan outlet and there are other system components, such as coils or branch takeoffs downstream of the fan. When the fan discharges into

PARALLEL-BLADE DAMPER ILLUSTRATING DIVERTED AIRFLOW

For a centrifugal fan, best air performance will usually be achieved by installing an opposed blade damper with its blades perpendicular to the fan shaft; however, other considerations, such as the need for thrust bearings, may require installation of the damper with its blades parallel to the fan shaft. When a damper is required, it is often furnished as accessory equipment by the fan manufacturer (see Figure 8.6). In many systems, a volume control damper will be located in the ductwork at or near the fan outlet. Published pressure drops for wide-open control dampers are based on uniform approach velocity profiles. When a damper is installed close to the outlet of a fan the approach velocity profile is nonuniform and much higher pressure losses through the damper can result. Figure 8.7 lists multipliers that should be applied to the damper manufacturer's catalog pressure drop when the damper is installed at the outlet of a centrifugal fan. These multipliers should be applied to all types of fan outlet dampers.

OPPOSED-BLADE DAMPER ILLUSTRATING NON-DIVERTED AIRFLOW

Figure 8.6 - Parallel Blade vs. Opposed Blade Damper

Fans and Systems | 85

VOLUME CONTROL DAMPER

BLAST AREA OUTLET AREA

PRESSURE DROP MULTIPLIER

0.4

7.5

0.5

4.8

0.6

3.3

0.7

2.4

0.8

1.9

0.9

1.5

1.0

1.2

Figure 8.7 - Pressure Drop Multipliers for Volume Control Dampers on a Fan Discharge 86 | Fans and Systems

8.6 Duct branches Standard procedures for the design of duct systems are based on the assumption of uniform airflow profiles in the system.

In Figure 8.8 branch takeoffs or splits are located close to the fan outlet. Non-uniform airflow conditions will exist and pressure loss and airflow may vary widely from the design intent. Wherever possible a length of straight duct should be installed between the fan outlet and any split or branch takeoff.

Note: Avoid location of split or duct branch close to fan discharge. Provide a straight section of duct to allow for air diffusion. Figure 8.8 - Branches Located Too Close to Fan Fans and Systems | 87

loss of energy, or even a flat flange (e) on the end of the duct or fan will reduce the loss to about one half of the loss through an un-flanged entry.

9. Inlet System Effect Factors Fan performance can be greatly affected by nonuniform or swirling inlet flow. Fan rating and catalog performance is typically obtained with unobstructed inlet flow. Any disruption to the inlet airflow will reduce a fan’s performance. Restricted fan inlets located close to walls, obstructions or restrictions caused by a plenum or cabinet will also decrease the performance of a fan. The fan performance loss due to inlet airflow disruption must be considered as a System Effect.

ANSI/AMCA 210 limits an inlet duct to a crosssectional area no greater than 112.5% or less than 92.5% of the fan inlet area. The slope of transition elements is limited to 15° converging and 7° diverging.

9.2 Inlet duct elbows Non-uniform airflow into a fan inlet is a common cause of deficient fan performance. An elbow located at, or in close proximity to the fan inlet will not allow the air to enter the impeller uniformly. The result is less than cataloged air performance.

9.1 Inlet ducts Fans intended primarily for use as "exhausters" may be tested with an inlet duct in place, or with a special bell-mouthed inlet to simulate the effect of a duct. Figure 9.1 illustrates variations in inlet airflow that will occur. The ducted inlet condition is shown as (a), and the effect of the bell-mouth inlet as (b).

A word of caution is required with the use of inlet elbows in close proximity to fan inlets. Other than the incurred System Effect Factor, instability in fan operation may occur as evidenced by an increase in pressure fluctuations and sound power level. Fan instability, for any reason, may result in serious structural damage to the fan. Axial fan instabilities were experienced in some configurations tested with inlet elbows in close proximity to the fan inlet. Pressure fluctuations approached ten (10) times the magnitude of fluctuations of the same fan with good inlet and outlet conditions. It is strongly advised that inlet elbows be installed a minimum of three (3) diameters away from any axial or centrifugal fan inlet.

Fans that do not have smooth entries (c), and are installed without ducts, exhibit airflow characteristics similar to a sharp edged orifice that develops a vena contracta. A reduction in airflow area is caused by the vena contracta and the following rapid expansion causes a loss that should be considered as a System Effect. If it is not practical to include such a smooth entry, a converging taper (d) will substantially diminish the

a.

c.

b. BELL MOUTH INLET PRODUCES FULL FLOW INTO FAN

IDEAL SMOOTH ENTRY TO DUCT ON A DUCT SYSTEM

d. CONVERGING TAPERED ENTRY INTO FAN OR DUCT SYSTEM

VENA CONTRACTA AT INLET REDUCES EFFECTIVE FAN INLET AREA

e. FLANGED ENTRY INTO FAN OR DUCT SYTEM

Figure 9.1 Typical Inlet Connections for Centrifugal and Axial Fans 88 | Fans and Systems

9.2.1 Axial fans - inlet duct elbows. The System Effect Curves shown in Figure 9.2 for tubeaxial and vaneaxial fans are the result of tests run with two and four piece mitered inlet elbows at or in close proximity to the fan inlets. Other variables tested included hubto-tip (H/T) ratio and blade solidity. The number of blades did not have a significant affect on the inlet elbow SEF.

are listed on Figure 9.4, and the System Effect Curves for various square duct elbows of given radius/diameter ratios are listed on Figure 9.5. The SEF for a particular elbow is found in Figure 7.1 at the intersection of the average fan inlet velocity and the tabulated System Effect Curve. This pressure loss should be added to the friction and dynamic losses already determined for the particular elbow. Note that when duct turning vanes and/or a suitable length of duct is used (three to eight diameters long, depending on velocities) between the fan inlet and the elbow, the SEF is not as great. These improvements help maintain uniform airflow

9.2.2 Centrifugal fans - inlet duct elbows. Nonuniform airflow into a fan inlet, Figure 9.3A, is a common cause of deficient fan performance. The System Effect Curves for mitered 90° round section elbows of given radius/diameter (R/D) ratios

TUBEAXIAL FAN SHOWN

DUCT LENGTH

DUCT LENGTH

VANEAXIAL FAN SHOWN H/T

90° Elbow

Tubeaxial Fan

.25

2 piece

U

Tubeaxial Fan

.25

4 piece

Tubeaxial Fan

.35

Vaneaxial Fan Vaneaxial Fan

[1][2]

[1][2]

1.0D [1][2]

3.0D

V

W

---

X

---

---

---

2 piece

V

W

X

.61

2 piece

Q-R

Q-R

S-T

T-U

.61

4 piece

W

W-X

---

---

No Duct

0.5D

Notes: [1] Instability in fan operation may occur as evidenced by an increase in pressure fluctuations and sound level. Fan instability, for any reason, may result in serious structural damage to the fan. [2] The data presented in Figure 9.2 is representative of commercial type tubeaxial and vaneaxial fans, i.e. 60% to 70% fan static efficiency. Figure 9.2 - System Effect Curves for Inlet Duct Elbows - Axial Fans Fans and Systems | 89

into the fan inlet and thereby approach the airflow conditions of the laboratory test setup. Occasionally, where space is limited, the inlet duct will be mounted directly to the fan inlet as shown on Figure 9.3B. The many possible variations in the width and depth of a duct influence the reduction in performance to varying degrees and makes it impossible to establish reliable SEF. Note: Capacity losses as high as 45% have been observed in poorly designed inlets such as in Figure 9.3B. This inlet condition should be AVOIDED. Existing installations can be improved with guide vanes or the conversion to square or mitered elbows with guide vanes, but a better alternative would be a specially designed inlet box similar to that shown in Figure 9.6. 9.2.3 Inlet boxes. Inlet boxes are added to centrifugal and axial fans instead of elbows in order to provide more predictable inlet conditions and to maintain stable fan performance. They may also be used to protect fan bearings from high temperature, or corrosive / erosive gases. The fan manufacturer should include the effect of any inlet box on the fan performance, and when evaluating a proposal it should be established that an appropriate loss has been incorporated in the fan rating. Should this information not be available from the manufacturer, refer to Section 10.4 for an approximate System Effect.

A counter-rotating vortex at the inlet may result in a slight increase in the pressure-volume curve but the power will increase substantially. There are occasions, with counter-rotating swirl, when the loss of performance is accompanied by a surging airflow. In these cases, the surging may be more objectionable than the performance change. Inlet spin may arise from a great variety of approach conditions and sometimes the cause is not obvious.

D

LENGTH OF DUCT

R

Figure 9.3A - Non-Uniform Airflow Into a Fan Inlet Induced by a 90°, 3-Piece Section Elbow-No Turning Vanes

9.3 Inlet vortex (spin or swirl) Another major cause of reduced performance is an inlet duct design or fan installation that produces a vortex or spin in the airstream entering a fan inlet. An example of this condition is illustrated in Figure 9.7. An ideal inlet condition allows the air to enter uniformly without spin in either direction. A spin in the same direction as the impeller rotation (pre-rotation) reduces the pressure- volume curve by an amount dependent upon the intensity of the vortex. The effect is similar to the change in the pressure-volume curve achieved by variable inlet vanes installed in a fan inlet; the vanes induce a controlled spin in the direction of impeller rotation, reducing the airflow, pressure and power (see Section 10.6).

90 | Fans and Systems

Figure 9.3B - Non-Uniform Airflow Induced Into Fan Inlet by a Rectangular Inlet Duct

SYSTEM EFFECT CURVES

LENGTH OF DUCT

D

R/D

NO DUCT

—

N

+ R

2D 5D DUCT DUCT P

R-S

Figure 9.4A - Two Piece Mitered 90° Round Section Elbow - Not Vaned

SYSTEM EFFECT CURVES

LENGTH OF DUCT

D

R/D

NO DUCT

2D 5D DUCT DUCT

0.5

O

Q

S

0.75

Q

R-S

T-U

1.0

R

S-T

U-V

2.0

R-S

T

U-V

3.0

S

T-U

V

+ R

Figure 9.4B - Three Piece Mitered 90° Round Section Elbow - Not Vaned SYSTEM EFFECT CURVES

LENGTH OF DUCT

R/D

NO DUCT

2D 5D DUCT DUCT

0.5

P-Q

R-S

T

0.75

Q-R

S

U

1.0

R

S-T

U-V

2.0

R-S

T

U-V

3.0

S-T

U

V-W

D

+ R

Figure 9.4C - Four or More Piece Mitered 90° Round Section Elbow - Not Vaned

DETERMINE SEF BY USING FIGURE 7.1 Figure 9.4 - System Effect Curves for Various Mitered Elbows without Turing Vanes Fans and Systems | 91

SYSTEM EFFECT CURVES H R/D

NO DUCT

2D 5D DUCT DUCT

0.5

O

Q

S

0.75

P

R

S-T

1.0

R

S-T

U-V

1.0

S

T-U

V

H

LENGTH OF DUCT + R

Figure 9.5A - Square Elbow with Inlet Transition - No Turning Vanes

H

SYSTEM EFFECT CURVES R/D

NO DUCT

0.5

S

T-U

V

1.0

T

U-V

W

2.0

V

V-W

W-X

H

LENGTH OF DUCT +

2D 5D DUCT DUCT

R

Figure 9.5B - Square Elbow with Inlet Transition - 3 Long Turning Vanes

SYSTEM EFFECT CURVES

H

R/D

NO DUCT

0.5

S

T-U

V

1.0

T

U-V

W

2.0

V

V-W

W-X

H

LENGTH OF DUCT

R

+

2D 5D DUCT DUCT

Figure 9.5C - Square Elbow with Inlet Transition - Short Turning Vanes D = Diameter of the inlet collar The inside area of the square duct (H x H) should be equal to the inside area of the fan inlet collar. * The maximum permissible angle of any converging element of the transition is 15°, and for a diverging element, 7°. DETERMINE SEF BY USING FIGURE 7.1 Figure 9.5 - System Effect Curves for Various Square Duct Elbows 92 | Fans and Systems

Figure 9.6 - Improved Flow Conditions with a Special Designed Inlet Box

IMPELLER ROTATION

COUNTER-ROTATING SWIRL Figure 9.7 - Example of a Forced Inlet Vortex

IMPELLER ROTATION

PRE-ROTATING SWIRL

IMPELLER ROTATION

COUNTER-ROTATING SWIRL

Figure 9.8 - Inlet Duct Connections Causing Inlet Spin Fans and Systems | 93

airflow entering a duct elbow with turning vanes will leave the duct elbow with non-uniform airflow.

9.4 Inlet turning vanes Where space limitations prevent the use of optimum fan inlet conditions, more uniform airflow can be achieved by the use of turning vanes in the inlet elbow (see Figure 9.9). Numerous variations of turning vanes are available, from a single curved sheet metal vane to multi-bladed "airfoil" vanes. The pressure drop (loss) through these devices must be added to the system pressure losses. The amount of loss for each device is published by the manufacturer, but it should be realized that the cataloged pressure loss will be based upon uniform airflow at the entry to the elbow. If the airflow approaching the elbow is significantly non-uniform because of a disturbance farther upstream in the system, the pressure loss through the elbow will be higher than the published figure. A non-uniform

9.5 Airflow straighteners Figure 9.10 shows two airflow straighteners used in testing setups to reduce fan swirl before measuring stations. Figure 9.10A is the egg-crate straightener used in ANSI/AMCA 210; larger cell sizes made proportionately longer could be used. Figure 9.10B shows the star straightener used in the ISO standard. A single splitter sheet may be used to eliminate swirl in some cases. Straighteners are intended to reduce swirl before or after a fan or a process station. Do not install straighteners where the air profile is known to be non-uniform, the device will carry the non-uniformity further downstream.

TURNING VANES

TURNING VANES IMPELLER ROTATION CORRECTED PREROTATING SWIRL

TURNING VANES

CORRECTED COUNTERROTATING SWIRL

Figure 9.9 - Corrections for Inlet Spin

94 | Fans and Systems

IMPELLER ROTATION

0.45D

D

0.075D DUCT 0.075D

Figure 9.10A - ANSI/AMCA Standard 210 Egg-Crate Straightener

DUCT

DUCT

D

2D Figure 9.10B - ISO 5801 Star Straightener

Figure 9.10 - Test Standard Airflow Straighteners Fans and Systems | 95

one-half impeller diameter between an enclosure wall and the fan inlet. Adjacent inlets of multiple double width centrifugal fans located in a common enclosure should be at least one impeller diameter apart if optimum performance is to be expected. Figure 9.11 illustrates fans with restricted inlets and their applicable System Effect Curves.

9.6 Enclosures (plenum and cabinet effects) Fans within plenums and cabinets or next to walls should be located so that air may flow unobstructed into the inlets. Fan performance is reduced if the space between the fan inlet and the enclosure is too restrictive. It is common practice to allow at least

2L

L

EQUAL

INLET DIA.

L

EQUAL

DIAMETER OF INLET

Figure 9.11A - Fans and Plenum

L

L

Figure 9.11B - Axial Fan Near Wall

L

DWDI

L

SWSI

Figure 9.11C - Centrifugal Fan Near Wall(s)

Figure 9.11D - DWDI Fan Near Wall on One Side

L - DISTANCE INLET TO WALL

For Figures 9.11A, B & C SYSTEM EFFECT CURVES

0.75 x DIA. OF INLET 0.50 x DIA. OF INLET 0.40 x DIA. OF INLET 0.30 x DIA. OF INLET

V-W U T S

For Figures 9.11D SYSTEM EFFECT CURVES

X V-W V-W U

Determine SEF by calculating inlet velocity and using Figure 7.1

Figure 9.11 - System Effect Curves for Fans Located in Plenums and Cabinet Enclosures and for Various Wall-to-inlet Dimensions 96 | Fans and Systems

L

The manner in which the air stream enters an enclosure in relation to the fan inlets also affects fan performance. Plenum or enclosure inlets or walls that are not symmetrical with the fan inlets will cause uneven airflow and/or inlet spin. Figure 9.12A illustrates this condition that must be avoided to achieve maximum performance from a fan. If this is not possible, inlet conditions can usually be improved with a splitter sheet to break up the inlet vortex as illustrated in Figure 9.12B.

common inlet obstructions. Some accessories such as fan bearings, bearing pedestals, inlet vanes, inlet dampers, drive guards and motors may also cause inlet obstruction and are discussed in more detail in Section 10. Obstruction at the fan inlet may be defined in terms of the unobstructed percentage of the inlet area. Because of the shape of the inlet cones of many fans it is sometimes difficult to establish the area of the fan inlet. Figure 9.14 illustrates the convention adopted for this purpose. Where an inlet collar is provided, the inlet area is calculated from the inside diameter of this collar. Where no collar is provided, the inlet plane is defined by the points of tangency of the fan housing side with the inlet cone radius.

For proper performance of axial fans in parallel installations minimum space of one impeller diameter should be allowed between fans, as shown in Figure 9.13. Placing fans closer together can result in erratic or uneven airflow into the fans.

The unobstructed percentage of the inlet area is calculated by projecting the profile of the obstruction on the profile of the inlet. The adjusted inlet velocity obtained is then used to enter the System Effect Curve chart and the SEF determined from the curve listed for that unobstructed percentage of the fan inlet area.

9.7 Obstructed inlets A reduction in fan performance can be expected when an obstruction to airflow is located in the fan inlet. Building structural members, columns, butterfly valves, blast gates and pipes are examples of more

SPLITTER SHEET

Figure 9.12A - Enclosure Inlet Not Symmetrical with Fan Inlet. PreRotational Vortex Induced

Figure 9.12B - Flow Condition of Figure 9.12A Improved with a Splitter Sheet. Substantial Improvement Would Be To Relocate Enclosure Inlet as Shown in Figure 9.11A

Figure 9.12 - Fan in Plenum with Non-Symmetrical Inlet

1 DIA. MIN

Figure 9.13 - Parallel Installation of Axial Flow Fans Fans and Systems | 97

ER

ET

AM

DE

SI

IN

DI

T

AR

LL

CO

E NL

I

INLET PLANE

FREE INLET AREA PLANE - FAN WITH INLET COLLAR POINT OF TANGENT WITH FAN HOUSING SIDE AND INLET CONE RADIUS

R TE E NT AM GE N DI TA OF

INLET PLANE

FREE INLET AREA PLANE - FAN WITHOUT INLET COLLAR

Table for Figure 9.14 System Effect Curve (Figure 7.1) Distance from obstruction to inlet plane Percentage of unobstructed inlet area

0.75 Inlet diameter

0.5 Inlet diameter

0.33 Inlet diameter

0.25 Inlet diameter

At Inlet plane

100

-

-

-

-

-

95

-

-

X

W

V

90

-

X

V-W

U-V

T-U

85

X

W-X

V-W

U-V

S-T

75

W-X

V

U

S-T

R-S

50

V-W

U

S-T

R-S

Q

25

U-V

T

S-T

Q-R

P

Figure 9.14 - System Effect Curves for Inlet Obstructions (Table based on Fans and Fan Systems, Thompson & Trickler, Chem Eng MAR83, p. 60) 98 | Fans and Systems

10. Effects of Factory Supplied Accessories Unless the manufacturer's catalog clearly states to the contrary, it should be assumed that published fan performance data does not include the effects of any accessories supplied with the fan.

If possible, the necessary information should be obtained directly from the manufacturer. The data presented in this section are offered only as a guide in the absence of specific data from the fan manufacturer. See Figure 10.1 for terminology.

Cone Type Variable Inlet Vanes

Figure 10.1 - Common Terminology for Centrifugal Fan Appurtenances Fans and Systems | 99

10.1 Bearing and supports in fan inlet

10.3 Belt tube in axial fan inlet or outlet

Arrangement 3 and 7 fans (see Figure 3.5) require that the fan shaft be supported by a bearing and bearing support in the fan inlet or just adjacent to it.

With a belt driven axial flow fan it is usually necessary that the fan motor be mounted outside the fan housing (see Figure 3.7 Arrangement 9, and Annex B Figure B.7).

These components may have an effect on the flow of air into the fan inlet and consequently on the fan performance, depending upon the size of the bearings and supports in relation to the fan inlet opening. The location of the bearing and support, that is, whether it is located in the actual inlet or "spaced out" from the inlet, will also have an effect. In cases where manufacturer's performance ratings do not include the effect of the bearings and supports, it will be necessary to compensate for this inlet restriction. Use the fan manufacturer's allowance for bearings in the fan inlet if possible. If no better data are available, use the procedures described in Section 9.7 as an approximation.

To protect the belts from the airstream, and also to prevent any air leakage through the fan housing, manufacturers in many cases provide a belt tube. Most manufacturers include the effects of an axial fan belt tube in their rating tables. In cases where the effect is not included, the appropriate SEF is approximated by calculating the percentage of unobstructed area of air passage way and using Figure 9.14.

10.4 Inlet box When an inlet box configuration is supplied by the fan manufacturer, the fan performance should include the effect of the inlet box.

10.2 Drive guards obstructing fan inlet All fans have moving parts that require guarding for safety in the same way as other moving machinery. Fans located less than 2.1 m (7 ft) above the floor require special consideration as specified in the United States’ Occupational Safety and Health Act. National, federal, state and local rules, regulations, and codes should be carefully considered and followed. Arrangement 3 and 7 fans may require a belt drive guard in the area of the fan inlet. Depending on the design, the guard may be located in the plane of the inlet, along the casing side sheet, or it may be "spaced out" due to "spaced out" bearing pedestals. In any case, depending on the location of the guard, and on the inlet velocity, the fan performance may be significantly affected by this obstruction. It is desirable that a drive guard located in this position be furnished with as much opening as possible to allow maximum flow of air to the fan inlet. If available, use the fan manufacturer's allowance for drive guards obstructing the fan inlet. SEF for drive guard obstructions situated at the inlet of a fan may be approximated using Figure 9.14. Where possible, open construction on guards is recommended to allow free air passage to the fan inlet. Guards and sheaves should be designed to obstruct, as little of the fan inlet as possible and in no case should the obstruction be more than 1/3 of the fan inlet area. 100 | Fans and Systems

The System Effect of fan inlet boxes can vary widely depending upon the design. This data should be available from the fan manufacturer. In the absence of fan manufacturer's data, a well-designed inlet box should approximate System Effect Curves "S" or "T" of Figure 7.1.

10.5 Inlet box dampers Inlet box dampers may be used to control the airflow through the system. Either parallel or opposed blades may be used (see Figure 10.1). The parallel blade type is installed with the blades parallel to the fan shaft so that, in a partially closed position, a forced inlet vortex will be generated. The effect on the fan characteristics will be similar to that of a variable inlet vane control. The opposed blade type is used to control airflow by the addition of pressure loss created by the damper in a partially closed position. If possible, complete data should be obtained from the fan manufacturer giving the System Effect of the inlet box and damper pressure drop over the range of application. If data are not available, System Effect Curves "S" or "T" from Figure 7.1 should be applied for the inlet box and pressure loss from the damper manufacturer for the damper in making the fan selection.

When variable inlet vanes are supplied by the fan manufacturer, the performance should include the effects of the variable inlet vane unit.

10.6 Variable inlet vane (VIV) Variable inlet vanes are mounted on the fan inlet to maintain fan efficiency at reduced airflow. They are arranged to generate an inlet vortex (pre-rotation) that rotates in the same direction as the fan impeller. Variable inlet vanes may be of two different basic types: 1) cone type integral with the fan inlet, 2) cylindrical type add-on (Figures 10.1 and 10.2).

VANE TYPE

The System Effect of a wide-open VIV (see Figure 10.2) must be accounted for in the original fan selection. If data are not available from the fan manufacturer the following System Effect Curves should be applied in making the fan selection.

SYSTEM EFFECT CURVE (100% Open)

a) Cone type, integral b) Cylindrical type

“Q” or “R” “S”

Determine SEF by calculating inlet velocity and using Figure 7.1

FAN PERFORMANCE W/OUT VARIABLE INLET VANES

120

CONE TYPE VARIABLE INLET VANES

CYLINDRICAL TYPE VARIABLE INLET VANES

PERCENT OF SHUT-OFF PRESSURE

100

VARIABLE INLET VANES 100% OPEN

75% OPEN

80 75% OPEN

60

40 75% OPEN

20

0

20

40

60

80

100

120

PERCENT OF WIDE OPEN VOLUME

Figure 10.2 - Typical Variable Inlet Vanes for a Backward Inclined Fan

Fans and Systems | 101

Annex A. SI / I-P Conversion Table (Informative) Taken from AMCA 99-0100

Quantity

I-P to SI

SI to I-P

Length

(ft) 0.3048 = m

(m) 3.2808 = ft

Mass (weight)

(lbs) 0.4536 = kg

(kg) 2.2046 = lbs.

Time

The unit of time is the second in both systems

Velocity

(ft-s) 0.3048 = ms (ft/min) 0.00508 = ms

(ms) 3.2808 = ft-s (ms) 196.85 = ft/min

Acceleration

(in./s2) 0.0254 = m/s2

(m/s2) 39.370 = in./s2

Area

(ft2) 0.09290 = m2

(m2) 10.764 = ft2

Volume Flow Rate

(cfm) 0.000471948 = m3/s

(m3/s) 2118.88 = cfm

Density

(lb/ft3) 16.01846 = kg/m3

(kg/m3) 0.06243 = lb/ft3

Pressure

(in. wg) 248.36 = Pa (in. wg) 0.24836 = kPa (in. Hg) 3.3864 = kPa

(Pa) 0.004026 = in. wg (kPa) 4.0264 = in. wg (kPa) 0.2953 = in. Hg

Viscosity: Absolute Kinematic

(lbm/ft-s) 1.4882 = Pa s (ft2/s) 0.0929 = m2/s

(Pa s) 0.6719 = (lbm/ft-s) (m2/s) 10.7639 = ft2/s

Gas Constant

(ft lb/lbm-°R) 5.3803 = J-kg/K

(j-kg/K) 0.1858 = (ft lb/lbm-°R)

Temperature

(°F - 32°)/1.8 = °C

(1.8 × °C) + 32° = °F

Power

(BHP) 746 = W (BHP) 0.746 = kW

(W)/746 = BHP (kW)/0.746 = BHP

102 | Fans and Systems

Annex B. Dual Fan Systems - Series and Parallel It is sometimes necessary to install two or more fans in systems that require higher pressures or airflow than would be attainable with a single fan. Two fans may offer a space, cost, or control advantage over a single larger fan, or it may be simply a field modification of an existing system to boost pressure or airflow.

B.1 Fans operating in series To obtain a system pressure boost, fans are often installed in series. The fans may be mounted as close as the outlet of one fan directly attached to the inlet of the next fan, or they may be placed in remote locations with considerable distance between fans. The fans must handle the same mass airflow, assuming no loss or gains between stages. The combined total pressure will then be the sum of each fan’s total pressure (Figure B.1). The velocity pressure corresponds to the air velocity at the outlet of the last fan stage. The static pressure for the combination is the total pressure minus the velocity pressure and is not the sum of the individual fan static pressures. In practice there is some reduction in airflow due to the increased air density in the later fan stage(s). There can also be significant loss of airflow due to non-uniform airflow into the inlet of the next fan. Sometimes multiple impellers are assembled in a single housing and this assembly is known as a “multi-stage” fan. This combination is seldom used in conventional ventilating and air conditioning systems but it is not uncommon in special industrial systems. It is advisable to request the fan manufacturer to review the proposed system design and make some estimate of its installed performance.

These types of systems normally have common inlet and outlet sections, or they may have individual ducts of equal resistance that join together at equal velocities. In either case, the characteristic curve is the sum of the separate airflows for a given static or total pressure (Figure B.2). The total performance of the multiple fans will be less than the theoretical sum if inlet conditions are restricted or the airflow into the inlets is not straight (see Section 9.6). Also, adding a parallel fan to an existing system without modifying the resistance (larger ducts, etc.) will result in lower than anticipated airflow due to increased system resistance. Fans that have a “positive” slope in the pressurevolume curve to the left of the peak pressure curve, typical of some axial and forward curved centrifugal fans (see Figure 4.2), can experience unstable operation under certain conditions. If fans are operated in parallel in the region of this “positive” slope, multiple operating conditions may occur. Figure B.2 illustrates the combined pressure-volume curve of two such fans operating in parallel. The closed loop to the left of the peak pressure point is the result of plotting all the possible combinations of volume airflow at each pressure. If the system curve intersects the combined volume-pressure curve in the area enclosed by the loop, more than one point of operation is possible. This may cause one of the fans to handle more of the air and could cause a motor overload if the fans are individually driven. This unbalanced airflow condition tends to reverse readily with the result that the fans will intermittently load and unload. This "pulsing" often generates noise and vibration and may cause damage to the fans, ductwork or driving motors. Aileron controls in forward curved fan outlets or dampers near the inlets or outlets may be used to correct unbalanced airflow or to eliminate pulsations or reversing operation (See Figure B.3).

B.2 Fans operating in parallel Suppliers of air handling equipment and designers of custom systems commonly incorporate two identical, in parallel fans to deliver large volumes of air while taking advantage of the space savings offered by using two smaller fans.

Fans and Systems | 103

PERCENT OF FAN STATIC PRESSURE

SYSTEM RESISTANCE

200%

SERIES FAN COMBINED PRESSURE CURVE

100% SINGLE FAN PRESSURE CURVE

100% PERCENT OF FAN AIRFLOW

Figure B.1 - Typical Characteristic Curve of Two Fans Operating in Series

104 | Fans and Systems

STA BL ES YS TE M

UN STA BL ES YS TE M

PERCENT OF FAN STATIC PRESSURE

FAN OPERATION NOT RECOMMENDED IN THIS RANGE

100 PARALLEL FANS - FAN PRESSURE AT COMBINED VOLUME

SINGLE FAN PRESSURE CURVE

200 PERCENT OF FAN AIRFLOW

Figure B.2 - Parallel Fan Operation

AILERON

Figure B.3 - Aileron Control

Fans and Systems | 105

Annex C. Definitions and Terminology C.1 The air C.1.1 Air velocity. The velocity of an air stream is its rate of motion, expressed in m/s (fpm). The velocity at a plane (Vx) is the average velocity throughout the entire area of the plane. C.1.2 Airflow. The airflow at a plane (Qx) is the rate of airflow, expressed in m3/s (cfm) and is the product of the average velocity at the plane and the area of the plane. C.1.3 Barometric pressure. Barometric pressure (pb) is the absolute pressure exerted by the atmosphere at a location of measurement (per AMCA 99-0066). C.1.4 Pressure-static. Static pressure is the portion of the air pressure that exists by virtue of the degree of compression only. If expressed as gauge pressure, it may be negative or positive (per AMCA 99-0066). Static pressure at a specific plane (Psx) is the arithmetic average of the gauge static pressures as measured at specific points in the traverse of the plane. C.1.5 Pressure-velocity. Velocity pressure is that portion of the air pressure which exists by virtue of the rate of motion only. It is always positive (per AMCA 99-0066). Velocity pressure at a specific plane (Pvx) is the square of the arithmetic average of the square roots of the velocity pressures as measured at specific points in the traverse plane. C.1.6 Pressure-total. Total pressure is the air pressure that exists by virtue of the degree of compression and the rate of motion. It is the algebraic sum of the velocity pressure and the static pressure at a point. Thus if the air is at rest, the total pressure will equal the static pressure (per AMCA 990066). Total pressure at a specific plane (Ptx) is the algebraic sum of the static pressure and the velocity pressure at that plane. kg/m3

C.1.7 Standard air density. A density of 1.2 (0.075 lbm/ft3) corresponding approximately to air at 20°C (68°F), 101.325 kPa (29.92 in. Hg) and 50% relative humidity (per AMCA 99-0066).

106 | Fans and Systems

C.1.8 Temperature. The dry-bulb temperature (td) is the air temperature measured by a dry temperature sensor. Temperatures relating to air density are usually referenced to the fan inlet. The wet-bulb temperature (tw) is the temperature measured by a temperature sensor covered by a water-moistened wick and exposed to air in motion. Readings shall be taken only under conditions that assure an air velocity of 3.6 to 10.2 m/s (700 to 2000 ft/min) over the wet-bulb and only after sufficient time has elapsed for evaporative equilibrium to be attained. Wet bulb depression is the difference between drybulb and wet-bulb temperatures (td - tw) at the same location.

C.2 The fan C.2.1 Blast area. The blast area of a centrifugal fan is the fan outlet area less the projected area of the cutoff; see Figure B.6 (per AMCA 99-0066). C.2.2 Inlet area. The fan inlet area (A1) is the gross inside area of the fan inlet (see Figure 9.14). C.2.3 Outlet area. The fan outlet area (A2) is the gross inside area of the fan outlet. C.2.4 Fan. (1) A device, which utilizes a power-drive rotating impeller for moving air or gases. The internal energy (enthalpy) increase imparted by a fan to a gas does not exceed 25 kJ/kg (10.75 BTU/lbm). (2) A device having a power-driven rotating impeller without a housing for circulating air in a room (per AMCA 99-0066). The volume airflow of a fan (Q) is the rate of airflow in m3/s (cfm) expressed at the fan inlet conditions. C.2.5 Fan impeller diameter. The fan impeller diameter is the maximum diameter measured over the impeller blades. C.2.6 Fan total pressure. Fan total Pressure (Pt) is the difference between the total pressure at the fan outlet and the total pressure at the fan inlet. Pt = Pt1 Pt2 (Algebraic). Ignoring the losses that exist between the planes of measurement and the fan, Figures C.1, C.2 and C.3 illustrate fan total pressures for three basic arrangements for fans connected to external systems.

Where the fan inlet is open to atmospheric air or where an inlet bell, as shown in the Figure C.1 is used to simulate an inlet duct, the total pressure at the fan inlet (Pt1) is considered to be the same as the total pressure in the region near the inlet (Pta) where no energy has been imparted to the air. This is the location of "still air". The following equations apply:

Where the fan outlet is open to atmospheric air or where an outlet duct three diameters or less in length is used to simulate a fan with an outlet duct and the outlet duct is open to atmospheric air, the total pressure at the fan outlet is equal to the fan velocity pressure (Pv). The following equations apply: Pt = Pt2 - Pt1 Pt2 = Pv Pt = Pv - Pt1

Pta = 0 Pt = Pt2 - Pt1 Pt1 = Pta = 0 Pt = Pt2

PLANE 1

PLANE 2

Pt2

Pt = Pt2

Figure C.1 - Fan Total Pressure for Installation Type B: Free Inlet, Ducted Outlet

Fans and Systems | 107

PLANE 1

PLANE 2

Pt1 Pt = Pv2 - Pt1

Figure C.2 - Fan Total Pressure for Installation Type C: Ducted Inlet, Free Outlet

PLANE 1

Pt1

PLANE 2

Pt

Pt = Pt2 - Pt1

Figure C.3 - Fan Total Pressure for Installation Type D: Ducted Inlet, Ducted Outlet

108 | Fans and Systems

Pt2

C.2.7 Fan velocity pressure. Fan velocity pressure (Pv) is the pressure corresponding to the average air velocity at the fan outlet. Pv = Pv2 Assuming no change in air density or area between the plane of measurement and the fan outlet, Figure C.4 illustrates fan velocity pressure. C.2.8 Fan static pressure. The difference between the fan total pressure and the fan velocity pressure. Therefore, fan static pressure is the difference between the static pressure at a fan outlet and the total pressure at a fan inlet (per AMCA 99-0066). Ps = Pt - Pv Ignoring losses between the planes of measurement and the fan, Figure C.5 illustrates the fan static pressure for a fan with ducted inlet and outlet. Ps = Ps2 - Ps1 - Pv1 (Algebraic) Where the fan inlet is open to atmospheric air, (free inlet, ducted outlet), the fan static pressure (Ps) is equal to the static pressure at the fan outlet. Ps = Ps2

PLANE 1

Pv = Pv2

Where the fan outlet is open to atmospheric air (ducted inlet, free outlet), ignoring the SEF, the fan static pressure (Ps) is equal to the inlet static pressure (Ps1) less the inlet velocity pressure (Pv1). Ps = -Ps1 - Pv1 Ps = -(-Ps1) - Pv1 Ps = Ps1 - Pv1

C.3 The system C.3.1 Equivalent duct diameter. The diameter of a circle having the same area as another geometric shape. For a rectangular cross-section duct with width (a) and height (b), the equivalent diameter is: (4ab/π)0.5 (per AMCA 99-0066). C.3.2 Fan performance. Fan performance is a statement of the volume airflow, static or total pressure, speed and power input at a stated inlet density and may include total and static efficiencies. C.3.3 Fan performance curve. Of the many forms of fan performance curves, generally all convey information sufficient to determine fan performance as defined above. In this manual, ‘fan performance curve’ refers to the constant speed performance

PLANE 2

Pv2

Figure C.4 - Fan Velocity Pressure for Installation Type B: Free Inlet, Ducted Outlet

Fans and Systems | 109

curve. This is a graphical representation of static or total pressure and power input over a range of volume airflow at a stated inlet density and fan speed. It may include static or total efficiency curves. The range of volume airflow that is covered generally extends from shutoff (zero airflow) to free delivery (zero fan static pressure). The pressure curves that appear are generally referred to as the pressurevolume curves. C.3.4 Normalized fan curve. A normalized fan curve is a constant speed curve in which the fan performance values appear as percentages, with 100% airflow at free delivery, 100% fan static pressure at shutoff, and 100% power at the maximum power input point. C.3.5 Point of duty. Point of duty is a statement of air volume flow rate and static or total pressure at a stated density and is used to specify the point on the system curve at which a fan is to operate. C.3.6 Point of operation. The relative position on a fan or air curtain performance curve corresponding to a particular airflow, pressure, power and efficiency (per AMCA 99-0066).

PLANE 1

Ps1

C.3.7 Point of rating. The specified fan operating point on its characteristic curve (per AMCA 99-0066). C.3.8 System. A series of ducts, conduits, elbows, branch piping, etc., designed to guide the flow of air, gas or vapor to and from one or more locations. A fan provides the necessary energy to overcome the resistance to flow of the system and causes air or gas to flow through the system. Some components of a typical system are louvers, grills, diffusers, filters, heating and cooling coils, air pollution control devices, burner assemblies, sound attenuators, the ductwork and related fittings. C.3.9 System curve. A graphic representation of the pressure versus volume airflow characteristics of a particular system. C.3.10 System Effect Factor (SEF). A pressure loss, which recognizes the effect of fan inlet restrictions, fan outlet restrictions, or other conditions influencing fan performance when installed in the system (per AMCA 99-0066).

PLANE 2

Pv1

Ps2 Ps = Ps2 - Ps1 - Pv1 (algebraic)

Figure C.5 - Fan Static Pressure for Installation Type D: Ducted Inlet, Ducted Outlet

110 | Fans and Systems

HOUSING

DIVERTER CU

TO

FF

CENTER PLATE BLAST AREA DISCHARGE OUTLET AREA SIDE SHEET BACKPLATE

FF

BLADE

TO

CU

INLET

SCROLL IMPELLER FRAME RIM BEARING SUPPORT INLET COLLAR

Figure C.6 - Terminology for Centrifugal Fan Components

Fans and Systems | 111

CASING

BACKPLATE RIM HUB

MOTOR GUIDE VANE

INLET

BLADE IMPELLER

INLET BELL

Figure C.7A - Tubular Centrifugal FanDirect Drive CASING

BLADE DIFFUSER HUB

MOTOR

IMPELLER CASING

Figure C.7B - Tubeaxial Fan-Direct Drive (Impeller Downstream)

BEARING CASING BELT TUBE BLADE

HUB

GUIDE VANE IMPELLER

Figure C.7C - Vaneaxial Fan-Belt Drive

Figure C.7 - Terminology for Axial and Tubular Centrifugal Fans

112 | Fans and Systems

The Ps required at the fan outlet (C) will be equal to the pressure drop at the desired airflow. Since there are no inlet obstructions and the duct near the fan outlet is the same as used in the test setup, the published fan performance can be used with no additional system effect factors applied.

Annex D. Examples of the Convertibility of Energy from Velocity Pressure to Static Pressure SI CONVERSION was done using 249 Pa = 1 in. wg, 1 m3/s = 2118 cfm, 1m/s = .00508 ft/min

D.1 Example of fan (tested with free inlet, ducted outlet) applied to a duct system The overall friction of the duct system results in a 747 Pa (3.0 in. wg) pressure drop at an airflow of 1.42 m3/s (3000 cfm). SI

I-P

A

Free inlet

0.00 Pa

(no SEF)

0.0 in. wg

B-C

Outlet with straight duct attached for two or more diameters.

0.00 Pa

(no SEF)

0.0 in. wg

(duct design)

3.0 in. wg

C-D

Duct friction at Q = 1.42 m3/s (3000 cfm).

REQUIRED FAN Ps

747.00 Pa 747.00 Pa

3.0 in. wg

Select a fan for Q = 1.42 m3/s (3000 cfm) and Ps = 747 Pa (3.0 in. wg). Use manufacturer's data for rpm (N) and power (H).

NO OBSTRUCTION AT FAN INLET

Pv = 124 Pa (0.5 in.wg)

FRICTION 747 Pa (3.0 in.wg) AT 1.42 m3/s (3000 cfm)

(I-P) in.wg (SI) Pa 996

4

747

3 Pt

498

2

249

1

0

0

Pv

Ps

A

B C

ATMOSPHERIC PRESSURE

124 Pa (0.5 in.wg)

D

Figure D.1 - Pressure Gradients - Fan as Tested Fans and Systems | 113

D.2 Example of fan (tested with free inlet, ducted outlet), connected to a duct system and then a plenum This example includes the same duct system as described in Example C.1. However, there is a short outlet duct on the fan followed by a plenum chamber with cross-sectional area more than 10 times larger than the area of the duct. The velocity in the duct from E to F is 14.4 m/s (2830 fpm), equal to a velocity pressure of 124.5 Pa (0.5 in. wg). At point "F" the Pv is 124.5 Pa (0.5 in. wg), the Ps is 0.0 Pa (0.0 in. wg), and the Pt is 124.5 Pa (0.5 in. wg). The friction of duct will cause a gradual increase in Ps and Pt back to point E. If the duct has a uniform cross-sectional area the Pv will be constant through this part of the system. Since there is an energy loss of 49.8 Pa (0.2 in. wg) as a result of the abrupt contraction from the plenum

to the duct, the Pt requirement in the plenum is 871.15 Pa (3.5 in. wg), Pt at duct entrance = 49.8 Pa (0.2 in. wg) in contraction loss, or 921.3 Pa (3.7 in. wg) Pt. Air flowing across the plenum from D to E will have a relatively low velocity and the Pv in the plenum will be 0.0 Pa (0.0 in. wg) since the velocity is negligible. At point D, there is an abrupt expansion energy loss equal to the entire Pv in the duct discharging into the plenum. The outlet duct between the fan and the plenum is 2.5 equivalent diameters long. It is the same as used during the fan rating test. The Ps in the outlet duct (also the Ps in the plenum) is the same as the Ps as measured during the rating test. This example requires a fan to be selected for 921.30 Pa (3.7 in. wg) at 1.42 m3/s (3000 cfm). Compare this with the previous selection of 747 Pa (3.0 in. wg) Ps at 1.42 m3/s (3000 cfm).

SI C-D

Outlet duct on fan as tested

0.00 Pa

D

Pv loss (also Pt loss) as result of air velocity decrease. Ps does not change from duct to plenum at D.

0.00 Pa

E

E

E-F

Contraction loss - plenum to duct

I-P (no SEF)

0.0 in. wg

0.0 in. wg

49.80 Pa

(part of duct system)

0.2 in. wg

Ps energy required to create velocity at E

124.50 Pa

(part of duct system)

0.5 in. wg

Duct friction at Q = 1.42 m3/s (3000 cfm)

747.00 Pa

3.0 in. wg

921.30 Pa

3.7 in. wg

REQUIRED FAN Ps Solution:

Select a fan for Q = 1.42 m3/s (3000 cfm) and Ps = 921.30 Pa (3.7 in. wg) Use manufacturer's data for rpm (N) and power (H).

114 | Fans and Systems

2.5 DIA.

NEGLIGIBLE LOSS

Pv = 124 Pa (0.5in.wg)

FRICTION 747 Pa (3.0 in.wg) AT 1.42 m3/s (3000 cfm)

(I-P) in.wg (SI) Pa

1046 Pa (3.7 in.wg)

1245

5

996

4

747

3

498

2

922 Pa (3.7 in.wg)

922 Pa (3.7 in.wg)

Pt 747 Pa (3.0 in.wg)

249

1

0

0 A

B C

D

E

Pv 124 Pa (0.5 in.wg)

Ps

ATMOSPHERIC PRESSURE

F

Figure D.2 - Pressure Gradients - Plenum Effect

Fans and Systems | 115

D.3 Example of fan with free inlet, free outlet - fan discharges directly into plenum and then to duct system (abrupt expansion at fan outlet)

energy is lost. In these applications, the energy loss and the System Effect Factor may exceed the fan outlet velocity pressure as defined in terms of "fan outlet area".

This example is similar to the plenum effect example except the duct at the fan outlet has been omitted. The fan discharges directly into the plenum.

The SEF for fans without outlet duct was obtained as follows: GIVEN:

It may seem unreasonable that the System Effect loss at the fan outlet is greater than the defined fan outlet velocity. Fans with cutoffs must generate higher velocities at the cutoff plane (blast area) than in the outlet duct (outlet area). This higher velocity (at cutoff) is partially converted to Ps when outlet ducts are used as on fan tests. When fans with cutoffs are "bulk-headed" into plenums or discharge directly into the atmosphere as with exhausters, all the velocity

B-C

SEF (see above)

B-C

Pv loss (also Pt loss) as result of air velocity decrease. Ps does not change from duct to plenum at C

D

D

D-E

contraction loss - plenum to duct

Fan

Blast Area = 0 .6 Outlet Area

Fan outlet velocity = 14.4 m/s (2830 fpm) No outlet duct System Effect Curve = R-S, (from Figure 8.3) SEF = 149.4 Pa (0.6 in. wg), (from Figure 7.1) at 14.4 m/s (2830 fpm) velocity and system curve R)

SI

I-P

149.40 Pa

0.6 in. wg

0.00 Pa

0.0 in. wg

49.80 Pa

(part of duct system)

0.2 in. wg

Ps energy required to create velocity at D

124.50 Pa

(part of duct system)

0.5 in. wg

duct friction at Q = 1.42 m3/s (3000 cfm)

747.00 Pa

(duct design)

3.0 in. wg

REQUIRED FAN Ps

1070.70 Pa

Solution: Select a fan for 1.42 m3/s (3000 cfm) Q and 1070.70 Pa (4.3 in. wg) Ps. Use manufacturer's data for rpm (N) and power (H).

116 | Fans and Systems

4.3 in. wg

Pv = 124 Pa (0.5 in.wg)

(I-P) in.wg

FRICTION 747 Pa (3.0 in.wg) AT 1.42 m3/s (3000 cfm)

149 Pa (0.6 in.wg) SEF

(SI) Pa

922 Pa (3.7 in.wg)

1245

5

996

4

747

3

498

2

872 Pa (3.5 in.wg)

Pt 747 Pa (3.0 in.wg)

249

1

0

0 A

B C

D

Pv 124 Pa (0.5 in.wg)

Ps

E

ATMOSPHERIC PRESSURE

Figure D.3 - Pressure Gradients - Abrupt Expansion at Fan Outlet

Fans and Systems | 117

Three SEFs are shown in this example:

D.4 Example of fan used to exhaust with obstruction in inlet, inlet elbow, inlet duct, free outlet

1) System Effect Curve R (see Figure 9.5 for a 3 piece inlet elbow with R/D ratio of 1 and no duct between the elbow and the fan inlet).

This example is an exhaust system. Note the entry loss at point A. An inlet bell will reduce this loss.

2) System Effect Curve U (see Figure 9.14 for a bearing in the fan inlet which obstructs 10% of the inlet).

On the suction side of the fan, Ps will be negative, but Pv is always positive.

3) System Effect Curve R (from Figure 8.3 for a fan discharging to atmosphere with no outlet duct).

Fan Pv = 124.5 Pa (0.5 in. wg)

SI A

Entrance loss - sharp edge duct

I-P

99.60 Pa

(duct design)

0.4 in. wg

(duct design)

3.0 in. wg

A-B

Duct friction at 1.42 m3/s (3000 cfm)

747.00 Pa

B

SEF 1

149.40 Pa

0.6 in. wg

C

SEF 2

49.80 Pa

0.2 in. wg

E

Fan Pv

124.50 Pa

0.5 in. wg

E

SEF 3

149.40.Pa

0.6 in. wg

1319.70 Pa

5.3 in. wg

REQUIRED FAN Pt

Fan Ps = fan Pt - fan Pv Fan Ps (SI) = 1319.70 Pa – 124.5 Pa = 1195.2 Pa Fan Ps (I-P) = 5.3 in. wg - 0.5 in. wg = 4.8 in wg Solution: Select a fan for 1.42 m3/s (3000 cfm) Q and 1195.2 Pa (4.8 in. wg) Ps Use manufacturer's data for rpm (N) and power (H).

118 | Fans and Systems

ABRUPT DISCHARGE SEF 149 Pa (0.6 in.wg)

Pv = 124 Pa (0.5 in.wg)

FRICTION 747 Pa (3.0 in.wg) AT 1.42 m3/s (3000 cfm)

(I-P) in.wg ELBOW SEF 149 Pa (0.6 in.wg)

(SI) Pa

OBSTRUCTION SEF 50 Pa (0.2 in.wg)

+249 +1

ATMOSPHERIC PRESSURE 0

0

-249

-1

-498

-2

-747

-3

100 Pa (0.4 in.wg) 149 Pa (0.6 in.wg) REQUIRED

Pt Pv

-847 Pa (-3.4 in.wg)

Ps

-996 Pa (4.0 in.wg)

224 Pa (0.9 in.wg)

-996

-4

-1245 -5 -1171 Pa (4.7 in.wg)

-971 Pa (3.9 in.wg)

A

B

C

D

E

-1121 Pa (4.5 in.wg)

FAN INLET

Figure D.4 - Pressure Gradients - Exhaust System

Fans and Systems | 119

Annex E. References These references contain additional information related to the subject of this manual: 1. ANSI/AMCA 210-99, Laboratory Methods of Testing Fans for Aerodynamic Performance Rating, Air Movement and Control Association International, Inc., 30 West University Drive, Arlington Heights, IL, 60004-1893 U.S.A., 1999. 2. AMCA Publication 200-95, Air Systems, Air Movement and Control Association International, Inc., 30 West University Drive, Arlington Heights, IL, 60004-1893 U.S.A., 1995. 3. AMCA Publication 202-98, Troubleshooting, Air Movement and Control Association International, Inc., 30 West University Drive, Arlington Heights, IL, 60004-1893 U.S.A., 1997. 4. ASHRAE Handbook, HVAC Systems and Equipment, 1996, The American Society of Heating, Refrigerating and Air Conditioning Engineers, Inc., 1791 Tullie Circle N.E., Atlanta, GA, 30329 U.S.A., 1996, (Chapter 18 Fans). 5. Traver, D. G., System Effects on Centrifugal Fan Performance, ASHRAE Symposium Bulletin, Fan Application, Testing and Selection, The American Society of Heating, Refrigerating and Air Conditioning Engineers, Inc., 1791 Tullie Circle N.E., Atlanta, GA, 30329 U.S.A., 1971. 6. Christie, D. H., Fan Performance as Affected By Inlet Conditions, ASHRAE Transactions, Vol. 77, The American Society of Heating, Refrigerating and Air Conditioning Engineers, Inc., 1791 Tullie Circle N.E., Atlanta, GA, 30329 U.S.A., 1971. 7. Zaleski, R. H., System Effect Factors For Axial Flow Fans, AMCA Paper 2011-88, AMCA Engineering Conference, Air Movement and Control Association International, Inc., 30 West University Drive, Arlington Heights, IL, 60004-1893 U.S.A., 1988. 8. Roslyng, O., Installation Effect on Axial Flow Fan Caused Swirl and Non-Uniform Velocity Distribution, Institution of Mechanical Engineers (IMechE), 1 Birdcage Walk, London SW1H 9JJ, England, 1984. 9. Clarke, M. S., Barnhart, J. T., Bubsey, F. J., Neitzel, E., The Effects of System Connections on Fan Performance, ASHRAE RP-139 Report, The American Society of Heating, Refrigerating and Air Conditioning Engineers, Inc., 1791 Tullie Circle N.E., Atlanta, GA, 30329 U.S.A., 1978. 10. Madhaven, S., Wright, T., J. DiRe, Centrifugal Fan Performance With Distorted Inflows, The American Society of Mechanical Engineers, 345 East 47th Street, New, York, NY, 10017 U.S.A., 1983. 11. Cory, W. T. W., Fan System Effects Including Swirl and Yaw, AMCA Paper 1832-84-A5, AMCA Engineering Conference, Air Movement and Control Association International, Inc., 30 West University Drive, Arlington Heights, IL, 60004-1893 U.S.A., 1984. 12. Cory, W. T. W., Fan Performance Testing and Effects of the System, AMCA Paper 1228-82-A5, AMCA Engineering Conference, Air Movement and Control Association International, Inc., 30 West University Drive, Arlington Heights, IL, 60004-1893 U.S.A., 1984. 13. Galbraith, L.E., Discharge Diffuser Effect on Performance Axial Fans, AMCA Paper 1950-86-A6, AMCA Engineering Conference, Air Movement and Control Association International, Inc., 30 West University Drive, Arlington Heights, IL, 60004-1893 U.S.A., 1986. 14. Industrial Ventilation –23rd Edition, American Conference of Governmental Industrial Hygienists, 1330 Kemper Meadow Drive, Cincinnati, OH 45240-1634 U.S.A., 1998. 15. Fans and Systems, John E. Thompson and C. Jack Trickler, The New York Blower Company, Chemical Engineering, March 21, 1983, pp. 48-63 16. AABC National Standards, Chapter 8, Volume Measurements, Associated Air Balance Council, 1518 K Street NW, Suite 503, Washington, DC 20005 U.S.A. 120 | Fans and Systems

Troubleshooting

202

1. Introduction After the installation of an air moving system is completed, a system sometimes fails to achieve its designed performance. This part of the AMCA Fan Application Manual will help you identify what is wrong and decide how to correct it.

2. Procedure for Troubleshooting 2.1 Look in the "Master Troubleshooting Appendices" for a subject which corresponds with the apparent problem. 2.2 Check each of the "Probable Causes" listed. 2.3 If the cause of the trouble is not found proceed through the "System Checklist" (see Section 4). 2.4 If the problem has still not been solved, it is now advisable to contact the representative of the fan manufacturer. He should be given the results of the "System Checklist" and the additional information listed in Section 5.1. 2.5 The fan manufacturer or his representative will analyze the information submitted, as outlined in Section 5.2. With this information and, if necessary, an on-site inspection, he may be able to explain why the system is not achieving its design performance and may recommend changes in the system or the fan installation which will overcome the problem.

3. Safety Precautions Before checking the fan and system it will be necessary to shut down the fan. During inspection the fan must be electrically isolated and all disconnect switches and other controls LOCKED in the "OFF" position. Where these are in locations remote from the fan, prominent DO NOT START signs should also be in place. CAUTION - Even when LOCKED out electrically, fans located outdoors or in a parallel or series fan system may be subject to "wind-milling." Therefore, as an added precaution, the impeller should be secured to physically restrict rotational movement.

4. System Checklist Poor system performance may arise from a number of causes including: • • • • • • •

Improper installation or assembly of the fan Damage in handling or transit System design error Deterioration of the system Faulty controls Poor fan selection A combination of several factors.

A systematic check of the items listed here should help identify the problem - or problems - and allow suitable corrective action to be taken.

SYSTEM CHECKLIST A) While the impeller is coasting to a stop, see if it is rotating in the proper direction (see Figures 4.1, 4.2 and 4.3). B) Make certain the impeller rotation is correct for the housing (guide vanes of vaneaxial and tubular centrifugal fans) and not installed backwards (see Figures 4.1, 4.2, and 4.3). Note: Tubeaxial fan rotation is the same as shown in Figure 4.2 except without the guide vanes.

ROTATION

RADIAL BLADE

BACKWARD INCLINED

AIRFOIL

RADIAL TIP

BACKWARD CURVED

FORWARD CURVED

Figure 4.1 - Types of Centrifugal Fan Impellers

Note: Fan manufacturers describe the rotation of the fan impeller as being "clockwise" or "counterclockwise”. •

For AXIAL fans when viewing the INLET or OUTLET as specified by the Fan Manufacturer

•

For CENTRIFUGAL fans when viewing the DRIVE SIDE (see AMCA Standard 99-2406)

•

For TUBULAR CENTRIFUGAL fans when viewing the OUTLET (see AMCA Standard 99-2410).

122 | Troubleshooting

ROTATION

ROTATION

INLET GUIDE VANES

OUTLET GUIDE VANES

AIRFLOW

AIRFLOW

Figure 4.2 - Vaneaxial Fan Impellers and Guide Vanes

ROTATION

GUIDE

VANES

AIRFLOW

Figure 4.3 - Tubular Centrifugal Impeller

Troubleshooting | 123

C) If the fan is belt driven: 1) Are the drive pulley (Motor Sheave) and the driven pulley (Fan Sheave) in alignment? Improper alignment of the sheaves can cause excessive power consumption (high amperage), squealing belts, shortened belt and sheave life and high axial vibration.

FAN

FAN

MOTOR

IMPROPER SHEAVE ALIGNMENT

MOTOR

PROPER SHEAVE ALIGNMENT

Figure 4.4 - Sheave Alignment 2) Are the belts loose? Loose belts that flap or slip can cause excessive noise and vibration. Slipping belts will also affect fan speed and reduce belt and sheave life. Belts should be tensioned to the belt manufacturer's recommendations. Tension of the drive belts should be adjusted for stretching after the first 48 hours of operation. Caution! Excessive belt tension will reduce fan and motor bearing life. 3) Are the belts and/or sheaves worn? If so, an immediate replacement could save down time at a later date. Where more than one belt is used, replace with a new set of matched belts.

IMPROPER BELT TENSION

PROPER BELT TENSION (SEE BELT MANUFACTURER’S SPECIFICATIONS)

Figure 4.5 - Belt Tension D) Check the flow surfaces (passages between the inlets, impeller blades and inside of housing) for cleanliness. A 2 mm (0.0625 in.) buildup of dirt on the flow surface could impair fan performance and/or cause vibration. E) Are there any gouges, tears, holes, erosion or corrosion in the impeller blades, rims or backplate: inlets and/or housing? If so, report the approximate size and location to the fan manufacturer and discontinue operation until repairs are made. F) Is any foreign matter trapped in the impeller, housing or ductwork (loose insulation, papers, ice, etc.)? If so, remove. 124 | Troubleshooting

G) Are coils, heaters, filters, ducts, etc. dirt laden? If so, clean or replace. Remove any non-essential obstructions to airflow in elbows, shutters, transformations, dampers, bird-screens, etc. Verify that all dampers (control, backdraft, fire, etc.) are adjusted to the proper settings. H) Have all the parts supplied with the fan been installed? I)

Are there any obstructions to airflow near the fan inlets? Objects such as pipes, airflow measurement stations, ductwork, columns, belt guards, belt drives, etc. could adversely affect the output of the fan. For more information, see AMCA Publication 201 Fans and Systems.

J)

Are the fan outlet connections correctly designed and installed? Duct takeoffs, or obstructions in the fan outlet could adversely affect the output of the fan (see AMCA Publication 201).

K) See Figure 4.6A for typical centrifugal fan inlet-impeller relationships and Figure 4.6A for typical axial fan housing-impeller relationships. A few simple measurements as indicated on these figures can tell the manufacturer if a problem exists in this area. Note: several measurements should be taken around the entire inlet/housing circumference to determine the average, maximum and minimum values.

I.S. = INLET SPACING I.G. = INLET GAP

I.G.

R.C. R.C. = RUNNING CLEARANCE (INLET CENTERED ON IMPELLER)

I.S.

I.S.

I.S. I.S. SHOULD BE SAME BOTH SIDES FOR DOUBLE WIDTH FAN

Figure 4.6A - Typical Centrifugal Fan Inlet-Impeller Relationships

R.C.

C.

R.C. = RUNNING CLEARANCE (IMPELLER CENTERED IN HOUSING)

C. = VANE CLEARANCE

Figure 4.6B - Typical Axial Fan Housing-Impeller Relationships Troubleshooting | 125

L)

On a double-width fan, is the approach to both inlets identical? Airflow should be symmetrical about the centerline of the fan housing (see Figure 4.7A). Non-symmetrical airflow can lead to decreased air performance. Belt drives, belt guards and motors can cause non-symmetrical airflow to the inlets if too severely restricted (see AMCA Publication 201).

L

L

P

AIRFLOW

P

Figure 4.7A - Symmetrical Inlet Airflow

L

L NOT EQUAL TO M N NOT EQUAL TO R

M

N

AIRFLOW

R

Figure 4.7B - Non-Symmetrical Inlet Airflow

Figure 4.8A - Turning Vanes in Elbow Adjacent to Centrifugal Fan

M) Are turning vanes installed in elbows that are too close to the fan inlet or discharge (see Figures 4.8A and 4.8B). Published pressure losses through elbows are based on a uniform velocity profile. Turning vanes help achieve this uniform flow (see AMCA Publication 201).

126 | Troubleshooting

Figure 4.8B - Turning Vanes in Elbow Adjacent to Axial Fan

N) If the fan is equipped with an inlet vane damper, check the operation as follows: 1)

Do not rely on the control arm position alone for locating the position of the inlet vane damper blades without first checking visually to see that the inlet vane damper position agrees with the position of the control arm.

2)

If the unit is a double-width fan equipped with inlet vanes or inlet vane damper control, both inlet vane dampers must be synchronized (the inlet vane dampers must be in the same relative position with respect to the impeller on both inlets). If the inlet vane dampers are not synchronized, there will be unbalanced airflow between inlets resulting in deficient air performance, unbalanced thrust on the bearings and/or a surge condition in the fan.

3)

Make certain that inlet vane dampers are of the proper rotation with respect to the impeller. As the vanes close they should cause the entering air to spin in the same direction as the impeller.

4) Are the inlet vane dampers correctly positioned for the design operating conditions? If not, the desired pressure-volume of the fan will not be realized (see Figure 4.9).

100% Open

Percent of No Delivery Pressure

100

80

75% Open 50% Open

60 25% Open 40

20

0

20

40

60

80

100

Percent of Free Delivery Volume

Figure 4.9 - Typical Pressure-Volume Curve for Operation with Inlet Vane Damper Control

Troubleshooting | 127

O) Inspect any ductwork or plenums approaching the fan inlets for the possibility of inducing swirl of air into the inlet. Pre-swirl of air entering the fan inlet can reduce the fan performance (see Figures 4.10A and 4.10B). (See AMCA Publication 201.) IMPELLER ROTATION

AIR SWIRL IN SAME DIRECTION AS IMPELLER ROTATION

Figure 4.10A - Pre-Rotation

IMPELLER ROTATION

AIR SWIRL IN OPPOSITE DIRECTION TO IMPELLER ROTATION

Figure 4.10B - Counter-Rotation

P) After completing the above steps and making sure the fan and system are safe to start, remove all DO NOT START signs on disconnect switches and override systems and put the unit back into operation. Q) Inspect the entire system including the fan, fan plenum and all ductwork for significant air leaks. Air leaks may be detected by sound, smoke, feel, soapy solution, etc. Some common air leak sources are: access doors, coils, duct seams, fan outlet connection, etc. Significant air leaks must be sealed.

5. Fan Manufacturer's Analysis If the cause of the trouble has still not been found after completing the "System Checklist," the fan manufacturer should be consulted. The fan manufacturer will review the information provided concerning the system and recommend an appropriate course of action.

5.1 Data required for analysis To make a complete analysis of the problem, in addition to the results of the "System Checklist," the manufacturer will need: A)

Complete plans (drawings) including all ductwork, location, size, model and manufacturer of all fans, motors, coils, dampers, etc. with all pertinent dimensions for the complete system as actually installed.

B)

If the fan/air handling system fails to achieve the design performance, the measured performance and the design performance figures should be supplied.

C)

A copy of the system design calculations.

D)

A copy of the specifications and any addendums.

128 | Troubleshooting

E)

If a separate air performance test has been conducted on the installed fan, a statement of measured fan performance along with a copy of the test data, the type of test and instrumentation, and the measurement location of the airflow rate and pressure determinations should be supplied. A statement of fan performance should contain: 1) Fan total pressure (Pt) rise or fan static pressure (Ps) 2) Airflow rate (Q) 3) Power (H) 4) Fan speed (N) 5) Air density (ρ)

5.2 Probable manufacturer action Among other actions, the fan manufacturer will: A)

Assess the probable accuracy of the field performance measurements (see AMCA Publication 203 Field Performance Measurements of Fan Systems).

B)

Examine the system drawings (plans) for any System Effect losses (see AMCA Publication 201) which were not allowed for in the system design calculations or the original fan selection.

C)

Reassess the fan performance, accounting for System Effect losses established in step B to the designed fan/system performance (see Figure 5.1).

D)

Check whether the fan selection is correct for the application.

SYSTEM EFFECT

DUCT SYSTEM CURVE DESIGN FAN OPERATING POINT

DESIGN PRESSURE

FAN CATALOG PRESSURE-VOLUME CURVE FAN OPERATING POINT W/SYSTEM EFFECT THEORETICAL PRESSURE-VOLUME CURVE ACCOUNTING FOR SYSTEM EFFECT LOSSES (FIELD PERFORMANCE MEASUREMENTS TO BE COMPARED AGAINST THIS CURVE).

DESIGN VOLUME

Figure 5.1

Troubleshooting | 129

The information obtained through the checklists in this manual should help in defining the necessary corrective action. In most cases, if the troubleshooting procedure has been followed carefully and impartially it will be apparent whether the system has been built and installed in accordance with the design drawings, whether the fan was properly selected and suitable allowances made for the appropriate System Effect Factors, or whether the fan is not performing up to its published ratings.

6. Conclusion By intelligent application of the procedures outlined in this manual it should be possible to find the cause of a performance problem in any air moving system. Identification of a problem associated directly with the fan may require the assistance of the fan manufacturer. Recognition of the cause of the trouble will usually be a major step toward curing it. Corrective measures may include alterations to the system, modifications to the fan outlet or inlet connections, adjustments to the fan, etc. In many cases an increase in the fan speed may be decided upon but it is extremely important that the fan shall not be operated above its cataloged maximum speed or the maximum speed recommended by the manufacturer. Excessive speed may result in catastrophic impeller failure. If a speed increase proves to be an acceptable alternative, then the motor should also be checked for its capacity to handle the increased fan power.

130 | Troubleshooting

MASTER TROUBLESHOOTING APPENDICES Annex A. Noise SOURCE

PROBABLE CAUSE

A-1

IMPELLER HITTING INLET OR HOUSING . . . . . . . . . . . . . a. b. c. d. e. f. g. h.

Impeller not centered in inlet or housing. Inlet or housing damage. Crooked or damaged impeller. Shaft loose in bearing. Impeller loose on shaft. Bearing loose in bearing support. Bent shaft. Misaligned shaft and bearings.

A-2

IMPELLER HITTING CUTOFF . . . . . . . . . . . . . . . . . . . . . . . a. b. c. d.

Cutoff not secure in housing. Cutoff damaged. Cutoff improperly positioned. Impeller improperly positioned.

A-3

DRIVE . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a. Sheave not tight on shaft (motor or fan). b. Belts hitting belt tube or belt guard. c. Belts too loose. Adjust for belt stretching after 48 hours of operation. d. Belts too tight. e. Belts wrong cross-section. f. Belts not "Matched" in length on multi-belt drive. g. Variable pitch sheaves not adjusted so each groove has same pitch diameter (multi-belt drive). h. Misaligned sheaves. i. Belts worn. j. Motor, motor base or fan not securely anchored. k. Belts oily or dirty. l. Improper drive selection. m. Loose key.

A-4

COUPLING . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a. Coupling unbalanced, misaligned, loose or may need lubricant. b. Loose key.

A-5

BEARING . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a. b. c. d. e. f. g. h. i.

A-6

SHAFT SEAL SQUEAL . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a. b. c. d.

Defective bearing. Needs lubrication. Loose on bearing support. Loose on shaft. Seals misaligned. Foreign material inside bearing. Worn bearing. Fretting corrosion between inner race and shaft. Bearing not sitting on flat surface. Needs lubrication. Misaligned. Bent shaft. Bearing loose on support. Troubleshooting | 131

A-7

IMPELLER . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a. Loose on shaft. b. Defective impeller. DO NOT OPERATE FAN. LOCK OUT THE FAN ELECTRICALLY AND CONTACT THE MANUFACTURER. c. Unbalance. d. Coating loose. e. Worn as result of abrasive or corrosive material moving through airflow passages. f. Blades rotating close to structural member. g. The number of blades may coincide with an equal number of structural members. h. Loose key.

A-8

HOUSING . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a. Foreign material in housing. b. Cutoff or other housing part loose (rattling during operation).

A-9

MOTOR . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a. b. c. d. e. f. g. h.

A-10

SHAFT . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a. Bent b. Undersized. May cause noise at impeller, bearings or sheave.

A-11

HIGH AIR VELOCITY . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a. Ductwork too small for application. b. The installed fan may be too small for the application. c. Registers or grilles too small for application. d. Heating or cooling coil with insufficient face area for application. OBSTRUCTION IN HIGH VELOCITY AIR STREAM MAY CAUSE RATTLE, OR PURE TONE WHISTLE . . . . . . . . . . . . . . . . . . . . . . . . . . . . a. Dampers. b. Registers. c. Grilles. d. Sharp elbows. e. Sudden expansion in ductwork. f. Sudden contraction in ductwork. g. Turning vanes.

A-12

A-13

Lead-in cable not secure. AC hum in motor or relay. Starting relay chatter. Noisy motor bearings. Single phasing a three phase motor. Low voltage. Cooling fan striking shroud. Poor motor/inverter match, more noticeable at low speeds.

PULSATION OR SURGE . . . . . . . . . . . . . . . . . . . . . . . . . . . a. Restricted system causes fan to operate at poor point of rating. b. Fan too large for application. c. Ducts vibrate at same frequency as fan pulsations. d. Rotating stall.

132 | Troubleshooting

e. Inlet vortex surge. f. Distorted inlet airflow. A-14

AIR LEAKS AND/OR OBSTRUCTED FLOW . . . . . . . . . . . . a. Air leaks in ductwork. 1) Bad joint connections; 2) Holes or tears; 3) Obstructions inside duct. b. Fins on coils. c. Registers or grilles.

A-15

RATTLES AND/OR RUMBLES . . . . . . . . . . . . . . . . . . . . . . . a. Vibrating ductwork. b. Vibrating cabinet parts. c. Vibrating parts not isolated from building.

Troubleshooting | 133

Annex B. Insufficient Air Flow SOURCE B-1

PROBABLE CAUSE

FAN . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a. b. c. d. e. f.

g. h. i. j. B-2

DUCT SYSTEM . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a. b. c. d. e.

Impeller installed backwards. Impeller running backwards. Impeller blade angle setting. Cutoff missing or improperly installed. Impeller not centered with inlet collar(s). Fan rpm below design: 1) Incorrect sheave or sheave setting; 2) Incorrect motor rpm; 3) Low voltage to motor; 4) Speed controller set too low; CAUTION! DO NOT INCREASE FAN SPEED BEYOND THE FAN MANUFACTURER'S RECOMMENDATIONS. ALSO, WHEN INCREASING FAN SPEED, MONITOR MOTOR AMPS SO AS NOT TO EXCEED MOTOR NAMEPLATE AMPS. Impeller/inlet dirty or clogged. Improper running clearance. Improper inlet cone to impeller fit. Improperly set inlet vane or damper. Actual system is more restrictive (more resistance to airflow) than expected. Dampers closed. Registers closed. Air leaks in supply ducts. Insulating duct liner loose.

B-3

FILTERS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a. b.

Dirty or clogged. Replacement filter with greater than specified pressure drop.

B-4

COILS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a. b.

Dirty or clogged. Incorrect fin spacing.

B-5

RECIRCULATION . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .a.

Internal cabinet air leaks in bulkhead separating fan outlet (pressure zone) from inlet(s) (suction zone). Air leaks around fan outlet at connection through cabinet bulk-head. Elbows, cabinet walls or other obstructions restrict airflow. Inlet obstructions cause more restrictive systems but do not cause increased negative pressure readings near the fan inlet(s) (see System Effects in AMCA Publication 201). Fan speed may be increased to counteract the effect of restricted fan inlet(s). CAUTION! DO NOT INCREASE FAN SPEED BEYOND THE FAN MANUFACTURER'S RECOMMENDATIONS. ALSO, WHEN INCREASING FAN

b. B-6

OBSTRUCTED FAN INLETS . . . . . . . . . . . . . . . . . . . . . . . .a.

134 | Troubleshooting

SPEED, MONITOR MOTOR AMPS SO AS NOT TO EXCEED MOTOR NAMEPLATE AMPS. B-7

NO STRAIGHT DUCT AT FAN OUTLET . . . . . . . . . . . . . . . a.

Fans which are normally used in duct system are tested with a length of straight duct at the fan outlet. If there is no straight duct at the fan outlet, decreased performance may result. If it is not practical to install a straight section of duct at the fan outlet, the fan speed may need to be increased to overcome this pressure loss (see System Effects in AMCA Publication 201). CAUTION! DO NOT INCREASE FAN SPEED BEYOND THE FAN MANUFACTURER'S RECOMMENDATIONS. ALSO, WHEN INCREASING FAN SPEED, MONITOR MOTOR AMPS SO AS NOT TO EXCEED MOTOR NAMEPLATE AMPS.

B-8

OBSTRUCTION IN HIGH VELOCITY AIR STREAM . . . . . . a. b. c. d.

Obstruction near fan outlet or inlet(s). Sharp elbows near fan outlet or inlet(s). Improperly designed turning vanes. Projections, dampers or other obstruction in a part of the system where air velocity is high.

Troubleshooting | 135

Annex C. Airflow High - (Too Much Airflow) SOURCE

PROBABLE CAUSE

C-1

SYSTEM . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a. b. c. d. e. f.

Oversized ductwork. Access door open. Registers or grilles not installed. Dampers set to by-pass coils. Filter(s) not in place. System resistance low.

C-2

FAN . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a. b.

Fan speed too fast. Improper impeller blade angle.

136 | Troubleshooting

Annex D. Static Pressure Wrong SOURCE D-1

PROBABLE CAUSE

SYSTEM, FAN OR INTERPRETATION OF MEASUREMENT . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . GENERAL DISCUSSION: The velocity pressure at any point of measurement is a function of the velocity of the air and its density. The static pressure at a point of measurement in the system is a function of the system design (resistance to airflow), air density and the amount of air flowing through the system. The static pressure measured in a "loose" or oversized system will be less than the static pressure in a "tight" or undersized system for the same airflow rate. In most systems, pressure measurements are indicators of how the installation is operating. These measurements are the result of airflow and as such are useful indicators in defining system characteristics. Field static pressure measurements rarely correspond with laboratory static pressure measurements unless the fan inlet(s) and fan outlet conditions of the installation are exactly the same as the inlet(s) and outlet conditions in the laboratory. Also see D-2 through D-6, E-2, F-1 and G-1 for specific cases. Static Pressure Low, Airflow Correct

D-2

GAS DENSITY . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Pressures will be less with high temperature gases or at high altitudes. Static Pressure Low, Airflow High

D-3

SYSTEM . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . System has less resistance to airflow than expected. This is a common occurrence. Fan speed may be reduced to obtain desired airflow rate. This will reduce power consumption (operating cost). Static Pressure Low, Airflow Low

D-4

SYSTEM . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a.

Fan inlet(s) and/or outlet conditions not same as tested. See System Effect Factors, in AMCA Publication 201.

Troubleshooting | 137

D-5

FAN . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a. b. c. d. e. f. g. h. i. j.

Impeller installed backwards. Impeller running backwards. Improper impeller blade angle. Cutoff missing or improperly installed. Impeller not centered with inlet collar(s). Fan speed too slow. Impeller/inlet dirty or clogged. Improper running clearance. Improper inlet cone to impeller fit. Improperly set inlet vanes or damper.

Static Pressure High, Airflow Low D-6

DUCT SYSTEM . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a. b. c. d.

Actual system is more restrictive (more resistance to airflow) than designed. Dampers closed. Registers closed. Insulating duct liner loose.

D-7

FILTERS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a. b.

Dirty or clogged. Replacement filter with greater than specified pressure drop.

D-8

COILS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a. b.

Dirty or clogged. Fin spacing too close.

138 | Troubleshooting

Annex E. Power High SOURCE E-1

PROBABLE CAUSE

FAN . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a. b. c.

d. E-2

SYSTEM . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a. b.

c. d.

Backward Inclined impeller installed backwards. Fan speed too high. Forward Curve or Radial Blade impeller operating below design pressures. Incorrect impeller blade angle. Oversized ductwork. Face and bypass dampers oriented so coil dampers are open at same time bypass dampers are open. Filter(s) left out. Access door open.

Note: The causes listed under E-2 pertain primarily to radial blade, radial tip and forward curve centrifugal fans, i.e., fans that exhibit rising power curves. Normally, backward inclined, backward curve or backward inclined airfoil centrifugal fans and axial flow fans do not fall into this category. E-3

GAS DENSITY . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .a.

Calculated power requirements based on light gas (e.g. high temperature) but actual gas is heavy (e.g. cold start-up).

E-4

FAN SELECTION . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .a.

Fan not selected at efficient point of rating.

Troubleshooting | 139

Annex F. Fan Does Not Operate SOURCE F-1

PROBABLE CAUSE

ELECTRICAL OR MECHANICAL . . . . . . . . . . . . . . . . . . . . . Mechanical and electrical problems are usually straight-forward and are normally analyzed in a routine manner by service personnel. In this category are such items: a. b. c. d. e. f. g. h. i. j.

140 | Troubleshooting

Blown fuses. Broken belts. Loose pulleys. Electricity turned off. Impeller touching housing. Wrong voltage. Motor too small and overload protector has broken circuit. Low voltage, excessive line drop or inadequate wire size. Load inertia too large for motor. Seized bearing.

Annex G. Premature Failure SOURCE

PROBABLE CAUSE

G-1

BELTS, BEARINGS, SHEAVES IMPELLERS HUBS, ETC.GENERAL DISCUSSION . . . . . . . . . . . . . . . . . Each fan component is designed to operate satisfactorily for a reasonable lifetime. Fans intended for heavy duty service are made especially for that type of service. For example, Class I fans are intended for operation below certain limits of pressure and outlet velocity. Class II fans are designed for higher operating limits (see AMCA Standard 99-2408). Not all components are limited by the same factors, e.g. limiting factors may be power, RPM, temperature, impeller tip speed, torque, corrosive atmospheres, expected life, etc. Also see A-2, A-5, A-6, A-7.

G-2

COUPLINGS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a. b.

Coupling unbalanced, misaligned, loose or may need lubricant. Loose key.

G-3

SHAFT . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a. b.

Bent. Undersized.

G-4

BEARINGS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a. b. c. d. e. f.

Over-lubrication or under- lubrication. Locking collar or set screw loose. Excessive belt tension. False brinelling. Wrong lubricant. Undersized shaft.

G-5

DRIVES . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a. b. c. d. e.

Over-tensioned or under-tensioned belts. Mismatched belts on multi-belt drive. Misalignment of motor and fan sheaves. Excessive start-stop cycles. Set screw on sheave loose.

Troubleshooting | 141

Annex H. Vibration AERODYNAMIC VIBRATION (PULSATION) SOURCE H-1

PROBABLE CAUSE

AIRSTREAM . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a. b.

c. d. e. f.

Operating fan left of peak in unstable (stall) region. Poor inlet conditions which generate air turbulence. 1. Partially obstructed inlet(s). 2. Sharp elbow at fan inlet(s) and/or outlet. System pulsation which is transmitted to the fan. Blade Pass frequency -- number of blades x fan RPM. Guide vane frequency -- number of vanes x fan RPM. Fan support frequency -- motor or bearing supports, belt tube, on axial fans.

MECHANICAL VIBRATION H-2

UNBALANCED IMPELLERS . . . . . . . . . . . . . . . . . . . . . . . . a. b. c. d. e. f. g.

h. i. j. k. l.

H-3

DRIVE OR COUPLING . . . . . . . . . . . . . . . . . . . . . . . . . . . . a.

b. c. d. e. f. g. H-4

LOOSE FASTENERS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a. b. c. d. e. f.

142 | Troubleshooting

Material build-up on impeller. Impeller components subject to abrasion, corrosion or impact. Moisture inside hollow airfoil blades. Lost balance weight. Excessive temperature. Impeller (blades) not tracking evenly. Eccentricity: 1. Bore off center. 2. Bore on angle (see H-2.f) Improper key-to-keyway length. Impeller rubbing stationary equipment. Shaft seal rub. Inverter drives. Motor torque pulses (on single phase motors). Alignment: 1. Improper assembly. 2. Shift during handling or shipment. Worn, loose, or mismatched belts. Eccentric sheaves or couplings. Belt resonance. Worn or chipped sheaves. Improper key-to-keyway length. Unbalanced sheave(s) or coupling. Impeller set screws. Bearing set screws. Drive component set screws. Fan mounting bolts. Bearing bolts. Motor bolts.

H-5

FAN SUPPORT STRUCTURE . . . . . . . . . . . . . . . . . . . . . . . a. b. c. d. e. f. g. h.

Support structure natural frequencies. Insufficient mass or stiffness. Large unsupported mass. Lack of lateral support, particularly with fans mounted on isolators. Fan skewed (bolted down to uneven surface). Broken support members. Short-circuited or improperly adjusted isolators. Fan (mounted on isolators) with rigid inlet(s) and outlet connections.

H-6

SHAFT . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a. b.

Bent. Undersized.

H-7

BEARINGS . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a. b. c.

Misalignment. Worn out; loose. Too much, too little, or incorrect lubricant.

H-8

BACKGROUND EXCITATION (FLANKING TRANSMISSION) . . . . . . . . . . . . . . . . . . . . . . .a. b. c. d.

Interconnected piping. Heavy machinery transmitting vibration through foundation. Fan(s) mounted on floating roofs. Wind-loading of fan mounted on isolators.

ELECTRICALLY INDUCED VIBRATION H-9

ELECTRIC MOTOR . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . a. b. c. d. e. f.

Loose stator laminations. Broken rotor bar. Stator problems. Phase unbalance. Eccentric rotor. Capacitor-start motor causes "shudder" during coast-down.

Troubleshooting | 143

Field Performance Measurements of Fan Systems

203

1. Introduction

2. Scope

Performance ratings of fans are developed from laboratory tests made according to specified procedures on standardized test setups. In North America, the standard is ANSI/AMCA Standard 210 / ANSI/ASHRAE 51 Laboratory Methods of Testing Fans for Rating.

The recommendations and examples in this publication may be applied to all types of centrifugal, axial, and mixed flow fans in ducted or nonducted installations used for heating, ventilating, air conditioning, mechanical draft, industrial process, exhaust, conveying, drying, air cleaning, dust collection, etc. Although the word air is used when reference is made in the general sense to the medium being handled by the fan, gases other than air are included in the scope of this publication.

In actual systems in the field, very few fans are installed in conditions reproducing those specified in the laboratory standard. This means that, in assessing the performance of the installed fansystem, consideration must be given to the effect on the fan’s performance of the system connections, including elbows, obstructions in the path of the airflow, sudden changes of area, etc. The effects of system conditions on fan performance is discussed in Section 5, and more completely in AMCA Publication 201, Fans and Systems. A major problem of testing in the field is the difficulty of finding suitable locations for making accurate measurements of flow rate and pressure. Sections 9.3 and 10.3 outline the requirements of suitable measurement sections. Because these problems and others will require special consideration on each installation, it is not practical to write one standard procedure for the measurement of the performance of all fan-systems in the field. This publication offers guidelines to making performance measurements in the field which are practical and flexible enough to be applied to a wide range of fan and system combinations. Because of the wide variety of fan types and systems encountered in the field, Annex A includes examples of a number of different field tests. In most cases, these examples are based on actual tests which have been conducted in the field. Before performing any field test, it is strongly recommended that the following AMCA publications be carefully reviewed: AMCA Publication 200 - Air Systems AMCA Publication 201 - Fans and Systems AMCA Publication 202 - Troubleshooting AMCA Standard 210 - Laboratory Methods of Testing Fans for Rating

Measurement of sound, vibration, and stress levels are not within the scope of this publication.

3. Types of Field Tests There are three general categories of field tests: A) General Fan System Evaluation A measurement of the fan-system’s performance to use as the basis of modification or adjustment of the system. B) Acceptance Test -A test specified in the sales agreement to verify that the fan is achieving the specified performance. C) Proof of Performance Test -A test in response to a complaint to demonstrate that the fan is meeting the specified performance requirement. As acceptance and proof of performance tests are related to contract provisions, they are usually subject to more stringent requirements and are usually more costly than a general evaluation test. In the case of large fans used in industrial applications and of mechanical draft fans used in the electrical power generation industry the performance of a field test may be part of the purchase agreement between the fan manufacturer and the customer. In addition to Publication 203, AMCA Standard 803 Site Performance Test Standard-Power Plant and Industrial Fans defines the conditions which must be met to achieve higher accuracy of measurement. In new installations of this type, it is desirable to include a suitable measuring section in the design. Agreement must be reached on the test method to be used prior to performance of the test.

4. Alternatives to Field Tests In some cases, considerations such as cost and problems of making accurate measurements may make the following alternative methods of testing worth investigation: A) Testing the fan before installation in a laboratory equipped to perform tests in accordance with AMCA Standard 210. Limitations in laboratory test facilities may preclude tests on full size fans. In this case, the full size fan can be tested at the installation site in accordance with AMCA Standard 210. This will usually require the installation of special ductwork. B) Testing a reduced scale model of the fan in accordance with AMCA Standard 210 and determining the performance of the full size fan as described in AMCA Publication 802, Power Plant Fans – Establishing Performance Using Laboratory Methods. C) Testing a reduced scale model of the complete fan and system using the test methods outlined in this publication. Tests conducted in accordance with AMCA Standard 210 will verify the performance characteristics of the fan but will not take into account the effect of the system connections on the fan’s performance (see Section 5).

5. System Effect Factors AMCA Publication 201, Fans and Systems, deals in detail with the effect of system connections on fan performance. It gives system effect factors for a wide variety of obstructions and configurations which may affect a fan’s performance. System Effect Factor (SEF) is a pressure loss which recognizes the effect of fan inlet restrictions, fan outlet restrictions, or other conditions influencing fan performance when installed in the system. SYSTEM EFFECT FACTORS (SEFs) ARE INTENDED TO BE USED IN CONJUNCTION WITH THE SYSTEM RESISTANCE CHARACTERISTICS IN THE FAN SELECTION PROCESS. Where SEFs are not applied in the fan selection process, SEFs must be applied in the calculations of the results of field tests. This is done for the purpose of allowing direct comparison of the test results to the design static pressure calculation. Thus, for a field test, the fan static pressure is defined as: Ps = Ps2 - Ps1 – Pv1 + SEF 1 + SEF 2 + + SEF n 146 | Field Performance Measurement

Examples of the application of SEFs in determining the results of field tests are included in Annex A. In field tests of fan-system installations in which system effects have not been accounted for, it is important that their sources be recognized and their magnitudes be established prior to testing. The alternative to dealing with a large magnitude SEF is to eliminate its source. This requires revisions to the system. This alternative course of action is recommended when swirl exists at the fan inlet (see Publication 201, Figure 9.8). The effect on fan performance as a result of swirl at the inlet is impossible to estimate accurately as the system effect is dependent upon the degree of swirl. The effect can range from a minor amount to an amount that results in the fan-system performance being completely unacceptable.

6. Fan Performance Fan performance is a statement of fan flow rate, fan total or static pressures, and fan power input at stated fan speed and fan air density. Fan total or static efficiencies may be included. The fan air density is the density at the fan inlet. The fan flow rate is the volume flow rate at the fan inlet density.

7. Referenced Planes Certain locations within a fan-system installation are significant to field tests. These locations are designated as follows: Plane 1: Plane of fan inlet Plane 2: Plane of fan outlet Plane 3: Plane of Pitot-static tube traverse for purposes of determining flow rate Plane 4: Plane of static pressure measurement upstream of fan Plane 5: Plane of static pressure measurement downstream of fan The use of the numerical designations as subscripts indicate that the values pertain to those locations.

8. Symbols and Subscripts SYMBOL A D De FLA H HL Hmo kW L N NLA NPH NPV Ps Psx Pt Ptx Pv Pvx pb pe pp px Q Qi Qx SEF T td tw V ∆Px,x’ ∆Ps ρ ρx Σ

UNIT

9.1 General

Area of cross-section Diameter Equivalent diameter Full load amps Fan power input Power transmission loss Motor power output Electrical power Length Speed of rotation No load amps Nameplated horsepower Nameplated volts Fan static pressure Static pressure at Plane x Fan total pressure Total pressure at Plane x Fan velocity pressure Velocity pressure at Plane x Barometric pressure Saturated vapor pressure at tw Partial vapor pressure Absolute pressure at Plane x Fan flow rate Interpolated flow rate Flow rate at Plane x System effect factor Torque Dry-bulb temperature Wet-bulb temperature Velocity Pressure loss between Planes x and x’ Pressure loss across damper Fan gas density Gas density at Plane x Summation sign

ft2 ft ft amps hp hp hp kilowatts ft rpm amps hp volts in. wg in. wg in. wg in. wg in. wg in. wg in. Hg in. Hg in. Hg in. Hg cfm cfm cfm in. wg lb-in. °F °F fpm

Determine fan flow rate using the area, velocity pressure, and density at the traverse plane and the density at the fan inlet. The velocity pressure at the traverse plane is the root mean square of the velocity pressure measurements made in a traverse of the plane. The flow rate at the traverse plane is calculated by converting the velocity pressure to its equivalent velocity and multiplying by the area of the traverse plane.

Airflow direction

---

SUBSCRIPT c r x 1 2 3 4 5

DESCRIPTION

9. Fan Flow Rate

in. wg in. wg lbm/ft3 lbm/ft3 ---

DESCRIPTION

Value converted to specified conditions Reading Plane 1, 2, 3, ..., as appropriate Plane 1 (fan inlet) Plane 2 (fan outlet) Plane 3 (plane of Pitot-static traverse for purpose of determining flow rate Plane 4 (plane of static pressure measurement upstream of fan) Plane 5 (plane of static pressure measurement downstream of fan)

9.2 Velocity measuring instruments Use a Pitot-static tube of the proportions shown in Annex B or a double reverse tube, shown in Annex C, and an inclined manometer to measure velocity pressure. The velocity pressure at a point in a gas stream is numerically equal to the total pressure diminished by the static pressure. The Pitot-static tube is connected to the inclined manometer as shown in Annex F. The double reverse tube is connected to the inclined manometer as shown in Annex C. 9.2.1 Pitot-static tube. The Pitot-static tube is considered to be a primary instrument and need not be calibrated if maintained in the specified condition. It is suited for use in relatively clean gases. It may be used in gases that contain moderate levels of particulate matter such as dust, water, or dirt, provided certain precautions are employed (see Section 15). 9.2.2 Double reverse tube. The double reverse tube is used when the amount of particulate matter in the gas stream impairs the function of the Pitot-static tube. The double reverse tube requires calibration. It is important that the double reverse tube be used in the same orientation as used during calibration. Mark the double reverse tube to indicate the direction of the gas flow used in its calibration. 9.2.3 Inclined manometers. Inclined manometers are available in both fixed and adjustable range types. Both types require calibration. The adjustable range type is convenient in that it may be adjusted at the test site to the range appropriate to the velocity pressures which are to be measured. It is adjusted by changing the slope to any of the various fixed settings and by changing the range scale accordingly. Each setting provides a different ratio of the length of the indicating column to its indicated height. Adjustable range type manometers in which the slope may be fixed at 1:1, 20:1, and intermediate ratios are available (see Figure 10 in Annex G). Field Performance Measurement | 147

The accuracy of the manometer used in the measurement of velocity pressures is of prime importance. Select a manometer that will provide an acceptable degree of accuracy; consider the range, slope, quality, scale graduations, indicating fluid of the instrument and the range of the velocity pressures to be measured. The graph in Annex G indicates the effect of expected resolution of manometer readings on the accuracy of velocity determinations. The basis for this graph is described in Section 9.6. Determine velocities in the very low range more accurately by using a manometer with a slope of 20:1. Due to practical limitations in length, its use is restricted to measurements where the velocities are very low. Also, errors in velocity determinations made by using a Pitot-static tube and manometer exceed normally acceptable values at velocity pressure readings less than 0.023 in. wg. This corresponds to a velocity of approximately 600 fpm for air of 0.075 lbm/ft3 density. 9.2.4 Low velocity instruments. Normally, velocities encountered in the field test situations are well in excess of 600 fpm. Therefore, recommendations regarding alternate test procedures and instrumentation for use for velocities less than 600 fpm are not presented in this publication. Descriptions of various types of instruments used to determine range velocities are presented in Annex J. Most of the instruments require frequent calibration, and some are not suited for use in high temperature, dirty, wet, corrosive, or explosive atmospheres. If it is necessary to use one of these instruments, the procedure for its use, its calibration, and the expected accuracy of results should be agreed upon by all interested parties.

9.3 Location of traverse plane For field tests, suitable test measurement station locations must be provided in the system. When suitable locations are not available, consider making temporary or permanent alterations to the ducting for improved test accuracy. For free inlet, free outlet fans, convert a free inlet, free outlet fan to a ducted inlet, free outlet fan by the addition of a temporary duct. Estimate free inlet, free outlet fan flow rate by measuring other parameters and interpreting certified ratings performance (see Section 17.1).

than 75% of the velocity pressure measurements are greater than 1/10 of the maximum measurement (see Figure 9.1) 2) The flow streams should be at right angles to the traverse plane. Variations from this flow condition as a result of swirl or other mass turbulence are considered acceptable when the angle between the flow stream and the traverse plane is within 10 degrees of a right angle. The angle of the flow stream in any specific location is indicated by the orientation of the nose of the Pitot-static tube that produces the maximum velocity pressure reading at the location. 3) The cross-sectional shape of the airway in which the traverse plane is located should not be irregular. Proper distribution of traverse points and accurate determination of the area of the traverse plane are difficult to achieve when the airway does not conform closely to a regular shape. 4) The cross-sectional shape and area of the airway should be uniform throughout the length of the airway in the vicinity of the traverse plane. When the divergence or convergence of the airway is irregular or more than moderate in degree, significantly nonuniform flow conditions may exist. 5) The traverse plane should be located to minimize the effects of gas leaks between the traverse plane and the fan. 6) When it is necessary to locate the traverse plane in a converging or diverging airway (not recommended), note that the traverse plane and area is located at the tip of the Pitot-static tube. A location well downstream in a long, straight run of uniform cross-section duct will usually provide acceptable conditions for the Pitot traverse plane. When locating the traverse plane close to the fan, as is often done in order to minimize the effect of leakage, flow conditions upstream of the fan are usually more suitable. In some installations, more than one traverse plane may be required in order to account for the total flow (Annex A contains examples).

A Pitot traverse plane suitable for the measurements used to determine flow rate are as follows:

When a field test is anticipated, particularly when the requirement for a field test is an item in the specifications, the system designer should provide a suitable traverse plane location in the system.

1) The velocity distribution should be uniform throughout the traverse plane. The uniformity of distribution is considered acceptable when more

When the fan is ducted outlet and the traverse plane is to be located downstream from the fan, the

148 | Field Performance Measurement

Pv MAX

Pv MAX 10

A: IDEAL Pv DISTRIBUTION

Pv MAX

Pv MAX 10

B: GOOD Pv DISTRIBUTION (ALSO SATISFACTORY FOR FLOW INTO FAN INLETS. MAY BE UNSATISFACTORY FOR FLOW INTO INLET BOXES - MAY PRODUCE SWIRL IN BOXES)

Pv MAX

Pv MAX 10

Pv MAX

Pv MAX 10

60% 80%

C: SATISFACTORY Pv DISTRIBUTION - MORE THAN 75% OF Pv READINGS GREATER THAN: Pv MAX 10 (UNSATISFACTORY FOR FLOW INTO FAN INLETS OR INLET BOXES)

Pv MAX 10

D: DO NOT USE UNSATISFACTORY Pv DISTRIBUTION LESS THAN 75% OF Pv READINGS GREATER THAN: Pv MAX 10 (UNSATISFACTORY FOR FLOW INTO FAN INLETS OR INLET BOXES)

Pv MAX

Pv MAX 10

Pv MAX

40%

35%

20%

35%

E: DO NOT USE UNSATISFACTORY Pv DISTRIBUTION LESS THAN 75% OF Pv READINGS GREATER THAN: P MAX v

10 (UNSATISFACTORY FOR FLOW INTO FAN INLETS OR INLET BOXES)

F: DO NOT USE UNSATISFACTORY Pv DISTRIBUTION LESS THAN 75% OF Pv READINGS GREATER THAN: P MAX v

10 (UNSATISFACTORY FOR FLOW INTO FAN INLETS OR INLET BOXES)

Figure 9.1 - Typical Velocity Pressure Distributions Encountered in Velocity Pressure Measurement Planes in Fan-System Installations Field Performance Measurement | 149

MEASUREMENT PLANE De MIN. 2 12 in. MIN.

Y

WHERE: De =

4YZ π

INLET BOX DAMPERS

Z

Note: The measurement plane should be located a minimum of ½ De from the inlet cone, but not less than 12 in. from the leaving edge of the damper blades. Figure 9.2

STACK

VELOCITY PROFILE

Note: Spiral vortex may form when fan discharges directly into a stack or similar arrangement. Figure 9.3 150 | Field Performance Measurement

traverse plane should be situated a sufficient distance downstream from the fan to allow the flow to diffuse to a more uniform velocity distribution and to allow the conversion of velocity pressure to static pressure. Annex P provides guidance for the location of the traverse plane in these cases. The location of the traverse plane on the inlet side of the fan should not be less than ½ equivalent diameter from the fan inlet. Regions immediately downstream from elbows, obstructions and abrupt changes in airway area are not suitable traverse plane locations. Regions where unacceptable levels of swirl are usually present, such as the region downstream from an axial flow fan that is not equipped with straightening vanes, should be avoided. Swirl may form when a fan discharges directly into a stack or similar arrangement (see Figure 9.2). 9.3.1 Inlet box location. When the traverse plane must be located within an inlet box, the plane should be located a minimum of 12 inches downstream from the leaving edges of the damper blades and not less than ½ equivalent diameter upstream from the edge of the inlet cone (see Figure 9.3). Do not locate traverse points in the wake of individual damper blades. In the case of double inlet fans, traverses must be conducted in both inlet boxes in order to determine the total flow rate. 9.3.2 Alternative locations. On occasion, an undesirable traverse plane location is unavoidable, or each of a limited number of prospective locations lacks one or more desirable qualities. In such cases, the alternatives are: 1) Accept the most suitable location and evaluate the effects of the undesirable aspects of the location on the accuracy of the test results. In some instances, the estimated accuracy may indicate that the results of the test would be meaningless, particularly in acceptance tests and proof of performance tests. 2) Provide a suitable location by modifying the system. This course of action is recommended for acceptance tests and proof of performance tests. The modifications may be temporary, permanent, minor or extensive, depending on the specific conditions encountered. When the inlet side of the fan is not ducted but is designed to accept a duct, consider installing a short length of inlet duct to provide a suitable traverse plane location. This duct should be of a size and shape to fit the fan inlet, a minimum of 2 equivalent diameters long and equipped with a bell shaped or flared fitting at its inlet. The traverse plane should be located a minimum of ½ equivalent diameters from the fan inlet and not less than 1½

equivalent diameters from the inlet of the duct. Where the duct is small, its length may necessarily be greater than 2 equivalent diameters in order to ensure that the tip of the Pitot-static tube is a minimum of 1½ equivalent diameters from the duct inlet. This short length of duct should produce no significant addition to the system resistance, but in some cases it may alter the pattern of flow into the fan impeller, and thereby affect the performance of the fan slightly.

9.4 The traverse Annex H contains recommendations for the number and distribution of measurement points in the traverse plane. If the flow conditions at the traverse plane are less than satisfactory, increase the number of measurement points in the traverse to improve accuracy. Since the flow at a traverse plane is never strictly steady, the velocity pressure measurements indicated by the manometer will fluctuate. Each velocity pressure measurement should be mentally averaged on a time-weighted basis. Any velocity pressure measurement that appears as a negative reading is to be considered a velocity pressure measurement of zero and included as such in the calculation of the average velocity pressure. When it is necessary to locate the traverse plane in a converging or diverging airway, orient the nose of the Pitot-static tube such that it coincides with the anticipated line of the flow stream. This is particularly important at measurement points near the walls of the airway (see Annex A-1A). No appreciable effect on Pitot-static tube readings occur until the angle of misalignment between the airflow and the tube exceeds 10 degrees.

9.5 Flow rate calculations 9.5.1 Flow rate at traverse plane. The flow rate at the traverse plane is calculated as follows: Q3 = V3A3 Where: A3 = the area of the traverse plane V3 = the average velocity at the traverse plane = 1096 (Pv3/ρ3)0.5 ρ3 = the density at the traverse plane Pv3 = the root mean square velocity pressure at the traverse plane = [∑(Pv3r)0.5 / number of readings]2 Field Performance Measurement | 151

Pv3r is the velocity pressure reading, corrected for manometer calibration and where applicable, corrected for the calibration of the double reverse tube. It is important that the calibration of the double reverse tube be applied correctly. The use of the calibration of the double reverse tube is described in Annex C. 9.5.2 Continuity of mass. The calculations of fan flow rate are based on considerations of continuity of mass, and as such, it is assumed that no mass is added or removed from the gas stream between the traverse plane and the fan inlet. In the general application, having determined the flow rate and density at the traverse plane, the flow rate at any location, (x), in the fan-system installation may be calculated, providing the density at this location is known and the assumption noted above is valid, i.e.: Qx = Q3 (ρ3/ρx) 9.5.3 Fan flow rate, single traverse plane. Where a single traverse plane is used, the calculation of the fan flow rate is: Q = Q1 = Q3 (ρ3/ρ1) Where: Q3 and ρ3 = as described in Section 9.5.1

ρ1 = the density at the fan inlet 9.5.4 Fan flow rate, multiple traverse planes. When it is necessary to use more than one traverse plane in order to account for the total flow: Q = Q1 = Q3a (ρ3a/ρ1) + Q3b (ρ3b/ρ1) + ... + Q3n (ρ3n/ρ1)

9.6 Accuracy The performance item of major concern in most fansystem installations is the flow rate. Every effort should be made to improve the accuracy of the flow rate determination. The uncertainty analysis presented in Annex T indicates that the uncertainties in flow rate determinations will range from 2% to 10%. This range is based on considerations of the conditions that are encountered in most field test situations. This includes instances in which the conditions at the Pitot traverse plane do not conform to all of the qualifications indicated in Section 9.3. The graph in Annex G provides guidance for improving the accuracy of the flow rate 152 | Field Performance Measurement

determinations. This graph indicates the effect of expected resolution of velocity determinations. This effect is shown for several manometer slope ratios. For all ratios, the expected resolution used as a basis for the graph is the length of indicating column equivalent to 0.05 in. wg in a manometer with slope ratio of 1:1. As indicated in the graph, reading resolution uncertainty can be significant. However, this uncertainty can be controlled by selecting a manometer with a slope suited to the velocity pressures to be measured and by avoiding regions of very low velocity in the selection of the traverse plane location. Reading resolution uncertainties exceed normally acceptable values at velocity pressures less than 0.023 in. wg. This corresponds to a velocity of approximately 600 fpm for air of 0.075 lbm/ft3 density. Generally, ducts are sized for velocities considerably in excess of 600 fpm. Velocities less than 600 fpm may exist in certain sections of the system in some installations, but these sections can usually be avoided. Do no use a Pitot-static tube and manometer to determine velocities in the low ranges associated with filters and cooling coils in air conditioning, heating, and ventilating units. In some instances, the uncertainties incurred in the determinations of low velocity flows may be acceptable. For example, an uncertainty of 15% in the determination of the flow rate in a branch duct that accounts for 20% of the total flow rate for the system affects the accuracy of the total flow rate determination by only 3%. In addition to low range velocities, other conditions may exist at the traverse plane which can significantly affect the accuracy of the flow rate determination. These include nonuniform velocity distribution, swirl, and other mass turbulence. Improve the accuracy of the flow rate determination by avoiding these conditions in the selection of the traverse plane location, or improve the conditions by modifying the system.

10. Fan Static Pressure 10.1 General Determine fan static pressure by using the static pressures at the fan inlet and outlet, the velocity pressure at the fan inlet, and applicable System Effect Factors. The use of System Effect Factors in the determination of fan static pressure is described in Section 5. The velocity pressure at the fan inlet is the calculated average velocity pressure at this location, and as such, its determination is based on the fan flow rate, the density at the fan inlet, and the fan inlet area. The static pressures at the fan inlet and outlet may be obtained directly by making pressure measurements at these locations; or they may be

determined by making pressure measurements at other locations, upstream and downstream of the fan. In the latter case, the determinations must account for the effects of velocity pressure conversions and pressure losses, as may occur between the measurement planes and the planes of interest.

10.2 Pressure measuring instruments This section describes only the instruments for use in measuring static pressure. Instruments for use in the other measurements involved in the determination of fan static pressure are described in Section 13. Use a Pitot-static tube of the proportions shown in Annex B, a double reverse tube as shown in Annex C, or a side wall pressure tap as shown in Annex E, and a manometer to measure static pressure. 10.2.1 Pitot-static tube. The comments that appear in Section 9.2 regarding the use and calibration of the Pitot-static tube are applicable to its use in the measurement of static pressures. 10.2.2 Double reverse tube. The double reverse tube cannot be used to measure static pressure directly. It must be connected to two manometers and the static pressure for each point of measurement must be calculated. Both the manometer connections and the method of calculation are shown in Annex C. 10.2.3 Pressure tap. The pressure tap does not require calibration. Use no fewer than four taps located 90 degrees apart. In rectangular ducts, a pressure tap should be installed near the center of each wall. It is important that the inner surfaces of the duct in the vicinities of the pressure taps be smooth and free from irregularities, and that the velocity of the gas stream does not influence the pressure measurements. 10.2.4 Manometers. A manometer with either vertical or inclined indicating column may be used to measure static pressure. Inclined manometers used to measure static pressures require calibration and should be selected for the quality, range, slope, scale graduations, and indicating fluid necessary to minimize reading resolution errors.

10.3 Static pressure measurements It is important that all static pressure measurements be referred to the same atmospheric pressure, and this atmospheric pressure be that for which the barometric pressure is determined. Make static pressure measurements near the fan inlet and the fan outlet, and where the airway

between the measurement plane and the plane of interest is straight and without change in crosssectional area. Then the duct friction loss between the measurement plane and the plane of interest is usually insignificant, and considerations of velocity pressure conversions and calculations of pressure losses for duct fitting and other system components can be avoided. When a system component is situated between the measurement plane and the plane of interest, the pressure loss of the component must be calculated and credited to the fan. The calculation of the pressure loss is usually based on the component’s performance ratings, which may be obtained from the manufacturer of the item. If there is a change in area between the measurement plane and the plane of interest, then the calculation of the static pressure at the plane of interest must account for velocity pressure conversion and include any associated pressure loss. When the change in area is moderate and gradual, the conversion of velocity pressure is considered to occur without loss and the static pressure is calculated on the basis of no change in total pressure between the measurement plane and the plane of interest. This assumes that the duct friction loss between the two planes is negligible. When the change in area is an abrupt and sizable enlargement, as in a duct leading into a large plenum, the loss is considered to be equivalent to the velocity pressure in the smaller area, and the static pressure at the plane of interest is considered to be the same as the static pressure at the measurement plane. This assumes that the velocity pressure in the larger area and the duct friction loss are negligible. 10.3.1 Location of the measuring plane. When the fan is ducted outlet, the static pressure measurement plane downstream of the fan should be situated a sufficient distance from the fan outlet to allow the flow to diffuse to a more uniform velocity distribution and to allow the conversion of velocity pressure to static pressure. See Annex P for guidance in locating the measurement plane in these cases. In general, pressure taps should be used if it is necessary to measure static pressure in the immediate vicinity of the fan outlet. The static pressure at this location is difficult to measure accurately with a Pitot-static tube due to the existence of turbulence and localized high velocities. If the surface conditions or the velocities at the duct walls are unsuited for the use of pressure taps, then a Pitot-static tube must be used with extreme care, particularly in aligning the nose of the tube with the lines of the flow streams. The location of the static pressure measurement Field Performance Measurement | 153

plane upstream of the fan should not be less than ½ equivalent diameter from the fan inlet. In the event that static pressure measurements must be made in an inlet box, the measurement plane should be located as indicated in Figure 9.2. In the case of double inlet fans, static pressure measurements must be made in both inlet boxes in order to determine the average static pressure on the inlet side of the fan. In general, the qualifications for a plane well suited for the measurement of static pressure are the same as those for the measurement of velocity pressure, as indicated in Section 9.3:

negative. By definition, positive values are those measured as being greater than atmospheric pressures; negative values are those measured as being less than atmospheric pressure. In all of the equations in this publication, the values of static pressures must be entered with their proper signs and combined algebraically. 10.4.1 Static pressure at measuring planes. The static pressure at a plane of measurement (x) is calculated as follows:

Psx = 1) The velocity distribution should be uniform throughout the traverse plane.

∑P

sxr

number of readings

Where: 2) The flow streams should be at right angles to the plane. 3) The cross-sectional shape of the airway in which the plane is located should not be irregular. 4) The cross-sectional shape and area of the airway should be uniform throughout the length of the airway in the vicinity of the plane. 5) The plane should be located such as to minimize the effects of leaks in the portion of the system that is located between the plane and the fan. A long, straight run of duct upstream of the measurement plane will usually provide acceptable conditions at the plane. Regions immediately downstream from elbows, obstructions, and abrupt changes in airway area are generally unsuitable locations. Regions where unacceptable levels of turbulence are present should be avoided. If in any fan-system installation the prospective locations for static pressure measurement planes lack one or more desirable qualities, the alternatives are to accept the best qualified locations and evaluate the effects of the undesirable aspects of the conditions on the accuracy of the test results or provide suitable locations by modifying the system. 10.3.2 When using a Pitot-static tube or a double reverse tube to measure static pressure, a number of measurements must be made throughout the plane. Use Annex H to determine the number and distribution of the measurement points. When using pressure taps, a single measurement at each of the taps located at the plane is sufficient.

Psxr = the static pressure reading, corrected for manometer calibration 10.4.2 Static pressure at fan inlet or outlet. The static pressure at the fan inlet, Ps1, and the static pressure at the fan outlet, Ps2, may be measured directly in some cases. In most cases, the static pressure measurements for use in determining fan static pressure will not be made directly at the fan inlet and outlet, but at locations a relatively short distance upstream from the fan inlet and downstream from the fan outlet. These static pressure measurements are designated Ps4 and Ps5, respectively. Static pressure at the fan inlet, Ps1, is derived as follows: Pt4 = Pt1 + ∆P4,1 Where: Pt4 = the total pressure plane of measurement Pt1 = the total pressure at the fan inlet ∆P4,1 = the sum of the pressure losses between the two planes These losses (∆P) include those attributable to duct friction, duct fittings, other system components, and changes in airway area. Although ∆P represents a loss in all cases, it is considered a positive value as used in the equations in this publication. By substitution and rearrangement: Ps1 = Ps4 + Pv4 - Pv1 - ∆P4,1 Similarly, for static pressure at the fan outlet, Ps2:

10.4 Static pressure calculations

Pt2 = Pt5 + ∆P2,5

Static pressure measurements may be positive or

Ps2 = Ps5 + Pv5 - Pv2 + ∆P2,5

154 | Field Performance Measurement

Where:

10.5 Accuracy

The velocity pressures at the various planes can be determined from the following general equations for the velocity pressure at a plane of measurement (x):

The uncertainty analyses in Annex T indicate that the uncertainties in fan static pressure determinations are expected range from 2% to 8%. This range is based on considerations of the conditions expected to be encountered in most field test situations.

Pvx = Pv3 (A3/Ax)2 (ρ3/ρx) Or: Pvx = (Qx/1096Ax)2 ρx Locate the static pressure measurement planes such that the pressure losses between the measurement planes and the planes of interest are insignificant. This will eliminate the uncertainties involved in the determination of the pressure losses, and the equations for Ps1 and Ps2 reduce to the following: Ps1 = Ps4 + Pv4 - Pv1 Ps2 = Ps5 + Pv5 - Pv2 These equations may be used when changes in area between the measurement planes and the planes of interest are moderate and gradual, and the pressure losses associated with conversions of velocity pressure to static pressure are negligible.

Improve the accuracy of the fan static pressure determination by avoiding static pressure measurement plane locations where turbulence or other unsteady flow conditions will produce significant uncertainties in the mental averaging of pressure readings. Other reading resolution uncertainties are not as significant in the fan static pressure determination as in the determination of flow rate. Generally, static pressure measurements are much greater in magnitude than velocity pressure measurements, and the selection of a manometer that will provide reasonably good accuracy is not usually a problem. The uncertainty analyses in Annex T and the resulting anticipated uncertainty range do not account for uncertainties that may occur in the following: •

Determinations of velocity pressure conversions occurring between the measurement planes and the planes of the fan inlet or fan outlet. The area and density values that are involved in these determinations are usually obtained without significant uncertainties. However, pressure losses associated with velocity pressure conversions are often difficult to determine accurately.

•

Determinations of other pressure losses occurring between the measurement planes and the fan inlet or fan outlet. This includes pressure losses in ducts, duct fittings, and other system components. The calculations of these losses are based on the assumption of uniform flow conditions. This assumption may not be valid, and the calculated pressure loss values may be significantly inaccurate.

•

Determinations of the values of System Effect Factors. These determinations are based on limited information, and as such, are subject to uncertainty.

If, in addition to the losses being negligible there are no changes in the areas between the measurement planes and the respective planes of interest, then the equations are further reduced to: Ps1 = Ps4 Ps2 = Ps5 These equations may also be used when the only losses between the measurement planes and the planes of interest are those associated with changes in area that are abrupt and sizable enlargements in the direction of flow. This assumes that the velocity pressure in the larger area is negligible. 10.4.3 Fan static pressure. The equation for fan static pressure is: Ps = Ps2 - Ps1 - Pv1 + SEF 1 + SEF 2 + ... + SEF n Where: SEF 1, SEF 2, ... SEF n = System Effect Factors that account for the various System Effects that are uncorrected and exist at the time of the field test.

Avoid situations requiring these determinations, thereby eliminating them as sources for uncertainties. The uncertainties involved in determining the values of System Effect Factors can be avoided only by correcting the causes of the System Effects. This requires alterations to the system. Field Performance Measurement | 155

11. Fan Power Input 11.1 General Fan power input data included as part of the fan performance ratings are normally defined and limited to either: •

power input to the fan shaft

•

the total of the power input to the fan shaft and the power transmission loss

The losses in fan shaft bearings are included in either case. Since the results of field tests are usually compared to the rated performance characteristics of the fan, field test values of fan power input should be determined on the same basis as that used in the fan ratings. For belt driven fans, the rated fan power input may or may not include belt drive losses. The information regarding the basis of the rated fan power input accompanies the rating data or is otherwise available from the fan manufacturer. In most instances, when a power transmission loss occurs, the loss will have to be determined and subtracted from the motor output in order to obtain the fan power input.

11.2 Power measurement methods In view of the fact that accuracy requirements for field test determinations of fan power input vary considerably, a number of test methods are recommended. These methods are intended to provide economical and practical alternatives for dealing with various levels of accuracy requirements. 11.2.1 Phase current method. This method for estimating the power output of three phase motors is based on the relationship of motor current and motor power output. The method, described in Annex K, requires measurements of the phase currents and voltages supplied to the motor while driving the fan. Depending on the operating load point of the motor, it may also involve the measurements of the no load phase currents. The phase current method is convenient and sufficiently accurate for most field tests. In this method, the closer the actual phase current is to the motor nameplate value of full load amps, the greater the accuracy. Since fan motors are normally selected for operation at or near the full load point, this method provides a reasonably accurate estimate of the power output of the fan motor. Determine fan power input by using the motor power output and, where applicable, the power transmission loss. 156 | Field Performance Measurement

11.2.2 Typical motor performance data. Typical motor performance data may be used to determine fan power input. These data, which are referred to as typical in that the data and the actual performance of the motor are expected to correspond closely, can usually be obtained from the motor manufacturer. The data provided can be in a variety of forms, but are sufficient to determine motor power output based on electrical input measurements. It is important that the power supplied to the motor during the field test be consistent with that used as the basis for the motor performance data. The phase voltage should be stable and balanced, and the average should be withing 2% of the voltage indicated in the performance data. Depending on the form of the typical motor performance data, motor power output is determined by one of the following methods: 1) Given the typical motor performance chart of watts input versus motor power output at a stated voltage. Hmo, is the value in the typical motor performance data that corresponds to the field test measurement of watts input to the motor. 2) Given the typical motor performance chart of watts input versus torque output and speed at a stated voltage. Use the field test measurement of watts input and the corresponding typical motor performance data values of torque output and speed; the motor power output is calculated as: Hmo =

T ×N 63025

3) Given the typical motor performance chart of watts input versus motor efficiency at a stated voltage. Use the field test measurement of watts input and the corresponding typical motor performance data value of motor efficiency, the motor power output is calculated as: Hmo =

watts input × motor efficiency 746

4) Given the typical motor performance chart of amps versus power factor and motor efficiency at a stated voltage. Use the field test measurements of amps input and volts, and the typical motor performance data values of power factor (pf) and motor efficiency, corresponding to the measured amps input; the motor power output is calculated as:

Hmo =

amps × volts × pf × motor efficiency 746

Or, for three phase motors: Hmo =

(3)0.5 × amps × volts × pf × motor efficiency 746

In both equations, amps and volts are the field test measurement values and, in the case of three phase motors, are the averages of the measured phase values. The fan power input is the motor power output minus the power transmission loss, where applicable. 11.2.3 Calibrated motors. A calibrated motor may be used to determine fan power input. When intending to use this method, it is usually necessary to specify in the motor purchase arrangements that the motor be calibrated since an additional cost is normally involved. Calibration data are similar to typical motor performance data with the exception that, instead of being merely typical, the calibration data represent the performance of a specific motor, based on a test of the motor. The motor is calibrated over a range of operation. Electrical input data and other data sufficient for the determination of power output are obtained in the calibration. The calibration normally provides data for operation at nameplate voltage, but may include data for operation at voltages 10% greater and 10% less than nameplate voltage. It is important that the power supplied to the motor during the field test be consistent with that used in its calibration. The phase voltage should stable and balanced, and the average should be within 2% of the voltage at which the motor was calibrated. The field test measurements and the calculations involved in the determination of motor power output are the same as those described in Section 11.2.2 for use with typical motor performance data. The fan power input is the motor power output minus the power transmission loss, where applicable. A calibrated motor provides accurate data to determine motor power output. However, the cost of the calibration is a limiting factor in the use of this method in field tests. For low horsepower applications, the fan manufacturer may be able to calibrate a motor. 11.2.4. Torquemeters. Another method to determine fan power input involves the use of a torquemeter installed between the fan and the driver. The use of a torquemeter requires some prearrangement with the purchaser, who would normally have specified such equipment, so that site conditions can be altered to

accommodate its installation. The torquemeter is extremely limited in field test application. This is due mainly to is high cost and the cost of its installation. In addition, the length of the shut down time and the revisions to site conditions required for its installation are usually undesirable. For practical considerations, it is not normally used in cases where the fan is belt driven and where the fan impeller is installed directly on the motor shaft.

11.3 Power measuring instruments Measurement of current, voltage, watts, and power factor can be obtained by using an industrial type power analyzer of good quality. This type of instrument is available with accuracies of 1% full scale for volts, amps and power factor, and 2% full scale for watts. Normally, the higher levels of accuracy requirements can be met by using this type of instrument, providing the measurements are well up on the scales. In many cases, accuracy level requirements will permit the use of a clip-on type ammeter-voltmeter. Clip-on instruments with accuracies of 3% full scale are available.

11.4 Power transmission losses Several types of power transmission equipment are used in driving fans. Those in which power transmission losses should be considered in the determination of fan power input include belt drives, gear boxes, fluid drives, and electromechanical couplings. Information as to whether the fan power input ratings include power transmission losses is included in the published performance ratings or is otherwise available from the fan manufacturer. It is important that this be established and that the fan power input be determined accordingly in order to provide a valid comparison of field test results to the fan performance ratings. In most cases, fan power input ratings do not include power transmission losses. 11.4.1 Estimating belt drive losses. In view of the lack of published information available for use in calculating belt drive losses, a graph is included in Annex L for this purpose. As indicated in the graph, belt drive loss, expressed as a percentage of motor power output, decreases with increasing motor power output and increases with increasing speed. This graph is based on the results of over 400 drive loss tests provided to AMCA by drive manufacturers. The graph serves as a reasonable guide in evaluating belt drive losses. The calculation of belt drive loss, using this graph, is included in many of the examples in Annex A. Field Performance Measurement | 157

11.4.2 Estimating other transmission losses. For other types of power transmission equipment, consult the fan manufacturer to establish whether transmission losses are included in the fan ratings, and if so, request the magnitudes of the losses allowed in the ratings. Otherwise, it will be necessary to consult the manufacturer of the power transmission equipment for the information regarding transmission losses.

11.5 Accuracy The uncertainty analyses presented in Annex T indicate that the uncertainties in fan power input determinations are expected to range from 4% to 8%. This range is based on considerations of the conditions encountered in most field test situations, estimated accuracies of the various test methods presented in this publication and allowances for uncertainties in the determinations of power transmission losses.

12. Fan Speed 12.1 Speed measuring instruments Measure speed with a revolution counter and chronometer, a stroboscopic tachometer, an electronic counter-timer, or any other precision type tachometer which has a demonstrated accuracy of 0.5% of the measured value. Friction driven and magnetic type pickups should not be used in low fan power ranges where they can influence speed and fan power input measurements.

12.2 Speed measurements Establish the speed by averaging a minimum of three measurements made during the test determination period. The variation in the measurements should not exceed 1% for any single point of operation.

13. Densities 13.1 Locations of density determinations Determine the densities of the gas stream for Plane 1, the fan inlet; and for Plane 3, the velocity pressure measurement plane. In addition, the density at Plane 2, the fan outlet, must be determined whenever the fan total pressure, the fan velocity pressure, or an SEF at the outlet side of the fan is required.

13.2 Data required at each location The pressure and temperature of the gas stream must be obtained for each plane at which a density 158 | Field Performance Measurement

determination is required. The pressures at Planes 1 and 2 are based on the static pressure measurements made for the purpose of determining the fan static pressure. The pressure at Plane 3 is obtained by averaging static pressure measurements made concurrent with the velocity pressure measurements made in a traverse of Plane 3. The absolute pressure at a plane is calculated by using the static pressure at the plane and the barometric pressure. For this reason, it is important that the barometric pressure be determined for the atmosphere to which static pressure measurements are referred. The temperatures used in density determinations are measured at the planes of interest.

13.3 Additional data Additional data required in the determination of density depends on the gas stream as indicated below: 1) For air, the wet-bulb temperature is required unless it is otherwise known that the air is saturated with water vapor or that the water vapor content of the air is insignificant. It should be noted that incorrect assumptions as to whether the air is dry or saturated can result in substantial errors in density determinations. 2) For gases other than air, the normal procedure is to rely on process personnel for the data necessary to determine the density of the gas. The information provided will include density or data sufficient to calculate the density, which should be for stated conditions of temperature and pressure.

13.4 Density values Gas stream density can be established when the pressure, temperature, and additional data, as indicated in Section 13.3, have been obtained. Procedures for establishing density are described in the examples in Annex M and are further illustrated in the field test examples in Annex A. Although the pressure and temperature of the gas stream must be obtained for each plane at which a density value is required, it is usually necessary to obtain additional data, such as the wet-bulb temperature, for only one plane in order to establish the densities at all planes. The densities at the planes for which the additional data is not obtained can be calculated, providing the gas stream does not change composition or undergo a change in phase between planes. The calculation is based on density being directly proportional to absolute pressure and

inversely proportional to absolute temperature. 13.4.1 Example calculation - ρ3 from ρ1. Use Figure N.1 of Annex N to establish the density of air at Plane 1 based on the test determinations of barometric pressure, pb, and the following Plane 1 values: Ps1, static pressure, in. wg td1, dry-bulb temperature, °F tw1, wet-bulb temperature, °F The following data are obtained for Plane 3: Ps3, static pressure, in. wg td3, dry-bulb temperature, °F Calculate the density at Plane 3 as follows: ⎛ P + 13.6 pb ⎞ ⎛ t d1 + 460 ⎞ ρ3 = ρ1 ⎜ s3 ⎟ ⎟⎜ ⎝ 13.6 p1 ⎠ ⎝ t d3 + 460 ⎠

thermometer should be accurate within 5°F of the measured value and readable to 5°F or finer. The temperature determination should be representative of the average temperature of the gas stream throughout the plane of interest. When the temperature varies with time or temperature stratification exists at the measurement plane, several temperature measurements may be necessary in order to obtain a representative average. At elevated temperatures, the thermometer may have to be shielded to prevent radiation effects from exposed heat sources. Locate the wet-bulb thermometer downstream from the dry-bulb thermometer in order to prevent the drybulb temperature measurement from being adversely affected. The wet-bulb thermometer wick should be clean, closely fitted, and wetted with fresh water. The velocity of the air over the wick should be between 700 and 2000 fpm. Use a sling psychrometer to obtain dry and wet-bulb air temperature measurements at the fan inlet for free inlet fans.

Where:

13.6 Barometric pressure p1 = the absolute pressure, in. Hg at Plane 1, calculated as follows: p1 = pb + (Ps1/13.6) In this manner, ρ3 can be calculated without having to measure the wet-bulb temperature at Plane 3. These equations can be used for gases other than air and can be adapted for use in calculations involving any two planes, subject to the limitations noted earlier. In the example calculation of ρ3, pb is determined for the atmosphere to which the measurements of Ps1 and Ps3 are referred. Refer static pressure measurements to a common atmosphere. When the pressures cannot be referred to a common atmosphere, the absolute pressure for each plane is calculated by using the static pressure measurement at the plane and the barometric pressure for the atmosphere to which the static pressure measurement is referred. However, for the purposes of accuracy, static pressure measurements that are used in the determination of fan static pressure must be referred to a common atmosphere.

13.5 Temperatures Measure temperatures with mercury-in-glass, dial, or thermocouple type thermometers. For temperatures through 220°F, the thermometer should be accurate within 2°F of the measured value and readable to 1°F or finer. For temperatures above 220°F, the

Use a portable aneroid barometer for field test determinations of barometric pressure when an acceptable site barometer is not available. The barometer should be accurate within 0.05 in. Hg of the measured value. Determine the test value of barometric pressure by averaging measurements made at the beginning and end of the test period. When the test value of barometric pressure is to be based on data obtained from a nearby airport, it is important that the data include the barometric pressure for the airport site and the elevation for which the pressure was determined (often the barometric pressure is corrected to sea level). This pressure value must then be corrected to the test site elevation. Barometric pressure decreases approximately 0.1 in. Hg for every 100 ft increase in elevation

13.7 Accuracy As indicated in Annex T, uncertainties in density determinations are expected to be less than 3%. However, care must be exercised in obtaining representative test measurements in order to prevent the uncertainties from exceeding this value.

14. Conversion Calculations Generally, the test fan will be operating at a speed and inlet density that are somewhat different from the Field Performance Measurement | 159

fan performance rating values of fan speed and inlet density. In order to provide a common basis for comparing the field test results to the fan performance ratings, each of these two items must be the same in both sets of data. This can be accomplished by converting the results of the field test to the speed and density conditions of the fan performance ratings. The equations for the conversion are as follows. Qc = Q (Nc / N) Psc = Ps (Nc / N)2 (ρc / ρ) Ptc = Pt (Nc / N)2 (ρc / ρ) Pvc = Pv (Nc / N)2 (ρc / ρ) Hc = H (Nc / N)3 (ρc / ρ) Where the subscript c designates values converted to specified conditions, and items without the subscript c are field test values. These conversion equations do not account for the effect of the compressibility of the gas stream. However, since the test fan usually operates at conditions of speed and inlet density that are reasonably close to the quoted fan performance, the conversion calculations usually result in small changes from field test values and the effect of the compressibility of the gas stream is considered to be negligible. Where test conditions are considerably different than design conditions, the effect of compressibility may need to be considered.

Work required to measurements (drilling installation of static thermometer wells, etc.) prior to the test date.

accommodate test of traverse holes, pressure taps and should be completed

4) System Effect Factors, if any, must be established prior to the conduct of the test. 5) The expected test uncertainties must be agreed upon prior to the test (see Annex T). 6) Responsibility for the cost of the test or any fansystem modifications required as a result of the test should be established. 7) Prior to testing, an inspection must be made to ensure that the fan is installed in accordance with the fan manufacturer’s recommendations. The duct system should also be inspected for compliance with design specifications, conditions of filters, abnormal duct restrictions, etc. 8) The majority of fan field performance tests cover a single point of operation, namely, the design duty. If it is deemed necessary to cover several points of operation, provision must be made in advance for changing the system resistance. The means used to vary the system resistance must not cause adverse flow conditions in the vicinities of the fan and measurement planes. If the fan cannot be tested at the quoted system design point, then it is sufficient for the evaluation of fan field performance to establish the proximity of the field test point to any portion of the fan performance rating curve within the limitations of the uncertainty analysis (see Annex T).

15. Test Preparation 15.1 The following items should be agreed upon by all interested parties prior to the start of a field performance test: 1) AMCA Publication 200, Air Systems, AMCA Publication 201, Fans and Systems, and AMCA Publication 202, Troubleshooting, should be reviewed and implemented before starting the field test. 2) Personnel conducting field tests on fans must be technically competent and fully conversant with all four parts of the AMCA Fan Application Manual. The person responsible for conducting the test should be designated and agreed upon by all parties. 3) The test instrumentation and locations of test measurement planes should be established. 160 | Field Performance Measurement

9) It must be established that the system remains constant for the duration of the test. Modulating dampers should be set in a fixed position, no process changes shall be undertaken, etc. Variable inlet vane controls or inlet box dampers must be set in the full open position for the duration of the test, except when testing for control characteristics. 10) All precautions to ensure the safety of test personnel must be observed. 11) The fan-system should be operated for a length of time sufficient to ensure steady state conditions prior to the start of the test. 12) It is advisable that representatives of all parties interested in the test results be present at the time of the test to cover their areas of responsibility.

15.2 It is recommended that as a minimum, the following equipment be taken to or be otherwise available at the job site: 1) Pitot-static tubes of suitable lengths for the maximum duct size to be traversed. Considerations should be given to the use of a double reverse tube in dirty atmospheres. 2) Manometers suitable for measuring static pressures. Manometer fluids other than water are acceptable, provided the specific gravity is known. A spare bottle of manometer fluid is advisable. 3) Inclined manometer suitable for measuring velocity pressures. 4) Flexible tubing of suitable length to enable manometers to be installed at a convenient location. 5) Tubing couplings and “T” type tubing connectors. 6) Thermometers to cover the range of anticipated temperatures. 7) Sling psychrometer for obtaining dry-bulb and wet-bulb temperatures. 8) Clip-on ammeter-voltmeter, power analyzer, or other suitable electrical measurement instruments for the determination of fan power input. 9) Fan speed measurement instrument. 10) Aneroid barometer. 11) Flashlight, tape, measuring rule, hand tools, coveralls, etc. 12) Test data sheets, calculator, and necessary drawings. 13) Complete AMCA Fan Application Manual containing Publications 200, 201, 202, and 203.

2) Static and total pressure manometer tubing must be “pinched off” prior to inserting or removing the Pitot-static tube from the test duct. Release both legs of the tubing simultaneously after the Pitotstatic tube is inside the test duct and properly oriented. Failure to release simultaneously may result in manometer fluid being blown from the manometer. 3) Loop the manometer tubing well above the manometer so that any fluid which is inadvertently blown from the gauge will drain back into the manometer. 4) The Pitot-static tube is intended for measuring pressures in relatively clean gases. When using Pitot-static tubes in dirty, wet, or corrosive atmospheres, both legs of the Pitot-static tube must be cleaned out frequently during the test. Since fan pressure readings are never strictly steady, absence of fluctuations is an indication of a plugged Pitot-static tube. Consider using a double reverse tube in these situations. 5) When making measurements in wet gas streams, continually check for the presence of moisture in the tubing. Clear plastic tubing is ideal from this standpoint. If moisture collects in the tubing, immediately remove the Pitot-static tube and clean the inside of the tubing and Pitotstatic tube before proceeding with the test. 6) Before performing any work inside a fan, ductwork, or other system components, make certain that the fan motor starter is “locked out.” 7) The area at the plane of flow measurement should be measured internally to account for internal insulation or other obstructions. 8) Do not rely on damper control indicators to ensure that dampers are fully open. Check visually. 9) Measure temperatures on both sides of double inlet fans as temperature differences may exist between each side.

16. Precautions The following precautions should be observed when conducting a field test: 1) Connect the Pitot-static tube to the manometers according to anticipated pressures, i.e., whether the pressures are positive or negative, and the magnitudes of pressures.

10) When measuring in high temperature, corrosive or explosive atmospheres, instruments should be selected for suitability for such atmospheres.

17. Typical Fan-System Installations A fan assembly may include any number of appurtenances: variable inlet vanes, inlet boxes, inlet box dampers, outlet dampers, inlet screens, belt Field Performance Measurement | 161

guards, inlet bells, diffusers (evasés). Alternately, these items may be included in the fan-system installation, but not be a part of the fan assembly. In order to determine the proper field test procedure and to provide a valid basis for comparing field test results to the fan performance ratings, it is important to establish which of these items are considered a part of the fan and which are considered a part of the system. The fan performance ratings may be assumed to include the appurtenances that are established as being a part of the fan assembly. The locations of the fan inlet and fan outlet depend on whether specific appurtenances are considered to be a part of the fan assembly. If the assembly includes an inlet box, the fan inlet is the inlet to the inlet box. For a fan assembly that includes a diffuser, the fan outlet is the outlet of the diffuser. In the case of heating, ventilating, and airconditioning equipment, the field test procedure will depend on whether the equipment is a factory assembled central station unit, a built-up unit, or a packaged unit (see Section 17.4). The performance ratings for a fan that includes inlet box dampers, variable inlet vanes or outlet dampers cover operation of the fan with these items in the full open positions. In order to be able to compare the field test results to the fan performance ratings, it is essential that these items be fixed in their full open positions for the duration of the test. In addition, when the loss through a damper must be calculated, it is essential that the damper blades be fixed in their full open positions during the test since this is the condition on which the damper pressure loss ratings are based. This consideration arises when a damper, which is not considered a part of the fan is located between a static pressure measurement plane and the fan. In order to determine the fan static pressure, the loss through the damper must be calculated. In these cases, the calculation of the loss is based on the performance ratings for the damper.

17.1 Free inlet, free outlet fans It is difficult to achieve an accurate field test of a free inlet, free outlet fan. The most obvious problem is the lack of a suitable location for the velocity pressure measurement plane. In addition, in the case of ventilators that supply or exhaust air from a buildingthe most commonly encountered applications of free inlet, free outlet fans-it is extremely difficult to define, set, and maintain for the duration of the test the “normal” system condition. Items affecting the system include:

162 | Field Performance Measurement

a) The operations of ovens, furnaces, paint booths, air conditioning equipment, other fans, and similar items that may supply or exhaust air from the building in intermittent or modulating fashions. b) The use of doors providing access to the building. The effect is most significant when large doors that are normally closed are kept open for extended periods such as in loading operations. c) The velocity and direction of the wind outside the building, particularly in conjunction with the item immediately above and as it may affect the flow of air from the outlet of the ventilator. d) The use of interior doors that my restrict the flow of air from areas normally expected to be ventilated. Assuming that these difficulties can be resolved and the desired system is fixed for the duration of the test, determine the fan performance by using one of the following methods: 1) Make field test measurements sufficient for determining fan static pressure, fan power input, fan speed, and the density of the air at the fan inlet. In this method for testing a free inlet, free outlet fan, the fan static pressure is calculated as the static pressure on the outlet side of the fan less the static pressure on the inlet side of the fan: Ps = Ps2 - Ps1. The static pressure measurements involved must be referred to the same atmospheric pressure and made at locations sufficiently distant from the fan inlet and outlet so as to be unaffected by the velocity of the air entering and leaving the fan. Using the fan manufacturer’s certified performance ratings, draw a performance curve for the fan for operation at the test values of fan speed and entering air density. Determine the fan air flow rate by entering this curve at the test values of fan static pressure and fan power input (see Example 5C in Annex A). 2) Use the method as described above with the exception that the performance curve is established by a laboratory test of the fan, conducted in accordance with AMCA Standard 210. For the laboratory test, the fan must be set up in a manner that duplicates the field installation conditions. That is, all appurtenances must be in place and any restrictions or obstructions to the free flow of air into the fan inlet and away from the fan outlet must be accurately duplicated in the laboratory test setup.

3) Install a duct on the inlet side of the fan for the purpose of providing a location for the velocity pressure measurement plane. All of the test measurements and calculations in this method for testing a free inlet, free outlet fan are the same as those required for a fan with a ducted inlet and a free outlet. The cross-sectional shape and area of the duct, which is temporarily installed for purposes of the test, should be selected on the basis of minimizing its interference with the flow of air into the fan inlet while providing velocity pressure of magnitudes that can be accurately measured. The length of the duct should be a minimum of twice its diameter or equivalent diameter, and the entrance to the duct should be flared in order to reduce the entrance loss. The velocity pressure measurement plane should be located a minimum of 1.5 diameters or equivalent diameters downstream from the duct inlet. The effect of this duct on the system is negligible, but in changing the pattern of the flow of air into the fan inlet, it may affect the performance of the fan slightly. Applications of this method of test are presented in Examples 5A and 5B in Annex A. The equation for calculating fan static pressure for this configuration is: Ps = Ps2 - (Ps1 + Pv1)

17.2 Free inlet, ducted outlet fans In the calculation of fan static pressure for this type of fan-system configuration, the sum of the static pressure at the fan inlet, Ps1, and the velocity pressure at the fan inlet, Pv1, is considered to be equal to the sum of the static pressure, Psx, and the velocity pressure, Pvx, at a point sufficiently distant from the fan inlet as to be in still air. At this point, the static pressure is zero, and the velocity pressure in still air is zero. Ps1 + Pv1 = Psx + Pvx = 0 This consideration, which is the same as that used in the methods for testing fans for performance rating purposes, charges to the fan the losses incurred in accelerating the air into the fan inlet and eliminates inaccuracies which may occur in any attempt to measure velocity pressure and static pressure at the fan inlet. Since Ps1 + Pv1 = 0, the equation for calculating fan static pressure for this configuration is: Ps = Ps2 + SEF 1 +SEF 2 + ... + SEF n

17.3 Ducted inlet, ducted outlet fans In this type of fan-system configuration, there is no special consideration in the calculation of fan static pressure. The equation for this calculation is: Ps = Ps2 - Ps1 - Pv1 + SEF 1 + SEF 2 + ... + SEF n In this configuration, the flow conditions on the inlet side of the fan are usually more favorable for the location of the velocity pressure measurement plane.

17.4 Ducted inlet, free outlet fans In this type of fan-system configuration, the static pressure at the fan outlet, Ps2, is zero gauge pressure, referred to the atmospheric pressure in the region of the fan outlet. However, the gas stream may be discharging from the fan into a region in which the atmospheric pressure is somewhat different from that to which all other pressure measurements are referred. When this possibility exists, it is essential that the static pressure measurements in the region of the fan outlet be referred to the same atmospheric pressure as used in all other pressure measurements. Ps = -Ps1 - Pv1 + SEF 1 + SEF 2 + ... + SEF n

17.5 Air handling units This category consists of draw-through and blowthrough types of equipment assemblies used in heating, ventilating, and air-conditioning applications. In addition to fans, these equipment assemblies may include any number of combinations of coils, filters, access sections, humidifiers, mixing boxes, dampers, etc. Air handling units include packaged units, factory assembled central station units, and built-up units. The basis used in establishing the air performance ratings for each of these unit types is described below. It is important that the field test method correspond to the rating method in each case. 17.5.1 Packaged units. This type of unit is supplied and rated by the manufacturer as an assembly. The static pressures at the inlet and outlet to the assembly and the velocity pressure at the inlet to the assembly are used in calculating the static pressure generated by this type of air handling unit. See Examples 4C and 4D in Annex A. 17.5.2 Factory assembled central station units. The air performance ratings for this type of unit are based on the operation of the fan section assembly only, but include the effects of the air flow conditions entering and leaving the fan section which are created by accessory equipment such as plenums, Field Performance Measurement | 163

coils, filters, mixing boxes, etc. The fan section assembly includes the fan and the cabinet in which the fan has been installed. The accessory items are considered to be included in the system in which the fan section operates. The static pressure and the velocity pressure at the inlet of the fan section and the static pressure at the fan section outlet, which coincides with the fan outlet, are used in calculating the static pressure generated by the fan section assembly. See examples 4B and 4E in Annex A. 17.5.3 Built-up units. Built-up units are similar to factory assembled central station units, except that in built-up units, the components are normally obtained from a number of equipment suppliers and the unit is assembled at the installation site. The fans which are used in built-up units are rated as free-standing, unencumbered by the cabinets in which they are installed. In the field test determination of the performance of the fan, the static pressure and velocity pressure at the fan inlet and the static pressure at the fan outlet are used in calculating the fan static pressure. An SEF that accounts for the effect of the cabinet is normally included in this calculation, and it may be necessary to include an SEF to account for the conditions at the fan outlet. See Example 4A in Annex A.

164 | Field Performance Measurement

Annex A. Field Test Examples This annex contains examples of field tests. The examples are presented in detail and cover several types of fansystem combinations. Field test procedures are illustrated in a variety of situations. Portions of the procedures are typical for all fan-system installations. Other portions of the procedures demonstrate methods for dealing with the more difficult features encountered in some installations. Not all of the possible fan-system combinations are included in the examples, but it is expected that the examples will provide sufficient guidance for dealing with those cases not covered.

EXAMPLES OF FANS, INSTALLATION TYPE B: FREE INLET, DUCTED OUTLET 1A: 1B: 1C: 1D:

Centrifugal Forced Draft Fan Centrifugal Forced Draft Fan with Inlet Silencers Axial Forced Draft Fan with Inlet Silencers Centrifugal Fans in Parallel

EXAMPLE OF FANS, INSTALLATION TYPE D: DUCTED INLET, DUCTED OUTLET 2A: 2B: 2C: 2D: 2E: 2F: 2G:

Utility Fan in a Ventilating System Centrifugal Fan in a Sawdust Conveying System Axial Fan in a Dryer System Centrifugal Fan in a Scrubber System Centrifugal Fan in a Process System Axial Fan in a Ventilation System High Pressure Centrifugal Fans in Series

EXAMPLES OF FANS, INSTALLATION TYPE C: DUCTED INLET, FREE OUTLET 3A: 3B: 3C: 3D:

Centrifugal Fan in an Exhaust System Axial Fan in an Exhaust System Centrifugal Fan in a Scrubber System Centrifugal Roof Ventilator with Ducted Inlet

EXAMPLES OF AIR HANDLING UNITS 4A: 4B: 4C: 4D: 4E:

Centrifugal Fan in a Built-up Air conditioning Unit Central Station Air Conditioning Unit, Factory Assembled Draw-Through Type Packaged Air Conditioning Unit Packaged Air Conditioning Unit Central Station Air Conditioning Unit, Factory Assembled Blow-Through Type

EXAMPLES OF FANS, INSTALLATION TYPE A: FREE INLET, FREE OUTLET 5A: 5B: 5C:

Free Inlet, Free Outlet Roof Ventilator with temporary duct Free Inlet, Free Outlet Propeller Fan with temporary duct Free Inlet, Free Outlet Roof Ventilator as installed

Field Performance Measurement | 165

EXAMPLE 1A: CENTRIFUGAL FORCED DRAFT FAN

SEF 1 DIFFUSER

3

2

L

A2

VARIABLE INLET VANES SIDE VIEW

A3 OUTLET SIDE VIEW

LOCATIONS OF PLANES 2 AND 3

ORIENTATION OF PITOT TUBE COMMENTS 1. The variable inlet vanes are considered part of the fan. Performance ratings for fans with inlet vanes cover operation with the inlet vanes in their full open position. In order to be able to compare the test results to the fan performance ratings, it is essential that the inlet vanes be fixed in their full open positions for the duration of the test. 2. Determine Pv3 by using the root mean square of the velocity pressure measurements made in a traverse of Plane 3, located near the end of the fan diffuser (evasé). Determine Ps3 by averaging the static pressure measurements made in the same traverse. Procedures for the traverse are described in Section 9.4. These velocity pressure and static pressure measurements are susceptible to error due to the turbulence existing in the region of the fan outlet. In addition, it is undesirable to have Plane 3 located in a diverging airway. However, no other more suitable location for Plane 3 exists in this example. It is recommended that the Pitot-static tube be oriented so that its nose is aligned with the anticipated flow streams, particularly near the walls of the diffuser, as shown in the diagram. Determine the area of the traverse plane, A3, which is located at the tip of the Pitot-static tube, as shown in the diagram, not at the location of the Pitot-static tube access holes in the diffuser.

166 | Field Performance Measurement

3. Measure td1 and tw1 in the path of the air flowing into the fan inlets. Determine pb for the general vicinity of the fan. Measure td3 in Plane 3. All of these measurements are used in the determination of densities at the various planes of interest. 4. Measure the fan speed and the motor amps, volts, and, if possible, watts. Record all pertinent motor nameplate data, including volts (NPV) and full load amps (FLA). If the motor power output is to be estimated by using the phase current method described in Annex K, it is not necessary to measure motor watts; however, it may be necessary to disconnect the drive and measure the no load amps (NLA) if the motor is not operating at or near its full load point (refer to Annex K). 5. SEF 1 is due to the effect of insufficient length of duct at the fan outlet. In order to calculate the value of SEF 1, it is necessary to measure the length of the outlet duct, L; the outlet area of the fan, A2; and the blast area of the fan. 6. The sum of the static pressure, Ps1, and velocity pressure, Pv1, at the inlets of a fan with unrestricted inlets is considered to be equal to the sum of the static pressure, Psx, and the velocity pressure, Pvx, at a point sufficiently distant from the fan inlets as to be in still air. At this point, the static pressure is zero, and

the velocity pressure in still air is zero.

GENERAL

Ps1 + Pv1 = Psx + Pvx = 0

VIVs in full open positions. Fan direct connected to motor.

This consideration, which is the same as that used in the methods for testing fans for performance rating purposes, charges to the fan losses incurred in accelerating the air into the fan inlets and eliminates the inaccuracies which arise in any attempt to measure the velocity pressure and static pressure at the fan inlets. To calculate the fan static pressure:

DENSITIES

Ps = Ps2 - Ps1 - Pv1 + SEF 1 = Ps2 - (Ps1 + Pv1) + SEF1

td1 = tw1 = p1 = =

CALCULATIONS

For fan inlet conditions of: 85°F 63°F pb 28.91 in. Hg

Since: Ps1 + Pv1 = 0 Ps = Ps2 + SEF 1

Use Figure N.1 in Annex N to obtain ρ1 = 0.0701 lbm/ft3 The density at Plane 3:

7. In order to compare the test results to the quoted fan curve drawn for operation at 1780 rpm and 0.0701 lbm/ft3 density, it is necessary to convert the results to the specified conditions. In this case, the test conditions are identical to the specified conditions and no calculations are required.

OBSERVATIONS

⎛ P + 13.6 pb ⎞ ⎛ t d1 + 460 ⎞ ρ3 = ρ1 ⎜ s3 ⎟ ⎟⎜ ⎝ 13.6 p1 ⎠ ⎝ t d3 + 460 ⎠ ⎛ 14.4 + 13.6 × 28.91 ⎞ ⎛ 545 ⎞ = 0.0701⎜ 13.6 × 28.91 ⎟⎠ ⎜⎝ 556 ⎟⎠ ⎝ = 0.0712 lbm/ft 3

SITE MEASUREMENTS

In this case, ρ2 is considered to be equal to ρ3.

pb = 28.91 in. Hg td1 = 85°F tw1 = 63°F td3 = 96°F Ps3 = 14.4 in. wg Pv3 = 1.52 in. wg N = 1780 rpm A2 = 11.94 ft2 A3 = 11.3 ft2 Blast Area = 7.76 ft2 L = 3 ft.

FLOW RATES

MEASURED MOTOR DATA Volts = = Amps = =

570, 560, 572 567 av. 160, 166, 163 163 av.

MOTOR NAMEPLATE DATA 200 hp, 3 phase, 60 hertz 575 volts, 1800 rpm, 181 FLA

V3 = 1096 (Pv3/ρ3)0.5 = 1096 (1.52/0.0712)0.5 = 5064 fpm Q3 = V3A3 = 5064 × 11.3 = 57223 cfm Q = = = =

Q1 Q3 (ρ3/ρ1) 57223 (0.0712/0.0701) 58121 cfm

FAN POWER INPUT Measured amps/FLA = (163/181) = 0.90 = 90% Annex K indicates that Equation A will provide a reasonably accurate estimate of motor power output for a 200 hp motor operating at 90% FLA.

Field Performance Measurement | 167

Hmo = 200 (163/181) (567/575) = 178 hp Since the fan is direct connected to the motor: H

= Hmo = 178 hp

SYSTEM EFFECT FACTOR AMCA Publication 201-90, Figures 7.1 and 8.3 indicate the following calculations: Q2 = Q3 (ρ3/ρ2) = 57223 (0.0712/0.0712) = 57223 cfm V2 = (Q2/A2) = (57223/11.94) = 4793 fpm Duct diameter equivalent to the diffuser outlet area: De2 = 4 A2 / π =

( 4 × 11.94 ) / π

= 3.9 ft. Figure 8.3 shows that for velocities over 2500 fpm, 100% effective duct length is one duct diameter per 1000 fpm, = De2 (V2/1000) = 3.9 (4793/1000) = 18.7 ft L in % effective duct length = (L/18.7) 100 = (3/18.7) 100 = 16% Blast area ratio = Blast Area/A2 = 7.76/11.94 = 0.65 For blast area ratio of 0.65, and 16% effective duct length, Figure 8.3 shows System Effect Curve U applies. For 4793 fpm velocity and curve U, Figure 7.1 shows SEF 1 = 0.6 in. wg at 0.075 lbm/ft3. At 0.0712 lbm/ft3. SEF 1 = 0.6 (0.0712/0.075) = 0.57 in. wg

168 | Field Performance Measurement

FAN STATIC PRESSURE Since A2 is greater than A3, there may be some conversion of velocity pressure to static pressure between Planes 3 and 2. However, the amount of conversion will be very small relative to the static pressure measured at Plane 3 and ignoring any change in static pressure from Plane 3 to Plane 2 will have no appreciable effect on the test results. Therefore, Ps2 is considered equal to Ps3. Ps = Ps2 + SEF 1 = 14.4 + 0.57 = 14.97 in. wg CONVERSION TO SPECIFIED CONDITIONS Qc = = Psc = = Hc = =

Q 58121 cfm Ps 14.97 in. wg H 178 hp

EXAMPLE 1B: CENTRIFUGAL FORCED DRAFT FAN WITH INLET SILENCERS

TEMPORARY DUCT

DIFFUSER STATIC PRESSURE TAPS

3a

0.5 De

3b SILENCERS

1 A2

SEF 1

VARIABLE INLET VANES SIDE VIEW

OUTLET SIDE VIEW

2 COMMENTS

1. This fan, as supplied and rated by the manufacturer, includes the variable inlet vanes and inlet boxes, but does not include the silencers. Performance ratings for fans with inlet vanes cover operation with the inlet vanes in the full open positions. In order to be able to compare the test results to the fan performance ratings, it is essential that the inlet vanes be fixed in their full open positions for the duration of the test. 2. Determine Pv3a and Pv3b by using the root mean square of the velocity pressure measurements made in traverses of Planes 3a and 3b. A3a and A3b are the areas traversed. Determine Ps3a and Ps3b by averaging each of the two sets of static pressure measurements made in the same traverses. Procedures for traverses are described in Section 9.4. Ps3a and Ps3b are used in determining the density at the traverse plane. A location for Plane 3 measurements may be obtained by installing ducts on each silencer inlet, as shown in the diagram. The ducts should be a minimum of one equivalent diameter in length, and have flared inlets to reduce entrance losses and provide more uniform velocity profiles at the pressure measurement planes. 3. Measure Ps1a and Ps1b at locations close to the entrances to the inlet boxes and in planes which are substantially equal in area to the planes of the

entrances to the inlet boxes (Plane 1). Determine Ps2 by averaging the pressure measurements at each of four static pressure taps located near the end of the fan diffuser (evasé). See Annex E for details of static pressure taps. 4. Measure td3 and tw3 near the inlet ducts. Determine pb for the general vicinity of the fan. Measure td2 in Plane 2. All of these measurements are used in the determination of densities at the various planes of interest. 5. Measure the fan speed and the motor amps, volts, and if possible, watts. Record all pertinent motor nameplate data, including volts (NPV) and amps (FLA). If the motor power output is to be estimated by using the phase current method described in Annex K, it is not necessary to measure motor watts; however, it may be necessary to disconnect the drive and measure the no load amps (NLA) if the motor is not operating at or near its full load point. Refer to Annex K. 6. SEF 1 is due to the effect of there being no duct at the fan outlet. In order to calculate the value of SEF 1, it is necessary to measure the fan outlet area, A2, and the blast area of the fan. 7. To calculate the fan static pressure: Field Performance Measurement | 169

Ps = Ps2 - Ps1 - Pv1 + SEF 1

CALCULATIONS

Where:

DENSITIES

Pv1 = (Q/1096A1)2 ρ1

For Plane 3 conditions of:

8. In order to compare the test results to the quoted fan curve drawn for operation at 1180 rpm and 0.075 lbm/ft3 density, it is necessary to convert the results to the specified conditions. The basis for the calculations is described in Section 14.

td3 = 85°F tw3 = 58°F

OBSERVATIONS SITE MEASUREMENTS = 29.31 in. Hg = 93°F = 85°F = 58°F = -1.20 in. wg = -1.30 in. wg = 10.1 in. wg = -0.65 in. wg = -0.70 in. wg = 0.61 in. wg = 0.62 in. wg = 1180 rpm = A1b = 12.5 ft2 A2 = 18 ft2 A3a = A3b = 12.5 ft2 Blast Area = 13.5 ft2 pb td2 td3 tw3 Ps1a Ps1b Ps2 Ps3a Ps3b Pv3a Pv3b N A1a

MEASURED MOTOR DATA Volts = = Amps = =

460, 455, 465 460 av 257, 256, 258 257 av

MOTOR NAMEPLATE DATA 200 HP, 3 phase 60 hertz 460 volts, 1180 rpm, 285 FLA GENERAL VIVs in full open positions. Fan direct connected to motor. The motor manufacturer advises that this motor type has a peak efficiency of 91% at a power factor of approximately 0.89.

Ps3 = = = p3 = = =

(Ps3a + Ps3b)/2 (-0.65 - 0.70)/2 -0.675 in. wg pb + (Ps3/13.6) 29.31 + (-0.675/13.6) 29.26 in. Hg

Use Figure N.1 in Annex N to obtain ρ3 = 0.0712 lbm/ft3 It is assumed that the temperature at Plane 1 are the same as those at Plane 3. The density at Plane 1: ⎛ P + 13.6 pb ⎞ ⎛ t d3 + 460 ⎞ ρ1 = ρ3 ⎜ s1 ⎟⎜ ⎟ ⎝ 13.6 p3 ⎠ ⎝ t d1 + 460 ⎠ ⎛ −1.25 + 13.6 × 29.31 ⎞ ⎛ 545 ⎞ = 0.0712 ⎜ ⎟ ⎜ 545 ⎟ 13.6 × 29.26 ⎝ ⎠⎝ ⎠ = 0.0711 lbm/ft 3 The density at Plane 2: ⎛ P + 13.6 pb ⎞ ⎛ t d3 + 460 ⎞ ρ2 = ρ3 ⎜ s2 ⎟⎜ ⎟ ⎝ 13.6 p3 ⎠ ⎝ t d2 + 460 ⎠ ⎛ 10.1 + 13.6 × 29.31 ⎞ ⎛ 545 ⎞ = 0.0712 ⎜ ⎟⎜ ⎟ ⎝ 13.6 × 29.26 ⎠ ⎝ 553 ⎠ = 0.0721 lbm/ft 3 FLOW RATES V3a = 1096 (Pv3a/ρ3)0.5 = 1096 (0.61/0.0712)0.5 = 3208 fpm Q3a = V3aA3a = 3208 × 12.5 = 40100 cfm V3b = 1096 (Pv3b/ρ3)0.5 = 1096 (0.62/0.0712)0.5 = 3234 cfm Q3b = V3bA3b = 3234 × 12.5 = 40425 cfm

170 | Field Performance Measurement

Q3 = Q3a + Q3b = 40100 + 40425 = 80525 cfm Q = = = =

Q1 Q3 (ρ3/ρ1) 80525 (0.0712/0.0711) 80638 cfm

SYSTEM EFFECT FACTOR AMCA Publication 201-90, Figures 7.1 and 8.3 indicate the following calculations: Q3 (ρ3/ρ2) = 80525 (0.0712/0.0721) = 79520 cfm (Q2/A2)

FAN POWER INPUT

= (79520/18) = 4418 fpm

Measured amps/FLA = (257/285) = 0.90 = 90%

Blast area ratio = Blast Area/A2 = 13.5/18 = 0.75

Annex K indicates that Equation A will provide a reasonably accurate estimate of motor power output for a 250 hp motor operating at 90% FLA.

For a blast area ratio of 0.75, and no duct, Figure 8.3 shows System Effect Curve T applies. For 4418 fpm velocity and curve T, Figure 7.1 shows SEF 1 = 0.65 in. wg at 0.075 lbm/ft3. At 0.0720 lbm/ft3:

Hmo = 250 (257/285) (460/460) = 225 hp As a check of this value, using the motor efficiency data and the appropriate equation in Section 11.2.2:

Hmo

3 × 257 × 460 × 0.89 × 0.91 = 746 = 222 hp

Since the motor is not fully loaded, the power factor and efficiency may be less, which would reduce Hmo as calculated using the second method. However, this is a reasonable check. The value of Hmo is selected to be the average of the two results:

SEF 1 = 0.65 (0.0721/0.075) = 0.62 in. wg FAN STATIC PRESSURE Pv1 = (Q1/1096 A1)2 = (80638/1096 × 25)2 0.0711 = 0.62 in. wg Ps = Ps2 - Ps1 - Pv1 + SEF 1 = 10.1 - (-1.25) - 0.62 + 0.62 = 11.33 in. wg CONVERSION TO SPECIFIED CONDITIONS

Hmo = 224 hp

Qc = Q = 80638 cfm

Since the fan is direct-connected to the motor, there is no drive loss, and:

Psc = 11.33 (0.075/0.0711) = 11.95 in. wg

H = Hmo = 224 hp

Hc = 224 (0.075/0.0711) = 236 hp

Field Performance Measurement | 171

EXAMPLE 1C: AXIAL FORCED DRAFT FAN WITH INLET SILENCER

PLANE 3 LOCATION 3

TEMPORARY SHORT DUCT STATIC PRESSURE TAPS SILENCER

TRANSITION

0.5 De

5 INLET BOX

1

DIFFUSER SECTION

2

INNER CYLINDER

L

SIDE VIEW

GUIDE VANES COMMENTS 1. This is a variable pitch axial flow fan. The fan assembly, as supplied and rated by the manufacturer, includes the inlet box and diffuser section, but does not include the silencer. It is essential that the blade pitch angle be fixed for the duration of the test. This blade angle should be agreed upon by all interested parties. 2. A temporary short duct is installed upstream of the silencer to establish Plane 3 in which more uniform pressures can be obtained. The duct should be a minimum of one equivalent diameter in length, and have a flared inlet to reduce entrance losses and provide a more uniform velocity profile at the pressure measurement plane. Determine Pv3 by using the root mean square of the velocity pressure measurements made in a traverse of Plane 3. Ps3 is determined by averaging the static pressure measurements made in the same traverse. Procedures for traverses are described in Section 9.4. Ps3 is used in determining the density at the traverse plane. 3. Measure Ps1 at a location close to the entrance to the inlet box and in a plane which is substantially equal in area to the plane of the entrance to the inlet box (Plane 1). Determine Ps5 by averaging the pressure measurements at each of four static pressure taps located near the end of the fan diffuser. 172 | Field Performance Measurement

See Annex E for details of static pressure taps. In this example, Ps2 is considered to be equal to Ps5. 4. Measure td3 and tw3 near the entrance to the short inlet duct. Determine pb for the general vicinity of the fan. Measure td5 in Plane 5. All of these measurements are used in the determination of densities at the various planes of interest. 5. Measure the fan speed and the motor amps, volts, and if possible, watts. Record all pertinent motor nameplate data, including volts (NPV) and full load amps (FLA). If the motor power output is to be estimated by using the phase current method described in Annex K, it is not necessary to measure motor watts; however, it may be necessary to disconnect the drive and measure the no load amps (NLA) if the motor is not operating at or near its full load point. Motor performance data, supplied by the motor manufacturer, are used in the determination of motor power output for this example. 6. SEF 1 is due to the effect of insufficient length of duct between the diffuser outlet and the elbow downstream of the diffuser. In order to calculate the value of SEF 1, it is necessary to measure the length of the transition, L, and the outlet area of the diffuser, A2.

7. To calculate the Fan Static Pressure:

CALCULATIONS

Ps = Ps2 - Ps1 - Pv1 + SEF 1

DENSITIES

Where:

For Plane 3 conditions of:

Pv1 = Pv3 (A3/A1)2 (ρ3/ρ1)

td3 = 68°F tw3 = 62°F

8. Axial fans are often rated in Fan Total Pressure. Computation of Fan Total Pressure is illustrated in the CALCULATIONS section of this example. 9. In order to compare the test results to the quoted fan curve drawn for operation at 880 rpm and 0.0740 lbm/ft3 density, it is necessary to convert the results to the specified conditions. In this case, the test conditions are identical to the specified conditions and no calculations are required. OBSERVATIONS SITE MEASUREMENTS pb td3 tw3 td5 Ps1 Ps3 Ps5 Pv3 N A1 A2 A3 A5 L

= 29.8 in. Hg = 68°F = 62°F = 88°F = -1.80 in. wg = -1.40 in. wg = 20.8 in. wg = 1.30 in. wg = 880 rpm = 170.3 ft2 = 176 ft2 = 170.3 ft2 = A2 = 15 ft

MEASURED MOTOR DATA Volts = = Amps = =

4000, 4000, 4100 4033 av 450, 445, 448 448 av

MOTOR NAMEPLATE DATA 4000 hp, 3 phase 60 hertz 4000 volts, 900 rpm, 520 FLA GENERAL

p3 = pb + (Ps3/13.6) = 29.8 + (-1.40/13.6) = 29.70 in. Hg Use Figure 20 in Annex N to obtain ρ3 = 0.0744 lbm/ft3 It is assumed that td1 = td3. The density at Plane 1: ⎛ P + 13.6 pb ⎞ ⎛ t d3 + 460 ⎞ ρ1 = ρ3 ⎜ s1 ⎟⎜ ⎟ ⎝ 13.6 p3 ⎠ ⎝ t d1 + 460 ⎠ ⎛ −1.8 + 13.6 × 29.8 ⎞ ⎛ 528 ⎞ = 0.0744 ⎜ ⎟⎜ ⎟ ⎝ 13.6 × 29.70 ⎠ ⎝ 528 ⎠ = 0.0743 lbm/ft 3 The density at Plane 2:

ρ 2 = ρ5 ⎛ P + 13.6 pb ⎞ ⎛ t d3 + 460 ⎞ = ρ3 ⎜ s5 ⎟⎜ ⎟ ⎝ 13.6 p3 ⎠ ⎝ t d5 + 460 ⎠ ⎛ 20.8 + 13.6 × 29.8 ⎞ ⎛ 528 ⎞ = 0.0744 ⎜ ⎟⎜ ⎟ ⎝ 13.6 × 29.70 ⎠ ⎝ 548 ⎠ = 0.0756 lbm/ft 3 FLOW RATE V3 = 1096 (Pv3/ρ3)0.5 = 1096 (1.3/0.0744)0.5 = 4581 fpm Q3 = V3A3 = 4581 × 170.3 = 780144 cfm Q = = = =

Q1 Q3 (ρ3/ρ1) 780144 (0.0744/0.0743) 781194 cfm

Fan direct connected to motor. Motor performance data at operating load, as supplied by motor manufacturer: 0.88 power factor, 95% efficiency.

Field Performance Measurement | 173

FAN POWER INPUT Hmo =

3 × volts × amps × power factor × efficiency 746

3 × 4033 × 448 × 0.88 × 0.95 746 = 3507 hp =

SEF 1 = 0.32 (0.0756/0.075) = 0.32 in. wg FAN STATIC PRESSURE Pv1 = Pv3 (A3/A1)2 (ρ3/ρ1) = 1.3 (170.3/170.3)2 (0.0744/0.0743) = 1.30 in. wg

Since the fan is direct connected to the motor, there is no drive loss, and:

Ps2 = Ps5 = 20.8 in. wg

H = Hmo = 3507 hp

Ps = Ps2 - Ps1 - Pv1 + SEF 1 = 20.8 - (-1.80) - 1.30 + 0.32 = 21.62 in. wg

SYSTEM EFFECT FACTOR FAN TOTAL PRESSURE AMCA Publication 201-90, Figures 7.1, 8.1, and 8.4 indicate the following calculations: Q2 = Q3 (ρ3/ρ2) = 780144 (0.0744/0.0756) = 767761 cfm V2 = (Q2/A2) = (767761/176) = 4362 Duct diameter equivalent to the diffuser outlet area: De2 = 4 A2 / π =

( 4 × 176 ) / π

= 15 ft. Figure 8.1 shows that for velocities over 2500 fpm, 100% effective duct length is one duct diameter for every 1000 fpm: = De2 (V2/1000) = 15 (4362/1000) = 65.43 ft. L in % effective duct length = (L/65.43) 100 = (15/65.43) 100 = 23% For 23% effective duct length and a vaneaxial fan with a 2 piece elbow, Figure 8.4 shows System Effect Curve V applies. For 4362 fpm velocity and curve V, Figure 7.1 shows SEF 1 = 0.32 in. wg at 0.075 lbm/ft3. At 0.0756 lbm/ft3.

174 | Field Performance Measurement

Pt1 = Ps1 +Pv1 = -1.8 + 1.30 = -0.50 in. wg Pv2 = Pv3 (A3/A2)2 (ρ3/ρ2) = 1.3 (170.3/176)2 (0.0744/0.0756) = 1.20 in. wg Pt2 = Ps2 + Pv2 = 20.8 + 1.20 = 22.00 in. wg Pt = Pt2 - Pt1 + SEF 1 = 22.00 - (-0.50) + 0.32 = 22.82 in. wg Also: Pt = Pv = = Pt = =

Ps + Pv Pv2 1.20 in. wg 21.62 + 1.20 22.82 in. wg

CONVERSION TO SPECIFIED CONDITIONS Qc = = Psc = = Ptc = = Hc = =

Q 781194 cfm Ps 21.62 in. wg Pt 22.82 in. wg H 3507 hp

EXAMPLE 1D: CENTRIFUGAL FANS IN PARALLEL

3

STATIC PRESSURE TAPS OUTLET DAMPER

2

SEF 1 PLENUM 1

PLAN VIEW

1

SIDE VIEW

COMMENTS 1. Each of the fans, as supplied and rated by the manufacturer, includes an outlet damper. Performance ratings for fans with outlet dampers cover operation with the outlet damper in the full open position. In order to be able to compare the test results to the fan performance ratings it is essential that the outlet dampers be fixed in the full open positions for the duration of the test. 2. In this example, there are no suitable locations for traverse planes for use in determining directly the flow rate for each fan. The alternative is to determine the total flow rate and since the fans and their operating speeds are alike, assume that each fan delivers a flow rate proportional to its actual speed. Determine Pv3 by using the root mean square of the velocity pressure measurements made in a traverse of Plane 3, located near the end of a straight run of duct, such as shown in the diagram. Determine Ps3 by averaging the static pressure measurements made in the same traverse. Procedures for traverses are described in Section 9.4. Ps3 is used in determining the density at the traverse plane. Measure the area of traverse plane, A3, which is located at the tip of the Pitot-static tube. 3. Determine Ps2 for each fan by averaging the pressure measurements at each of four static pressure taps located in the short length of duct

between the outlet damper and the plenum. See Annex E for details of static pressure taps. Measure td2 in Plane 2 for each fan. 4. For each fan, measure td1 and tw1 in the path of the air flowing into the fan inlet. Determine pb for the general vicinity of the fans. Measure td3 in Plane 3. All of these measurements are used in the determination of densities at the various planes of interest. 5. Measure the fan speed and the motor amps, volts, and if possible, watts for each fan. Record all pertinent motor nameplate data, including volts (NPV) and full load amps (FLA). If the motor power outputs are to be estimated by using the phase current method described in Annex K, it is not necessary to measure motor watts; however, it may be necessary to disconnect the drives and measure the no load amps (NLA) if the motors are not operating at or near their full load points. Refer to Annex K. 6. SEF 1 is due to the effect of insufficient length of duct between the outlet of each fan and the plenum. In this case, the duct length is so short as to be judged equivalent to there being no duct at all. In order to calculate the value of SEF 1, it is necessary to measure the outlet areas of the fans, A2, and their blast areas. Field Performance Measurement | 175

7. The sum of the static pressure, Ps1, and the velocity pressure, Pv1, at the inlet of a fan with an unrestricted inlet is considered to be equal to the sum of the static pressure, Psx, and the velocity pressure, Pvx, at a point sufficiently distant from the inlet as to be in still air. At this point, the static pressure is zero, and the velocity pressure in still air is zero. Ps1 + Pv1 = Psx + Pvx =0 This consideration, which is the same as that used in the methods for testing fans for performance rating purposes, charges to the fan losses incurred in accelerating the air into the fan inlet and eliminates the inaccuracies which arise in any attempt to measure the velocity pressure and static pressure at the fan inlet. To calculate the Fan Static Pressure: Ps = Ps2 - Ps1 - Pv1 + SEF 1 = Ps2 - (Ps1 + Pv1) + SEF 1 Since Ps1 + Pv1 = 0: Ps = Ps2 + SEF 1 8. In order to compare the test results to the quoted fan curve drawn for operation at 1900 rpm and 0.075 lbm/ft3 density, it is necessary to convert the results to the specified conditions. The basis for the calculations is described in Section 14. OBSERVATIONS SITE MEASUREMENTS pb = td3 = Ps3 = Pv3 = A3 =

29.05 in. Hg 78°F 5.6 in. wg 0.47 in. wg 7.4 ft2

LH Fan td1 = 75°F tw1 = 57°F td2 = 79°F Ps2 = 6.4 in. wg N = 1910 rpm, LH fan speed A2 = 3.2 ft2 Blast Area = 2.25 ft2 RH Fan td1 = 75°F tw1 = 57°F td2 = 79°F 176 | Field Performance Measurement

Ps2 = 6.4 in. wg N = 1890 rpm, RH fan speed A2 = 3.2 ft2 Blast Area = 2.25 ft2 MEASURED MOTOR DATA LH Fan Volts = = Amps = = NLA =

575, 572, 578 575 av 16, 17, 17 16.7 av 7.0

RH Fan Volts = = Amps = = NLA =

575, 574, 573 574 av 15, 16, 16 15.7 av 7.0

MOTOR NAMEPLATE DATA LH Fan 25 hp, 3 phase, 60 hertz 575 volts, 1780 rpm, 23 FLA RH Fan 25 hp, 3 phase, 60 hertz 575 volts, 1780 rpm, 23 FLA GENERAL Outlet dampers in full open positions. Fans connected to motors through belt drives. CALCULATIONS DENSITIES For inlet conditions for both fans of: td1 = 75°F tw1 = 57°F p1 = pb = 29.05 in. Hg Use Figure N.1 in Annex N to obtain ρ1 = 0.0718 lbm/ft3 The density at Plane 2:

⎛ P + 13.6 pb ⎞ ⎛ t d1 + 460 ⎞ ρ2 = ρ1 ⎜ s2 ⎟ ⎟⎜ ⎝ 13.6 p1 ⎠ ⎝ t d2 + 460 ⎠ ⎛ 6.4 + 13.6 × 29.05 ⎞ ⎛ 535 ⎞ = 0.0718 ⎜ ⎟⎜ ⎟ ⎝ 13.6 × 29.05 ⎠ ⎝ 539 ⎠ = 0.0724 lbm/ft 3

Annex K indicates that the average of the results of Equation A and Equation B will provide a reasonably accurate estimate of motor power output for a 25 hp motor operating at approximately 70% FLA. LH Fan Eqn A = 25 (16.7/23) (575/575) = 18.15 hp

The density at Plane 3: ⎛ P + 13.6 pb ⎞ ⎛ t d1 + 460 ⎞ ρ3 = ρ1 ⎜ s3 ⎟ ⎟⎜ ⎝ 13.6 p1 ⎠ ⎝ t d3 + 460 ⎠ ⎛ 5.6 + 13.6 × 29.05 ⎞ ⎛ 535 ⎞ = 0.0718 ⎜ ⎟⎜ ⎟ ⎝ 13.6 × 29.05 ⎠ ⎝ 538 ⎠ = 0.0724 lbm/ft 3

Eqn B = 25 [(16.7 - 7)/(23 - 7)] (575/575) = 15.16 hp Hmo

= (18.15 + 15.16)/2 = 16.66 hp

RH Fan Eqn A = 25 (15.7/23) (574/575) = 17.04 hp

FLOW RATES V3 = 1096 (Pv3/ρ3)0.5 = 1096 (0.47/0.0724)0.5 = 2792 fpm

Eqn B = 25 [(15.7 - 7)/(23 - 7)] (574/575) = 13.57 hp Hmo

= (17.04 + 13.57)/2 = 15.31 hp

Q3 = V3A3 = 2792 × 7.4 = 20661 cfm

Figure L.1 in Annex L indicates estimated belt drive loss of 5% for each fan.

Q = Q1 = Q3 (ρ3/ρ1) = 20661 (0.0724/0.0718) = 20834 cfm Assume that the air flow rate for each fan is proportional to its speed.

LH Motor HL = 0.05 Hmo = 0.05 × 16.66 = 0.83 hp H = Hmo - HL = 16.66 - 0.83 = 15.83 hp

LH Fan Q = Q1 = 20834 [1910/(1910 + 1890)] = 10472 cfm RH Fan Q = Q1 = 20834 [1890/(1910 + 1890)] = 10362 cfm FAN POWER INPUT LH Fan Measured amps/FLA = (16.7/23) = 0.73 = 73% RH Fan Measured amps/FLA = (15.7/23) = 0.68 = 68%

RH Motor HL = 0.05 Hmo = 0.05 × 15.31 = 0.77 hp H = Hmo - HL = 15.31 - 0.77 = 14.54 hp SYSTEM EFFECT FACTOR AMCA Publication 201-90, Figures 7.1 and 8.3 indicate the following calculations: LH Fan Q2 = Q1 (ρ1/ρ2) = 10472 (0.0718/0.0724) = 10385 cfm V2 = (Q2/A2) = (10385/3.2) = 3245 fpm Field Performance Measurement | 177

Blast area ratio = Blast Area/A2 = 2.25/3.2 = 0.70 RH Fan Q2 = Q1 (ρ1/ρ2) = 10362 (0.0718/0.0724) = 10276 cfm V2 = (Q2/A2) = (10276/3.2) = 3211 fpm Blast area ratio = Blast Area/A2 = 2.25/3.2 = 0.70 For a blast area ratio of 0.7 and no duct, Figure 8.3 shows System Effect Curve S applies. For each fan with velocities of 3245 fpm and 3211 fpm and curve S, Figure 7.1 shows SEF 1 = 0.5 in. wg at 0.075 lbm/ft3. At 0.0724 lbm/ft3: SEF 1 = 0.5 (0.0724/0.075) = 0.48 in. wg FAN STATIC PRESSURE Ps = Ps2 + SEF 1 LH Fan Ps = 6.4 + 0.48 = 6.88 in. wg RH Fan Ps = 6.4 + 0.48 = 6.88 in. wg

178 | Field Performance Measurement

CONVERSION TO SPECIFIED CONDITIONS LH Fan Qc = 10472 (1900/1910) = 10417 cfm Psc = 6.88 (1900/1910)2 (0.075/0.0718) = 7.11 in. wg Hc = 15.83 (1900/1910)3 (0.075/0.0718) = 16.28 hp RH Fan Qc = 10362 (1900/1890) = 10417 cfm Psc = 6.88 (1900/1890)2 (0.075/0.0718) = 7.26 in. wg Hc = 14.54 (1900/1890)3 (0.075/0.0718) = 15.43 hp

EXAMPLE 2A: UTILITY FAN IN A VENTILATION SYSTEM

3 STATIC PRESSURE TAPS 1

2

PLAN VIEW

L

3-PIECE ELBOW R/D = 1 SEF 1

SEF 2

SIDE VIEW

OUTLET SIDE VIEW

COMMENTS 1. Determine Pv3 by using the root mean square of the velocity pressure measurements made in a traverse of Plane 3, located near the end of a straight run of duct, such as shown in the diagram. Determine Ps3 by averaging the static pressure measurements made in the same traverse. Procedures for traverses are described in Section 9.4. Ps3 is used in determining the density at the traverse plane. Measure the area of the traverse plane, A3, which is located at the tip of the Pitot-static tube. 2. Determine Ps1 by averaging the pressure measurements at each of four static pressure taps in the collar connection at the fan inlet. Determine Ps2 by averaging the pressure measurements at each of four static pressure taps located near the fan outlet. 3. Measure td3 and tw3 in the traverse plane. Assume td1 is equal to td3. Determine pb for the general vicinity of the fan. Measure td2 in Plane 2. All of these measurements are used in determining densities at the various planes of interest. 4. Measure the fan speed and the motor amps, volts, and if possible, watts. Record all pertinent motor nameplate data, including volts (NPV) and full load amps (FLA). If the motor power output is to be estimated by using the phase current method

described in Annex K, it is not necessary to measure motor watts; however, it may be necessary to disconnect the drive and measure the no load amps (NLA) if the motor is not operating at or near its full load point. Refer to Annex K. 5. SEF 1 is due to the effect of the elbow located at the fan inlet. SEF 2 is due to the effect of insufficient length of duct between the fan outlet and the elbow downstream of the fan. In order to calculate the values of the SEFs, it is necessary to measure the inlet area and the outlet area of the fan, A1 and A2; the length of the outlet duct, L; and the blast area of the fan. 6. To calculate the Fan Static Pressure: Ps = Ps2 - Ps1 - Pv1 + SEF 1 + SEF 2 Where: Pv1 = Pv3 (A3/A1)2 (ρ3/ρ1) 7. In order to compare the test results to the quoted fan curve drawn for operation at 1880 rpm and 0.075 lbm/ft3 density, it is necessary to convert the results to the specified conditions. The basis for the calculations is described in Section 14.

Field Performance Measurement | 179

OBSERVATIONS SITE MEASUREMENTS pb = 29.20 in. Hg td2 = 72°F td3 = 72°F tw3 = 66°F Ps1 = -2.18 in. wg Ps2 = 0.35 in. wg Ps3 = -1.95 in. wg Pv3 = 0.45 in. wg N = 1730 rpm A1 = 1.07 ft2 A2 = 1.17 ft2 A3 = 1.07 ft2 Blast Area = 0.7 ft2 L = 0.83 ft MEASURED MOTOR DATA Volts = = Amps = = NLA =

227, 229, 228 228 av 10.2, 10.3, 10.4 10.3 av 7.1

MOTOR NAMEPLATE DATA 5 hp, 3 phase, 60 hertz 230 volts, 1750 rpm, 14 FLA

The density at Plane 1: ⎛ P + 13.6 pb ⎞ ⎛ t d3 + 460 ⎞ ρ1 = ρ3 ⎜ s1 ⎟⎜ ⎟ ⎝ 13.6 p3 ⎠ ⎝ t d1 + 460 ⎠ ⎛ −2.18 + 13.6 × 29.20 ⎞ ⎛ 532 ⎞ = 0.0719 ⎜ ⎟ ⎜ 532 ⎟ 13.6 × 29.06 ⎝ ⎠⎝ ⎠ = 0.0718 lbm/ft 3 The density at Plane 2: ⎛ P + 13.6 pb ⎞ ⎛ t d3 + 460 ⎞ ρ2 = ρ3 ⎜ s2 ⎟⎜ ⎟ ⎝ 13.6 p3 ⎠ ⎝ t d2 + 460 ⎠ ⎛ 0.35 + 13.6 × 29.20 ⎞ ⎛ 532 ⎞ = 0.0719 ⎜ ⎟ ⎜ 532 ⎟ 13.6 × 29.06 ⎝ ⎠⎝ ⎠ = 0.0723 lbm/ft 3 FLOW RATES V3 = 1096 (Pv3/ρ3)0.5 = 1096 (0.45/0.0719)0.5 = 2742 fpm Q3 = V3A3 = 2742 × 1.07 = 2934 cfm

GENERAL

Q = = = =

Fan connected to motor through belt drive.

FAN POWER INPUT

CALCULATIONS DENSITIES For Plane 3 conditions of: td3 = 72°F tw3 = 66°F p3 = pb + (Ps3/13.6) = 29.20 + (-1.95/13.6) = 29.06 in. Hg Use Figure N.1 in Annex N to obtain ρ3 = 0.0719 lbm/ft3

Measured amps/FLA = 10.3/14 = 0.74 = 74% Annex K indicates that the average of the results of Equation A and Equation B will provide a reasonably accurate estimate of motor power output for a 5 hp motor operating at 74% FLA. Eqn A = 5 (10.3/14) (228/230) = 3.65 hp Eqn B = 5 [(10.3 - 7.1)/(14 - 7.1)] (228/230) = 2.30 hp Hmo

It is assumed that td1 = td3

180 | Field Performance Measurement

Q1 Q3 (ρ3/ρ1) 2934 (0.0719/0.0718) 2938 cfm

= (3.65 + 2.30)/2 = 2.98 hp

Figure L.1 in Annex L indicates estimated belt drive loss of 6.5%. HL = 0.065 Hmo = 0.065 × 2.98 = 0.19 hp H = Hmo - HL = 2.98 - 0.19 = 2.79 hp SYSTEM EFFECT FACTORS To determine the value of SEF 1, calculate the velocity at the fan inlet: V1 = Q1/A1 = 2938/1.07 = 2746 fpm

L in % effective duct length = (L/3.05) 100 = (0.83/3.05) 100 = 27% Blast area ratio = Blast Area/A2 = 0.7/1.17 = 0.6 For blast area ratio of 0.6, 27% effective duct length and elbow position C, Figure 8.5 shows System Effect Curve P - Q applies. For 2494 fpm velocity and curve P - Q, Figure 7.1 shows SEF 2 = 0.7 in. wg at 0.075 lbm/ft3. At 0.0723 lbm/ft3: SEF 2 = 0.7 (0.0723/0.075) = 0.67 in. wg FAN STATIC PRESSURE

AMCA Publication 201-90, Figure 9.5 indicates that for a three piece elbow with radius to diameter ratio of 1, and with no duct between the elbow and the fan inlet, System Effect Curve R applies. For 2746 fpm velocity and curve R, Figure 7.1 shows SEF 1 = 0.55 in. wg at 0.075 lbm/ft3. At 0.0718 lbm/ft3: SEF 1 = 0.55 (0.0718/0.075) = 0.53 in. wg

Pv1 = Pv3 (A3/A1)2 (ρ3/ρ1) = 0.45 (1.07/1.07)2 (0.0719/0.0718) = 0.45 in. wg Ps = Ps2 - Ps1 - Pv1 + SEF 1 + SEF 2 = 0.35 - (-2.18) - 0.45 + 0.53 + 0.67 = 3.28 in. wg CONVERSION TO SPECIFIED CONDITIONS

For SEF 2, AMCA Publication 201-90, Figures 7.1, 8.1, and 8.5 indicate the following calculations: Q2 = Q3 (ρ3/ρ2) = 2934 (0.0719/0.0723) = 2918 cfm V2 = (Q2/A2) = 2918/1.17 = 2494 fpm

Qc = 2938 (1880/1730) = 3193 cfm Psc = 3.28 (1880/1730)2 (0.075/0.0718) = 4.05 in. wg Hc = 2.79 (1880/1730)3 (0.075/0.0718) = 3.74 hp

Duct diameter equivalent to the fan outlet area: De2 = (4A2/π)0.5 = (4 × 1.17/π)0.5 = 1.22 ft Figure 8.1 shows that for velocities of 2500 fpm or less, the 100% effective outlet duct length is 2.5 duct diameters, = 2.5 × 1.22 = 3.05 ft

Field Performance Measurement | 181

EXAMPLE 2B: CENTRIFUGAL FAN IN A SAWDUST CONVEYING SYSTEM

SEF 2 2

1 SEF 1 4-PIECE ELBOW R/D = 1

L2 L1

3 SIDE VIEW

OUTLET SIDE VIEW

COMMENTS 1. Determine Pv3 by using the root mean square of the velocity pressure measurements made in a traverse of Plane 3, located near the end of a straight run of duct, such as shown in the diagram. Determine Ps3 by averaging the static pressure measurements made in the same traverse. Procedures for traverses are described in Section 9.4. Ps3 is used in determining the density at the traverse plane. Measure the area of the traverse plane, A3, which is located at the tip of the Pitot-static tube. 2. Determine Ps1 by using a Pitot-static tube or static pressure taps in the duct connection at the fan inlet. If a Pitot-static tube is used, it should not project into the upstream elbow but be located well within the length of the duct connection as shown in the diagram. The friction loss in the short length of outlet duct is assumed to be negligible, and Ps2 is considered to be equal to the static pressure at the duct outlet. The static pressure at the outlet of the duct is zero gauge pressure, referred to the atmospheric pressure in the region of the duct outlet. In situations such as this example, the air may be discharging from the duct into a region in which the atmospheric pressure is somewhat different from that to which all other pressure measurements are referred. When this possibility exists, it is essential that the static pressure in the region of the discharging air be measured, referred to the same 182 | Field Performance Measurement

atmospheric pressure as used in all other pressure measurements. In this case, the pressure was measured as 0.1 in. wg. 3. Measure td3 and tw3 in the traverse plane. Determine pb for the general vicinity of the fan. Measure td1 and td2. All of these measurements are used in determining densities at the various planes of interest. 4. Measure the fan speed and the motor amps, volts, and if possible, watts. Record all pertinent motor nameplate data, including volts (NPV) and full load amps (FLA). If the motor power output is to be estimated by using the phase current method described in Annex K, it is not necessary to measure motor watts; however, it may be necessary to disconnect the drive and measure the no load amps (NLA) if the motor is not operating at or near its full load point. Refer to Annex K. 5. SEF 1 is due to the effect of insufficient length of duct between the fan inlet and the elbow upstream of the fan. SEF 2 is due to the effect of insufficient length of duct at the fan outlet. In order to calculate the values of the SEFs, it is necessary to measure the inlet area and the outlet area of the fan, A1 and A2; the lengths of the inlet connection and the outlet duct, L1 and L2; and the blast area of the fan.

6. To calculate the Fan Static Pressure:

CALCULATION

Ps = Ps2 - Ps1 - Pv1 + SEF 1 + SEF 2

DENSITIES

Where:

For Plane 2 conditions of:

Pv1 = Pv3 (A3/A1)2 (ρ3/ρ1)

td2 = 91.3°F tw2 = 70.4°F

7. In order to compare the test results to the quoted fan curve drawn for operation at 2075 rpm and 0.075 lbm/ft3, it is necessary to convert the results to the specified conditions. The basis for the calculations is described in Section 14. OBSERVATIONS SITE MEASUREMENTS pb td1 td2 tw2 td3 Ps1 Ps2 Ps3 Pv3 N A1 A2 A3

= = = = = = = = = = = = =

29.82 in. Hg 86.6°F 91.3°F 70.4°F 86°F -11.4 in. wg 0.1 in. wg -8.9 in. wg 1.24 in. wg 2120 rpm, fan speed 1.40 ft2 1.40 ft2 1.57 ft2

Blast Area = 1.26 ft2 L1 = 1.33 ft L2 = 3.0 ft MEASURED MOTOR DATA Volts = = Amps = = NLA =

460, 460, 459 460 av 26.5, 25.5, 26 26 av 11.3

p2 = pb + (Ps2/13.6) = 29.82 + (0.1/13.6) = 29.83 in. Hg Use Figure N.1 in Annex N to obtain ρ2 = 0.0714 lbm/ft3 The density at Plane 1: ⎛ P + 13.6 pb ⎞ ⎛ t d2 + 460 ⎞ ρ1 = ρ2 ⎜ s1 ⎟ ⎟⎜ ⎝ 13.6 p2 ⎠ ⎝ t d1 + 460 ⎠ 4 + 13.6 × 29.82 ⎞ ⎛ 551.3 ⎞ ⎛ −11.4 = 0.0714 ⎜ ⎟ ⎜ 546.6 ⎟ 13.6 × 29.83 ⎝ ⎠⎝ ⎠ = 0.0700 lbm/ft 3 The density at Plane 3: ⎛ P + 13.6 pb ⎞ ⎛ t d2 + 460 ⎞ ρ3 = ρ2 ⎜ s3 ⎟⎜ ⎟ ⎝ 13.6 p3 ⎠ ⎝ t d3 + 460 ⎠ ⎛ −8.9 + 13.6 × 29.82 ⎞ ⎛ 551.3 ⎞ = 0.0714 ⎜ ⎟ ⎜ 546 ⎟ 13.6 × 29.83 ⎝ ⎠⎝ ⎠ = 0.0705 lbm/ft 3 FLOW RATES V3 = 1096 (Pv3/ρ3)0.5 = 1096 (1.24/0.0705)0.5 = 4596 fpm Q3 = V3A3 = 4596 × 1.57 = 7216 cfm

MOTOR NAMEPLATE DATA 30 hp, 3 phase, 60 hertz 460 volts, 1750 rpm, 36 FLA GENERAL Fan connected to motor through belt drive.

Q = = = =

Q1 Q3 (ρ3/ρ1) 7216 (0.0705/0.0700) 7268 cfm

FAN POWER INPUT Measured amps/FLA = (26/36) = 0.72 = 72%

Field Performance Measurement | 183

Annex K indicates that the average of the results of Equation A and Equation B will provide a reasonably accurate estimate of motor power output for a 30 hp motor operating at 72% FLA. Eqn A = 30 (26/36) (460/460) = 21.67 hp Eqn B = 30 [(26 - 11.3)/(36 - 11.3)] (460/460) = 17.85 hp Hmo

= (21.67 + 17.85)/2 = 19.76 hp

For SEF 2, AMCA Publication 201-90, Figure 8.3 indicates the following calculations: Q2 = Q3 (ρ3/ρ2) = 7216 (0.0705/0.0714) = 7125 cfm V2 = (Q2/A2) = (7125/1.40) = 5089 fpm Duct diameter equivalent to the fan outlet area:

Figure L.1 in Annex L indicates estimated belt drive loss of 4.8%.

De2 = (4A2/π)0.5 = (4 × 1.40/π)0.5 = 1.34 ft

HL = 0.048 Hmo = 0.048 × 19.76 = 0.95 hp

Figure 8.3 shows that for velocities over 2500 fpm, 100% effective duct length is one duct diameter per 1000 fpm:

H = Hmo - HL = 19.76 - 0.95 = 18.81 hp SYSTEM EFFECT FACTORS To determine the value of SEF 1, calculate the velocity at the fan inlet: V1 = (Q1/A1) = (7268/1.40) = 5191 fpm

= D2 (V2/1000) = 1.34 (5089/1000) = 6.82 ft The length of the outlet duct in % effective duct length: = (L2/6.82) 100 = (3.0/6.82) 100 = 44%

The diameter of the fan inlet:

Blast ratio area = Blast Area/A2 = 1.26/1.40 = 0.9

D1 = (4A1/π)0.5 = (4 × 1.40/π)0.5 = 1.34 ft.

For blast area ratio of 0.9 and 44% effective duct length, Figure 8.3 shows no System Effect Curve applies and SEF 2 = 0.

The length of the duct between the elbow and the fan inlet in terms of D1:

FAN STATIC PRESSURE

= (L1/D1) = (1.33/1.34) = 1.0 AMCA Publication 201-90, Figure 9.5 indicates that for a four piece elbow with a radius to diameter ratio of 1, and with a length of duct between the elbow and the fan inlet equal to 1 equivalent diameter, System Effect Curve S applies. For 5191 fpm velocity and curve S, Figure 7.1 shows SEF 1 = 1.3 in. wg at 0.075 lbm/ft3. At 0.0700 lbm/ft3: SEF 1 = 1.3 (0.0700/0.075) = 1.2 in. wg 184 | Field Performance Measurement

Pv1 = Pv3 (A3/A1)2 (ρ3/ρ1) = 1.24 (1.57/1.40)2 (0.0705/0.0700) = 1.57 in. wg Ps = Ps2 - Ps1 - Pv1 + SEF 1 + SEF 2 = 0.1 - (-11.4) - 1.57 + 1.2 + 0 = 11.13 in. wg CONVERSIONS TO SPECIFIED CONDITIONS Qc = 7268 (2075/2120) = 7114 cfm Psc = 11.13 (2075/2120)2 (0.075/0.0700) = 11.42 in. wg

Hc = 18.81 (2075/2120)2 (0.075/0.0700) = 18.90 hp

Field Performance Measurement | 185

EXAMPLE 2C: AXIAL FAN IN A DRYER SYSTEM

5

4 1

STRAIGHTENING VANES 3

2

SEF 2

STATIC PRESSURE TAPS

A3

SEF 1

PLAN VIEW

INNER CYLINDER LOCATION OF PLANE 3

SIDE VIEW

COMMENTS 1. This type of installation is normally classified as one in which a satisfactory test cannot be conducted. Due to the configurations of the airways, there are no locations at which reasonably accurate pressure measurements can be made. In addition, the judgments required in determining the values of the SEFs are susceptible to error. The purpose of presenting this example is to illustrate the not uncommon instance in which a test must be conducted in order to provide performance information, even though the results will be innaccurate to a degree which is not normally acceptable. 2. Determine Pv3 by using the root mean square of the velocity pressure measurements made in a traverse of Plane 3, located as shown in the diagram. Determine Ps3 by averaging the static pressure measurements made in the same traverse. Procedures for traverses are described in Section 9.4. These velocity pressure and static pressure measurements are susceptible to error due to the turbulence existing in the region of the fan outlet. In addition, it is undesirable to have Plane 3 located in a diverging airway. However, no other more suitable location for Plane 3 exists in this example. It is recommended that the Pitot-static tube be oriented so that its nose is aligned with the anticipated flow streams, particularly near the walls of the diffuser. Determine the area of the traverse plane, A3, which is 186 | Field Performance Measurement

located at the tip of the Pitot-static tube, as shown in the diagram, not at the location of the Pitot-static tube access holes. 3. Determine Ps4 by averaging the pressure measurements at each of four static pressure taps located near the fan inlet. In the same manner, determine Ps5 at a location near the fan outlet. It is undesirable to have pressure measurement planes located in converging and diverging airways, but there are no other more suitable locations for these planes in this installation. Measure A4 and A5, the cross-sectional areas of the airways at Planes 4 and 5. 4. Measure td3, tw3, and td4. Determine pb for the general vicinity of the fan. These measurements are used in the determination of densities at the various planes of interest. 5. Measure the fan speed and the motor amps, volts, and if possible, watts. Record all pertinent motor nameplate data, including volts (NPV) and full load amps (FLA). If the motor power output is to be estimated by using the phase current method described in Annex K, it is not necessary to measure motor watts; however, it may be necessary to disconnect the drive and measure the no load amps (NLA) if the motor is not operating at or near its full load point. Refer to Annex K.

6. Although an elbow is located a short distance upstream of the fan, it is considered to produce no system effect since it is equipped with turning vanes and the average velocity through the elbow will be relatively low due to its large cross-sectional area. Therefore, SEF 1 = 0. In judging SEF 2, the rapidly diverging transition fitting downstream of the fan is considered equivalent to no duct at the fan outlet. In order to calculate the value of SEF2, it is necessary to measure the outlet area of the fan, A2.

MOTOR NAMEPLATE DATA

7. To calculate the Fan Static Pressure,

DENSITIES

Ps = Ps2 - Ps1 - Pv1 + SEF 1 + SEF 2

For Plane 3 conditions of:

Where:

td3 = 86.5°F tw3 = 75.5°F

25 hp, 3 phase, 60 hertz 460 volts, 1750 rpm, 31 FLA GENERAL Fan connected to motor through belt drive CALCULATIONS

Pv1 = Pv3 (A3/A1)2 (ρ3/ρ1) Ps1 and Ps2 are calculated on the basis of total pressure considerations, using Ps4, Ps5, and the calculated velocity pressures at Planes 1, 2, 4, and 5. 8. In order to compare the test results to the quoted fan curve drawn for operation at 1580 rpm and 0.0690 lbm/ft3 density, it is necessary to convert the results to the specified conditions. The basis for the calculations is described in Section 14. OBSERVATIONS SITE MEASUREMENTS pb td3 tw3 td4 Ps3 Pv3 Ps4 Ps5 N A1 A3 A4 A5

= 28.90 in. Hg = 86.5°F = 75.5°F = 85°F = 1.5 in. wg = 0.044 in. wg = -1.57 in. wg = 1.22 in. wg = 1590 rpm = A2 = 8.0 ft2 = 29.8 ft2 = 12.4 ft2 = 9.6 ft2

MEASURED MOTOR DATA Volts = = Amps = = NLA =

450, 449, 448 449 av 25.0, 24.5, 25.0 24.8 av 9.4

p3 = pb + (Ps3/13.6) = 28.90 + (1.5/13.6) = 29.01 in. Hg Use Figure N.1 from Annex N to obtain ρ3 = 0.0694 lbm/ft3 The density at Plane 4: ⎛ P + 13.6 pb ⎞ ⎛ t d3 + 460 ⎞ ρ 4 = ρ3 ⎜ s4 ⎟⎜ ⎟ ⎝ 13.6 p3 ⎠ ⎝ t d4 + 460 ⎠ ⎛ −1.57 + 13.6 × 28.90 ⎞ ⎛ 546.5 ⎞ = 0.0694 ⎜ ⎟ ⎜ 545 ⎟ 13.6 × 29.01 ⎝ ⎠⎝ ⎠ = 0.0691 lbm/ft 3 It is assumed that td1 = td4 and at the low pressure levels which exist at Planes 1 and 4, the difference between these pressures will be small, and assuming ρ1 = ρ4, will result in an error which is considered negligible. By similar reasoning, it is assumed that ρ5 = ρ2 = ρ3. FLOW RATES V3 = 1096 (Pv3/ρ3)0.5 = 1096 (0.044/0.0694)0.5 = 873 fpm Q3 = V3A3 = 873 × 29.8 = 26015 cfm Q = = = =

Q1 Q3 (ρ3/ρ1) 26015 (0.0694/0.0691) 26128 cfm Field Performance Measurement | 187

FAN POWER INPUT

FAN STATIC PRESSURE

Measured amps/FLA = (24.8/31) = 0.80 = 80%

Pv4 = Pv3 (A3/A4)2 (ρ3/ρ4) = 0.044 (29.8/12.4)2 (0.0694/0.0691) = 0.26 in. wg

Annex K indicates that the average of the results of Equation A and Equation B will provide a reasonably accurate estimate of motor power output for a 25 hp motor operating at 80% FLA.

Pv1 = Pv3 (A3/A1)2 (ρ3/ρ1) = 0.044 (29.8/8.0)2 (0.0694/0.0691) = 0.61 in. wg

Eqn A = 25 (24.8/31) (449/460) = 19.52 hp

Ps1 + Pv1 = Ps4 + Pv4 Ps1 = Ps4 + Pv4 - Pv1 = -1.57 + 0.26 - 0.61 = -1.92 in. wg

Eqn B = 25 [(24.8 - 9.4)/(31 - 9.4)] (449/460) = 17.40 hp Hmo

= (19.52 + 17.40)/2 = 18.46 hp

Figure L.1 in Annex L indicates estimated belt drive loss of 4.9%. HL = 0.049 Hmo = 0.049 × 18.46 = 0.90 hp H = Hmo - HL = 18.46 - 0.90 = 17.56 hp SYSTEM EFFECT FACTORS SEF 1 = 0 See item 6 under COMMENTS. To determine the value of SEF 2, AMCA Publication 201-90, Figure 8.2 indicates that a vaneaxial fan with no outlet duct will use System Effect Curve U. Q2 = Q3 (ρ3/ρ2) = 26015 (0.0694/0.0694) = 26015 cfm V2 = (Q2/A2) = (26015/8.0) = 3252 fpm From Figure 7.1, using 3252 fpm and curve U, SEF 2 = 0.26 in. wg at 0.075 lbm/ft3. At 0.0694 lbm/ft3: SEF 2 = 0.26 (0.0694/0.075) = 0.24 in. wg

188 | Field Performance Measurement

Pv5 = Pv3 (A3/A5)2 (ρ3/ρ5) = 0.044 (29.8/9.6)2 (0.0694/0.0694) = 0.42 in. wg Pv2 = Pv3 (A3/A2)2 (ρ3/ρ2) = 0.044 (29.8/8.0)2 (0.0694/0.0694) = 0.61 in. wg Ps2 + Pv2 = Ps5 + Pv5 Ps2 = Ps5 + Pv5 - Pv2 = 1.22 + 0.42 - 0.61 = 1.03 in. wg Ps = Ps2 - Ps1 - Pv1 + SEF 1 + SEF 2 = 1.03 - (-1.92) - 0.61 + 0 + 0.24 = 2.58 in. wg Losses between Planes 1 and 4 and between Planes 2 and 5 have been ignored. CONVERSION TO SPECIFIED CONDITIONS Qc = 26128 (1580/1590) = 25964 cfm Psc = 2.58 (1580/1590)2 (0.0690/0.0691) = 2.54 in. wg Hc = 17.56 (1580/1590)3 (0.0690/0.0691) = 17.21 hp

EXAMPLE 2D: CENTRIFUGAL FAN IN A SCRUBBER SYSTEM

INLET BOX DAMPER STATIC PRESSURE TAPS

SEF 1

3 1 L INLET BOX 2 DIFFUSER SIDE VIEW

OUTLET SIDE VIEW

COMMENTS 1. This fan, as supplied and rated by the manufacturer, includes the inlet box damper and the inlet box. Performance ratings for fans with inlet box dampers cover operation with the dampers in the full open positions. In order to be able to compare the test results to the fan performance ratings, it is essential that the damper be fixed in the full open position for the duration of the test. 2. Determine Pv3 by using the root mean square of the velocity pressure measurements made in a traverse of Plane 3, located shortly upstream of the inlet damper. Determine Ps3 by averaging the static pressure measurements made in the same traverse. Procedures for traverses are described in Section 9.4. Measure A3, the area of the traverse plane, located at the tip of the Pitot-static tube and A1, the area of the inlet to the damper. 3. Determine Ps2 by averaging the pressure measurements at each of four static pressure taps located near the end of the fan outlet. See Annex E for details of static pressure taps. 4. Measure td3 and tw3 in the traverse plane. Determine pb for the general vicinity of the fan. Measure td2 in Plane 2. These measurements are used in the determination of densities at the various planes of interest.

5. Measure the fan speed and the motor amps, volts, and if possible, watts. Record all pertinent motor nameplate data, including volts (NPV), and full load amps (FLA). If the motor power output is to be estimated by using the phase current method described in Annex K, it is not necessary to measure motor watts; however, it may be necessary to disconnect the drive and measure the no load amps (NLA) if the motor is not operating at or near its full load point. Refer to Annex K. 6. SEF 1 is due to the effect of insufficient length of duct at the fan outlet. In order to calculate the value of SEF 1, it is necessary to measure the length of the outlet duct, L; the fan outlet area, A2; and the blast area of the fan. 7. To calculate the Fan Static Pressure: Ps = Ps2 - Ps1 - Pv1 + SEF 1 Since Plane 1 is located shortly downstream of Plane 3 in an airway of uniform cross-section (A1 = A3), the conditions which exist at Plane 3 are assumed to exist at Plane 1. Therefore, Ps1 = Ps3 and Pv1 = Pv3. 8. In order to compare the test results to the quoted fan curve drawn for operation at 1780 rpm and 0.059 lbm/ft3 density, it is necessary to convert the results Field Performance Measurement | 189

to the specified conditions. The basis for the calculations is described in Section 14. OBSERVATIONS

pe = 0.5603 in. Hg pp = pe - [p3 (td3 - tw3)/2700] = 0.5603 - [24.28 (63 - 62)/2700] = 0.5513 in. Hg

SITE MEASUREMENTS pb = 29.44 in. Hg td2 = 97°F td3 = 63°F tw3 = 62°F Ps2 = 1.1 in. wg Ps3 = -70.2 in. wg Pv3 = 0.64 in. wg N = 1790 rpm A1 = 6.5 ft2 A2 = 5.32 ft2 A3 = 6.5 ft2 Blast Area = 1.89 ft2 L = 2.50 ft MEASURED MOTOR DATA Volts = 4160, 4150, 4150 = 4153 av Amps = 50, 51, 52 = 51 av NLA = 14 MOTOR NAMEPLATE DATA 500 hp, 3 phase, 60 hertz 4160 volts, 1785 rpm, 61 FLA GENERAL Inlet box damper in full open position. Fan direct connected to motor. CALCULATIONS DENSITIES For Plane 3 conditions of: td3 = 63°F tw3 = 62°F p3 = pb + (Ps3/13.6) = 29.44 + (-70.2/13.6) = 24.28 in. Hg Use the modified Apjohn equation, described in Section M.2.3 in Annex M, and the table in Figure N.2 in Annex N to calculate the density at Plane 3.

190 | Field Performance Measurement

ρ3 = =

1.3257( p3 − 0.378 pp ) t d3 + 460 1.3257 ( 24.28 − 0.378 × 0.5513 )

63 + 460 = 0.0610 lbm/ft 3

Consider ρ1 to be equal to ρ3. The density at Plane 2: ⎛ P + 13.6 pb ⎞ ⎛ t d3 + 460 ⎞ ρ2 = ρ3 ⎜ s2 ⎟⎜ ⎟ ⎝ 13.6 p3 ⎠ ⎝ t d2 + 460 ⎠ ⎛ 1.1 + 13.6 × 29.44 ⎞ ⎛ 523 ⎞ = 0.0610 ⎜ ⎟⎜ ⎟ ⎝ 13.6 × 24.28 ⎠ ⎝ 557 ⎠ = 0.0696 lbm/ft 3 FLOW RATES V3 = 1096 (Pv3/ρ3)0.5 = 1096 (0.64/0.0610)0.5 = 3550 fpm Q3 = = = Q = =

V3A3 3550 × 6.5 23075 cfm Q1 = Q3 23075 cfm

FAN POWER INPUT Measured amps/FLA = 51/61 = 0.84 = 84% Annex K indicates that the average of the results of Equation A and Equation B will provide a reasonably accurate estimate of motor power output for a 500 hp motor operating at 84% FLA. Eqn A = 500 (51/61) (4153/4160) = 417 hp Eqn B = 500 [(51 - 14)/(61 - 14)] (4153/4160) = 393 hp Hmo

= (417 + 393)/2 = 405 hp

Since the fan is direct-connected to the motor, there is no drive loss, and: H = Hmo = 405 hp SYSTEM EFFECT FACTOR AMCA Publication 201-90, Figures 7.1 and 8.3 indicate the following calculations. Q2 = Q3 (ρ3/ρ2) = 23075 (0.0610/0.0696) = 20224 cfm V2 = Q2/A2 = 20224/5.32 = 3802 fpm Duct diameter equivalent to the diffuser outlet area: De2 = (4A2/π)0.5 = (4 × 5.32/π)0.5 = 2.60 ft Figure 8.3 shows that for velocities over 2500 fpm 100% effective duct length is one duct diameter for every 1000 fpm: = De2 (V2/1000) = 2.60 (3802/1000) = 9.89 ft. L in % effective duct length:

Blast area ratio = Blast Area/A2 = 1.89/5.32 = 0.36 For a blast area ratio of 0.36, and 25% effective duct length, Figure 8.3 shows System Effect Curve U applies. For 3802 fpm velocity and curve U, Figure 7.1 shows SEF 1 = 0.36 in. wg at 0.075 lbm/ft3. At 0.0696 lbm/ft3: SEF 1 = 0.36 (0.0696/0.075) = 0.33 in. wg FAN STATIC PRESSURE Ps1 = = Pv1 = = Ps = = =

Ps3 - 70.2 in. wg Pv3 0.64 in. wg Ps2 - Ps1 - Pv1 + SEF 1 1.1 - (-70.2) - 0.64 + 0.33 71.0 in. wg

CONVERSION TO SPECIFIED CONDITIONS Qc = 23075 (1780/1790) = 22946 cfm Psc = 71.0 (1780/1790)2 (0.059/0.0610) = 67.9 in. wg Hc = 405 (1780/1790)3 (0.059/0.0610) = 385 hp

= (L/9.89) 100 = (2.50/9.89) 100 = 25%

Field Performance Measurement | 191

EXAMPLE 2E: CENTRIFUGAL FAN IN A PROCESS SYSTEM

STATIC OUTLET DAMPER PRESSURE TAPS 5 2

INLET BOXES

INLET BOX DAMPERS

1a

1b

3a

3b

OPPOSITE OUTLET SIDE VIEW

SIDE VIEW

COMMENTS 1. This fan, as supplied and rated by the manufacturer, includes the inlet box dampers and the inlet boxes, but does not include the outlet damper. Performance ratings for fans with inlet box dampers cover operation with the dampers in the full open positions. Also, performance ratings for items such as the outlet damper are for operation in the full open position. In order to be able to compare the test results to the fan performance ratings, it is essential that the outlet damper and the inlet dampers be fixed in their full open positions. 2. Determine Pv3a and Pv3b by using the root mean square of the velocity pressure measurements made in Planes 3a and 3b. Determine Ps3a and Ps3b by averaging each of the two sets of static pressure measurements made in the same traverses. Procedures for traverses are described in Section 9.4. Measure A3a and A3b, the areas of the traverse planes and A1a and A1b, the areas of the inlets to the inlet dampers. 3. Determine Ps5 by averaging the pressure measurements of each of four static pressure taps located downstream of the outlet damper.

192 | Field Performance Measurement

4. Measure td3a, td3b, and td5. Since flue gas is being handled by the fan, the Orsat apparatus is used by process personnel to determine the density of the gas. Determine pb for the general vicinity of the fan. These data are used in the determination of densities at the various planes of interest. 5. Measure the fan speed and the motor amps, volts, and if possible, watts. Record all pertinent motor nameplate data, including volts (NPV) and full load amps (FLA). If the motor power output is to be estimated by using the phase current method described in Annex K, it is not necessary to measure motor watts; however, it may be necessary to disconnect the drive and measure the no load amps (NLA) if the motor is not operating at or near its full load point. Motor performance data, supplied by the motor manufacturer, are used in the determination of motor power output for this example. 6. In this example, the duct downstream of the outlet damper is of sufficient length, and no SEF applies. 7. To calculate the Fan Static Pressure: Ps = Ps2 - Ps1 - Pv1

Ps2 is calculated on the basis of total pressure considerations using Ps5, the outlet damper pressure loss, and the calculated velocity pressures at Planes 2 and 5. Since the inlets to the inlet dampers (Planes 1a and 1b) are located shortly downstream of the traverse planes (Planes 3a and 3b) in an airway of uniform cross-section, the conditions which exist at the traverse planes are assumed to exist at the inlets to the inlet dampers. Ps1 = Ps3 = (Ps3a + Ps3b)/2 Pv1 is calculated using the total flow rate and the total area at the inlets to the inlet dampers. Pv1 = (Q1/1096A1

)2

ρ1

8. In order to compare the test results to the quoted fan curve drawn for operation at 880 rpm and 0.049 lbm/ft3 density, it is necessary to convert the results to the specified conditions. The basis for calculations is described in Section 14. OBSERVATIONS SITE MEASUREMENTS = 30.12 in. Hg = 345°F = 359°F = 363°F = -18.8 in. wg = -18.3 in. wg = 2.053 in. wg = 2.028 in. wg = -1.6 in. wg = 892 rpm = A1b = 60.7 ft2 A2 = 115 ft2 A3a = A3b = 60.7 ft2 A5 = 140 ft2 Blast Area = 80 ft2 pb td3a td3b td5 Ps3a Ps3b Pv3a Pv3b Ps5 N A1a

The density of the gas, as determined by Orsat analysis, is 0.0725 lbm/ft3 at 29.92 in. Hg and 70°F. MEASURED MOTOR DATA Volts = = Amps = = kW =

4300, 4250, 4200 4250 av 378, 376, 380 378 av 2519

MOTOR NAMEPLATE DATA 3000 hp, 3 phase, 60 hertz 4000 volts, 880 rpm, 385 FLA GENERAL Inlet box dampers and outlet damper in full open positions. Fan direct connected to motor. Motor efficiency data supplied by motor manufacturer. Pressure loss data supplied by manufacturer of outlet damper. CALCULATIONS DENSITIES The densities at Planes 3a and 3b are: ⎛ P + 13.6 pb ⎞ ⎛ 70 + 460 ⎞ ρ3a = 0.0725 ⎜ s3a ⎟ ⎟⎜ ⎝ 13.6 × 29.92 ⎠ ⎝ t d3a + 460 ⎠ ⎛ −18.8 + 13.6 × 30.12 ⎞ ⎛ 530 ⎞ = 0.0725 ⎜ ⎟ ⎜ 805 ⎟ 13.6 × 29.92 ⎝ ⎠⎝ ⎠ 3 = 0.0458 lbm/ft ⎛ P + 13.6 pb ⎞ ⎛ 70 + 460 ⎞ ρ3b = 0.0725 ⎜ s3b ⎟ ⎟⎜ ⎝ 13.6 × 29.92 ⎠ ⎝ t d3b + 460 ⎠ ⎛ −18.3 + 13.6 × 30.12 ⎞ ⎛ 530 ⎞ = 0.0725 ⎜ ⎟ ⎜ 819 ⎟ 13.6 × 29.92 ⎝ ⎠⎝ ⎠ 3 = 0.0451 lbm/ft It is assumed that ρ1a = ρ3a and ρ1b = ρ3b. The density at Plane 5: ⎛ P + 13.6 pb ⎞ ⎛ 70 + 460 ⎞ ρ5 = 0.0725 ⎜ s5 ⎟ ⎟⎜ ⎝ 13.6 × 29.92 ⎠ ⎝ t d5 + 460 ⎠ ⎛ −1.6 + 13.6 × 30.12 ⎞ ⎛ 530 ⎞ = 0.0725 ⎜ ⎟ ⎜ 823 ⎟ 13.6 × 29.92 ⎝ ⎠⎝ ⎠ = 0.0468 lbm/fft 3 It is assumed that ρ2 = ρ5. FLOW RATES V3a = 1096 (Pv3a/ρ3a)0.5 = 1096 (2.053/0.0458)0.5 = 7338 fpm Q3a = V3aA3a = 7338 × 60.7 = 445417 cfm Field Performance Measurement | 193

V3b = 1096 (Pv3b/ρ3b)0.5 = 1096 (2.028/0.0451)0.5 = 7349 fpm Q3b = V3bA3b = 7349 × 60.7 = 446084 cfm Q3 = Q3a + Q3b = 445417 + 446084 = 891501 cfm Since the air is divided evenly between the two inlet boxes:

FAN STATIC PRESSURE Pv1 = (Q1/1096A1)2 ρ1 = (891501/1096 × 121.4)2 0.0455 = 2.04 in. wg Pv2 = Pv1 (A1/A2)2 (ρ1/ρ2) = 2.04 (121.4/115)2 (0.0455/0.0468) = 2.21 in. wg Pv5 = Pv1 (A1/A5)2 (ρ1/ρ5) = 2.04 (121.4/140)2 (0.0455/0.0468) = 1.49 in. wg

ρ1 = ρ3 = (ρ3a + ρ3b)/2 = (0.0458 + 0.0451)/2 = 0.0455 lbm/ft3

Ps2 + Pv2 = Ps5 + Pv5 + Damper Loss

Q = = = =

Ps1 = = = =

Q1 Q3 (ρ3/ρ1) 891501 (0.0455/0.0455) 891501 cfm

FAN POWER INPUT Measured amps/FLA = (378/385) = 0.98 = 98% Annex K indicates that Equation A will provide a reasonably accurate estimate of motor power output for a 3000 hp motor operating at 98% FLA. Hmo = 3000 (378/385) (4250/4000) = 3130 hp The data supplied by the motor manufacturer indicate motor efficiency of 94% at the measured power input of 2519 kW. Using this information: Hmo = (2519 × 0.94)/0.746 = 3174 hp The more accurate method of estimating the motor power output is assumed to be the latter. Since the fan is direct connected to the motor, there is no drive loss, and: H = Hmo = 3174 hp

194 | Field Performance Measurement

Ps2 = Ps5 + Pv5 + Damper Loss - Pv2 = -1.6 + 1.49 + 0.75 - 2.21 = -1.57 in. wg Ps3 (Ps3a + Ps3b)/2 (-18.8 - 18.3)/2 -18.55 in. wg

Ps = Ps2 - Ps1 - Pv1 = -1.57 - (-18.55) - 2.04 = 14.94 in. wg CONVERSION TO SPECIFIED CONDITIONS Qc = 891501 (880/892) = 879508 cfm Psc = 14.94 (880/892)2 (0.049/0.0455) = 15.66 in. wg Hc = 3174 (880/892)3 (0.049/0.0455) = 3282 hp

EXAMPLE 2F: AXIAL FAN IN A VENTILATION SYSTEM

3 GUIDE VANES

STATIC PRESSURE TAPS 5

4 SEF 1

2-PIECE ELBOW (TYPICAL) INNER CYLINDER

L1 1

SEF 2

L2 2

COMMENTS 1. The unusual duct arrangement in this example makes it very difficult to obtain accurate pressure measurements, and this fact should be understood before testing begins. Also, the use of a diverging inlet fitting and a converging outlet fitting with this fan can pose additional problems. Unless the degrees of divergence and convergence are moderate, as they are in this example, the fan performance will be adversely affected. 2. Determine Pv3 by using the root mean square of the velocity pressure measurements made in a traverse of Plane 3, located well downstream in a straight run of duct, such as shown in the diagram. Determine Ps3 by averaging the static pressure measurements made in the same traverse. Procedures for traverses are described in Section 9.4. Ps3 is used in determining the density at the traverse plane. Measure the area of the traverse plane, A3. 3. Determine Ps5 by averaging the pressure measurements at each of four static pressure taps located near the end of the duct connection at the fan outlet. Determine Ps4 by using static pressure taps in the duct connection at the fan inlet. Measure A4 and A5, the cross-sectional areas of the duct connections at the static pressure taps.

4. Measure td3 and tw3 in the traverse plane. Determine pb for the general vicinity of the fan. Measure td4. These measurements are used in determining densities at the various planes of interest. 5. Measure the fan speed, motor amps, volts, and if possible, watts. Record all pertinent motor nameplate data, including volts (NPV) and full load amps (FLA). If the motor power output is to be estimated by using the phase current method described in Annex K, it is not necessary to measure motor watts; however, it may be necessary to disconnect the drive and measure the no load amps (NLA) if the motor is not operating at or near its full load point. Motor performance data, supplied by the motor manufacturer, are used in the determination of motor power output for this example. 6. SEF 1 is due to the effect of insufficient length of duct between the fan inlet and the elbow upstream of the fan. SEF 2 is due to the effect of insufficient length of duct between the fan outlet and the elbow downstream of the fan. In order to calculate the values of the SEFs, it is necessary to measure the inlet area and the outlet area of the fan, A1 and A2; and the lengths of the inlet and outlet duct connections, L1 and L2.

Field Performance Measurement | 195

7. To calculate the Fan Static Pressure:

460 volts, 1760 rpm, 24.6 FLA

Ps = Ps2 - Ps1 - Pv1 + SEF 1 + SEF 2

GENERAL

Where:

Fan direct connected to motor. Motor efficiency data supplied by motor manufacturer. Fan speed measurement was not obtained due to the closed duct arrangements on both sides of the fan. The measured amps indicate that the motor is operating very close to the full load condition, so the rpm was assumed to be the motor nameplate value of 1760.

Pv1 = Pv3 (A3/A1)2 (ρ3/ρ1) Ps2 and Ps1 are calculated using measured static pressure values and constant total pressure considerations. Ps1 + Pv1 = Ps4 + Pv4 Ps2 + Pv2 = Ps5 + Pv5 Where each velocity pressure is calculated in a manner similar to the calculation of Pv1, shown above. 8. In order to compare the test results to the quoted fan curve drawn for operation at 1750 rpm and 0.075 lbm/ft3 density, it is necessary to convert the results to the specified conditions. The basis for the calculations is described in Section 14. OBSERVATIONS SITE MEASUREMENTS pb td3 tw3 td4 Ps3 Pv3 Ps4 Ps5 A1 A3 A4 L1 L2

= = = = = = = = = = = = = = =

29.76 in. Hg 82.8°F 57.2°F 80°F 0.5 in. wg 0.783 in. wg -1.1 in. wg 0.82 in. wg A2 7.1 ft2 A5 4.91 ft2 6.2 ft2 3.0 ft 3.5 ft

CALCULATIONS DENSITIES For Plane 3 conditions of: td3 = 82.8°F tw3 = 57.2°F p3 = pb + (Ps3/13.6) = 29.76 + (0.5/13.6) = 29.80 in. Hg Use Figure N.1 in Annex N to obtain ρ3 = 0.0728 lbm/ft3. It is assumed ρ2 = ρ5 = ρ3. The density at Planes 1 and 4:

ρ1 = ρ 4 ⎛ P + 13.6 pb ⎞ ⎛ t d3 + 460 ⎞ = ρ3 ⎜ s4 ⎟⎜ ⎟ ⎝ 13.6 p3 ⎠ ⎝ t d4 + 460 ⎠ ⎛ −1.1 + 13.6 × 29.76 ⎞ ⎛ 542.8 ⎞ = 0.0728 ⎜ ⎟ ⎜ 540 ⎟ 13.6 × 29.80 ⎝ ⎠⎝ ⎠ = 0.0729 lbm/ft 3 FLOW RATES V3 = 1096 (Pv3/ρ3)0.5 = 1096 (0.783/0.0728)0.5 = 3594 fpm

MEASURED MOTOR DATA Volts = = Amps = = kW =

460, 461, 459 460 av 25.0, 25.0, 24.8 24.9 av 18.0

MOTOR NAMEPLATE DATA 20 hp, 3 phase, 60 hertz 196 | Field Performance Measurement

Q3 = V3A3 = 3594 × 4.91 = 17647 cfm Q = = = =

Q1 Q3 (ρ3/ρ1) 17647 (0.0728/0.0729) 17623 cfm

FAN POWER INPUT

Diameter of the fan outlet:

The data supplied by the motor manufacturer indicate motor efficiency of 87.5% at the measured power input of 18.0 kW. Using this information:

D2 = (4A2/π)0.5 = (4 × 7.1/π)0.5 = 3.01 ft

Hmo = (18.0 × 0.875)/0.746 = 21.1 hp

Figure 8.1 shows that for velocities of 2500 fpm or less, the 100% effective duct length is 2.5 diameters:

Since the fan is direct connected to the motor, there is no drive loss, and: H = Hmo = 21.1 hp SYSTEM EFFECT FACTORS To determine the value of SEF 1, calculate the velocity at the fan inlet: V1 = (Q1/A1) = (17623/7.1) = 2482 fpm

= 2.5 × 3.01 = 7.53 ft The length of the outlet duct in % effective duct length: = (L2/7.53) 100 = (3.5/7.53) 100 = 46% From Figure 8.4, for a vaneaxial fan with a 46% effective duct length between its discharge and a two piece elbow, System Effect Curve W applies. From Figure 7.1 for 2485 fpm velocity and curve W, SEF 2 is less than 0.1 in. and is considered negligible.

Diameter of the fan inlet: SEF 2 = 0.00 D1 = (4A1/π)0.5 = (4 × 7.1/π)0.5 = 3.01 ft The length of the duct between the elbow and the fan inlet in terms of the fan inlet diameter: = (L1/D1) = (3.0/3.01) = 1.00 AMCA Publication 201-90, Figure 9.2 indicates that for a two piece elbow with a length of duct between the elbow and the fan inlet equal to 1.00 diameter System Effect Curve S-T applies. For a velocity of 2482 fpm and curve S-T, Figure 7.1 shows SEF 1 = 0.25 in. wg at 0.075 lbm/ft3. At 0.0729 lbm/ft3: SEF 1 = 0.25 (0.0729/0.075) = 0.24 in. wg For SEF 2, AMCA Publication 201-90, Figures 7.1, 8.1, and 8.4 indicate the following calculations: Q2 = Q3 (ρ3/ρ2) = 17647 (0.0728/0.0728) = 17647 cfm V2 = Q2/A2 = 17647/7.1 = 2485 fpm

FAN STATIC PRESSURE Pv5 = Pv3 (A3/A5)2 (ρ3/ρ5) = 0.783 (4.91/4.91)2 (0.0728/0.0728) = 0.783 in. wg Pv2 = Pv3 (A3/A2)2 (ρ3/ρ2) = 0.783 (4.91/7.1)2 (0.0728/0.0728) = 0.37 in. wg Ps2 + Pv2 = Ps5 + Pv5 Ps2 = Ps5 + Pv5 - Pv2 = 0.82 + 0.783 - 0.37 = 1.23 in. wg Pv4 = Pv3 (A3/A4)2 (ρ3/ρ4) = 0.783 (4.91/6.2)2 (0.0728/0.0729) = 0.49 in. wg Pv1 = Pv3 (A3/A1)2 (ρ3/ρ1) = 0.783 (4.91/7.1)2 (0.0728/0.0729) = 0.37 in. wg Ps1 + Pv1 = Ps4 + Pv4 Ps1 = Ps4 + Pv4 - Pv1 = -1.1 + 0.49 - 0.37 = -0.98 in. wg

Field Performance Measurement | 197

Ps = Ps2 - Ps1 - Pv1 + SEF 1 + SEF 2 = 1.23 - (-0.98) - 0.37 + 0.24 + 0 = 2.08 in. wg CONVERSION TO SPECIFIED CONDITIONS Qc = 17623 (1750/1760) = 17523 cfm Psc = 2.08 (1750/1760)2 (0.075/0.0729) = 2.12 in. wg Hc = 21.1 (1750/1760)3 (0.075/0.0729) = 21.3 hp

198 | Field Performance Measurement

EXAMPLE 2G: HIGH PRESSURE CENTRIFUGAL FAN IN A SERIES

3

2b

STATIC PRESSURE TAPS 1b 2a

FAN B

1a

DAMPER

INLET BOX

INLET BOX

FAN A SIDE VIEW

COMMENTS 1. The two single inlet fans in this example have been rated by the manufacturer as a two stage assembly. Although rated as an assembly, sufficient measurements are made to provide performance data for each fan. The damper downstream of the second fan is not included as part of the rated assembly. In virtually all cases in which an air flow control damper, such as the one shown in the diagram, is included in the system, the point of operation of major interest and for which the fan has been selected is at the maximum air flow rate. This example is no exception. Therefore, it is essential that the damper be fixed in its full open position for the duration of the test. 2. Determine Pv3 by using the root mean square of the velocity pressure measurements made in a traverse of Plane 3. Determine Ps3 by averaging the static pressure measurements made in the same traverse. Procedures for traverses are described in Section 9.4. Ps3 is used in determining the density at the traverse plane, A3, which is located at the tip of the Pitot-static tube. 3. Determine the static pressures at Planes 1a, 1b2a, and 2b. As shown in the diagram, these planes are located shortly downstream of the inlets and outlets of the fans, which are the planes of interest. In each case, the conditions which exist at the plane of

measurements are assumed to exist at the respective plane of interest because of the close proximity and the fact that the two planes are equal in area. The static pressure at each plane may be determined by averaging the static pressure measurements at each of four static pressure taps, or by averaging the static pressure measurements made in a Pitot-static tube traverse of the plane. However, due to the turbulence existing in the regions of the outlets of the fans, it is recommended that static pressure taps be used at Planes 1b-2a and 2b. 4. Measure td3, tw3, td1b, and td2b; td1a is assumed to be equal to td3. Determine pb for the general vicinity of the fan. These measurements are used in determining densities at the planes of interest. 5. Measure the fan speed and the motor amps, volts, and if possible, watts for each fan. Record all pertinent motor nameplate data, including volts (NPV), and full load amps (FLA). If the motor power outputs are to be estimated by using the phase current method described in Annex K, it is not necessary to measure motor watts; however, it may be necessary to disconnect the drives and measure the no load amps (NLA) if the motors are not operating at or near their full load points. In this example, a watts input measurement is made for Field Performance Measurement | 199

each motor and motor performance data, supplied by the motor manufacturer, are used in determining motor power outputs. 6. The SEF which would normally be attributed to insufficient length of duct at the outlet of the first stage fan does not apply in this case because the fans have been rated as an assembly. 7. To calculate the static pressure for the two stage assembly:

Second Stage Volts = 4080, 4040, 4020 = 4047 av Amps = 44, 44.5, 45 = 44.5 av kW = 272 MOTOR NAMEPLATE DATA Data identical for each stage: 350 hp, 3 phase, 60 hertz 4000 volts, 1790 rpm, 44.5 FLA

Ps = Ps2b - Ps1a - Pv1a GENERAL Where: Pv1a = Pv3 (A3/A1a

)2

(ρ3/ρ1a)

8. In order to compare the test results to the performance quoted for the two stage assembly for operation at 1780 rpm and 0.045 lbm/ft3 density, it is necessary to convert the results to the specified conditions. The basis for the calculations is described in Section 14. OBSERVATIONS SITE MEASUREMENTS = = = = = td2b = Pv3 = Ps3 = Ps1b = = Ps2b = Na = Nb = A1a = = A3 = pb td3 tw3 td1b

28.64 in. Hg 35°F 33°F td2a 95°F 147°F 0.745 in. wg -150 in. wg Ps2a -79.5 in. wg 0.5 in. wg 1790 rpm, first stage fan speed 1790 rpm, second stage fan speed A2a = A1b = A2b 5.6 ft2 4.92 ft2

MEASURED MOTOR DATA First Stage Volts = = Amps = = kW =

4000, 4040, 4080 4040 av 44.5, 45, 45.5 45 av 278

200 | Field Performance Measurement

Fans direct connected to motors. Motor efficiency data supplied by motor manufacturer. CALCULATIONS DENSITIES For Plane 3 conditions of: td3 = 35°F tw3 = 33°F p3 = pb + (Ps3/13.6) = 28.64 + (-150/13.6) = 17.61 in. Hg Use the modified Apjohn equation for partial vapor pressure and the density equation based on perfect gas relationships, both of which are described in Annex M, and the data in Figure N.2 in Annex N to calculate the density at Plane 3. pe = 0.1879 in. Hg p3 (t d3 − t w3 ) 2700 17.61(35 − 33) = 0.1879 − 2700 = 0.1749 in. Hg

pp = pe −

ρ3 = =

1.3257( p3 − 0.378 pp ) t d3 + 460 1.3257 (17.61 − 0.378 × 0.1749 )

35 + 460 = 0.0470 lbm/ft 3 Any conversion of velocity pressure to static pressure which may occur between Planes 3 and 1a can be ignored with no significant effect on the accuracy of the test results. Therefore:

Ps1a = Ps3 = -150 in. wg

FAN POWER INPUT

Assuming no change in temperature between Planes 3a and 1a:

At the measured power input values of 278 kW and 272 kW, the data supplied by the motor manufacturer indicate efficiency of 95% for each motor.

ρ1a = ρ3 = 0.0470 lbm/ft3

Hmoa = (278 × 0.95)/0.746 = 354 hp

To provide information regarding the flow rates between stages and leaving the second stage, additional density values are calculated as follows:

Hmob = (272 × 0.95)/0.746 = 346 hp

ρ1b = ρ2a ⎛ P + 13.6 pb ⎞ ⎛ t d3 + 460 ⎞ = ρ3 ⎜ s1b ⎟⎜ ⎟ ⎝ 13.6 p3 ⎠ ⎝ t d4 + 460 ⎠ ⎛ −79.5 + 13.6 × 28.64 ⎞ ⎛ 495 ⎞ = 0.0470 ⎜ ⎟ ⎜ 555 ⎟ 13.6 × 17.61 ⎝ ⎠⎝ ⎠

Since each fan is direct connected to its motor, there are no drive losses and: Ha

= Hmoa = 354 hp

Hb

= Hmob = 346 hp

= 0.0543 lbm/fft 3

FAN STATIC PRESSURE ⎛ P + 13.6 pb ⎞ ⎛ t d3 + 460 ⎞ ρ2b = ρ3 ⎜ s2b ⎟⎜ ⎟ ⎝ 13.6 p3 ⎠ ⎝ t d2b + 460 ⎠ ⎛ 0.5 + 13.6 × 28.64 ⎞ ⎛ 495 ⎞ = 0.0470 ⎜ ⎟⎜ ⎟ ⎝ 13.6 × 17.61 ⎠ ⎝ 607 ⎠ = 0.0624 lbm/ftt 3 FLOW RATES V3 = 1096 (Pv3/ρ3)0.5 = 1096 (0.745/0.0470)0.5 = 4364 fpm Q3 = V3A3 = 4364 × 4.92 = 21471 cfm Q

= = = =

Q1a Q3 (ρ3/ρ1a) 21471 (0.0470/0.0470) 21471 cfm

Q1b = = = =

Q2a Q3 (ρ3/ρ2a) 21471 (0.0470/0.0543) 18584 cfm

Pv1a = Pv3 (A3/A1a)2 (ρ3/ρ1a) = 0.745 (4.92/5.6)2 (0.0470/0.0470) = 0.575 in. wg The static pressure for the two stage assembly: Ps

= Ps2b - Ps1a - Pv1a = 0.5 - (-150) - 0.575 = 149.9 in. wg

CONVERSION TO SPECIFIED CONDITIONS Qc

= 21471 (1780/1790) = 21351 cfm

Psc

= 149.9 (1780/1790)2 (0.045/0.0470) = 141.9 in. wg

Hac = 354 (1780/1790)3 (0.045/0.0470) = 333 hp Hbc = 346 (1780/1790)3 (0.045/0.0470) = 326 hp

Q2b = Q3 (ρ3/ρ2b) = 21471 (0.0470/0.0624) = 16172 cfm

Field Performance Measurement | 201

EXAMPLE 3A: CENTRIFUGAL FAN IN AN EXHAUST SYSTEM

AIR INTAKE VENTS BACKDRAFT DAMPER SEF 1 3a

2 3c

3b

1 STATIC PRESSURE TAPS PLAN VIEW

COMMENTS 1. This fan, as supplied and rated by the manufacturer, does not include the backdraft damper. 2. Normally, velocity pressure measurements would be made in a single plane, located in a duct common to all branches. In this example, a measurement plane which provides a satisfactory velocity profile cannot be located within the short length of duct between the point of connection of the branch ducts and the fan inlet. The alternative, as indicated in the diagram, is to make a velocity pressure measurement traverse in the longest available duct run of each branch. The velocity pressure for each branch is determined by using the root mean square of the velocity pressure measurements made in the traverse. The static pressure at each traverse plane is determined by averaging the static pressure measurements made in the same traverse. These static pressure values are used in determining the densities at the traverse planes. Procedures for traverses are described in Section 9.4. In order to determine the air flow rates it is necessary to measure the area of each traverse point. 3. Ps1, the static pressure at the fan inlet may be determined by averaging the static pressure measurements at each of four static pressure taps or by averaging the static pressure measurements made in a Pitot-static tube traverse of Plane 1. If a 202 | Field Performance Measurement

Pitot-static tube is used, it should be positioned well within the inlet collar in which Plane 1 is located. Measure the area of Plane 1 for use in calculating Pv1. The static pressure at the outlet of the backdraft damper is zero gauge pressure, referred to the atmospheric pressure in the region of the outlet of the backdraft damper. In situations such as this example, the air may be discharging from the damper into a region in which the atmospheric pressure is somewhat different from that to which all other pressure measurements are referred. When this possibility exists, it is essential that the static pressure in the region of the discharging air be measured, referred to the same atmospheric pressure as used in all other pressure measurements. 4. Measure the dry-bulb and wet-bulb temperatures at each velocity traverse plane and the dry-bulb temperature at Plane 1. In this example, td2 is assumed to be equal to td1. Determine pb for the general vicinity of the fan. These measurements are used in determining densities at the planes of interest. 5. Measure the fan speed and the motor amps, volts, and if possible, watts. Record all pertinent motor nameplate data, including volts (NPV) and full load amps (FLA). If the motor power output is to be

estimated by using the phase current method described in Annex K, it is not necessary to measure motor watts; however, it may be necessary to disconnect the drive and measure the no load amps (NLA) if the motor is not operating at or near its full load point. Refer to Annex K. 6. SEF 1 is due to the effect of there being no duct at the fan outlet. In order to calculate the value of SEF 1, it is necessary to measure the outlet area of the fan, A2, and the blast area of the fan. 7. Determine the backdraft damper pressure loss by using the performance ratings supplied by the manufacturer and the pressure loss multiplier data in Figure 8.7 of AMCA Publication 201-90. The use of the multiplier is indicated because the damper is mounted directly to the fan outlet.

Pv3a Pv3b Pv3c N A1 A2 A3a A3b

= 0.765 in. wg = 0.88 in. wg = 0.86 in. wg = 800 rpm = 16.8 ft2 = 13.8 ft2 = 5.4 ft2 = A3c = 3.0 ft2 Blast Area = 11.0 ft2 MEASURED MOTOR Volts = = Amps = = NLA =

460, 458, 462 460 av 28, 27, 26 27 av 14.7

8. To calculate the Fan Static Pressure: MOTOR NAMEPLATE DATA Ps = Ps2 - Ps1 - Pv1 + SEF 1 Where:

25 hp, 3 phase, 60 hertz 460 volts, 1760 rpm, 32 FLA

Pv1 = (Q1/1096 A1)2 ρ1

GENERAL

Q1 = Q3a (ρ3a/ρ1) + Q3b (ρ3b/ρ1) + Q3c (ρ3c/ρ1)

Fan connected to motor through belt drive. Pressure loss data supplied by manufacturer of backdraft damper.

Ps2 is the sum of the static pressure in the region of the damper outlet, which was measured as zero, and the backdraft damper pressure loss. 9. In order to compare the test results to the quoted fan curve drawn for operation at 810 rpm and 0.075 lbm/ft3 density, it is necessary to convert the results to the specified conditions. The basis for the calculations is described in Section 14. OBSERVATIONS SITE MEASUREMENTS pb td1 tw1 td3a tw3a td3b tw3b td3c tw3c Ps1 Ps3a Ps3b Ps3c

= = = = = = = = = = = = =

29.8 in. Hg 72°F 62°F 77°F 67°F 65°F 56°F 70°F 62°F -1.00 in. wg -0.80 in. wg -0.45 in. wg -0.040 in. wg

CALCULATIONS DENSITIES Since the static pressure values at Planes 1, 3a, 3b, and 3c are very small, no appreciable error will occur by using the barometric pressure instead of the absolute pressure at each plane in the determination of the densities. The densities at these planes are obtained by using Figure N.1 in Annex N.

ρ1 ρ3a ρ3b ρ3c

= = = =

0.0739 0.0731 0.0750 0.0741

lbm/ft3 lbm/ft3 lbm/ft3 lbm/ft3

FLOW RATES V3a = 1096 (Pv3a/ρ3a)0.5 = 1096 (0.765/0.0731)0.5 = 3546 fpm V3b = 1096 (Pv3b/ρ3b)0.5 = 1096 (0.88/0.0750)0.5 = 3754 fpm Field Performance Measurement | 203

V3c = 1096 (Pv3c/ρ3c)0.5 = 1096 (0.86/0.0741)0.5 = 3734 fpm Q3a = V3aA3a = 3546 × 5.4 = 19148 cfm Q3b = V3bA3b = 3754 × 3.0 = 11262 cfm Q3c = V3cA3c = 3734 × 3.0 = 11202 cfm Q = Q1 = Q3a ( ρ3a / ρ1 ) + Q3 b ( ρ3 b / ρ1 ) + Q3c ( ρ3c / ρ1 ) ⎛ 0.0731 ⎞ ⎛ 0.0750 ⎞ ⎛ 0.0741 ⎞ = 19148 ⎜ ⎟ + 11262 ⎜ 0.0739 ⎟ + 11202 ⎜ 0.0739 ⎟ 0 0739 . ⎝ ⎠ ⎝ ⎠ ⎝ ⎠ = 41603 cfm

SYSTEM EFFECT FACTOR AMCA Publication 201-90, Figures 7.1 and 8.3 indicate the following calculations: Q2 = Q1 = 41603 cfm It is assumed that ρ2 = ρ1. V2 = (Q2/A2) = (41603/13.8) = 3015 fpm Blast area ratio = Blast area/A2 = 11.0/13.8 = 0.80 For a blast area ratio of 0.8 and no duct, Figure 8.3 shows System Effect Curve T-U applies. For 3015 fpm velocity and curve T-U, Figure 7.1 shows SEF 1 = 0.27 in. wg at 0.075 lbm/ft3 density. At 0.0739 lbm/ft3:

FAN POWER INPUT

SEF 1 = 0.27 (0.0739/0.075) = 0.27 in. wg

Measured amps/FLA = (27/32) = 0.84 = 84%

BACKDRAFT DAMPER LOSS MULTIPLIER

Annex K indicates that the average of the results of Equation A and Equation B will provide a reasonably accurate estimate of motor power output for a 25 hp motor operating at 84% FLA. Eqn A = 25 (27/32) (460/460) = 21.1 hp Eqn B = 25 [(27 - 14.7)/(32 - 14.7)] (460/460) = 17.8 hp Hmo

= (21.1 + 17.8)/2 = 19.45 hp

The data supplied by the manufacturer of the damper indicate that the pressure loss for the damper, ∆Ps, is 0.4 in. wg at the flow rate of 41603 cfm at 0.075 lbm/ft3 density. AMCA Publication 201-90, Figure 8.7 indicates a ∆Ps multiplier of 1.9 for a damper which is mounted directly to the outlet of a fan which has a blast area ratio of 0.8. Backdraft damper loss = ∆Ps × 1.9 × (ρ2/0.075) = 0.4 × 1.9 (0.0739/0.075) = 0.75 in. wg FAN STATIC PRESSURE

Figure L.1 in Annex L indicates estimated belt drive loss of 4.8%.

Pv1 = (Q1/1096 A1)2 ρ1 = [41603/(1096 × 16.8)]2 0.0739 = 0.38 in. wg

HL = 0.048 Hmo = 0.048 × 19.45 = 0.93 hp

Ps2 is equal to the static pressure at the outlet of the damper, which is zero, plus the damper loss.

H = Hmo - HL = 19.45 - 0.93 = 18.52 hp

204 | Field Performance Measurement

Ps2 = = = Ps = = =

0 + damper loss 0 + 0.75 0.75 in. wg Ps2 - Ps1 - Pv1 + SEF 1 0.75 - (-1.0) - 0.38 + 0.27 1.64 in. wg

CONVERSION TO SPECIFIED CONDITIONS Qc = 41603 (810/800) = 42123 cfm Psc = 1.64 (810/800)2 (0.075/0.0739) = 1.71 in. wg Hc = 18.52 (810/800)3 (0.075/0.0739) = 19.51 hp

Field Performance Measurement | 205

EXAMPLE 3B: AXIAL FAN IN AN EXHAUST SYSTEM 3 2-PIECE ELBOW

SEF 1 L1 1 STATIC PRESSURE TAPS

GUIDE VANES INNER CYLINDER 2 L2

SEF 2 5

PLAN VIEW

COMMENTS 1. Determine Pv3 by using the root mean square of the velocity pressure measurements made in a traverse of Plane 3, located near the end of a straight run of duct, as shown in the diagram. Determine Ps3 by averaging the static pressure measurements made in the same traverse. Procedures for traverses are described in Section 9.4. Ps3 is used in determining the density at the traverse plane. Measure the area of the traverse plane, A3, which is located at the tip of the Pitot-static tube. 2. Determine Ps5 by averaging the pressure measurements at each of four static pressure taps located near the end of the duct connection at the fan outlet. Determine Ps1 by using a Pitot-static tube or static pressure taps in the duct connection at the fan inlet. If a Pitot-static tube is used, it should not project into the upstream elbow but be located well within the length of the duct connection. 3. Measure td3 and tw3 in the traverse plane; td1 is assumed to be equal to td3. Determine pb for the general vicinity of the fan. Measure td5. These measurements are used in determining densities at the planes of interest. 4. Measure the fan speed and the motors amps, volts, and if possible, watts. Record all pertinent 206 | Field Performance Measurement

motor nameplate data, including volts (NPV), and full load amps (FLA). If the motor power output is to be estimated by using the phase current method described in Annex K, it is not necessary to measure motor watts; however, it may be necessary to disconnect the drive and measure the no load amps (NLA) if the motor is not operating at or near its full load point. Refer to Annex K. 5. SEF 1 is due to the effect of insufficient length of duct between the fan inlet and the elbow upstream of the fan. SEF 2 is due to the effect of insufficient length of duct between the fan outlet and the elbow downstream of the fan. In order to calculate the values of the SEFs, it is necessary to measure the inlet area and the outlet area of the fan, A1 and A2; and the lengths of the inlet and outlet duct connections, L1 and L2. 6. To calculate the Fan Static Pressure: Ps = Ps2 - Ps1 - Pv1 + SEF 1 + SEF 2 Where: Pv1 = Pv3 Since: A1 = A3

And:

CALCULATIONS

ρ1 = ρ3

DENSITIES

Due to the close proximity of Planes 2 and 5 and the fact that there is no change in area between the two planes, all conditions which exist at Plane 5 are assumed to exist at Plane 2.

For Plane 3 conditions of:

Therefore:

p3 = pb + (Ps3/13.6) = 29.20 + (-1.92/13.6) = 29.06 in. Hg

Ps2 = Ps5 7. In order to compare the test results to the quoted fan curve drawn for operation at 1730 rpm and 0.075 lbm/ft3 density, it is necessary to convert the results to the specified conditions. The basis for the calculations is described in Section 14. OBSERVATIONS SITE MEASUREMENTS = = = = = = = = = = = L1 = L2 = pb td3 tw3 td5 Ps1 Ps3 Pv3 Ps5 N A1

td3 = 72°F tw3 = 66°F

29.20 in. Hg 72°F 66°F 73°F -2.02 in. wg -1.92 in. wg 0.35 in. wg 0.10 in. wg 1710 rpm A2 = A3 = A5 2.64 ft2 1.5 ft, length of inlet duct 2.25 ft, length of the outlet duct

Use Figure N.1 in Annex N to obtain ρ3 = 0.0719 lbm/ft3. Assume that td1 = td3. ⎛ P + 13.6 pb ⎞ ⎛ t d3 + 460 ⎞ ρ1 = ρ3 ⎜ s1 ⎟⎜ ⎟ ⎝ 13.6 p3 ⎠ ⎝ t d1 + 460 ⎠ ⎛ −2.02 + 13.6 × 29.20 ⎞ ⎛ 532 ⎞ = 0.0719 ⎜ ⎟ ⎜ 532 ⎟ 13.6 × 29.06 ⎝ ⎠⎝ ⎠ 3 = 0.0719 lbm/ft Assume that td2 = td5 and Ps2 = Ps5.

ρ 2 = ρ5 ⎛ P + 13.6 pb ⎞ ⎛ t d3 + 460 ⎞ = ρ3 ⎜ s5 ⎟⎜ ⎟ ⎝ 13.6 p3 ⎠ ⎝ t d5 + 460 ⎠ ⎛ 0.10 + 13.6 × 29.20 ⎞ ⎛ 532 ⎞ = 0.0719 ⎜ ⎟ ⎜ 533 ⎟ 13.6 × 29.06 ⎝ ⎠⎝ ⎠ = 0.0721 lbm/ft 3

MEASURED MOTOR DATA

FLOW RATES

Volts = = Amps = = NLA =

V3 = 1096 (Pv3/ρ3)0.5 = 1096 (0.35/0.0719)0.5 = 2418 fpm

227, 229, 228 228 av 12.2, 12.3, 12.4 12.3 av 7

MOTOR NAMEPLATE DATA 5 hp, 3 phase, 60 hertz 230 volts, 1760 rpm, 14.0 FLA GENERAL Fan connected to motor through belt drive.

Q3 = V3A3 = 2418 × 2.64 = 6384 cfm Q = = = =

Q1 Q3 (ρ3/ρ1) 6384 (0.0719/0.0719) 6384 cfm

Q2 = = = =

Q5 Q3 (ρ3/ρ5) 6384 (0.0719/0.0721) 6366 cfm Field Performance Measurement | 207

FAN POWER INPUT Measured amps/FLA = (12.3/14.0) = 0.88 = 88% Annex K indicates that the average of the results of Equation A and Equation B will provide a reasonably accurate estimate of motor power output for a 5 hp motor operating at 88% FLA.

R-S, Figure 7.1 shows SEF 1 = 0.24 in. wg at 0.075 lbm/ft3 density. At 0.0719 lbm/ft3: SEF 1 = 0.24 (0.0719/0.075) = 0.23 in. wg For SEF 2, AMCA Publication 201-90, Figures 7.1, 8.1, and 8.4 indicate the following calculations:

Eqn A = 5 (12.3/14) (228/230) = 4.35 hp

V2 = (Q2/A2) = (6366/2.64) = 2411 fpm

Eqn B = 5 [(12.3 - 7)/(14 - 7)] (228/230) = 3.75 hp

The diameter of the fan outlet:

Hmo

= (4.35 + 3.75)/2 = 4.05 hp

Figure L.1 in Annex L indicates estimated belt drive loss of 6.3%. HL = 0.063 Hmo = 0.063 × 4.05 = 0.26 hp H = Hmo - HL = 4.05 - 0.26 = 3.79 hp SYSTEM EFFECT FACTORS To determine the value of SEF 1, calculate the velocity at the fan inlet: V1 = (Q1/A1) = (6384/2.64) = 2418 fpm

D2 = (4A2/π)0.5 = (4 × 2.64/π)0.5 = 1.83 ft Figure 8.1 shows that for velocities of 2500 fpm or less, the 100% effective duct length is 2.5 diameters: = 2.5 × 1.83 = 4.58 ft The length of the outlet duct in % effective duct length: = (L2/4.58) 100 = (2.25/4.58) 100 = 49% From Figure 8.4, for a vaneaxial fan with a 49% effective duct length between its discharge and a two piece elbow, System Effect Curve W applies. From Figure 7.1, for 2411 fpm velocity and curve W, SEF 2 is less than 0.1 in. wg, and is considered negligible. SEF 2 = 0.00

Calculate the diameter of the fan inlet: FAN STATIC PRESSURE D1 = (4A1/π)0.5 = (4 × 2.64/π)0.5 = 1.83 ft. Calculate the length of duct between the elbow and the fan inlet in terms of the fan inlet diameter: = (L1/D1) = (1.5/1.83) = 0.82 AMCA Publication 201-90, Figure 9.2, indicates that for a vaneaxial fan with a two piece elbow with a length of duct between the elbow and the fan inlet equal to 0.8 diameters, System Effect Curve R-S (estimated) applies. For 2418 fpm velocity and curve 208 | Field Performance Measurement

Since: A1 = A3 ρ1 = ρ3 Pv1 = Pv3 Ps = Ps2 - Ps1 - Pv1 + SEF 1 + SEF 2 = 0.10 - (-2.02) - 0.35 + 0.23 + 0.00 = 2.00 in. wg

CONVERSION TO SPECIFIED CONDITIONS Qc = 6384 (1730/1710) = 6459 cfm Psc = 2.00 (1730/1710)2 (0.075/0.0719) = 2.14 in. wg Hc = 3.79 (1730/1710)3 (0.075/0.0719) = 4.09 hp

Field Performance Measurement | 209

EXAMPLE 3C: CENTRIFUGAL FAN IN A SCRUBBER SYSTEM 3 WET CELL SCRUBBER

1

2 SEF 1

PLAN VIEW

SIDE VIEW

COMMENTS 1. Determine Pv3 by using the root mean square of the velocity pressure measurements made in a traverse of Plane 3, located in the duct connection at the fan inlet, as shown in the diagram. Determine Ps3 by averaging the static pressure measurements made in the same traverse. Procedures for traverses are described in Section 9.4. Ps3 is used in determining the density at the traverse plane. Measure the area of the traverse plane, A3, which is located at the tip of the Pitot-static tube. In locating Plane 3 downstream of the scrubber, changes in the composition of the air as a result of the action of the scrubber are properly taken into account in the determination of fan air flow rate. Due to the close proximity of Planes 1 and 3, and the fact that there is no change in area between the two planes, the conditions which exist at Plane 3 are assumed to exist at Plane 1.

amps (FLA). If the motor power output is to be estimated by using the phase current method described in Annex K, it is not necessary to measure motor watts; however, it may be necessary to disconnect the drive and measure the no load amps (NLA) if the motor is not operating at or near its full load point. Refer to Annex K.

2. Ps2, the static pressure at the fan outlet, is zero.

Pv1 = Pv3 Ps1 = Ps3 Ps2 = 0

3. Measure td3 and tw3 in the traverse plane. Determine pb for the general vicinity of the fan. Measure td2. These measurements are used in determining densities at the planes of interest. 4. Measure the fan speed and the motor amps, volts, and if possible, watts. Record all pertinent motor nameplate data, including volts (NPV), and full load 210 | Field Performance Measurement

5. SEF 1 is due to the effect of there being no duct at the fan outlet. In order to calculate the value of SEF 1, it is necessary to measure the outlet area of the fan, A2, and the blast area of the fan. 6. To calculate the Fan Static Pressure: Ps = Ps2 - Ps1 - Pv1 + SEF 1 Where:

7. In order to compare the test results to the quoted fan curve drawn for operation at 1700 rpm and 0.071 lbm/ft3 density, it is necessary to convert the results to the specified conditions. The basis for the calculations is described in Section 14.

OBSERVATIONS SITE MEASUREMENTS = 29.80 in. Hg = 65°F = 64°F = 70°F = -8.0 in. wg = 0.337 in. wg = 1672 rpm = A3 = 7.06 ft2 A2 = 5.15 ft2 Blast Area = 3.67 ft2

pb td3 tw3 td2 Ps3 Pv3 N A1

MEASURED MOTOR DATA Volts = = Amps = =

450, 458, 462 457 av 44, 45, 44.5 44.5 av

MOTOR NAMEPLATE DATA 40 hp, 3 phase, 60 hertz 460 volts, 1780 rpm, 49 FLA GENERAL Fan connected to motor through belt drive. CALCULATIONS

⎛ P + 13.6 pb ⎞ ⎛ t d3 + 460 ⎞ ρ2 = ρ3 ⎜ s2 ⎟⎜ ⎟ ⎝ 13.6 p3 ⎠ ⎝ t d2 + 460 ⎠ ⎛ 0 + 13.6 × 29.80 ⎞ ⎛ 525 ⎞ = 0.0732 ⎜ ⎟⎜ ⎟ ⎝ 13.6 × 29.21 ⎠ ⎝ 530 ⎠ = 0.0740 lbm/ft 3 FLOW RATES V3 = 1096 (Pv3/ρ3)0.5 = 1096 (0.337/0.0732)0.5 = 2352 fpm Q3 = V3A3 = 2353 × 7.06 = 16605 cfm Q = = = =

Q1 Q3 (ρ3/ρ1) 16605 (0.0732/0.0732) 16605 cfm

Q2 = Q3 (ρ3/ρ2) = 16605 (0.0732/0.0740) = 16425 cfm FAN POWER INPUT Measured amps/FLA = (44.5/49) = 0.91 = 91%

DENSITIES

Annex K indicates that Equation A will provide a reasonably accurate estimate of motor power output for a 40 hp motor operating at 91% FLA.

For Plane 3 conditions of: td3 = 65°F tw3 = 64°F

Hmo = 40 (44.5/49) (457/460) = 36.1 hp

p3 = pb + (Ps3/13.6) = 29.80 + (-8.0/13.6) = 29.21 in. Hg Use Figure N.1 in Annex N to obtain ρ3 = 0.0732 lbm/ft3. It is assumed that: td1 = td3 Ps1 = Ps3 ρ1 = ρ3

Figure L.1 in Annex L indicates estimate belt drive loss of 4.5%. HL = 0.045 Hmo = 0.045 × 36.1 = 1.6 hp H = Hmo - HL = 36.1 - 1.6 = 34.5 hp SYSTEM EFFECT FACTOR AMCA Publication 201-90, Figures 7.1 and 8.3, indicate the following calculations:

Field Performance Measurement | 211

V2 = (Q2/A2) = (16425/5.15) = 3189 fpm Blast area ratio = Blast area/A2 = 3.67/5.15 = 0.71 For a blast area ratio of 0.7 and no duct, Figure 8.3 shows System Effect Curve S applies. For 3189 fpm velocity and curve S, Figure 7.1 shows SEF 1 = 0.5 in. wg at 0.075 lbm/ft3 density. At 0.0740 lbm/ft3: SEF 1 = 0.5 (0.074/0.075) = 0.49 in. wg FAN STATIC PRESSURE Pv1 = Pv3 = 0.337 in. wg Ps = Ps2 - Ps1 - Pv1 + SEF 1 = 0 - (-8.0) - 0.337 + 0.49 = 8.15 in. wg CONVERSION TO SPECIFIED CONDITIONS Qc = 16605 (1700/1672) = 16883 cfm Psc = 8.15 (1700/1672)2 (0.071/0.0732) = 8.17 in. wg Hc = 34.5 (1700/1672)3 (0.071/0.0732) = 35.2 hp

212 | Field Performance Measurement

EXAMPLE 3D: CENTRIFUGAL ROOF VENTILATOR WITH DUCTED INLET

2 1

BACKDRAFT DAMPER 4 STATIC PRESSURE TAPS

3a

SIDE VIEW

3b

COMMENTS 1. This centrifugal roof ventilator, as supplied and rated by the manufacturer, does not include the backdraft damper. It is essential that the backdraft damper blades be fixed in their full open positions, otherwise uneven velocity distribution will occur at the inlet to the ventilator, adversely affecting its performance. 2. Normally, velocity pressure measurements would be made in a single plane, located in a duct common to all branches. In this example, a measurement plane which provides a satisfactory velocity profile cannot be located within the short length of duct between the point of connection of the branch ducts and the ventilator inlet. The alternative, as indicated in the diagram, is to make a velocity pressure measurement traverse in each branch. The velocity pressure for each branch is determined by using the root mean square of the velocity pressure measurements made in the traverse. The static pressure at each traverse plane is determined by averaging the static pressure measurements made in the same traverse. These static pressure values are used in determining the densities at the traverse planes. Procedures for traverses are described in Section 9.4. In order to determine the air flow rates, it is necessary to measure the area of each traverse plane.

3. Ps4 may be determined by averaging the static pressure measurements at each of four static pressure taps or by averaging the static pressure measurements made in a Pitot-static tube traverse of Plane 4. If a Pitot-static tube is used, it should be positioned well within the duct in which Plane 4 is located, and not project into the upstream elbows. Measure the area of Plane 1 for use in calculating Pv1. In this example, A4 = A1. Ps2, the static pressure at the outlet of the ventilator, is zero gauge pressure, referred to the atmospheric pressure in the region of the ventilator outlet. In situations such as this example, the air may be discharging from the ventilator into a region in which the atmospheric pressure is somewhat different from that to which all other pressure measurements are referred. When this possibility exists, it is essential that the static pressure in the region of the discharging air be measured, referred to the same atmospheric pressure as used in all other pressure measurements. In this case, Ps2 was measured as zero. 4. Measure the dry-bulb and wet-bulb temperatures at each velocity traverse plane. In this example, td1 and td4 are assumed to be equal to td3a. Determine pb for the general vicinity of the fan. These measurements are used in determining densities at the planes of interest. Field Performance Measurement | 213

5. Measure the fan speed and the motor amps, volts, and if possible, watts. Record all pertinent motor nameplate data, including volts (NPV) and full load amps (FLA). If the motor power output is to be estimated by using the phase current method described in Annex K, it is not necessary to measure motor watts; however, it may be necessary to disconnect the drive and measure the no load amps (NLA) if the motor is not operating at or near its full load point. Refer to Annex K.

MEASURED MOTOR DATA Volts = = Amps = =

450, 455, 460 455 av 5.7, 5.85, 5.9 5.82 av

MOTOR NAMEPLATE DATA 5 hp, 3 phase, 60 hertz 460 volts, 1780 rpm, 5.95 FLA

6. Determine the backdraft damper pressure loss by using the performance ratings supplied by the manufacturer.

GENERAL

7. To calculate the Fan Static Pressure:

Fan connected to motor through belt drive. Pressure loss data supplied by manufacturer of backdraft damper.

Ps = Ps2 - Ps1 - Pv1 CALCULATIONS Where: DENSITIES Pv1 = (Q1/1096 A1

)2

ρ1

Q1 = Q3a (ρ3a/ρ1) + Q3b (ρ3b/ρ1)

For Planes 3a and 3b conditions of:

Ps1 = Ps4 - backdraft damper pressure loss Ps2 = 0

td3a = = tw3a = =

td3b 72°F tw3b 66°F

8. In order to compare the test results to the quoted fan curve drawn for operation at 620 rpm and 0.075 lbm/ft3 density, it is necessary to convert the results to the specified conditions. The basis for the calculations is described in Section 14.

p3a = = = =

p3b pb + (Ps3a/13.6) 29.20 + (-0.85/13.6) 29.14 in. Hg

OBSERVATIONS SITE MEASUREMENTS pb = 29.20 in. Hg td3a = td3b = 72°F tw3a = tw3b = 66°F Ps2 = 0 in. wg Ps4 = -0.88 in. wg Ps3a = Ps3b = -0.85 in. wg Pv3a = 0.27 in. wg Pv3b = 0.275 in. wg N = 625 rpm A1 = A4 = 7.9 ft2 A3a = 3.4 ft2 A3b = 3.3 ft2

214 | Field Performance Measurement

Use Figure N.1 in Annex N to obtain:

ρ3a = ρ3b = 0.0721 lbm/ft3 It is assumed that: td1 = td4 = td3a = td3b Since the differences in the static pressures at Planes 1, 3a, and 4 are very small, no appreciable error will occur by assuming:

ρ1 = ρ4 = ρ3a = ρ3b FLOW RATES V3a = 1096 (Pv3a/ρ3a)0.5 = 1096 (0.27/0.0721)0.5 = 2121 fpm

V3b = 1096 (Pv3b/ρ3b)0.5 = 1096 (0.275/0.0721)0.5 = 2140 fpm Q3a = V3aA3a = 2121 × 3.4 = 7211 cfm Q3b = V3bA3b = 2140 × 3.3 = 7062 cfm

BACKDRAFT DAMPER LOSS The data supplied by the manufacturer of the damper indicate that the pressure loss for the damper, ∆Ps, is 0.22 in. wg at the flow rate of 14273 cfm at 0.075 lbm/ft3 density. Backdraft damper loss = ∆Ps (ρ4/0.075) = 0.22 (0.0721/0.075) = 0.21 in. wg FAN STATIC PRESSURE

Q

= = = =

Q1 Q3a (ρ3a/ρ1) + Q3b (ρ3b/ρ1) 7211 (0.0721/0.0721) + 7062 (0.0721/0.0721) 14273 cfm

FAN POWER INPUT Measured amps/FLA = (5.82/5.95) = 0.98 = 98% Annex K indicates that Equation A will provide a reasonably accurate estimate of motor power output for a 5 hp motor operating at 98% FLA.

Pv1 = (Q1/1096 A1)2 ρ1 = [14273/(1096 × 7.9)]2 0.0721 = 0.20 in. wg Ps1 = Ps4 - damper loss = -0.88 - 0.21 = -1.09 in. wg Ps = Ps2 - Ps1 - Pv1 = 0 - (-1.09) - 0.20 = 0.89 in. wg CONVERSION TO SPECIFIED CONDITIONS

Hmo = 5 (5.82/5.95) (455/460) = 4.84 hp

Qc = 14273 (620/625) = 14159 cfm

Figure L.1 in Annex L indicates estimated belt drive loss of 5.8%.

Psc = 0.89 (620/625)2 (0.075/0.0721) = 0.91 in. wg

HL = 0.058 Hmo = 0.058 × 4.84 = 0.28 hp

Hc = 4.56 (620/625)3 (0.075/0.0721) = 4.63 hp

H

= Hmo - HL = 4.84 - 0.28 = 4.56 hp

Field Performance Measurement | 215

EXAMPLE 4A: CENTRIFUGAL FAN IN A BUILT-UP AIR CONDITIONING UNIT 2 4

RETURN AIR

SPRAY SECTION

5

SEF 2

PLAN VIEW

3a FAN SECTION

+

+

OUTSIDE AIR

SEF 1 L

+

+ +

+

+ +

+

+

PREHEAT COILS FILTER SECTION

DIFFUSER PLATE

REHEAT COIL 3b

SIDE VIEW

COMMENTS 1. This is an air conditioning unit which has been assembled at the installation site. The subject of the test is the fan, which is rated by the manufacturer as free-standing, unencumbered by the cabinet in which it has been installed. The fan performance ratings are based on operation with the fan outlet ducted. Before proceeding with the test, it is essential that all dampers--outside air, return air, mixing box, multizone, face and bypass or volume control--be fixed in the positions agreed upon by all interested parties as being applicable for the installation. Also, the temperatures of the heating coils must be kept constant throughout the test period. It may be necessary to lock out, disconnect, or otherwise modify automatic control devices in order to prevent the positions of the dampers and temperatures of the coils from changing during the test. Refer to Section 17.4.3 for additional considerations affecting the test procedure for fans in this type of installation. 2. Normally, velocity pressure measurements would be made in a single plane, located in a duct common to all branches. In this example, a measurement plane which provides a satisfactory velocity profile cannot be located upstream of the fan or between the point of connection of the branch ducts and the fan outlet. The alternative, as indicated in the diagram, is to make a velocity pressure measurement traverse in each branch. The velocity pressure for each branch 216 | Field Performance Measurement

is determined by using the root mean square of the velocity pressure measurements made in the traverse. the static pressure at each traverse plane is determined by averaging the static pressure measurements made in the same traverse. These static pressure values are used in determining the densities at the traverse planes. Procedures for traverses are described in Section 9.4. In order to determine the air flow rates, it is necessary to measure the area of each traverse plane. 3. Determine Ps4 by averaging the static pressure measurements made in a traverse of Plane 4. Determine Ps5 in a similar manner. Pitot-static tube traverses are used in determining these static pressures because the installation of suitable pressure taps is usually prevented by the insulating material encountered in this type of equipment. Due to the abrupt expansion in area from Plane 2 to Plane 5, it is assumed that there is no conversion of velocity pressure at Plane 2 to static pressure at Plane 5. Therefore, it is assumed that Ps2 = Ps5. Measure the area of Plane 4 for use in calculating Pv4. 4. Measure the dry-bulb and wet-bulb temperatures at Plane 4 and the dry-bulb temperatures at Planes 3a, 3b, and 5. Determine pb for the general vicinity of the air conditioning unit. These measurements are used in determining densities at the planes of interest.

5. Measure the fan speed and motor amps, volts, and if possible, watts. Record all pertinent motor nameplate data, including volts (NPV), and full load amps (FLA). If the motor power output is to be estimated by using the phase current method described in Annex K, it is not necessary to measure motor watts; however, it may be necessary to disconnect the drive and measure the no load amps (NLA) if the motor is not operating at or near its full load point. Refer to Annex K. 6. SEF 1 is due to the effect of insufficient distance between the fan inlets and the side walls of the fan cabinet. SEF 2 is attributed to the high degree of divergence of the transition fitting at the fan outlet. The effect created by this fitting is considered to be equivalent to the effect created by having no duct at the fan outlet. In order to determine the values of the SEFs, it is necessary to measure the diameter of an inlet of the fan, the distance between a fan inlet and a side wall of the fan cabinet, and the outlet area and blast area of the fan. 7. To calculate the Fan Static Pressure: Ps = Ps2 - Ps1 - Pv1 + SEF 1 + SEF 2 = Ps2 - (Ps1 + Pv1) + SEF 1 + SEF 2 Where: Ps2 = Ps5 Ps1 + Pv1 = Ps4 + Pv4 Pv4 = (Q4/1096 A4)2 ρ4 Q4 = Q1 = Q3a (ρ3a/ρ1) + Q3b (ρ3b/ρ1) The calculation of Pv4 is often ignored in instances similar to this example on the basis that the calculated value of Pv4 is relatively small and its omission does not affect the test results significantly. 8. In order to compare the test results to the quoted fan curve drawn for operation at 1170 rpm and 0.075 lbm/ft3 density, it is necessary to convert the results to the specified conditions. The basis for the calculations is described in Section 14.

OBSERVATIONS SITE MEASUREMENTS pb = 28.72 in. Hg td3a = 59°F td3b = 90°F td4 = 56°F td5 = 58°F Ps4 = -1.75 in. wg Ps3a = 3.65 in. wg Ps3b = 3.45 in. wg Pv3a = 0.60 in. wg Pv3b = 0.47 in. wg Ps5 = 3.77 in. wg N = 1160 rpm A2 = 18.9 ft2 A3a = 7.2 ft2 A3b = 9.7 ft2 A4 = 93.2 ft2 Blast Area = 13.3 ft2 D1 = 3.92 ft, fan inlet diameter L = 2.83 ft MEASURED MOTOR DATA Volts = = Amps = =

462, 465, 465 464 av 82, 81, 83 82 av

MOTOR NAMEPLATE DATA 75 hp, 3 phase, 60 hertz 460 volts, 1780 rpm, 90.3 FLA GENERAL Fan connected to motor through belt drive. CALCULATIONS DENSITIES For Plane 4 conditions of: td4 = 56°F tw4 = 54°F p4 = pb + (Ps4/13.6) = 28.72 + (-1.75/13.6) = 28.59 in. Hg

Field Performance Measurement | 217

Use Figure N.1 in Annex N to obtain ρ4 = 0.0731 lbm/ft3. It is assumed that ρ1 = ρ4. ⎛ P + 13.6 pb ⎞ ⎛ t d4 + 460 ⎞ ρ5 = ρ 4 ⎜ s5 ⎟ ⎟⎜ ⎝ 13.6 p4 ⎠ ⎝ t d5 + 460 ⎠ ⎛ 3.77 + 13.6 × 28.72 ⎞ ⎛ 516 ⎞ = 0.0731⎜ ⎟ ⎜ 518 ⎟ 13.6 × 28.59 ⎝ ⎠⎝ ⎠ = 0.0739 lbm/ft 3 ⎛ P + 13.6 pb ⎞ ⎛ t d4 + 460 ⎞ ρ3a = ρ 4 ⎜ s3a ⎟ ⎟⎜ ⎝ 13.6 p4 ⎠ ⎝ t d3a + 460 ⎠ ⎛ 3.65 + 13.6 × 28.72 ⎞ ⎛ 516 ⎞ = 0.0731⎜ ⎟ ⎜ 519 ⎟ 13.6 × 28.59 ⎝ ⎠⎝ ⎠ = 0.0737 lbm/ft 3

FAN POWER INPUT Measured amps/FLA = (82/90.3) = 0.91 = 91% Annex K indicates that Equation A will provide a reasonably accurate estimate of motor power output for a 75 hp motor operating at 91% FLA. Hmo = 75 (82/90.3) (464/460) = 68.7 hp Figure L.1 in Annex L indicates estimated belt drive loss of 4.3%. HL = 0.043 Hmo = 0.043 × 68.7 = 2.95 hp H

⎛ P + 13.6 pb ⎞ ⎛ t d4 + 460 ⎞ ρ3b = ρ 4 ⎜ s3b ⎟ ⎟⎜ ⎝ 13.6 p4 ⎠ ⎝ t d3b + 460 ⎠ ⎛ 3.45 + 13.6 × 28.72 ⎞ ⎛ 516 ⎞ = 0.0731⎜ ⎟ ⎜ 550 ⎟ 13.6 × 28.59 ⎝ ⎠⎝ ⎠ = 0.0695 lbm/ft 3

= Hmo - HL = 68.7 - 2.95 = 68.75 hp

SYSTEM EFFECT FACTORS SEF 1 is due to the effect of insufficient distance between the fan inlets and the side walls of the fan plenum. The distance is 2.83 ft, or:

FLOW RATES V3a = 1096 (Pv3a/ρ3a = 1096 (0.60/0.0737)0.5 = 3127 fpm )0.5

V3b = 1096 (Pv3b/ρ3b)0.5 = 1096 (0.47/0.0695)0.5 = 2850 fpm Q3a = V3aA3a = 3127 × 7.2 = 22514 cfm Q3b = V3bA3b = 2850 × 9.7 = 27645 cfm Q = Q1 = Q3a (ρ3a/ρ1) + Q3b (ρ3b/ρ1) = 22514 (0.0737/0.0731) + 27645 (0.0695/0.0731) = 48982 cfm Q2 = Q1 (ρ1/ρ2) = 48982 (0.0731/0.0739) = 48452 cfm

218 | Field Performance Measurement

(2.83/3.92) = 0.72 = 72% Of the fan inlet diameter. The area of the fan inlets: A1 = 2 (π D12/4) = 2 (π × 3.922/4) = 24.1 ft2 The fan inlet velocity: V1 = (Q1/A1) = (48982/24.1) = 2032 fpm AMCA Publication 201-90, Figure 9.11A, indicates that for a plenum wall spacing of 72% of the fan inlet diameter System Effect Curve V applies. For 2032 fpm inlet velocity and curve V, Figure 7.1 shows SEF 1 = 0.06 in. wg at 0.075 lbm/ft3 density. At 0.0731 lbm/ft3: SEF 1 = 0.06 (0.0731/0.075) = 0.06 in. wg For SEF 2, AMCA Publication 201-90, Figures 7.1 and 8.3, indicate the following calculations:

V2 = (Q2/A2) = (48452/18.9) = 2564 fpm Blast area ratio = Blast Area/A2 = 13.3/18.9 = 0.70 For a blast area ratio of 0.7 and no duct, Figure 8.3 shows System Effect Curve S applies. For 2564 fpm velocity and curve S, Figure 7.1 shows SEF 2 = 0.33 in. wg at 0.075 lbm/ft3 density. At 0.0739 lbm/ft3: SEF 2 = 0.33 (0.0739/0.075) = 0.33 in. wg FAN STATIC PRESSURE Pv4 = (Q4/1096 A4)2 ρ4 Since:

ρ4 = ρ1 Q4 = Q1 Pv4 = (48982/1096 × 93.2)2 0.0731 = 0.02 in. wg Ps1 + Pv1 = Ps4 + Pv4 = -1.75 + 0.02 = -1.73 in. wg Ps = = = =

Ps2 - Ps1 - Pv1 + SEF 1 + SEF 2 Ps2 - (Ps1 + Pv1) + SEF 1 + SEF 2 3.77 - (-1.73) + 0.06 + 0.33 5.89 in. wg

CONVERSION TO SPECIFIED CONDITIONS Qc = 48982 (1170/1160) = 49404 cfm Psc = 5.89 (1170/1160)2 (0.075/0.0731) = 6.15 in. wg Hc = 65.75 (1170/1160)3 (0.075/0.0731) = 69.22 hp

Field Performance Measurement | 219

EXAMPLE 4B: CENTRAL STATION AIR CONDITIONING UNIT, FACTORY ASSEMBLED DRAWTHROUGH TYPE 1

PLAN VIEW

RETURN AIR

3

STATIC PRESSURE TAPS L 5

SEF 1

2 +

OUTSIDE AIR

+

+

+

FAN SECTION SIDE VIEW

FILTER SECTION

COIL SECTION COMMENTS

1. This is a factory assembled, draw-through central station unit. The subject of the test is the fan section, which is rated by the manufacturer as an assembly of the fan and the cabinet in which the fan has been installed. As a draw-through unit, the performance ratings for the fan section are based on operation with the fan outlet ducted. Before proceeding with the test, it is essential that all dampers--outside air, return air, mixing box, multizone, face and bypass, or volume control--be fixed in the positions agreed upon by all interested parties as being applicable for the installation. Also, the temperatures of heating and cooling coils must be kept constant throughout the test period. It may be necessary to lock out, disconnect, or otherwise modify automatic control devices in order to prevent the positions of the dampers and temperatures of the coils from changing during the test. Refer to Section 17.4.2 for additional considerations affecting the test procedure in this type of installation. 2. Determine Pv3 by using the root mean square of the velocity pressure measurements made in a traverse of Plane 3, located near the end of a straight run of duct, as shown in the diagram. Determine Ps3 by averaging the static pressure measurements made in the same traverse. This static pressure value is used to determine the density at the traverse plane. Procedures for traverses are described in 220 | Field Performance Measurement

Section 9.4. In order to determine the air flow rate, it is necessary to measure the area of the traverse plane. 3. Determine Ps1 by averaging the static pressure measurements made in a traverse of Plane 1. Ps5 may be determined in a similar manner or by averaging the pressure measurements at each of four static pressure taps. If it is possible to install suitable pressure taps, their use is preferred in the region of the fan outlet. due to the close proximity of Planes 2 and 5, and the fact that there is no change in area between the two planes, the conditions which exist at Plane 5 are assumed to exist at Plane 2. Measure the area of Plane 1 for use in calculating Pv1. 4. Measure the dry-bulb and wet-bulb temperatures at Plane 3 and the dry-bulb temperatures at Planes 1 and 5. Determine pb for the general vicinity of the air conditioning unit. These measurements are used to determine densities at the planes of interest. 5. Measure the fan speed and the motor amps, volts, and if possible, watts. Record all pertinent motor nameplate data, including volts (NPV), and full load amps (FLA). If the motor power output is to be estimated by using the phase current method described in Annex K, it is not necessary to measure

motor watts; however, it may be necessary to disconnect the drive and measure the no load amps (NLA) if the motor is not operating at or near its full load point. Refer to Annex K. 6. SEF 1 is due to the effect of insufficient length of duct between the fan outlet and the elbow downstream of the fan. In order to determine the value of SEF 1, it is necessary to measure the outlet area of the fan, A2; the length of the outlet duct, L; and the blast area of the fan.

MEASURED MOTOR DATA Volts = = Amps = =

440, 444, 442 442 av 47.4, 47.7, 48.0 47.7 av

MOTOR NAMEPLATE DATA 40 hp, 3 phase, 60 hertz 440 volts, 1770 rpm, 49.7 FLA

7. To calculate the Fan Section Static Pressure:

GENERAL

Ps = Ps2 - Ps1 - Pv1 + SEF 1

Fan connected to motor through belt drive. CALCULATIONS

Where: Ps2 = Ps5

DENSITIES

Pv1 = (Q1/1096A1)2 ρ1

For Plane 3 conditions of:

The calculation of Pv1 is often ignored in instances similar to this example on the basis that the calculated value of Pv1 is relatively small, and it omission does not affect the test results significantly. 8. In order to compare the test results to the quoted fan section curve drawn for operation at 1430 rpm and 0.075 lbm/ft3 density, it is necessary to convert the results to the specified conditions. The basis for the calculations is described in Section 14. OBSERVATIONS SITE MEASUREMENTS = 29.27 in. Hg = 47.5°F = 49.3°F = 47.3°F = 49°F = -0.847 in. wg = 1.31 in. wg = 0.294 in. wg = 1.39 in. wg = 1402 rpm = 147.2 ft2 = A3 = A5 = 15.42 ft2 Blast Area = 9.4 ft2 L = 2.0 ft, length of outlet duct pb td1 td3 tw3 td5 Ps1 Ps3 Pv3 Ps5 N A1 A2

td3 = 49.3°F tw3 = 47.3°F p3 = pb + (Ps3/13.6) = 29.27 + (1.31/13.6) = 29.37 in. Hg Use Figure N.1 in Annex N to obtain ρ3 = 0.0762 lbm/ft3. ⎛ P + 13.6 pb ⎞ ⎛ t d3 + 460 ⎞ ρ1 = ρ3 ⎜ s1 ⎟⎜ ⎟ ⎝ 13.6 p3 ⎠ ⎝ t d1 + 460 ⎠ ⎛ −0.847 + 13.6 × 29.27 ⎞ ⎛ 509.3 ⎞ = 0.0762 ⎜ ⎟ ⎜ 507.5 ⎟ 13.6 × 29.37 ⎝ ⎠⎝ ⎠ = 0.0760 lbm/ftt 3 ⎛ P + 13.6 pb ⎞ ⎛ t d3 + 460 ⎞ ρ5 = ρ3 ⎜ s5 ⎟⎜ ⎟ ⎝ 13.6 p3 ⎠ ⎝ t d5 + 460 ⎠ ⎛ 1.39 + 13.6 × 29.27 ⎞ ⎛ 509.3 ⎞ = 0.0762 ⎜ ⎟ ⎜ 509 ⎟ 13.6 × 29.37 ⎝ ⎠⎝ ⎠ = 0.0763 lbm/ft 3 It is assumed ρ2 = ρ5. FLOW RATES V3 = 1096 (Pv3/ρ3)0.5 = 1096 (0.294/0.0762)0.5 = 2153 fpm

Field Performance Measurement | 221

Q3 = V3A3 = 2153 × 15.42 = 33199 cfm Q = = = =

Q1 Q3 (ρ3/ρ1) 33199 (0.0762/0.0760) 33286 cfm

Q2 = = = =

Q5 Q3 (ρ3/ρ5) 33199 (0.0762/0.0763) 33155 cfm

FAN POWER INPUT Measured amps/FLA = (47.7/49.7) = 0.96 = 96% Annex K indicates that Equation A will provide a reasonably accurate estimate of motor power output for a 40 hp motor operating at 96% FLA. Hmo = 40 (47.7/49.7) (442/440) = 38.6 hp Figure L.1 in Annex L indicates estimated belt drive loss of 4.5%. HL = 0.045 Hmo = 0.045 × 38.6 = 1.74 hp H

= Hmo - HL = 38.6 - 1.74 = 36.86 hp

For velocities of 2500 fpm or less, the 100% effective outlet duct length is 2.5 duct diameters: = 2.5 × 4.43 = 11.1 ft The length of the outlet duct in % effective duct length: = (L/11.1) 100 = (2.0/11.1) 100 = 18% Blast area ratio = Blast Area/A2 = 9.4/15.42 = 0.61 For a blast area ratio of 0.6, 18% effective duct length and elbow position A, Figure 8.5 shows System Effect Curve R applies. For 2150 fpm velocity and curve R, Figure 7.1 shows SEF 1 = 0.34 in. wg at 0.075 lbm/ft3 density. At 0.0762 lbm/ft3: SEF 1 = 0.34 (0.0762/0.075) = 0.35 in. wg FAN SECTION STATIC PRESSURE Pv1 = (Q1/1096 A1)2 ρ1 = (33286/1096 × 147.2)2 0.0760 = 0.003 in. wg It is assumed that Ps2 = Ps5 Ps = Ps2 - Ps1 - Pv1 + SEF 1 = 1.39 - (-0.847) - 0.003 + 0.35 = 2.58 in. wg

SYSTEM EFFECT FACTOR

CONVERSION TO SPECIFIED CONDITIONS

To determine SEF 1, AMCA Publication 201-90, Figures 7.1 and 8.5, indicate the following calculations:

Qc = 33286 (1430/1402) = 33951 cfm

V2 = (Q2/A2) = (33155/15.42) = 2150 fpm Duct diameter equivalent to the fan outlet area: De2 = (4 A2/π)0.5 = (4 × 15.42/π)0.5 = 4.43 ft

222 | Field Performance Measurement

Psc = 2.58 (1430/1402)2 (0.075/0.0760) = 2.65 in. wg Hc = 36.86 (1430/1402)3 (0.075/0.0760) = 38.60 hp

EXAMPLE 4C: PACKAGED AIR-CONDITIONING UNIT

3 2 L

SEF 1 PLAN VIEW

4 1 INLET PLENUM

FILTERS

FANS

5

+

+

COOLING COIL

SIDE VIEW

COMMENTS 1. The subject of the test in this example is the air conditioning unit assembly. This assembly does not include the inlet plenum. The performance ratings for the unit assembly are based on operation with the outlets of the fans ducted. Before proceeding with the test, it is essential that all system dampers be fixed in the positions agreed upon by all interested parties as being applicable for the installation. Also, the temperature of the cooling coil must be kept constant throughout the test period. It may be necessary to lock out, disconnect or otherwise modify automatic control devices in order to prevent the positions of the dampers and the temperature of the coil from changing during the test. Refer to Section 17.4.1 for additional considerations affecting the test procedure in this type of installation.

3. Ps4 may be determined by averaging the pressure measurements at each of four static pressure taps or by averaging the static pressure measurements made in a Pitot-static tube traverse of Plane 4. Ps5 is determined in a similar manner. However, if it is possible to install suitable static pressure taps, their use is preferred in the regions of the outlets of the fans. Due to the close proximity of Planes 1 and 4 and the fact that there is no change in area between the two planes, the conditions which exist at Plane 4 are assumed to exist at Plane 1. Although Plane 5 is greater in area that Plane 2, the degree of divergence is relatively small. Therefore, Ps2 will be calculated based on Ps5 and the assumption that there is no change in total pressure from Plane 2 to Plane 5.

2. Determine Pv3 by using the root mean square of the velocity pressure measurements made in a traverse of Plane 3, located near the end of a straight run of duct, as shown in the diagram. Determine Ps3 by averaging the static pressure measurements made in the same traverse. This static pressure value is used to determine the density at the traverse plane. Procedures for traverses are described in Section 9.4. in order to determine the air flow rate, it is necessary to measure the area of the traverse plane.

4. Measure the dry-bulb and wet-bulb temperatures at Plane 4 and the dry-bulb temperatures at Planes 3 and 5. In this example, the cooling medium, normally circulated in the coil was shut off in order to maintain constant air temperatures during the test. In order to account for water vapor which may have been added to the air as a result of evaporation of moisture previously condensed on the coil, the wet-bulb temperature at Plane 3 was measured. Determine pb for the general vicinity of the air conditioning unit. These measurements are used in determining densities at the planes of interest.

Field Performance Measurement | 223

5. Measure the fan speed and the motor amps, volts, and if possible, watts. Record all pertinent motor nameplate data including volts (NPV), and full load amps (FLA). If the motor power output is to be estimated by using the phase current method described in Annex K, it is not necessary to measure motor watts; however, it may be necessary to disconnect the drive and measure the no load amps (NLA) if the motor is not operating at or near its full load point. Refer to Annex K.

A1 = A4 = 31.7 ft2 A2 = 11.5 ft2 A3 = 16.4 ft2 A5 = 14.3 ft2 Blast Area = 4.0 ft2 per fan L = 2.0 ft, length of outlet duct

6. Although an elbow is located shortly downstream of the fans, SEF 1 is judged to be more closely characterized as the effect due to insufficient lengths of duct on the outlets of the fans. In order to determine the value of SEF 1, it is necessary to measure the outlet area and the blast area of one of the fans and the length, L, of its outlet duct.

Volts = = Amps = =

7. To calculate the static pressure for the unit assembly:

MEASURED MOTOR DATA 460, 455, 465 460 av 38.2, 38, 37.9 38.0 av

MOTOR NAMEPLATE DATA 25 hp, 3 phase, 60 hertz 460 volts, 1760 rpm, 39.5 FLA GENERAL

Ps = Ps2 - Ps1 - Pv1 + SEF 1 Fans connected to motor through belt drive. Where: CALCULATIONS Ps1 = Ps4

DENSITIES

Pv1 = (Q1/1096A1)2 ρ1 Ps2 = Ps5 + Pv5 - Pv2 Pv2 and Pv5 are calculated in manners similar to the calculation of Pv1. 8. In order to compare the test results to the quoted unit assembly curve drawn for operation at 1050 rpm and 0.075 lbm/ft3 density, it is necessary to convert the results to the specified conditions. The basis for the calculations is described in Section 14. OBSERVATIONS SITE MEASUREMENTS pb td3 tw3 td4 tw4 td5 Ps3 Pv3 Ps4 Ps5 N

= = = = = = = = = = =

29.65 in. Hg 75.0°F 59.5°F 72.5°F 58.5°F 74.5°F 2.02 in. wg 0.35 in. wg -0.32 in. wg 2.11 in. wg 1025 rpm

224 | Field Performance Measurement

For Plane 3 conditions of: td3 = 75.0°F tw3 = 59.5°F p3 = pb + (Ps3/13.6) = 29.65 + (2.03/13.6) = 29.80 in. Hg Use Figure N.1 in Annex N to obtain ρ3 = 0.0736 lbm/ft3. For Plane 4 conditions of: td4 = 72.5°F tw4 = 58.5°F p4 = pb + (Ps4/13.5) = 29.65 + (-0.32/13.6) = 29.63 in. Hg Use Figure N.1 in Annex N to obtain ρ4 = 0.0735 lbm/ft3. It is assumed that ρ1 = ρ4.

⎛ P + 13.6 pb ⎞ ⎛ t d3 + 460 ⎞ ρ5 = ρ3 ⎜ s5 ⎟⎜ ⎟ ⎝ 13.6 p3 ⎠ ⎝ t d5 + 460 ⎠ ⎛ 2.11 + 13.6 × 29.65 ⎞ ⎛ 535 ⎞ = 0.0736 ⎜ ⎟ ⎜ 534.5 ⎟ 13.6 × 29.80 ⎝ ⎠⎝ ⎠ = 0.0737 lbm/ft 3

SYSTEM EFFECT FACTOR To determine SEF 1, AMCA Publication 201-90, Figures 7.1 and 8.3, indicate the following calculations:

It is assumed ρ2 = ρ5.

V2 = (Q2/A2) = (39143/11.5) = 3404 fpm

FLOW RATES

Duct diameter equivalent to the outlet area of one fan:

V3 = 1096 (Pv3/ρ3)0.5 = 1096 (0.35/0.0736)0.5 = 2390 fpm

De2 = (4A2/2π)0.5 = (4 × 11.5/2π)0.5 = 2.71 ft

Q3 = V3A3 = 2390 × 16.4 = 39196 cfm

Figure 8.3 shows that for velocities over 2500 fpm, 100% effective duct length is one diameter for every 1000 fpm:

Q2 = = = =

Q5 Q3 (ρ3/ρ5) 39196 (0.0736/0.0737) 39143 cfm

Q = = = =

Q1 = Q4 Q3 (ρ3/ρ4) 39196 (0.0736/0.0735) 39249 cfm

FAN POWER INPUT Measured amps/FLA = (38.0/39.5) = 0.96 = 96%

= De2 (V2/1000) = 2.71 (3404/1000) = 9.22 ft L in % effective duct length: = (L/9.22) 100 = (2.0/9.22) 100 = 22% Blast area ratio = Blast area/A2 = (2 × 4.0)/11.5 = 0.70

Annex K indicates that Equation A will provide a reasonably accurate estimate of motor power output for a 25 hp motor operating at 96% FLA.

For a blast area ratio of 0.7, and 22% effective duct length Figure 8.3 shows System Effect Curve W applies. For 3404 fpm velocity and curve W, Figure 7.1 shows SEF 1 = 0.13 in. wg at 0.075 lbm/ft3 density. At 0.0737 lbm/ft3:

Hmo = 25 (38.0/39.5) (460/460) = 24.1 hp

SEF 1 = 0.13 (0.0737/0.075) = 0.13 in. wg

Figure L.1 in Annex L indicates estimated belt drive loss of 4.8%.

STATIC PRESSURE OF UNIT

HL = 0.048 Hmo = 0.048 × 24.1 = 1.2 hp H = Hmo - HL = 24.1 - 1.2 = 22.9 hp

Pv5 = (Q5/1096 A5)2 ρ5 = (39143/1096 × 14.3)2 0.0737 = 0.46 in. wg Pv2 = (Q2/1096 A2)2 ρ2 = (39143/1096 × 11.5)2 0.0737 = 0.71 in. wg

Field Performance Measurement | 225

Ps2 + Pv2 = Ps5 + Pv5 Ps2 = Ps5 + Pv5 - Pv2 = 2.11 + 0.46 - 0.71 = 1.86 in. wg Pv1 = (Q1/1096 A1)2 ρ1 = (39249/1096 × 31.7)2 0.0735 = 0.09 in. wg Ps = Ps2 - Ps1 - Pv1 + SEF 1 = 1.86 - (-0.32) - 0.09 + 0.13 = 2.22 in. wg CONVERSION TO SPECIFIED CONDITIONS Qc = 39249 (1050/1025) = 40206 cfm Psc = 2.22 (1050/1025)2 (0.075/0.0735) = 2.38 in. wg Hc = 22.9 (1050/1025)3 (0.075/0.0735) = 25.1 hp

226 | Field Performance Measurement

EXAMPLE 4D: PACKAGED AIR-CONDITIONING UNIT

3a 3b

PLAN VIEW

STATIC PRESSURE TAPS 2

5

L

FILTER SECTION 1

SEF 1 +

+

INLET LOUVER

HEATING COIL SIDE VIEW

COMMENTS 1. The subject of the test in this example is the air conditioning unit assembly. This assembly includes the filter section and the inlet louver. The performance ratings for the unit assembly are based on operation with the outlets of the fans ducted. Before proceeding with the test, it is essential that all system dampers be fixed in the positions agreed upon by all interested parties as being applicable for the installation. Also, the temperature of the heating coil must be kept constant throughout the test period. It may be necessary to lock out, disconnect or otherwise modify automatic control devices in order to prevent the positions of the dampers and the temperature of the coil from changing during the test. Refer to Section 17.5.1 for additional considerations affecting the test procedure in this type of installation. 2. Normally, velocity pressure measurements would be made in a single plane, located in a duct common to all branches. In this example, a measurement plane which provides a satisfactory velocity profile cannot be located upstream of the fans or between the point of connection of the branch ducts and the outlets of the fans. The alternative, as indicated in the diagram, is to make a velocity pressure measurement traverse in each of two branches. the velocity pressure for reach branch is determined by using the root mean square of the velocity pressure measurements made in the traverse. The static

pressure at each traverse plane is determined by using the root mean square of the velocity measurement traverse in each of two branches. The velocity pressure for each branch is determined by using the root mean square of the velocity pressure measurements made in the traverse. The static pressure at each traverse plane is determined by averaging the static pressure measurements made in the same traverse. These static pressure values are used in determining the densities at the traverse planes. Procedures for traverses are described in Section 9.4. In order to determine the air flow rates, it is necessary to measure the area of each traverse plane. 3. Determine Ps5 by averaging the pressure measurements at each of four static pressure taps located in the duct fitting at the outlets of the fans. The conditions which exist at Plane 5, including the static pressure, are assumed to exist at Plane 2, based on their close proximity and the fact that there is no change in area between the two planes. In situations such as this example, it is important to be certain that all pressure measurements are referred to the same atmospheric pressure. 4. Measure the dry-bulb and wet-bulb temperatures at Plane 1 and the dry-bulb temperatures at Planes 3a, 3b, and 5. Determine pb for the general vicinity of Field Performance Measurement | 227

the air conditioning unit. These measurements are used to determine densities at the planes of interest. 5. Measure the fan speed and the motor amps, volts, and if possible, watts. Record all pertinent motor nameplate data, including volts (NPV), and full load amps (FLA). If the motor power output is to be estimated by using the phase current method described in Annex K, it is not necessary to measure motor watts; however, it may be necessary to disconnect the drive and measure the no load amps (NLA) if the motor is not operating at or near its full load point. Motor performance data, supplied by the motor manufacturer, are used in the determination of motor power output in this example. 6. SEF 1 is due to the effect of insufficient length of duct between the outlets of the fans and the elbow downstream of the fans. In order to determine the value of SEF 1, it is necessary to measure the outlet area and the blast area of one of the fans and the length of the duct, L, between the fan and the elbow. 7. The sum of the static pressure, Ps1, and velocity pressure, Pv1, at the inlet to the unit assembly is considered to be equal to the sum of the static pressure, Psx, and velocity pressure, Pvx, at a point sufficiently distant from the inlet as to be in still air. At this point, the static pressure is zero, and the velocity pressure in still air is zero. Ps1 + Pv1 = Psx + Pvx = 0 This consideration, which is the same as that used in the methods for testing this type of unit for performance rating purposes, charges to the unit losses incurred in accelerating the air into its inlet and eliminates the inaccuracies which arise in any attempt to measure the velocity pressure and static pressure at the inlet. To calculate the static pressure for the unit assembly: Ps = Ps2 - Ps1 - Pv1 + SEF 1 = Ps2 - (Ps1 + Pv1) + SEF 1 Since: Ps1 + Pv1 = 0 Ps = Ps2 + SEF 1 Where: Ps2 = Ps5

8. In order to compare the test results to the quoted performance curve for the packaged unit drawn for operation at 1720 rpm and 0.075 lbm/ft3 density, it is necessary to convert the results to the specified conditions. The basis for the calculations is described in Section 14. OBSERVATIONS SITE MEASUREMENTS = 29.65 in. Hg = 72°F = 61°F = 85°F = 82.5°F = 83°F = 1.25 in. wg = 1.15 in. wg = 1.22 in. wg = 0.56 in. wg = 0.60 in. wg = 1710 rpm = A5 = 5.64 ft2 A3a = 3.1 ft2 A3b = 2.2 ft2 Blast Area = 2.5 ft2 per fan L = 0.96 ft, length of outlet duct pb td1 tw1 td5 td3a td3b Ps5 Ps3a Ps3b Pv3a Pv3b N A2

MEASURED MOTOR DATA Volts = = Amps = =

460, 458, 462 460 av 10.0, 10.0, 9.8 9.9 av

MOTOR NAMEPLATE DATA 10 hp, 3 phase, 60 hertz 460 volts, 1750 rpm, 13.5 FLA GENERAL Fans connected to motor through belt drive. The following motor performance data was supplied by the motor manufacturer: Motor Efficiency: 82.5% at 1/2 load 84.5% at 3/4 load 84.5% at full load Power Factor = 0.85

228 | Field Performance Measurement

DENSITIES For Plane 1 conditions of: td1 = 72°F tw1 = 61°F p1 = pb = 29.65 in. Hg Use Figure N.1 in Annex N to obtain ρ1 = 0.0735 lbm/ft3. ⎛ P + 13.6 pb ⎞ ⎛ t d1 + 460 ⎞ ρ5 = ρ1 ⎜ s5 ⎟ ⎟⎜ ⎝ 13.6 p1 ⎠ ⎝ t d5 + 460 ⎠ ⎛ 1.25 + 13.6 × 29.65 ⎞ ⎛ 532 ⎞ = 0.0735 ⎜ ⎟ ⎜ 545 ⎟ 13.6 × 29.65 ⎝ ⎠⎝ ⎠ = 0.0720 lbm/ft 3 It is assumed that ρ2 = ρ5 ⎛ P + 13.6 pb ⎞ ⎛ t d1 + 460 ⎞ ρ3a = ρ1 ⎜ s3a ⎟ ⎟⎜ ⎝ 13.6 p1 ⎠ ⎝ t d3a + 460 ⎠ ⎛ 1.15 + 13.6 × 29.65 ⎞ ⎛ 532 ⎞ = 0.0735 ⎜ ⎟ ⎜ 542.5 ⎟ 13.6 × 29.65 ⎝ ⎠⎝ ⎠ = 0.0723 lbm/ft 3 ⎛ P + 13.6 pb ⎞ ⎛ t d1 + 460 ⎞ ρ3b = ρ1 ⎜ s3b ⎟ ⎟⎜ ⎝ 13.6 p1 ⎠ ⎝ t d3b + 460 ⎠ ⎛ 1.22 + 13.6 × 29.65 ⎞ ⎛ 532 ⎞ = 0.0735 ⎜ ⎟ ⎜ 543 ⎟ 13.6 × 29.65 ⎝ ⎠⎝ ⎠ = 0.0722 lbm/ft 3 FLOW RATES V3a = 1096 (Pv3a/ρ3a)0.5 = 1096 (0.56/0.0723)0.5 = 3050 fpm V3b = 1096 (Pv3b/ρ3b = 1096 (0.60/0.0722)0.5 = 3159 fpm )0.5

Q3a = V3aA3a = 3050 × 3.1 = 9455 cfm Q3b = V3bA3b = 3159 × 2.2 = 6950 cfm

Q = = = =

Q1 Q3a (ρ3a/ρ1) + Q3b (ρ3b/ρ1) 9455 (0.0723/0.0735) + 6950 (0.0722/0.0735) 16128 cfm

Q2 = = = =

Q5 Q1 (ρ1/ρ5) 16128 (0.0735/0.0720) 16464 cfm

FAN POWER INPUT Measured amps/FLA = (9.9/13.5) = 0.73 = 73% The data supplied by the motor manufacturer indicate power factor of 0.85 and motor efficiency of 84.5% for the motor operating at 73% FLA. Using the appropriate equation in Section 10.2.2: Hmo = (3)0.5 × 9.9 × 460 × 0.85 × 0.845/746 = 7.59 hp Figure L.1 in Annex L indicates estimated belt drive loss of 5.6%. HL = 0.056 Hmo = 0.056 × 7.59 = 0.43 hp H

= Hmo - HL = 7.59 - 0.43 = 7.16 hp

SYSTEM EFFECT FACTOR SEF 1 is due to the effect of insufficient lengths of duct between the outlets of the fans and the elbow downstream of the fans. AMCA Publication 201-90, Figures 7.1, 8.1, and 8.5 indicate the following calculations: V2 = (Q2/A2) = (16464/5.64) = 2919 fpm Duct diameter equivalent to the outlet area of one fan: De2 = (4A2/2π)0.5 = (4 × 5.64/2π)0.5 = 1.89 ft Figure 8.1 shows that for velocities over 2500 fpm 100% effective duct length is one diameter for every 1000 fpm: Field Performance Measurement | 229

= De2 (V2/1000) = 1.89 (2919/1000) = 17% L, in % effective duct length: = (L/5.52) 100 = (0.96/5.52) 100 = 17% Blast area ratio = Blast Area/A2 = (2 × 2.5)/5.64 = 0.89 For a blast area ratio of 0.89, 17% effective duct length and elbow position C, Figure 8.5 shows System Effect Curve S applies. For 2919 fpm velocity and curve S, Figure 7.1 shows SEF 1 = 0.43 in. wg at 0.075 lbm/ft3 density. At 0.0720 lbm/ft3: SEF 1 = 0.43 (0.0720/0.075) = 0.41 in. wg STATIC PRESSURE OF UNIT Ps2 = Ps5 = 1.25 in. wg Ps = Ps2 + SEF 1 = 1.25 + 0.41 = 1.66 in. wg CONVERSION TO SPECIFIED CONDITIONS Qc = 16128 (1720/1710) = 16222 cfm Psc = 1.66 (1720/1710)2 (0.075/0.0735) = 1.71 in. wg Hc = 7.16 (1720/1710)3 (0.075/0.0735) = 7.44 hp

230 | Field Performance Measurement

EXAMPLE 4E: CENTRAL STATION AIR CONDITIONING UNIT, FACTORY ASSEMBLED BLOWTHROUGH TYPE 2

1

STATIC PRESSURE TAPS

5

PLAN VIEW

RETURN AIR

3b

SPRAY SECTION

3a

HEATING COIL + +

+

OUTSIDE AIR

+ + + +

FILTER SECTION

+

FAN SECTION

COOLING COIL

SIDE VIEW

COMMENTS 1. This is a factory assembled, blow-through central station unit. The subject of the test is the fan section, which is rated by the manufacturer as an assembly of the fan and the cabinet in which the fan has been installed. As a blow-through unit, the performance ratings for the fan section are based on operation without the fan outlet ducted. Before proceeding with the test, it is essential that all dampers (outside air, return air, mixing box, multizone, face and bypass, or volume control) be fixed in the positions agreed upon by all interested parties as being applicable for the installation. Also, the temperatures of heating and cooling coils must be kept constant throughout the test period. It may be necessary to lock out, disconnect, or otherwise modify automatic control devices in order to prevent the positions of the dampers and temperatures of the coils from changing during the test. In instances in which a cooling coil is located between a velocity pressure traverse plane and the fan, as in this example, the flow of the cooling medium should be stopped or its temperature raised to a level sufficient to prevent condensation on the cooling coil, otherwise the moisture condensed will not be properly taken into account in the determination of fan air flow rate. Refer to Section 17.5.2 for additional considerations affecting the test procedure in this type of installation. 2. Normally, velocity pressure measurements would

be made in a single plane, located in a duct common to all branches. In this example, a measurement plane which provides a satisfactory velocity profile cannot be located upstream of the fan or between the point of connection of the branch ducts and the fan outlet. The alternative, as indicated in the diagram, is to make a velocity pressure measurement traverse in each branch. The velocity pressure for each branch is determined by using the root mean square of the velocity pressure measurements made in the traverse. The static pressure at each traverse plane is determined by averaging the static pressure measurements made in the same traverse. These static pressure values are used in determining the densities at the traverse plane. Procedures for traverses are described in Section 9.4. In order to determine the air flow rates it is necessary to measure the area of each traverse plane. 3. Determine Ps1 by averaging the static pressure measurements made in a traverse of Plane 1. Ps5 may be determined in a similar manner or by averaging the pressure measurements at each of four static pressure taps. If it is possible to install suitable pressure taps, their use is preferred in the regions of the fan outlet. Due to the abrupt expansion in area from Plane 2 to Plane 5, it is assumed that there is no conversion of velocity pressure at Plane 2 to static pressure at Plane 5. Therefore, it is assumed Field Performance Measurement | 231

that Ps2 = Ps5. Measure the area of Plane 1 for use in calculating Pv1. 4. Measure the dry-bulb and wet-bulb temperatures at Planes 1, 3a, and 3b. Determine pb for the general vicinity of the air conditioning unit. These measurements are used to determine densities at the planes of interest. The measurements of additional wet-bulb temperatures were made in this example in order to provide data which may be used to determine whether the moisture content of the air changed between Plane 1 and Planes 3a and 3b. 5. Measure the fan speed and the motor amps, volts, and if possible, watts. Record all pertinent motor nameplate data, including volts (NPV), and full load amps (FLA). If the motor power output is to be estimated by using the phase current method described in Annex K, it is not necessary to measure motor watts; however, it may be necessary to disconnect the drive and measure the no load amps (NLA) if the motor is not operating at or near its full load point. Refer to Annex K.

tw3a = td3b = tw3b = Ps1 = Ps5 = Ps3a = Ps3b = Pv3a = Pv3b = N = A1 = A3a = A3b =

71.5°F 60°F 58°F -2.43 in. wg 6.55 in. wg 5.35 in. wg 5.1 in. wg 0.53 in. wg 0.60 in. wg 1695 rpm 68.9 ft2 5.37 ft2 6.78 ft2

MEASURED MOTOR DATA Volts = = Amps = = NLA =

570, 575, 565 570 av 81.5, 82.5, 81 81.7 19

MOTOR NAMEPLATE DATA 6. Since the performance ratings for the fan section are based on operation without the fan outlet ducted, an SEF does not apply for the unducted position.

100 hp, 3 phase, 60 hertz 575 volts, 1790 rpm, 95 FLA

7. To calculate the Fan Section Static Pressure:

GENERAL

Ps = Ps2 - Ps1 - Pv1

Fan connected to motor through belt drive.

Where:

CALCULATIONS

Ps2 = Ps5 Pv1 = (Q1/1096 A1)2 ρ1 Q1 = Q3a (ρ3a/ρ1) + Q3b (ρ3b/ρ1)

DENSITIES

The calculation of Pv1 is often ignored in instances similar to this example on the basis that the calculated value of Pv1 is relatively small, and its omission does not affect the test results significantly.

td1 = 65°F tw1 = 60°F

8. In order to compare the test results to the quoted fan section curve drawn for operation at 1650 rpm and 0.075 lbm/ft3 density, it is necessary to convert the results to the specified conditions. The basis for the calculations is described in Section 14.

For Plane 1 conditions of:

p1 = pb + (Ps1/13.6) = 28.85 + (-2.43/13.6) = 28.67 in. Hg Use Figure N.1 in Annex N to obtain ρ1 = 0.0720 lbm/ft3. For Plane 3a conditions of:

OBSERVATIONS SITE MEASUREMENTS pb td1 tw1 td3a

= = = =

28.85 in. Hg 65°F 60°F 100°F

232 | Field Performance Measurement

td3a = 100°F tw3a = 71.5°F p3a = pb + (Ps3a/13.6) = 28.85 + (5.35/13.6) = 29.24 in. Hg

Use Figure N.1 in Annex N to obtain ρ1 = 0.0720 lbm/ft3.

Hmo

For Plane 3b conditions of:

Reference to Figure L.1 in Annex L indicates estimated belt drive loss of 4.2%.

td3b = 60°F tw3b = 58°F p3b = pb + (Ps3b/13.6) = 28.85 + (5.1/13.6) = 29.23 in. Hg Use Figure N.1 in Annex N to obtain ρ3b = 0.0741 lbm/ft3. FLOW RATES V3a = 1096 (Pv3a/ρ3a)0.5 = 1096 (0.53/0.0691)0.5 = 3035 fpm V3b = 1096 (Pv3b/ρ3b)0.5 = 1096 (0.60/0.0741)0.5 = 3119 fpm Q3a = V3aA3a = 3035 × 5.37 = 16298 fpm Q3b = V3bA3b = 3119 × 6.78 = 21147 cfm Q = Q1 = Q3a (ρ3a/ρ1) + Q3b (ρ3b/ρ1) = 16298 (0.0691/0.0720) + 21147 (0.0741/0.0720) = 37405 cfm

= (85.3 + 81.8)/2 = 83.6 hp

HL = 0.042 Hmo = 0.042 × 83.6 = 3.5 hp H = Hmo - HL = 83.6 - 3.5 = 80.1 hp FAN SECTION STATIC PRESSURE Pv1 = (Q1/1096 A1)2 ρ1 = (37405/1096 × 68.9)2 0.0720 = 0.02 in. wg It is assumed that Ps2 = Ps5 Ps = Ps2 - Ps1 - Pv1 = 6.55 - (-2.43) - 0.02 = 8.96 in. wg CONVERSION TO SPECIFIED CONDITIONS Qc = 37405 (1650/1695) = 36412 cfm Psc = 8.96 (1650/1695)2 (0.075/0.0720) = 8.84 in. wg Hc = 80.1 (1650/1695)3 90.075/0.0720) = 77.0 hp

FAN POWER INPUT Measured amps/FLA = (81.7/95) = 0.86 = 86% Annex K indicates that the average of the results of Equation A and Equation B will provide a reasonably accurate estimate of motor power output for a 100 hp motor operating at 86% of FLA. Eqn. A = 100 (81.7/95) (570/575) = 85.3 hp Eqn. B = 100 [(81.7 - 19)/(95 - 19)] (570/575) = 81.8 hp

Field Performance Measurement | 233

EXAMPLE 5A: FREE INLET, FREE OUTLET ROOF VENTILATOR

2

1

3 2 De

TEMPORARY DUCT WITH SQUARE CROSS-SECTION, De = EQUIVALENT DIAMETER OF DUCT

1.5 De

COMMENTS 1. The subject of the test in this example is the roof ventilator assembly. Before proceeding with the test, refer to Section 17.4 for considerations affecting the test procedure in this type of installation. 2. Determine Pv3 by using the root mean square of the velocity pressure measurements made in a traverse of Plane 3, located in the duct which has been installed on the inlet side of the ventilator. Determine Ps3 by averaging the static pressure measurements made in the same traverse. Procedures for traverses are described in Section 9.4. Measure the area of the traverse plane, A3, which is located at the tip of the Pitot-static tube. The duct, temporarily installed for purposes of the test, is square in cross-section. Its cross-sectional dimensions were selected as the maximum permissible for its installation into the opening in the ventilator mounting curb. The length of the duct is twice its equivalent diameter and the entrance to the duct is flared in oder to reduce inlet losses. The installation of a duct of this size and cross-sectional configuration is judged as creating no significant effect on the performance of the ventilator in this example. 3. Ps2, the static pressure at the outlet of the ventilator, is zero gauge pressure, referred to the atmospheric pressure in the region of the ventilator outlet. In situations such as this example, the air may 234 | Field Performance Measurement

be discharging from the ventilator into a region in which the atmospheric pressure is somewhat different from that to which all other pressure measurements are referred. When this possibility exists, it is essential that the static pressure in the region of the discharging air be measured, referred to the same atmospheric pressure as used in all other pressure measurements. In this example, Ps2 was measured, referred to the same atmospheric pressure as in the static pressure measurements made at Plane 3. 4. Measure the dry-bulb and wet-bulb temperatures at the velocity traverse plane. Determine pb for the general vicinity of the ventilator. These measurements are used to determine densities at the planes of interest. 5. Measure the fan speed and the motor amps and volts. Record all pertinent motor nameplate data. For the horsepower rating of the motor in this example, it is recommended that the fan power input be determined by using the measured watts input to the motor and motor performance data, obtained from the motor manufacturer. 6. To calculate the Fan Static Pressure: Ps = Ps2 - Ps1 - Pv1 = Ps2 - (Ps1 + Pv1)

Where:

FLOW RATE

Ps1 + Pv1 = Ps3 + Pv3

V3 = 1096 (Pv3/ρ3)0.5 = 1096 (0.077/0.0727)0.5 = 1128 fpm

7. In order to compare the test results to the quoted fan curve drawn for operation at 1180 rpm and 0.075 lbm/ft3 density, it is necessary to convert the results to the specified conditions. The basis for the calculations is described in Section 14. OBSERVATIONS SITE MEASUREMENTS pb td3 tw3 Ps2 Ps3 Pv3 N A3

= = = = = = = =

29.37 in. Hg 73.5°F 58.1°F 0.037 in. wg -0.085 in. wg 0.077 in. wg 1177 rpm 5.58 ft2

Q = = = =

Q1 = Q3 V3A3 1128 × 5.58 6294 cfm

FAN POWER INPUT At the measured power input value of 755 watts, the data supplied by the motor manufacturer indicate efficiency of 61% for the motor. Hmo = (755 × 0.61)/746 = 0.62 hp Since the fan is direct connected to the motor, there is no drive loss, and:

MEASURED MOTOR DATA

H = Hmo = 0.62 hp

Volts = 235, 230, 230 = 232 av Watts = 755

FAN STATIC PRESSURE

MOTOR NAMEPLATE DATA 1 hp, 3 phase, 60 hertz 230 volts, 1175 rpm, 3.6 FLA

Ps1 + Pv1 = Ps3 + Pv3 = -0.085 + 0.077 = -0.008 in. wg Ps = Ps2 - (Ps1 + Pv1) = 0.037 - (-0.008) = 0.045 in. wg

General CONVERSION TO SPECIFIED CONDITIONS Fan direct connected to motor. Motor efficiency data supplied by motor manufacturer.

Qc = 6294 (1180/1177) = 6310 cfm

CALCULATIONS DENSITIES For Plane 3 conditions of: td3 = 73.5°F tw3 = 58.1°F

Psc = 0.045 (1180/1177)2 (0.075/0.0727) = 0.047 in. wg Hc = 0.62 (1180/1177)3 (0.075/0.0727) = 0.64 hp

p3 = pb + (Ps3/13.6) = 29.37 + (-0.085/13.6) = 29.36 in. Hg Use Figure N.1 in Annex N to obtain ρ3 = 0.0727 lbm/ft3. It is assumed that ρ1 = ρ3. Field Performance Measurement | 235

EXAMPLE 5B: FREE INLET, FREE OUTLET PROPELLER FAN

2 De

2 3

1.5 De

D2

TEMPORARY DUCT WITH SQUARE CROSS-SECTION, De = EQUIVALENT DIAMETER OF DUCT

COMMENTS 1. The subject of the test in this example is the propeller fan assembly. Before proceeding with the test, refer to Section 17.4 for considerations affecting the test procedure in this type of installation. 2. Determine Pv3 by using the root mean square of the velocity pressure measurements made in a traverse of Plane 3, located in the duct which has been installed on the inlet side of the fan. Determine Ps3 by averaging the static pressure measurements made in the same traverse. Procedures for traverses are described in Section 9.4. Measure the area of the traverse plane, A3, which is located at the tip of the Pitot-static tube. The duct, temporarily installed for purposes of the test, is square in cross-section, with side dimension of 1.5 D2. The shape and area of the duct cross-section were selected on the basis of minimizing the effect of the duct on the performance of the fan while providing velocity pressure readings of measurable magnitudes. The length of the duct is twice its equivalent diameter, and the entrance to the duct is flared in order to reduce inlet losses. The installation of the duct is judged as creating no significant effect on the performance of the fan in this example.

such as this example, the air may be discharging from the fan into a region in which the atmospheric pressure is somewhat different from that to which all other pressure measurements are referred. When this possibility exists, it is essential that the static pressure in the region of the discharging air be measured, referred to the same atmospheric pressure as used in all other pressure measurements. In this example, Ps2 was measured, referred to the same atmospheric pressure as in the static pressure measurements made at Plane 3. 4. Measure the dry-bulb and wet-bulb temperatures at the velocity traverse plane. Determine pb for the general vicinity of the fan. These measurements are used to determine densities at the planes of interest. 5. Measure the fan speed and the motor amps and volts. Record all pertinent motor nameplate data. For the horsepower rating of the motor in this example, it is recommended that the fan power input be determined by using the measured watts input to the motor and motor performance data obtained from the motor manufacturer. 6. To calculate the Fan Static Pressure:

3. Ps2, the static pressure at the outlet of the fan, is zero gauge pressure, referred to the atmospheric pressure in the region of the fan outlet. In situations 236 | Field Performance Measurement

Ps = Ps2 - Ps1 - Pv1 = Ps2 - (Ps1 + Pv1)

Where:

FLOW RATES

Ps1 + Pv1 = Ps3 + Pv3

V3 = 1096 (Pv3/ρ3)0.5 = 1096 (0.025/0.0715)0.5 = 648 fpm

7. In order to compare the test results to the quoted fan curve drawn for operation at 1725 rpm and 0.075 lbm/ft3 density, it is necessary to convert the results to the specified conditions. The basis for the calculations is described in Section 14. OBSERVATIONS SITE MEASUREMENTS pb = td3 = tw3 = Ps2 = Ps3 = Pv3 = N = A3 =

29.65 in. Hg 85°F 74°F 0 in. wg -0.027 in. wg 0.025 in. wg 1775 rpm 5.06 ft2

Q = = = =

Q1 = Q3 V3A3 648 × 5.06 3279 cfm

FAN POWER INPUT At the measured power input value of 637 watts, the data supplied by the motor manufacturer indicate efficiency of 65% for the motor. Hmo = (637 × 0.65)/746 = 0.56 hp Since the fan is direct connected to the motor, there is no drive loss, and:

MEASURED MOTOR DATA

H = Hmo = 0.56 hp

Volts = 230, 225, 230 = 228 av Watts = 637

FAN STATIC PRESSURE

MOTOR NAMEPLATE DATA 3/4 hp, 3 phase, 60 hertz 230 volts, 1760 rpm, 4.8 FLA GENERAL Fan direct connected to motor. Motor efficiency data supplied by motor manufacturer.

Ps1 + Pv1 = Ps3 + Pv3 = -0.027 + 0.025 = -0.002 in. wg Ps = Ps2 - (Ps1 + Pv1) = 0 - (-0.002) = 0.002 in. wg This small value is attributed to the loss at the duct inlet, and the fan is considered to be operating at free delivery (Ps = 0).

CALCULATIONS CONVERSION TO SPECIFIED CONDITIONS DENSITIES For Plane 3 conditions of: td3 = 85°F tw3 = 74°F p3 = pb + (Ps3/13.6) = 29.65 + (-0.027/13.6) = 29.65 in. Hg

Qc = 3279 (1725/1775) = 3187 cfm Psc = 0 in. wg Hc = 0.56 (1725/1775)3 (0.075/0.0715) = 0.54 hp

Use Figure N.1 in Annex N to obtain ρ3 = 0.0715 lbm/ft3. It is assumed that ρ1 = ρ3 Field Performance Measurement | 237

EXAMPLE 5C: FREE INLET, FREE OUTLET ROOF VENTILATOR

2

1

3

COMMENTS 1. The subject of the test in this example is the roof ventilator assembly. Before proceeding with the test, refer to Section 17.1 for considerations affecting the test procedure in this type of installation. 2. Ps3, the static pressure in the vicinity of the ventilator inlet, would normally be determined by averaging the static pressure measurements made in a Pitot tube traverse. But in this example, a temporary duct was not installed and the Pitot tube traverse could not be accomplished. In this method for testing a nonducted fan, consider the fan static pressure (Ps) as the differential pressure, as read on a manometer, between the pressure measured inside the room (Ps3) and the pressure measured outside the room in the vicinity of the ventilator outlet (Ps2). These pressures are measured at a sufficient distance from the ventilator so as to be unaffected by the velocity of the entering or leaving air. 3. Ps2 is considered to be zero gauge pressure, but since this measurement is actually part of the differential pressure described in paragraph 2, it is necessary to make only one density correction; the correction is to the differential pressure, which is the fan static pressure.

238 | Field Performance Measurement

4. Measure the dry-bulb and wet-bulb temperatures in the region of the inside pressure measurement. Also, determine pb in the same vicinity. 5. Measure the fan speed and the motor amps and volts. Record all pertinent motor nameplate data. For the horsepower rating of the motor in this example, it is recommended that the fan power input be determined by using the measured watts input to the motor and motor performance data obtained from the motor manufacturer. 6. Airflow rates are determined from the fan manufacturer’s certified performance ratings. Draw a fan performance curve from these ratings converted to operation at the test values of fan speed and entering air density. The basis for these calculations is described in Section 14. The fan airflow rate is then determined by entering this curve at the test values of fan static pressure and fan power input. OBSERVATIONS SITE MEASUREMENTS pb = 29.19 in. Hg td3 = 79°F tw3 = 63°F Ps2 - Ps3 = 0.13 in. wg N = 1735 rpm

MEASURED MOTOR DATA

FAN STATIC PRESSURE

Volts = 229, 229, 232 = 230 av Watts = 1390

The fan static pressure is considered to be the differential static pressure.

MOTOR NAMEPLATE DATA

Ps = Ps2 - Ps3 = 0.13 in. wg

1.5 hp, 3 phase, 60 hertz 230 volts, 1740 rpm, 4.8 FLA

It is assumed that Ps1 = Ps3

GENERAL

CONVERSION OF MANUFACTURER’S RATINGS TO OPERATING CONDITIONS

Fan direct connected to motor. Motor efficiency data supplied by motor manufacturer.

Rating Point #1

Fan performance, at standard air density, as supplied by fan manufacturer for 1750 rpm. Point

CFM

Ps

HP

1) 2) 3)

8900 8520 8060

0 1/8 1/4

1.45 1.50 1.55

CALCULATIONS DENSITIES For Plane 3 conditions of: td3 = 79°F tw3 = 63°F pb3 = pb + (Ps2 - Ps1)/13.6 = 29.19 + (0.13/13.6) = 29.2 in. Hg Use Figure N.1 in Annex N to obtain ρ3 = 0.0715 lbm/ft3. It is assumed that ρ1 = ρ3. FAN POWER INPUT At the measured power input value of 1395 watts, the data supplied by the motor manufacturer indicate efficiency of 77% for the motor. Hmo = (1390 × 0.77)/746 = 1.43 hp Since the fan is direct connected to the motor, there is no drive loss, and: H

= Hmo = 1.43 hp

FAN

Q1c = 8900 (1735/1750) = 8824 cfm Ps1c = 0 H1c = 1.45 (1735/1750)3 (0.0715/0.0750) = 1.35 hp Rating Point #2 Q2c = 8520 (1735/1750) = 8447 cfm Ps2c = 0.125 (1735/1750)2 (0.0715/0.0750) = 0.117 in. wg H2c = 1.50 (1735/1750)3 (0.0715/0.0750) = 1.39 hp Rating Point #3 Q3c = 8060 (1735/1750) = 7991 cfm Ps3c = 0.25 (1735/1750)2 (0.0715/0.0750) = 0.234 in. wg H3c = 1.55 (1735/1750)3 (0.0715/0.0750) = 1.44 hp Draw a performance curve for these operating conditions. Enter the measured values for static pressure and horsepower on the appropriate curves. Ideally, these two points will coincide at the same cfm. However, usually they will not coincide and should be averaged to determine the fan airflow rate. If this difference is small, such as in this example, it is only a reflection of test inaccuracies. If, however, these differences exceed 10%, the system should be reanalyzed for SEFs that may have been overlooked, or for procedural errors in the initial testing. Field Performance Measurement | 239

Qa = 8070 cfm (based upon horsepower) Qb = 8400 cfm (based upon static pressure) Use: Q = 8235 cfm (average of above).

x

x BHP

x 1.25

.30

STATIC PRESSURE IN. WG (Ps)

x 1.00

.20

x .10

SP

x

0 7000

8000

9000 CFM(Q)

Fan Performance at 0.0715 Air Density

240 | Field Performance Measurement

BHP (H)

1.50

.40

Annex B. Pitot Static Tubes 16D

8D 0.8D

0.5D Radius

0.4D D

3D Radius

Head shall be free from nicks and burrs. 90° ± 0.1°

All dimensions shall be within ±2%. SECTION A-A

Static Pressure

8 holes - 0.13D, not to exceed 0.04 in., diameter equally spaced and free from burrs. Hole depth shall not be less than the hole diameter.

Note: Surface finish shall be 32 micro in. or better. The static orifices may not exceed 0.04 in. diameter. The minimum Pitot tube stem diameter recognized under this standard shall be 0.10 in. In no case shall the stem diameter exceed 1/30 of the test duct diameter.

Total Pressure

PITOT-STATIC TUBE WITH SPHERICAL HEAD All other dimensions are the same as for spherical head pitot-static tubes. 8D

D

X

0.2D Diameter V

X/D

V/D

X/D

V/D

0.000 0.237 0.336 0.474 0.622

0.500 0.496 0.494 0.487 0.477

1.602 1.657 1.698 1.730 1.762

0.314 0.295 0.279 0.266 0.250

0.741 0.936 1.025 1.134 1.228

0.468 0.449 0.436 0.420 0.404

1.796 1.830 1.858 1.875 1.888

0.231 0.211 0.192 0.176 0.163

1.313 1.390 1.442 1.506 1.538 1.570

0.388 0.371 0.357 0.343 0.333 0.323

1.900 1.910 1.918 1.920 1.921

0.147 0.131 0.118 0.109 0.100

ALTERNATE PITOT-STATIC TUBE WITH ELLIPSOIDAL HEAD Figure B.1 Field Performance Measurement | 241

Annex C. Double Reverse Tubes AIR FLOW TUBE ENDS MUST BE SMOOTH AND FREE FROM BURRS

IMPACT TUBE

REVERSE TUBE

SECTION VIEW

STAINLESS STEEL TUBING PREFERRED APPROX. 0.375 in. OD

TOTAL PRESSURE = READING A CORRECTED FOR MANOMETER CALIBRATION

READING A

FLEXIBLE TUBING

ING B

READ

VELOCITY PRESSURE = READING B CORRECTED FOR MANOMETER CALIBRATION AND CALIBRATION FACTOR FOR THE DOUBLE REVERSE TUBE. Notes: 1. For use in dirty or wet gas streams. 2. The double reverse tube must be calibrated and used in the same orientation as used in its calibration 3. Also referred to as impact reverse tube, combined reverse tube, and type S tube. Figure C.1 - Double Reverse Tube 242 | Field Performance Measurement

Annex D. Pitot-Static Tube Holder

0.312 in. DIA.

PITOT-STATIC TUBE SPLIT BRASS BUSHING PRESS TO FIT INTO TUBING

THERMOCOUPLE

DUCT WALL

1½ in. PIPE HALF-COUPLING WELDED TO DUCT BRASS BUSHINGS

1½ in. PIPE NIPPLE 12 in. LONG

STAINLESS STEEL TUBING 1 in. OUTSIDE DIA. × 8 ft. LONG SLIP FIT IN BRASS BUSHINGS

Notes: ¼ in. OUTSIDE DIA. STAINLESS STEEL TUBING FOR GAS SAMPLING

SPLIT BRASS BUSHING

1. Apparatus for mounting Pitot-static tube on duct 2. For use in large ducts or high velocity gas streams 3. 1 in. diameter tube slides inside 1.5 in. pipe, which can be unscrewed and moved to another traverse location 4. The gas sampling tube and thermocouple may be omitted if these data are obtained in other manners

CUT-OFF AND REBRAZE AFTER ASSEMBLY

Figure D.1 - Pitot-Static Tube Holder (Typical) Field Performance Measurement | 243

Annex E. Static Pressure Tap

DUCT WALL

MAXIMUM 0.125 in. DIAMETER FOR USE IN RELATIVELY CLEAN GASES. MAY BE NECESSARY TO INCREASE TO 0.312 in. DIAMETER FOR DIRTY OR WET GASES ½ in. PIPE HALF-COUPLING OR SIMILAR ARRANGEMENT

INSIDE SURFACE OF DUCT AND EDGE OF HOLE ARE TO BE SMOOTH AND FREE FROM BURRS

Figure E.1 - Static Pressure Tap

MINIMUM OF FOUR TAPS, LOCATED 90° APART AND NEAR THE CENTER OF EACH WALL

STATIC PRESSURE MEASUREMENT REQUIRED AT EACH TAP. USE THE AVERAGE OF THE MEASUREMENTS AS THE STATIC PRESSURE FOR THE PLANE

Figure E.2 - Locations of Static Pressure Taps

244 | Field Performance Measurement

Annex F. Pitot-Static Tube Connections

PLANE 2

PLANE 1

PLANE 4

PLANE 3

*SEF 1 Ps4 FAN STATIC PRESSURE Ps = - Ps1 - Pv1 + SEF 1 where Ps1 = Ps4 Pv1 = Pv3 Figure F.1 - Fan with Inlet Duct Only Ps2 = 0 PLANE 3

PLANE 5

Ps3

Ps3

P v3 *SEF 1 is due to no duct at fan outlet PLANE 2

PLANE 1

FAN STATIC PRESSURE Ps = Ps2 where Ps2 = Ps5 Pt1 = 0

Ps5

P v3

Figure F.2 - Fan with Outlet Duct Only ALTERNATE PLANE 5 PLANE 3

Ps5

PLANE 2

PLANE 1

FAN STATIC PRESSURE Ps = Ps2 - Ps1 - Pv1 where Ps2 = Ps5 Ps1 = Ps4 Pv1 = Pv3

PLANE 4

PLANE 3

Ps3

Ps4

P v3

Figure F.3 - Fan with Inlet Duct and Outlet Duct Field Performance Measurement | 245

Annex G. Manometer Data

10 in. wg 1:1 SLOPE RATIO

2 in. wg 5:1 SLOPE RATIO

0.5 in. wg 20:1 SLOPE RATIO

Figure G.1 - Manometer Data

246 | Field Performance Measurement

1 in. wg 10:1 SLOPE RATIO

PERCENT UNCERTAINTY IN VELOCITY DETERMINATION USING PITOT-STATIC TUBE AND MANOMETER DUE TO MANOMETER SLOPE Based on an uncertainty equivalent to an indicating column length of 0.05 in. wg in a vertical manometer (1:1 slope ratio)

.01

.02

VELOCITY PRESSURE READING, in. wg .04 .06 0.1 0.2 0.4 0.6 1 2

3 4

6 8 10

10.0 8.0 6.0 5.0

3.0

2.0

R TE ME TIO NO RA MA OPE SL 1:1

1.0 0.8 0.6 0.5 2:1

10 :1

:1

0.3

5:1

0.4

20

% UNCERTAINTY IN VELOCITY DETERMINATION

4.0

0.2

0.3

0.4

0.6

0.8

1

2

3

4

6

8

10

15

STANDARD AIR VELOCITY, fpm (×1000)

Figure G.2 - Uncertainty in Velocity Determination

Field Performance Measurement | 247

Annex H. Distribution of Traverse Points In order to obtain a representative average velocity in a duct, it is necessary to locate each traverse point accurately. It is recommended that the number of traverse points increase with increasing duct size. The distributions of traverse points for circular ducts, as indicated below, are based on log-linear Pitot traverse method.

X1

60º

X2

X3 X4

D

Xn

Xa = D × Ka Where: D is the inside diameter of the duct Ka is the factor corresponding to the duct size and the traverse point location as indicated in the table below NUMBER OF INSIDE TRAVERSE DIAMETER POINTS IN K1 OF DUCT EACH OF 3 DIAMETERS

K2

K3

K4

K5

K6

K7

K8

K9

K10

K11

K12

K13

K14

K15

K16

LESS THAN 8 ft.

8

.021 .117 .184 .345 .655 .816 .883 .979

8 ft. THROUGH 12 ft.

12

.014 .075 .114 .183 .241 .374 .626 .759 .817 .886 .925 .986

GREATER THAN 12 ft.

16

.010 .055 .082 .128 .166 .225 .276 .391 .609 .724 .775 .834 .872 .918 .945 .990

Figure H.1 - Distribution of Traverse Points for Circular Ducts 248 | Field Performance Measurement

The recommended minimum number of traverse points for rectangular ducts is indicated below in Figure H.3. For rectangular ducts with cross-sectional areas of 24 square feet and less, the recommended minimum number is 24. For cross-sectional areas greater than 24 square feet, the minimum number of points increases as indicated in Figure H.3. The points are to be located in the centers of equal areas with the areas as nearly square as practical (see Figure H.2). If the flow conditions at the traverse plane are less than satisfactory, the accuracy of the determination of flow rate may be improved by using more than the recommended minimum number of points. Fewer points may be used if the flow is very uniform; however, the maximum area covered per point should not exceed 3 square feet. Y

Y 2

X 2 X

Figure H.2 - Distribution of Traverse Points for Rectangular Duct

NUMBER OF TRAVERSE POINTS

100 90 80 70 60 50 40 30 25 20 15

10 10

15

20

25 30

40

50 60 70 80 100

150

200 250 300

DUCT CROSS-SECTIONAL AREA, ft2 Figure H.3 - Recommended Minimum Number of Traverse Points for Rectangular Ducts Field Performance Measurement | 249

Annex J. Instrumentation Characteristics

Table J.1 - Temperature Measurement

No. Measurement Means 1. Glass-stem thermometers Mercury-glass thermometer

Application Temp of gases and liquids by contact

Alcohol-glass thermometer Pentane-glass thermometers Jena or quartz mercury nitrogen thermometers 2. Gas thermometer 3. Resistance thermometers Platinum-resistance thermometer

Nickel-resistance thermometer

Precision F

Limitations

-38/575

Less than 0.1 to 10

In gases, accuracy affected by radiation

” ”

-100/100 -200/70

”

-38/1000 -459/1000

Primary standard

” ” ”

”

Requires considerable skill to use High cost; accuracy affected by radiation in gases Accuracy affected by radiation in gases

-320/1800

Less than 0.02 to 5

Remote readings; temp by contact

-150/300

0.3

Up to 600

0.1

”

” ”

Less than 0.01

Precision; remote readings; temp of fluids or solids by contact

Thermistors 4. Thermocouples Pt-Pt-Rh thermocouple

Approximate Range F

High cost; also, requires expensive measuring device Less accurate than above Subject to oxidation

500/3000

0.1 to 5

General testing of high temp; remote rapid readings by direct contact ” Same as above, especially suited for low temp

Up to 2200

0.1 to 15

Up to 1500 Up to 700

0.1 to 15 0.1 to 15

5. Beckman thermometers (metastatic)

For differential temp in same applications as in glass stem thermometer

9 diff

0.018

6. Bimetallic thermometers

For approx temp

0/1000

1, usually much more

7. Pressure-bulb thermometers Gas-filled bulb

Remote-testing

-100/1000

2

Caution must be exercised so that installation is correct

” ” For intensity of narrow spectra band of high temp radiation (remote)

20/500 -50/2100

2 2 15

” ”

9. Radiation pyrometers

For intensity of total high temp radiation (remote)

Any range

10. Seger cones (fusion pyrometers)

Approx temp (within temp source)

1000/3600

50

11. Indicating crayons

Approx temp (in surface)

125/900

12. Melting and boiling points of materials

Standards

All except extremely high temp

±1% Extremely precise

Chromel-alumel thermocouple Iron-constantain thermocouple Copper-constantan thermocouple Chromel-constantan thermocouple

Vapor-filled bulb Liquid-filled bulb 8. Optical pyrometers

Standard for thermocouples

1500 upward

Reprinted by permission from ASHRAE Handbook - 1989 Fundamentals

250 | Field Performance Measurement

Must be set for temp to be measured Time lag; unsuitable for remote use; unreliable

For laboratory use only

Table J.2 - Differential Pressure Measurement

No.

Range

Precision

1. Micromanometer

Measurement Means

Very low press. diff.

Application

0 to 6 in. H20

0.005 to 0.001 in. H20

Not readily portable; not easy to use with pulsating pressure

Limitations

2. Draft gauges

Moderately low press. diff.

0 to 10 in. H20

0.005 to 0.05 in. H20

Must be leveled carefully

3. Manometer

Medium press diff.

0 to 100 in. H20 or Hg

0.05 in.

Where used with liquid must be compensated for liquid density

4. Swinging-vane-type gauge

Moderately low press. diff.

0 to 0.5 in. H20 0 to 20 in. H20

5%

Generally usable to atmospheric pressure only

5. Bourdon-tube type

Medium to high press. diff., usually to atmosphere

Any

0.05 to 5%

Subject to damage due to over press-shock or pulsation

6. Pressure transducersstrain gauge, capacity, potentiometer, crystal, magnet

Remote reading, responds to rapid changes of pressure

0.05 to 50,000 psi

0.1 to 0.5%

Requires electronic amplifier and readout device

Table J.3 - Velocity Measurement No.

Measurement Means

Application

Range

Precision

5 to 50

10 to 20%

Air velocities in rooms, at outlets, etc; directional

30 to 24,000

5%

Not well suited for duct readings; needs periodic check calibration

3. Revolving-vane anemometer

Moderate air velocities in ducts and rooms; somewhat directional

100 to 3000

5 to 20%

Extremely subject to error with variations in velocities with space or time; easily damaged; needs periodic calibration

4. Pitot tube

Std instrument for measurement of duct velocities

180 to 10,000 with micromanometer 600 to 10,000 with draft gauges; 10,000 up with manometer

1 to 5%

Accuracy falls off at low end of range

5. Impact tube and sidewall or other static tap

High velocities, small tubes and where air direction may be variable

120 to 10,000 with micromanometer; 600 to 10,000 with draft gauges; 10,000 up with manometer

1 to 5%

Accuracy depends upon constancy of static pressure across stream section

6. Heated thermocouple anemometer

Air velocities in ducts, velocity distributions

10 to 2000

3 to 20%

Accuracy of some types not good at lower end of range; steady state measurements only

7. Hot-wire anemometer

(a) Low air velocities; directional and nondirectional available

1 to 1000

1 to 20%

Requires accurate calibration at frequent intervals; complex, costly

up to 60,000

1 to 20%

1. Smoke puff or airborne solid tracer

Low air velocities in rooms; highly directional

2. Deflecting-vane anemometer

(b) High air velocities

Limitations Awkward to use but valuable in tracing air movement

(c) Transient velocity and turbulence

Reprinted by permission from ASHRAE Handbook - 1989 Fundamentals

Field Performance Measurement | 251

Annex K. Phase Current Method for Estimating the Power Output of Three Phase Fan Motors

Use Equation A to estimate the Hmo for motors of 5 horsepower and greater, operating at 90% or more of FLA. The uncertainties will be less than 5%.

The power output of three phase motors can be estimated based on the relationship of motor current and motor power output. Two equations can be used in estimating the motor power output. The equations are as follows:

Use the average of Equation A and Equation B to estimate the Hmo for all motors operating at less than 90% of FLA and for 3 horsepower and smaller motors operating above 90% of FLA. An estimated Hmo less than 50% of NPH can contain 15% uncertainties or greater.

Equation A: ⎛ Measured amps ⎞ ⎛ Measured volts ⎞ Hmo = NPH ⎜ ⎟⎜ ⎟ FLA NPV ⎝ ⎠⎝ ⎠ Where: Hmo = motor power output NPH = nameplate horsepower FLA = full load amps NPV = nameplate volts measured volts = average of the measured phase volts measured amps = average of the measured phase amps Equation B: ⎛ Measured amps - NLA ⎞ ⎛ Measured volts ⎞ Hmo = NPH ⎜ ⎟⎜ ⎟ FLA - NLA NPV ⎝ ⎠⎝ ⎠

Where: NLA = average of the measured phase values of no load amps NPH = nameplate horsepower FLA = full load amps NPV = nameplate volts NLA can usually be obtained with the motor operating and the motor shaft coupling or belt drive disconnected. In the case where the fan impeller is mounted directly on the motor shaft, it will be necessary to remove the impeller in order to obtain NLA measurements.

252 | Field Performance Measurement

Figure K.1 represents the relationship of motor current and motor power output. The “dashed” lines between 0% NPH and 100% NPH for motor sizes shown represents Equation B. The solid lines between these same end points for the motor sizes shown represent the general shape of typical motor calibration amp/load curves. The solid line from 100% NPH and 100% FLA to 0% NPH and 0% FLA represents Equation A. These curves indicate that if you average the results of Equation A and Equation B for a specific measured amp draw, that your results approach the typical calibration curve. It also points out that the uncertainties are low if just Equation A is used above 90% FLA, especially in the larger integral motor horsepowers. Many fractional horsepower and small integral horsepower motors do not have a significant change in current from no load to full load. The actual ampsload characteristics for motors of the same horsepower rating can vary greatly from motor manufacturer to motor manufacturer. No load amperage (NLA) varies significantly for the same size motor between manufacturers. In addition, various motor design requirements result in different ampload characteristics even though the horsepower ratings of the motors are the same. These are some of the reasons that Figure K.1 cannot be used to determine the motor output directly. The chart is only intended to indicate the accuracy and suitability of using the above equations for estimating motor power output.

GENERALIZED CURVES ILLUSTRATING THE RELATIONSHIP OF HORSEPOWER TO AMPS FOR THREE PHASE MOTORS Do not use for determining actual motor horsepower DOTTED LINES PER EQUATION B: Hmo ∝ MEASURED AMPS - NLA/FLA - NLA 100

90

RATED HORSEPOWER 1 2

80

70 3 60 5 50 10 40

400 30 2500 20

10

0 0

10

20

30

40

50

60

70

80

90

100

% NAMEPLATE HORSEPOWER PER EQUATION A: Hmo ∝

MEASURED AMPS FLA

CAUTION: THIS CHART IS REPRESENTATIVE ONLY! SINCE THE AMP-LOAD CHARACTERISTICS OF THE SAME SIZE MOTOR WILL VARY BETWEEN THE VARIOUS MOTOR MANUFACTURERS, IT CANNOT BE USED TO DETERMINE THE HORSEPOWER OUTPUT OF A MOTOR. USE THE EQUATIONS AS DIRECTED ON THE PREVIOUS PAGE.

Field Performance Measurement | 253

Annex L. Estimated Belt Drive Loss Drive loss is defined as follows: Percent drive loss equals power to driving sheave minus power from driven sheaves times 100, divided by power to driving sheave. There are several things which can affect belt drive efficiencies. Some of these are: 1) Over-designed drives. This was considered good practice at one time because the drive would last longer. It will still last longer but it is more inefficient. 2) Multiple belts on subminimum diameter sheaves are less efficient than fewer belts on larger diameter sheaves. Both the National Electric Motor Association and the Rubber Manufacturer’s Association publish data dealing with minimum recommended sheave diameters. As these minimum sheave diameters are approached, the drive loss becomes greater. 3) A larger belt section than required will increase the drive loss. 4) A badly undertensioned drive will increase the drive loss. 5) Misaligned drives will increase the drive loss. Drive loss is manifested as heat in belt drives. Under ambient conditions of less than 100°F, well designed drives that operate efficiently will be warm to the touch immediately after being shut down. If the drive is uncomfortable to the touch (approximately 140°F or more), then the drive loss is high. Obviously poorly tensioned and misaligned drives should be corrected before estimating brake horsepowers and drive losses.

254 | Field Performance Measurement

100 80 60

DRIVE LOSS, % MOTOR POWER OUTPUT*

40 30 RANGE OF DRIVE LOSSES FOR STANDARD BELTS

20 15 10 8 6 4 3 2 1.5 1 0.3 0.4 0.6 0.8 1

2

3

4

6

8 10

20

30 40

60 80 100

200 300 400

600

MOTOR POWER OUTPUT, hp

HIGHER BELT SPEEDS TEND TO HAVE HIGHER LOSSES THAN LOWER BELT SPEEDS AT THE SAME HORSEPOWER *Drive losses are based on the conventional V-belt, which has been the “work horse” of the drive industry for several decades. EXAMPLE • Motor power output, Hmo, is determined to be 13.3 hp • The belts are the standard type and just warm to the touch immediately after shutdown • From chart, drive loss = 5.1% • Drive loss, HL = 0.051 × 13.3 = 0.7 hp • Fan power input, H = 13.3 - 0.7 = 12.6 hp Figure L.1 - Estimated Belt Drive Loss

Field Performance Measurement | 255

Annex M. Density Determinations

Since:

M.1 General

Ps1 = 0 p1 = pb = 28.60 in. Hg

This annex contains examples illlustrating the procedures for determining densities. Determinations of densities are shown for air and for gases other than air.

M.2 Determination of the density of air, general case Determine air density by using the Psychrometric Density Chart, shown in Figure N.1 in Annex N, the Psychrometric Density Table, shown in Annex N, or a calculation procedure which makes use of perfect gas relationships and the modified Apjohn equation for partial vapor pressure. Examples of the use of these procedures are included in this section. Each of the procedures requires knowledge of the pressure, dry-bulb temperature and wet-bulb temperature of the air. The Psychrometric Density Chart and the Psychrometric Density Table are limited to the temperatures and pressures normally encountered in fan applications. Limit the use of the calculation procedure that is based on perfect gas relationships and illustrated in Example M2.3, to instances in which the dry-bulb temperature is 180°F or less. Accurate wet-bulb temperature measurements are difficult to obtain when the dry-bulb temperature exceeds 180°F. When the dry-bulb temperature exceeds 180°F, it may be necessary to rely on site personnel for the water vapor content of the air. Alternately, commercially available instrumentation for dew point determination may be used. For the procedure required to determine density based on the data provided in either of the above cases, refer to Psychrometric Tables and Charts by Zimmerman and Lavine.1 EXAMPLE M2.1 The conditions that exist at the inlet of a fan that is not ducted on the inlet side are: td1 = 78°F tw1 = 62°F

The wet-bulb depression is: td1 - tw1 = 78 - 62 = 16°F For wet-bulb depression of 16°F, dry-bulb temperature of 78°F and absolute pressure of 28.60 in. Hg, obtain ρ1 = 0.0701 lbm/ft3 by using the Psychrometric Density Chart in Figure N.1 in Annex N. EXAMPLE M2.2 The conditions at a fan inlet, located at an elevation of 1000 ft above sea level are: Ps1 = -3.45 in. wg td1 = 85°F tw1 = 75°F Barometric pressure, obtained from a nearby airport, is 29.82 in. Hg at sea level. Using the data in Figure N.3 in Annex N, the barometric pressure at 1000 ft above sea level is: pb = 29.82 × 0.964 = 28.75 in. Hg The absolute pressure at the fan inlet is: p1 = pb + (Ps1/13.6) = 28.75 + (-3.45/13.6) = 28.50 in. Hg The wet-bulb depression is: td1 - tw1 = 85 - 75 = 10°F For dry-bulb temperature of 85°F, absolute pressure of 28.50 in. Hg and wet-bulb depression of 10°F, use the Psychrometric Density Table in Figures N.5 in Annex N to obtain:

ρ1 = 0.06829 + 10 × 0.000041 = 0.0687 lbm/ft3

1. O. T. Zimmerman and I. Lavine, Psychrometric Tables and Charts, 2nd ed. (Dover, N.H.: Industrial Research Service Inc., 1964)

256 | Field Performance Measurement

Example M2.3

EXAMPLE M3.1

The conditions at a fan inlet are:

Dry air is entering a fan inlet, located at an elevation of 1000 ft above sea level. The pressure and temperature at the inlet are:

Ps1 = -8.75 in. wg td1 = 146°F tw1 = 93°F The barometric pressure, pb, measured for the atmosphere to which Ps1 is referred, is 28.15 in. Hg. The absolute pressure at the fan inlet is: p1 = pb + (Ps1 /13.6) = 28.15 + (-8.75/13.6) = 27.51 in. Hg Use Figure N.2 in Annex N to obtain saturated vapor pressure, pe, of 1.562 in. Hg for the wet-bulb temperature of 93°F. Use the modified Apjohn equation for partial vapor pressure, pp, to obtain: pp = pe - p1 (td1 - tw1)/2700 = 1.562 - 27.51 (146 - 93)/2700 = 1.022 in. Hg

ρ1 is calculated by using perfect gas relationships:

ρ1 =

=

1.3257 ( p1 − 0.378 pp )

( td1 + 460 )

1.3257 ( 27.51 − 0.378 × 1.022 )

(146 + 460 )

= 0.0593 lbm/ft 3

M.3 Determination of the density of air, special cases The procedures for the determination of the density of air that are described in Section M.2 are valid for dry air, air that is saturated with water vapor and air that is partially saturated with water vapor. This section contains alternate procedures for cases in which it is known that the air is either dry or saturated. Knowledge that the air is either dry or saturated eliminates the usual requirement of the wet-bulb temperature determination; however, it should be noted that an incorrect assumption of either of these conditions can result in a significant uncertainty in the density determination.

Ps1 = -15 in. wg td1 = 95°F Barometric pressure, obtained from a nearby airport, is 29.24 in. Hg at sea level. Using the data in Figure N.3 in Annex N, the barometric pressure at 1000 ft above seal level is: pb = 29.24 × 0.964 = 28.19 in. Hg The absolute pressure at the fan inlet is: p1 = pb + (Ps1/13.6) = 28.19 + (-15/13.6) = 27.09 in. Hg Dry air at 29.92 in. Hg and 70°F has a density of 0.075 lbm/ft3. Consider the density of air to be directly proportional to absolute pressure and inversely proportional to absolute temperature. The density of the air at the fan inlet is calculated as follows:

ρ1 = 0.075 (p1/29.92) [(70 + 460)/(td1 + 460)] = 0.075 (27.09/29.92) [530/(95 + 460)] = 0.0648 lbm/ft3 EXAMPLE M3.2 Saturated air is enterting a fan inlet, located at an elevation of 1500 ft above sea level. The pressure and temperature at the inlet are: Ps1 = - 6.75 in. wg td1 = 103°F Barometric pressure, obtained from a nearby airport, is 29.66 in. Hg at sea level. Using the data in Figure N.3 in Annex N, the barometric pressure at 1500 ft above sea level is: pb = 29.66 × 0.947 = 28.09 in. Hg The absolute pressure at the fan inlet is: p1 = pb + (Ps1/13.6) Field Performance Measurement | 257

= 28.09 + (-6.75/13.6) = 27.59 in. Hg Refer to Figure N.4 in Annex N to obtain saturated air density of 0.06868 at 103°F and 29.92 in. Hg.

EXAMPLE M4.1 A gas is entering a fan inlet located at an elevation of 2000 ft above sea level. The pressure and temperature at the inlet are:

Assuming the density of saturated air to be directly proportional to absolute pressure, the density at the fan inlet is calculated as follows:

Ps1 = - 22 in. wg td1 = 230°F

ρ1 = 0.06868 (p1/29.92) = 0.06868 (27.59/29.92) = 0.0633 lbm/ft3

Barometric pressure, obtained from a nearby airport, is 29.92 in. Hg at sea level. The composition of the gas is 5.5% CO2, 1% CO, 15% O2, 1% H2, and 77.5% N2, by volume.

Assuming the density of saturated air to be directly proportional to absolute pressure is an approximation. The uncertainty in the density determination as a result of this approximation increases with increasing temperature and increases with increasing variation between the actual absolute pressure and 29.92 in. Hg, which is the stated pressure for the data in Figure N.4. The uncertainty will be approximately 1% or less under the following conditions: • • •

At 120°F and at an absolute pressure within 20% of 29.92 in. Hg At 150°F and at an absolute pressure within 10% of 29.92 in. Hg At 180°F and at an absolute pressure within 4% of 29.92 in. Hg

M.4 DETERMINATION OF THE DENSITY OF A GAS OTHER THAN AIR The determination of the density of a gas other than air may require the use of complex equipment. Unless specifically qualified, an expert should be consulted for the proper use of the equipment. If the gas is a complex mixture of various consitutuents, as found in certain industrial processes, it is suggested that the company chemist be consulted for the gas analysis. Particular care should be used if the gas is toxic, corrosive or explosive; and in these cases, consideration should be given to substituting air for the test. The first two examples in this section illustrate gas density determinations based on analyses that provide the relative amounts of the gas constituents. Typical flue gas density data, which is provided in Figure N.6 in Annex N, is illustrated in Example M4.3. Since the actual density may be significantly different from the density determined by using typical data, it is recommended that the typical data be used only in the even that more specific information is not available.

258 | Field Performance Measurement

The apparent molecular weight of the gas is determined as follows: Volume Molecular Component Fraction × Weight = lb/mole CO2 CO O2 H2 N2

0.055 0.01 0.15 0.01 0.775

44 28 32 2 28

1.00

2.42 0.28 4.80 0.02 21.70 29.22

Apparent molecular weight = (29.22/1.00) = 29.22 The density of the gas at 70°F and 29.92 in. Hg is calculated as follows: Apparent molecular weight 29.22 = 386.7 386.7 = 0.0756 lbm/ft 3

Using the data in Figure N.3 in Annex N, the barometric pressure at 2000 ft above sea level is: pb = 29.92 × 0.930 = 27.83 in. Hg The absolute pressure at the fan inlet is: p1 = pb + (Ps1/13.6) = 27.83 + (-22/13.6) = 26.21 in. Hg Consider the density of the gas to be directly proportional to absolute pressure and inversely proportional to absolute temperature. The density of the gas at the fan inlet is calculated as follows:

ρ1 = 0.0756 (p1/29.92)[(70 + 460)/(td1 + 460)] = 0.0756 (26.21/29.92) [530/(230 + 460)] = 0.0509 lbm/ft3 EXAMPLE M4.2 The conditions that exist at the inlet of a fan are Ps1 = -19.5 in. wg and td1 = 240°F. The barometric pressure, pb, measured for the atmospheric to which Ps1 is referred is 29.35 in. Hg. The composition of the gas is 5.5% CO2, 1% CO, 15% O2, 1% H2, and 77.5% N2 by weight. The apparent molecular weight of the gas is determined as follows: Volume Molecular Component Fraction × Weight = lb/mole CO2 CO O2 H2 N2

0.055 0.01 0.15 0.01 0.775

44 28 32 2 28

1.00

0.00125 0.00036 0.0047 0.005 0.0277 0.0390

Apparent molecular weight = 1/0.0390 = 25.6 The density of the gas at 70°F and 29.92 in. Hg is calculated as follows:

EXAMPLE M4.3 Flue gas is flowing at Plane 3, the Pitot traverse measurement plane. The flue gas is the result of using natural gas as the fuel. The conditions that exsit at Plane 3 are: Ps3 = 5.74 in. wg td3 = 680°F The barometric pressure, pb, measured for the atmosphere to which Ps3 is referred is 28.85 in. Hg. The absolute pressure at Plane 3 is: p3 = pb + (Ps3/13.6) = 28.85 + (5.74/13.6) = 29.27 in. Hg Refer to Figure N.6 in Annex N to obtain typical flue gas density when natural gas is used as the fuel of 0.0725 lbm/ft3 at 70°F and 29.92 in. Hg. Consider the density of the flue gas to be directly proportional to absolute pressure and inversely proportional to absolute temperature. The density of the gas at Plane 3 is calculated as follows:

ρ1 = 0.0725 (p3/29.92)[(70 + 460)/(td3 + 460)] = 0.0725 (29.27/29.92) [530/(680 + 460)] = 0.0330 lbm/ft3

Apparent molecular weight 25.6 = 386.7 386.7 = 0.0662 lbm/ft 3 The absolute pressure at the fan inlet is: p1 = pb + (Ps1/13.6) = 29.35 + (-19.5/13.6) = 27.92 in. Hg Consider the density of the gas to be directly proportional to absolute pressure and inversely proportional to absolute temperature. The density of the gas at the fan inlet is calculated as follows:

ρ1 = 0.0662 (p1/29.92)[(70 + 460)/(td1 + 460)] = 0.0662 (27.92/29.92) [530/(240 + 460)] = 0.0468 lbm/ft3

Field Performance Measurement | 259

76

74

72

70

68

66

64

62

60

58

56

54

52

50

48

46

44

42

0 DRY-BULB TEMPERATURE, °F

0.080

0.079 78

2 80

0.078

4

82

0.077

6

84

0.076

8

86

0.075

10

0.070

0.069

0.068

0.067

0.066

0.074

88

Wet-bulb depression = 4°F; proceed horizontally to 54°F dry-bulb temperature; read vertically to 29.9 in. Hg; read horizontally to the density -- ρ = 0.0769 lbm/ft3.

12

90

td = 54°F; tw = 50°F; pb = 29.9 in. Hg

Given: Solution:

• •

0.065

0.064

0.073

92

4.

14

94

Read vertically to the absolute pressure. Then read horizontally to the density.

3.

0.063

0.062

0.072

96

Proceed horizontally to the appropriate dry-bulb temperature.

2.

Example

Calculate wet-bulb depression. Enter chart at the left.

1.

0.061

0.060

16

g

0.071

98

28.0 28.2 28.4 28.6 28.8 29.0 29.2 29.4 29.6 29.8 30.0

in. H

18

20

22

24

26

28

30

32

34

36

38

OLU

ABS

40

TE

PRE SS

URE

260 | Field Performance Measurement

AIR DENSITY, lbm/ft3

WET-BULB DEPRESSION, °F

Figure N.1 - Psychrometric Density Chart

Field Performance Measurement | 261

tw °F

pe in. Hg

tw °F

pe in. Hg

tw °F

pe in. Hg

tw °F

pe in. Hg

tw °F

pe in. Hg

30 31 32 33 34

.1646 .1724 .1805 .1879 .1956

60 61 62 63 64

.5219 .5408 .5603 .5804 .6011

90 91 92 93 94

1.423 1.468 1.515 1.562 1.611

120 121 122 123 124

3.451 3.548 3.647 3.749 3.853

150 151 152 153 154

7.580 7.770 7.963 8.161 8.362

35 36 37 38 39

.2036 .2118 .2204 .2292 .2384

65 66 67 68 69

.6225 .6445 .6667 .6906 .7148

95 96 97 98 99

1.662 1.714 1.767 1.821 1.877

125 126 127 128 129

3.960 4.069 4.180 4.295 4.412

155 156 157 158 159

8.569 8.779 8.994 9.213 9.437

40 41 42 43 44

.2478 .2576 .2678 .2783 .2892

70 71 72 73 74

.7397 .7653 .7917 .8188 .8468

100 101 102 103 104

1.935 1.994 2.054 2.117 2.180

130 131 132 133 134

4.531 4.654 4.779 4.908 5.038

160 161 162 163 164

9.665 9.898 10.14 10.38 10.63

45 46 47 48 49

.3004 .3121 .3241 .3365 .3494

75 76 77 78 79

.8757 .9053 .9359 .9673 .9997

105 106 107 108 109

2.246 2.313 2.381 2.452 2.525

135 136 137 138 139

5.173 5.310 5.450 5.593 5.740

165 166 167 168 169

10.88 11.13 11.40 11.66 11.94

50 51 52 53 54

.3626 .3764 .3905 .4052 .4203

80 81 82 83 84

1.033 1.067 1.103 1.139 1.176

110 111 112 113 114

2.599 2.675 2.753 2.833 2.915

140 141 142 143 144

5.889 6.043 6.199 6.359 6.522

170 171 172 173 174

12.21 12.50 12.79 13.08 13.38

55 56 57 58 59

.4359 .4520 .4687 .4859 .5036

85 86 87 88 89

1.214 1.254 1.294 1.336 1.379

115 116 117 118 119

2.999 3.085 3.173 3.263 3.356

145 146 147 148 149

6.689 6.860 7.034 7.212 7.394

175 176 177 178 179 180

13.69 14.00 14.32 14.64 14.94 15.31

Adapted from ASHRAE Handbook - 1989 Fandamentals

Figure N.2 - Thermodynamic Properties of Water at Absolute Vapor Pressures, Inches of Mercury

262 | Field Performance Measurement

ALTITUDE ft.

SPECIFIC GRAVITY

PRESSURE in. Hg

ALTITUDE ft.

SPECIFIC GRAVITY

PRESSURE in. Hg

0 100 200 300 400

1.00 0.996 0.993 0.989 0.986

29.92 29.81 29.70 29.60 29.49

3000 3200 3400 3600 3800

0.896 0.890 0.833 0.877 0.870

26.82 26.62 26.42 26.23 26.03

500 600 700 800 900

0.982 0.979 0.975 0.971 0.968

29.38 29.28 29.17 29.07 28.96

4000 4200 4400 4600 4800

0.864 0.857 0.851 0.845 0.838

25.84 25.65 25.46 25.27 25.08

1000 1100 1200 1300 1400

0.964 0.961 0.957 0.954 0.950

28.86 28.75 28.65 28.54 28.44

5000 5200 5400 5600 5800

0.832 0.826 0.820 0.814 0.807

24.90 24.71 24.52 24.34 24.16

1500 1600 1700 1800 1900

0.947 0.944 0.940 0.937 0.933

28.33 28.23 28.13 28.02 27.92

6000 6500 7000 7500 8000

0.801 0.786 0.772 0.757 0.743

23.98 23.53 23.09 22.65 22.22

2000 2100 2200 2300 2400

0.930 0.926 0.923 0.920 0.916

27.82 27.72 27.62 27.52 27.42

8500 9000 9500 10000 15000

0.729 0.715 0.701 0.688 0.564

21.80 21.39 20.98 20.58 16.89

2500 2600 2700 2800 2900

0.913 0.909 0.906 0.903 0.899

27.32 27.21 27.11 27.01 26.91

20000 25000 30000 35000 40000

0.460 0.371 0.297 0.235 0.185

13.75 11.10 8.89 7.04 5.54

Note: Specific gravity of standard air at sea level and 29.92 in. Hg = 1.00 Figure N.3 - Relative Specific Gravity of Air at Various Altitudes1

1. Robert Jorgensen, ed., Fan Engineering, 7th ed. (Buffalo, NY, Buffalo Forge Co., 1970) p.8 - Reprinted by Permission

Field Performance Measurement | 263

PROPERTIES OF SATURATED AIR2

Temp °F

WEIGHT IN A CUBIC FOOT OF MIXTURE

WEIGHT OF THE VAPOR

VOLUME ft3/lb

OF TOTAL lb/lb lb/lb DRY AIR VAPOR WEIGHT DRY AIR OF OF lb lb lb DRY AIR MIXTURE

Temp °F

WEIGHT IN A CUBIC FOOT OF MIXTURE

VOLUME ft3/lb

WEIGHT OF THE VAPOR

OF TOTAL lb/lb lb/lb DRY AIR VAPOR WEIGHT DRY AIR OF OF lb lb lb DRY AIR MIXTURE

-25 -20 -15 -10 -5

.09134 .09025 .08922 .08820 .08723

.000018 .000024 .000031 .000041 .000053

.09136 .09027 .08925 .08824 .08728

10.95 11.07 11.21 11.34 11.46

.00020 .00027 .00035 .00046 .00061

.00020 .00027 .00035 .00046 .00061

46 47 48 49 50

.07768 .00750 .07731 .07714 .07694

.000509 .000527 .000545 .000567 .000587

.07819 .07803 .07785 .07771 .07753

12.87 12.90 12.93 12.96 12.99

.00655 .00680 .00705 .00734 .00762

.00651 .00675 .00700 .00728 .00756

0 5 10 15 20

.08625 .08529 .08434 .08340 .08247

.000068 .000087 .000110 .000140 .000176

.08632 .08538 .08445 .08354 .08264

11.59 11.72 11.85 11.99 12.12

.00080 .00102 .00130 .00168 .00213

.00080 .00102 .00130 .00168 .00213

51 52 53 54 55

.07676 .07657 .07637 .07620 .07600

.000608 .000632 .000651 .000675 .000700

.07737 .07720 .07702 .07687 .07670

13.02 13.06 13.09 13.12 13.15

.00792 .00823 .00854 .00884 .00921

.00786 .00819 .00845 .00877 .00913

21 22 23 24 25

.08230 .08210 .08193 .08173 .08156

.000185 .000193 .000202 .000213 .000222

.08248 .08229 .08213 .08194 .08178

12.15 12.18 12.20 12.23 12.26

.00225 .00235 .00246 .00260 .00272

.00224 .00234 .00245 .00259 .00271

56 57 58 59 60

.07582 .07562 .07544 .07524 .07506

.000723 .000749 .000775 .000801 .000829

.07654 .07637 .07622 .07604 .07589

13.19 13.22 13.25 13.29 13.32

.00952 .00989 .01026 .01063 .01103

.00943 .00980 .01016 .01052 .01091

26 27 28 29 30

.08136 .08117 .08099 .08083 .08063

.000233 .000243 .000254 .000264 .000277

.08159 .08141 .08124 .08109 .08090

12.29 12.32 12.34 12.37 12.40

.00285 .00300 .00314 .00328 .00345

.00284 .00299 .00313 .00327 .00344

61 62 63 64 65

.07486 .07468 .07447 .07429 .07408

.000857 .000886 .000916 .000947 .000979

.07572 .07557 .07539 .07524 .07506

13.35 13.39 13.42 13.46 13.49

.01143 .01185 .01229 .01273 .01320

.01130 .01171 .01214 .01257 .01303

31 32 33 34 35

.08043 .08025 .08006 .07989 .07970

.000290 .000303 .000315 .000327 .000339

.08072 .08055 .08038 .08022 .08004

12.43 12.46 12.49 12.51 12.54

.00362 .00378 .00393 .00409 .00426

.00361 .00376 .00392 .00408 .00425

66 67 68 69 70

.07390 .07369 .07350 .07330 .07310

.001012 .001045 .001080 .001115 .001152

.07491 .07473 .07458 .07441 .07425

13.53 13.57 13.60 13.64 13.68

.01368 .01417 .01468 .01520 .01576

.01349 .01397 .01447 .01497 .01551

36 37 38 39 40

.07952 .07933 .07916 .07897 .07880

.000353 .000364 .000380 .000394 .000409

.07987 .07969 .07954 .07936 .07921

12.57 12.60 12.63 12.66 12.69

.00444 .00460 .00480 .00499 .00519

.00442 .00458 .00478 .00496 .00516

71 72 73 74 75

.07290 .07270 .07250 .07229 .07208

.001189 .001229 .001268 .001310 .001352

.07409 .07393 .07377 .07360 .07343

13.71 13.75 13.79 13.83 13.87

.01630 .01691 .01748 .01812 .01876

.01604 .01662 .01717 .01780 .01841

41 42 43 44 45

.07860 .07843 .07825 .07805 .07788

.000425 .000440 .000456 .000473 .000491

.07902 .07887 .07871 .07852 .07837

12.72 12.75 12.78 12.81 12.84

.00541 .00561 .00583 .00606 .00630

.00538 .00558 .00579 .00602 .00626

76 77 78 79 80

.07188 .07166 .07144 .07124 .07104

.001395 .001439 .001485 .001532 .001579

.07328 .07310 .07293 .07277 .07262

13.91 13.95 13.99 14.03 14.08

.01941 .02008 .02079 .02150 .0223

.01904 .01968 .02036 .02106 .02174

Figure N.4 - Weights of Air, Water Vapor, and Saturated Mixture of Air and Water Vapor at Different Temperatures and 29.92 in. Hg

2. Jorgensen, op. cit., pp 15-17

264 | Field Performance Measurement

Reprinted by Permission

PROPERTIES OF SATURATED AIR2

Temp °F

WEIGHT IN A CUBIC FOOT OF MIXTURE

VOLUME ft3/lb

WEIGHT OF THE VAPOR

OF TOTAL lb/lb lb/lb DRY AIR VAPOR WEIGHT DRY AIR OF OF lb lb lb DRY AIR MIXTURE

Temp °F

WEIGHT IN A CUBIC FOOT OF MIXTURE

VOLUME ft3/lb

WEIGHT OF THE VAPOR

OF TOTAL lb/lb lb/lb DRY AIR VAPOR WEIGHT DRY AIR OF OF lb lb lb DRY AIR MIXTURE

81 82 83 84 85

.07081 .07059 .07038 .07015 .06993

.001629 .001680 .001733 .001785 .001840

.07244 .07227 .07211 .07193 .07177

14.12 14.16 14.21 14.26 14.30

.02301 .02380 .02462 .02545 .02631

.02249 .02325 .02403 .02482 .02566

116 117 118 119 120

.06186 .06154 .06124 .06092 .06060

.004427 .004548 .004669 .004794 .004921

.06629 .06609 .06591 .06571 .06552

16.16 16.24 16.32 16.41 16.50

.07157 .07390 .07625 .07869 .08121

.06678 .06882 .07084 .07296 .07511

86 87 88 89 90

.06970 .06947 .06925 .06902 .06880

.001898 .001954 .002014 .002072 .002139

.07160 .07142 .07126 .07109 .07094

14.34 14.39 14.44 14.48 14.53

.02723 .02813 .02908 .03002 .03109

.02651 .02736 .02826 .02915 .03015

121 122 123 124 125

.06027 .05995 .05960 .05927 .05892

.005049 .005183 .005319 .005456 .005598

.06532 .06513 .06492 .06473 .06452

16.58 16.68 16.77 16.87 16.96

.08376 .08646 .08925 .09204 .09502

.07729 .07958 .08194 .08428 .08677

91 92 93 94 95

.06855 .06832 .06809 .06785 .06760

.002201 .002267 .002334 .002404 .002474

.07075 .07058 .07042 .07025 .07007

14.58 14.63 14.69 14.73 14.79

.03211 .03318 .03428 .03543 .03660

.03111 .03212 .03314 .03422 .03531

130 135 140 145 150

.05713 .05524 .05319 .05100 .04865

.006355 .007195 .008128 .009162 .010303

.06349 .06244 .06132 .06016 .05895

17.49 18.10 18.79 19.60 20.55

.11125 .13026 .15280 .17966 .21178

.10010 .11523 .13255 .15230 .17478

96 97 98 99 100

.06736 .06711 .06688 .06660 .06634

.002546 .002620 .002692 .002770 .002853

.06991 .06973 .06957 .06931 .06919

14.84 14.90 14.95 15.01 15.07

.03780 .03904 .04025 .04159 .04300

.03642 .03757 .03870 .03993 .04124

155 160 165 170 175

.04612 .04340 .04048 .03734 .03398

.011547 .012937 .014436 .016118 .017926

.05767 .05634 .05492 .05346 .05191

21.67 23.03 24.69 26.77 29.43

.25038 .29810 .35660 .43168 .52750

.20022 .22962 .26285 .30150 .34530

101 102 103 104 105

.06610 .06583 .06557 .06530 .06504

.002937 .003019 .003106 .003193 .003283

.06904 .06885 .06868 .06849 .06832

15.12 15.18 15.25 15.31 15.37

.04443 .04586 .04737 .04890 .05048

.04255 .04385 .04523 .04662 .04806

180 185 190 195 200

.03035 .02645 .02228 .01779 .01297

.019905 .022062 .024393 .026957 .029730

.05036 .04851 .04667 .04475 .04270

32.94 37.78 44.85 56.20 77.11

.65580 .83410 1.0948 1.5153 2.2923

.39525 .45425 .52270 .60240 .69660

106 107 108 109 110

.06477 .06451 .06421 .06394 .06364

.003375 .003470 .003568 .003666 .003766

.06814 .06798 .06778 .06761 .06741

15.44 15.50 15.57 15.64 15.71

.05212 .05379 .05556 .05734 .05917

.04953 .05105 .05264 .05422 .05587

205 210 212

.00782 .00232 .00000

.032715 .035942 .037298

.04064 .03836 .03730

127.9 431.0 ____

4.1838 15.493 Inf.

.80500 .93700 1.0000

111 112 113 114 115

.06336 .06306 .06278 .06247 .06216

.003872 .003978 .004085 .004199 .004311

.06723 .06704 .06686 .06667 .06647

15.78 15.85 15.93 16.00 16.08

.06111 .06308 .06507 .06722 .06935

.05760 .05934 .06110 .06299 .06486

Figure N.4 - Weights of Air, Water Vapor, and Saturated Mixture of Air and Water Vapor at Different Temperatures and 29.92 in. Hg

2. Jorgensen, op. cit., pp 15-17

Reprinted by Permission

Field Performance Measurement | 265

Density of Saturated Air for Various Barometric Conditions - lbm/ft3 Dry-Bulb Temp. °F

Barometric Pressure in. Hg

Approximate average Increase in increase in density per density per °F wet-bulb 0.1 in. pressure depression

28.5

29.0

29.5

30.0

30.5

31.0

30 31 32 33 34

.07703 .07687 .07671 .07654 .07638

.07839 .07822 .07806 .07789 .07772

.07974 .07957 .07940 .07924 .07907

.08110 .08093 .08075 .08058 .08041

.08245 .08228 .08210 .08193 .08175

.08380 .08363 .08345 .08327 .08310

.00027 .00027 .00027 .00027 .00027

.000017 .000017 .000017 .000018 .000018

35 36 37 38 39

.07621 .07605 .07589 .07573 .07557

.07756 .07739 .07723 .07706 .07690

.07890 .07873 .07856 .07840 .07823

.08024 .07807 .07990 .07973 .07956

.08158 .08141 .08123 .08106 .08089

.08292 .08274 .08257 .08239 .08222

.00027 .00027 .00027 .00027 .00027

.000018 .000018 .000019 .000019 .000019

40 41 42 43 44

.07541 .07525 .07509 .07493 .07477

.07674 .07657 .07641 .07625 .07609

.07806 .07790 .07773 .07757 .07740

.07939 .07922 .09705 .07889 .07872

.08072 .08055 .08038 .08021 .08004

.08205 .08187 .08170 .08153 .08135

.00027 .00026 .00026 .00026 .00026

.000019 .000020 .000020 .000020 .000020

45 46 47 48 49

.07461 .07445 .07429 .07413 .07397

.07592 .07576 .07560 .07544 .07528

.07724 .07707 .07691 .07674 .07658

.07855 .07838 .07822 .07805 .07788

.07986 .07970 .07953 .07936 .07919

.08118 .08101 .08084 .08066 .08049

.00026 .00026 .00026 .00026 .00026

.000020 .000021 .000021 .000021 .000022

50 51 52 53 54

.07381 .07366 .07350 .07334 .07318

.07512 .07496 .07479 .07464 .07447

.07642 .07625 .07609 .07593 .07576

.07772 .07755 .07739 .07722 .07706

.07902 .07885 .07868 .07852 .07835

.08032 .08015 .07998 .07981 .07964

.00026 .00026 .00026 .00026 .00026

.000022 .000022 .000023 .000023 .000023

55 56 57 58 59

.07302 .07287 .07271 .07255 .07240

.07431 .07415 .07399 .07383 .07367

.07560 .07544 .07528 .07512 .07495

.07689 .07673 .07656 .07640 .07623

.07818 .07801 .07784 .07768 .07751

.07947 .07930 .07913 .07896 .07879

.00026 .00026 .00026 .00026 .00026

.000024 .000024 .000025 .000025 .000025

60 61 62 63 64

.07224 .07208 .07193 .07177 .07161

.07352 .07336 .07320 .07304 .07288

.07479 .07463 .07447 .07430 .07414

.07607 .07590 .07574 .07557 .07541

.07734 .07718 .07701 .07684 .07668

.07862 .07845 .07828 .07811 .07794

.00026 .00026 .00026 .00026 .00026

.000026 .000026 .000027 .000027 .000028

Note: Approximate average decrease in density per 0.1°F rise in dry-bulb temperature equals .000017 lbm/ft3. Figure N.5 - Psychrometric Density Table (I-P) 266 | Field Performance Measurement

Psychrometric Density Table (I-P) Density of Saturated Air for Various Barometric Conditions - lbm/ft3 Dry-Bulb Temp. °F

Barometric Pressure in. Hg

Approximate average Increase in increase in density per density per °F wet-bulb 0.1 in. pressure depression

28.5

29.0

29.5

30.0

30.5

31.0

65 66 67 68 69

.07145 .07130 .07114 .07098 .07083

.07272 .07256 .07240 .07224 .07208

.07398 .07382 .07366 .07350 .07333

.07525 .07508 .07492 .07475 .07459

.07651 .07634 .07618 .07601 .07584

.07770 .07760 .07744 .07727 .07710

.00026 .00026 .00026 .00026 .00026

.000028 .000029 .000029 .000030 .000030

70 71 72 73 74

.07067 .07051 .07035 .07020 .07004

.07192 .07176 .07160 .07144 .07128

.07317 .07301 .07285 .07268 .07252

.07442 .07426 .07410 .07393 .07377

.07568 .07551 .07534 .07517 .07501

.07693 .07676 .07659 .07642 .07625

.00026 .00025 .00025 .00025 .00025

.000031 .000031 .000032 .000033 .000033

75 76 77 78 79

.06988 .06972 .06956 .06940 .06925

.07112 .07096 .07080 .07064 .07048

.07236 .07220 .07203 .07187 .07171

.07360 .07343 .07327 .07310 .07294

.07484 .07467 .07451 .07434 .07417

.07603 .07591 .07574 .07557 .07540

.00025 .00025 .00025 .00025 .00025

.000034 .000034 .000035 .000036 .000036

80 81 82 83 84

.06909 .06893 .06877 .06861 .06845

.07032 .07015 .07000 .06983 .06967

.07155 .07138 .07122 .07105 .07089

.07277 .07261 .07244 .07227 .07211

.07400 .07383 .07366 .07349 .07333

.07523 .07506 .07489 .07472 .07454

.00025 .00025 .00024 .00024 .00024

.000037 .000038 .000039 .000039 .000040

85 86 87 88 89

.06829 .06812 .06796 .06780 .06764

.06950 .06934 .06917 .06901 .06885

.07072 .07056 .07039 .07022 .07005

.07194 .07177 .07160 .07143 .07126

.07316 .07299 .07281 .07264 .07247

.07437 .07420 .07403 .07385 .07368

.00024 .00024 .00024 .00024 .00024

.000041 .000042 .000043 .000043 .000044

90 91 92 93 94

.06748 .06731 .06715 .06698 .06682

.06868 .06852 .06835 .06818 .06801

.06989 .06972 .06955 .06938 .06921

.07109 .07092 .07075 .07058 .07041

.07230 .07213 .07195 .07178 .07161

.07351 .07333 .07316 .07298 .07280

.00024 .00024 .00024 .00024 .00024

.000045 .000046 .000047 .000048 .000049

95 96 97 98 99

.06665 .06648 .06632 .06615 .06598

.06785 .06768 .06751 .06734 .06717

.06904 .06887 .06870 .06853 .06835

.07024 .07006 .06989 .06972 .06954

.07143 .07126 .07108 .01091 .07073

.07263 .07245 .07227 .07209 .07191

.00024 .00024 .00024 .00024 .00024

.000050 .000051 .000052 .000053 .000054

100

.06581

.06700

.06818

.06937

.07055

.07174

.00024

.000055

Note: Approximate average decrease in density per 0.1°F rise in dry-bulb temperature equals .000017 lbm/ft3. Figure N.5 - Psychrometric Density Table (I-P) Field Performance Measurement | 267

FUEL

FLUE GAS DENSITY lbm/ft3

COAL

0.078

OIL

0.075

NATURAL GAS

0.0725

BAGASSE

0.070

BLAST FURNACE GAS

0.076

LIGNITE

0.073

WOOD

0.070

The above densities at 70°F and 29.92 in. Hg are based on average fuel analyses and moisture contents Figure N.6 - Typical Densities for Various Flue Gases

268 | Field Performance Measurement

Annex P. Diffusion at Fan Outlets

BLAST AREA DISCHARGE DUCT CUTOFF

OUTLET AREA

25% 50% 75% CENTRIFUGAL FAN 100% EFFECTIVE DUCT LENGTH AXIAL FAN

To calculate 100% effective duct length, assume a minimum of 2½ duct diameters for 2500 fpm or less. Add 1 duct diameter for each additional 1000 fpm. Example: 5000 fpm = 5 equivalent duct diameters If the duct is rectangular, with side dimensions equal to a and b, the equivalent duct diameter is equal to (4ab/π)0.5

Figure P.1 - Controlled Diffusion and Establishment of a Uniform Velocity Profile in a Straight Length of Outlet Duct

Field Performance Measurement | 269

Annex R. Terminology for Fans and Air Handling Units CASING

BACKPLATE RIM INLET

HUB

MOTOR GUIDE VANE

BLADE IMPELLER

INLET BELL

Tubular Centrifugal Fan - Direct Drive CASING

BLADE DIFFUSER HUB

MOTOR

IMPELLER CASING

Tubeaxial Fan-Direct Drive (Impeller Downstream)

BEARING CASING BELT TUBE BLADE

HUB

GUIDE VANE

Vaneaxial Fan-Belt Drive IMPELLER

INLET BOX

BEARINGS

FAN CASING

GUIDE VANES

MECHANISM FOR CONTROLLING BLADE ANGLE

INNER CYLINDER

IMPELLER DIFFUSER

Vaneaxial Mechanical Draft Fan

Figure R.1 - Common Terminology for Axial and Tubular Centrifugal Fans 270 | Field Performance Measurement

HOUSING

DIVERTER CU

TO

FF

CENTER PLATE BLAST AREA DISCHARGE OUTLET AREA SIDE SHEET BACKPLATE

FF

BLADE

TO

CU

INLET

SCROLL IMPELLER FRAME RIM BEARING SUPPORT INLET COLLAR

Figure R.2 - Common Terminology for Centrifugal Fan Field Performance Measurement | 271

Figure R.3 - Common Terminology for Centrifugal Fan Appurtenances 272 | Field Performance Measurement

HEATING AND VENTILATING DRAW-THROUGH UNIT FS

BELT GUARD FS CS

EXT F & BP

MB

FB

INT F & BP

HC

MB

FB

AS

+

+

HEATING AND VENTILATING BLOW-THROUGH UNIT ZONE DAMPERS

FS HC

BYPASS COLD DECK

+ +

HOT DECK

+

+

AIR-CONDITIONING DRAW-THROUGH UNIT FS

AS

MB

FB

CC

SS

HC

ELIM

+

+

+

MB

FB

+

+ +

+

+

DRIP TRAY

AIR-CONDITIONING BLOW-THROUGH UNIT DIFFUSER PLATE

ZONE DAMPERS

HC

CC

+

+

+

HOT DECK

FS

+

HC

FB

MB

+

COLD DECK CC

+

+

+

FLEXIBLE CONNECTION AS CS CC HC

ACCESS SECTION COIL SECTION COOLING COIL HEATING COIL

EXT F & BP INT F & BP ELIM

EXTERNAL FACE AND BYPASS DAMPER INTERNAL FACE AND BYPASS DAMPER ELIMINATORS

FS FB MB SS

FAN SECTION FILTER BOX MIXING BOX SPRAY SECTION

Figure R.4 - Common Terminology for Central Station Air-Handling Units Field Performance Measurement | 273

Annex S. Typical Format for Field Test Data Sheet

FIELD TEST DATA SHEET JOB DESCRIPTION: Location, User, Contractor, Engineer, . . . . . FAN DESCRIPTION: Mfgr., Size, Type, Ident. No., . . . . . MOTOR DESCRIPTION: Mfgr., Nameplate Data (Ident. No., hp, volts, FLA, . . . ), Performance Data Reference, . . . . . DRIVE DESCRIPTION: Type, Mfgr., Ident. No., Size, . . . . . REFERENCE DRAWINGS OR SKETCHES OF INSTALLATION: System Configuration with Dimensions, Measurement Plane Locations, . . . . . MEASUREMENTS AMBIENT DATA: Barometric Pressure, Dry-Bulb Temp., Wet-Bulb Temp, . . . . . MOTOR DATA: volts, amps, watts, rpm, . . . . . FAN SPEED GAS DENSITY DATA: GAS TEMPERATURES AT MEASUREMENT PLANES:

READING

Ps1 or Ps4

Ps2 or Ps5

Ps3

Pv3

Pv3

1 2 3 4 5

• • • • n TOTAL AVERAGE

CALCULATIONS: (Refer to the various sections of this publication for the appropriate calculation procedures.)

Figure S.1 - Typical Format for Field Test Data Sheet

274 | Field Performance Measurement

Annex T. Uncertainty Analysis T.1 Introduction In an attempt to determine the range of uncertainties likely to be encountered in field testing of fans, a statistical uncertainty analysis was undertaken. Maximum and minimum uncertainties were assigned to each quantity to be measured based on the degree of difficulty in measuring the quantity, the previously specified accuracies of instruments and the conditions expected to be encountered in field testing. These individual maximum and minimum uncertainties were then combined statistically to arrive at the probable range of overall uncertainties for the fan flow rate, fan static pressure, and fan power input. It would be unlikely, however, that any particular field installation would have all minimum or all maximum uncertainties occurring simultaneously. Therefore, an agreement by the parties as to acceptable measurement tolerances for a given installation should be established prior to testing. In Type A tests, it may be sufficient to accept the results of any field test without consideration of the probable uncertainties in the results. For Type B and Type C tests, it may be necessary to calculate the uncertainties. To do this, each measured quantity is assigned an estimated uncertainty by agreement of the parties involved and the overall uncertainty is calculated as outlined in this annex.

T.2 General This analysis is based on the assumption that fan perfomance can be treated as a statistical quantity and that the performances derived from repeated tests would have a normal distribution. The most probable performance would, therefore, be the mean results based on repeated observations at each point of operation. Only one set of observations is specified in this publication. This analysis deals, therefore, with the probable uncertainty in the results obtained from a single set of observations. The results of a fan field performance test for a single point of operation are a combination of variables which are normally presented graphically. Test results will be considered to be the fan static pressure versus flow rate and fan power input versus flow rate. The uncertainty in results will be expressed in terms of fan flow rate, fan static pressure, and fan power input. The accuracies specified in this publication are based upon two standard deviations. This means that there should be a 95% probability that the actual uncertainties will be less than the specified value.

This applies only to random uncertainties. Systematic uncertainties should be eliminated by the use of properly calibrated test instruments. This analysis considers only the uncertainties inherent in testing. This publication specifies uncertainties in percent. These are, of course, per unit uncertainties, multiplied by 100. Absolute uncertainties which bear the units of the quantity being measured or calculated, are equal to the per unit uncertainty multiplied by the measured or calculated quantity. Since the tolerance on measured values is specified on the basis of 95% confidence limits, the actual deviations in results will be less than the calculated deviations 95% of the time. For the purposes of a field test, an uncertainty range will be defined with minimum and maximum values. This range of possible uncertainty is necessary to cover the varying degrees of difficulty encountered in performing tests in field installations. Field test conditions range from near ideal to near impossible.

T.3 Symbols In the analysis that follows, certain symbols and notations are used in addition to those shown in Annex Q. Symbol

Quantity

ex ∆X R

Per Unit Uncertainty in X Absolute Uncertainty in X Gas Constant (ft-lb/lbm —°R)

Subscript

Description

A b d f g h H N P Q w x ρ

area Barometric Pressure Dry-bulb Temperature Velocity Pressure Static Pressure Power Input Fan Power Input Fan Speed Fan Static Pressure Fan Flow Rate Wet-bulb Depression Generalized Quantity (A, b, ..., ρ) Density

T.4 Measurement uncertainties The various measurement uncertainty ranges used in this publication are listed below. The considerations that led to their adoption include difficulties in field testing generally not encountered in laboratory testing. Field Performance Measurement | 275

T.4.1 Barometric pressure. The estimated uncertainty in measuring barometric pressure is between 0.3% minimum and 0.7% maximum. eb = 0.003 (min) to 0.007 (max) Barometric pressure is generally obtained by portable aneroid barometer, on-site barometer (mercury or aneroid) or by use of data obtained from a nearby airport. The uncertainty range above is estimated based on the use of portable or on-site instrumentation and applicable corrections. T.4.2 Dry-bulb temperature. The estimated uncertainty in measuring dry-bulb temperature is between 0.5% of absolute temperature minimum and 2.0% of absolute temperature maximum. ed = 0.005 (min) to 0.02 (max) The estimated uncertainty range is based on a broad temeprature range and the likelihood of stratification. T.4.3 Web-bulb depression. The estimated uncertainty in measuring wet-bulb depression is between 5°F minimum and 10°F maximum. ew = 5/(td - tw) (min) to 10/(td - tw) (max) The estimated uncertainty range is based on a broad temperature range with the associated difficulties in determining wet-bulb readings at high or low temperatures and the likelihood of stratification. T.4.4 Fan speed. The estimated uncertainty in measuring fan speed is between 0.5% minimum and 1.0% maximum. eN = 0.005 (min) to 0.01 (max) The uncertainty range in fan speed is estimated on the basis of portable instrumentation accuracy and an allowance for fluctuation in fan speed. T.4.5 Power input. The estimated uncertainty in measuring power input is betwen 3.0% minimum and 7.0% maximum. eh = 0.03 (min) to 0.07 (max) The estimated uncertainty range is based on the various measurement methods and their respective accuracies, estimated drive losses, and the broad horsepower range encountered in the field. T.4.6 Pitot traverse. A properly performed field traverse is estimated to have an accuracy of 1.5% minimum to 7.5% maximum. 276 | Field Performance Measurement

ec = 0.015 (min) to 0.075 (max) The uncertainty range in the Pitot traverse is estimated on the basis of traverse location, broad range of duct sizes, nonuniform velocity profiles, and turbulence. T.4.7 Flow measurement area. The estimated uncertainty in the flow measurement area is between 1.0% minimum to 2.0% maximum. eA = 0.010 (min) to 0.020 (max) The estimated uncertainty is based on a broad range of duct sizes, accessibility, and the rigidity of ducts under pressure. T.4.8 Velocity pressure. An allowance of 2.0% minimum to 5.0% maximum of the reading is estimated for the mental averaging performed on a fluctuating reading. An allowance of 1.0% minimum to 2.0% maximum of the reading is estimated for calibrated manometer uncertainty and relocation of the instrument after calibration. In addition, an allowance of 0.5% minimum to 10.0% maximum of the reading is estimated for instrument precision. No allowance is included for yaw on the assumption that the Pitot-static tube is aligned within 10 degrees of streamlines. A combined uncertainty can be written as: ef (min) = [(0.02)2 + (0.01)2 + (0.005)2]0.5 = 0.0229 ef (max) = [(0.05)2 + (0.02)2 + (0.10)2]0.5 = 0.1136 T.4.9 Static pressure. An allowance of 1.0% minimum to 5.0% maximum of the reading is estimated for the mental averaging performed on a fluctuating reading. An allowance of 1.0% minimum to 2.0% maximum of the reading is estimated for calibrated manometer uncertainty and relocation of the instrument after . In addition, a tolerance of 10% minimum to 20.0% maximum of the fan velocity pressure should cover the influence of Pitot-static tube yaw or velocity influence on static pressure taps and other possible effects. A combined uncertainty can be written as: eg (min) = {(0.01)2 + (0.01)2 + (0.005)2 + [0.1 Pv/(Ps2 - Ps1)]2}0.5 = {0.000225 + [0.1 Pv/(Ps2 - Ps1)]2}0.5 eg (max) = {(0.05)2 + (0.02)2 + (0.02)2 + [0.2 Pv/(Ps2 - Ps1)]2}0.5 = {0.0033 + [0.2 Pv/(Ps2 - Ps1)]2}0.5

Where the denominator in the final term in each equation will involve Ps2 or Ps5 and Ps1 or Ps4, whichever are measured.

Assuming ∆70.73 and ∆R are both zero:

The estimated uncertainty range is based on an allowance for fluctuation in the fan-system operation, lack of ideal measurement locations, turbulence, and the relocation of instrumentation after calibration.

It can be shown that: ev2 = [(0.00000725 tw - 0.0000542) ∆(td - tw)]2 Where:

T.5 Combined uncertainties The uncertainties in the test performance are the result of using various values, each of which contains a probable uncertainty. The combined uncertainty for each of the fan performance variables is given below. T.5.1 Density. Air density involves the various psychrometric measurements and the approximate formula:

ρ=

eρ = (eb2 + ev2 + ed2)0.5

70.73 pbV R ( t d + 460 )

∆(td - tw) = Absolute uncertainty in wet-bulb depression. Other methods for determining density are assumed to have equal accuracy. T.5.2 Fan flow rate. Fan flow rate directly involves the area at the flow measuring station, the Pitot traverse, the square root of the pressure measurement for flow, and the square root of the density. Uncertainties in fan speed will produce a first-power uncertainty in flow rate when making the fan law conversions. Combining: eQ = [ec2 + eA2 (ef/2)2 + (eρ/2)2 + eN2]0.5

Where: V = 1.0 - 0.378 {(pe/pb) - [(td - tw)/2700]} For random and independant uncertainties in products, the combined uncertainty is determined as follows: ∆ρ/ρ = {(∆70.73/70.73)2 + (∆pb/pb)2 + (∆V/V)2 + (∆R/R)2 + [∆td/(td + 460)]2}0.5

T.5.3 Fan static pressure. Fan static pressure directly involves static pressure measurements. Uncertainties in density will produce a first-power uncertainty in fan static pressure while uncertainties in fan speed will produce a second-power uncertainty in fan static pressure when making fan law conversions. Combining: ep = [eg2 + eρ2 + (2eN)2]0.5

Table T.1 Measurement eb ed** eW eN eh ec eA ef eg

Minimum

Maximum

0.003 0.005 5/(td - tw) 0.005 0.030 0.015 0.010 0.0229 {0.000225 + [0.1 Pv/(Ps2 - Ps1)]2}0.5

0.007 0.020 10/(td - tw) 0.010 0.070 0.075 0.020 0.1136 {0.0033 + [0.2 Pv/(Ps2 - Ps1)]2}0.5

* These uncertainties do not account for the effect of swirl at the fan inlet. This situation must be corrected in order to produce acceptable fan-system performance (see Section 5). ** Based on absolute temperature Field Performance Measurement | 277

In order to simplify the application of this uncertainty analysis to the results of field tests, the above equation was developed on the basis of tests in which static pressure measurements are made at a single plane, as would be the case in which a fan is ducted on one side only. However, the equation is reasonably accurate for all other fan-system configurations. Although in most cases the determination of fan static pressure involves Pv1, the uncertainty in determining Pv1 is not included in the above equation on the basis that it normally has a very small effect on the overall uncertainty in fan static pressure. For purposes of this publication, eP is applied directly to Psc, which may include System Effect Factors. T.5.4 Fan power input. Fan power input directly involves the power measurement; in addition, when making fan law conversions, density has a first-power effect and speed has a third-power effect on fan power input. Combining: eH = [eh2 + eρ2 + (3eN)2]0.5

T.6 Summary The minimum and maximum measurement uncertainties (See Table T.1) were defined earlier in Section T.4. Summarizing, the per unit uncertainties are as shown in Table T.1. The uncertainty calculations lead to absolute uncertainties in fan flow rate, fan static pressure, and fan power input that can be applied directly to the corresponding test results. The uncertainty results can then be plotted as rectangles around the test point. Intersection of the rectangles with the quoted fan performance within the limitations of a field test. See the examples in Section T.7.

T.7 Examples Two examples of the calculation of uncertainties and the method of comparison with the quoted fan curve are included in this section. Uncertainty calculations and comparisons have been developed for Examples 2B and 2C of Annex A. Uncertainty calculations for Example 2B utilize all minimum uncertainty tolerances. Uncertainty calculations for Example 2C utilize all maximum uncertainty tolerances. It would be unlikely that any field installation would lend itself to all minimum or all maximum measurement tolerances. Agreement of the parties as to acceptable measurement tolerances for a given installation should be established prior to testing. 278 | Field Performance Measurement

EXAMPLE 1: CALCULATION OF UNCERTAINTIES IN TEST RESULTS BASED ON MINIMUM MEASUREMENT UNCERTAINTY TEST VALUES Reference: Example 2B in Annex A SITE MEASUREMENTS td2 = tw2 = Ps1 = Ps2 = Pv3 = A2 = A3 = ρ2 = ρ3 =

91.3°F 70.4°F -11.4 in. wg 0.1 in. wg 1.24 in. wg 1.40 ft2 1.57 ft2 0.0714 lbm/ft3 0.0705 lbm/ft3

CONVERTED RESULTS Qc = 7114 cfm Psc = 11.42 in. wg Hc = 18.90 hp MEASUREMENT UNCERTAINTIES Reference: Minimum values per Section T.6 eb ed ew eN eh ec eA ef eg

= = = = = = = = =

0.003 0.005 5/(td2 - tw2) 0.005 0.030 0.015 0.010 0.0229 {0.000225 + [0.1 Pv/(Ps2 - Ps1)]2}0.5 CALCULATIONS

Pv = = = =

Pv2 Pv3 (A3/A2)2 (ρ3/ρ2) 1.24 (1.57/1.40)2 (0.0705/0.0714) 1.54 in. wg

eg = {0.000225 + [0.1 Pv/(Ps2 - Ps1)]2}0.5 = {0.000225 + [(0.1 × 1.54)/(0.1 + 11.4)]2}0.5 = 0.02011

ev2 = [(0.00000725 tw - 0.0000542) ∆(td - tw)]2 = [(0.00000725 × 70.4 - 0.0000542) 5]2 = 0.00000520 eρ = [eb2 + ev2 + ed2)0.5 = (0.0032 + 0.00000520 + 0.0052)0.5 = 0.006261 eP = [eg2 + eρ2 + (2eN)2]0.5 = [0.020112 + 0.0062612 + (2 × 0.005)2]0.5 = 0.0233 eQ = [ec2 + eA2 + (ef/2)2 + (eρ/2)2 + eN2]0.5 = [0.0152 + 0.0102 + (0.0229/2)2 + (0.006261/2)2 + 0.0052]0.5 = 0.0222 eH = [eh2 + eρ2 + (3eN)2]0.5 = [0.0302 + 0.0062612 + (3 × 0.005)2]0.5 = 0.0341 ∆P = ePPsc = 0.0233 × 11.42 = 0.27 in. wg Psc + ∆P = 11.42 + 0.27 = 11.69 in. wg Psc - ∆P = 11.42 - 0.27 = 11.15 in. wg ∆Q = eQQc = 0.0222 × 7114 = 158 cfm Qc + ∆Q = 7114 + 158 = 7272 cfm Qc - ∆Q = 7114 - 158 = 6956 cfm ∆H = eHHc = 0.0341 × 18.90 = 0.64 hp Hc + ∆H = 18.90 + 0.64 = 19.54 hp Hc - ∆H = 18.90 - 0.64 = 18.26 hp

Field Performance Measurement | 279

GRAPHICAL PRESENTATION

Psc

Psc + ∆P

TEST POINT MINIMUM UNCERTAINTY RANGE

Ps, FAN STATIC PRESSURE

Psc - ∆P

Qc = 7114 cfm ∆Q = 158 cfm Psc = 11.42 in. wg ∆P = 0.27 in. wg Hc = 18.90 hp ∆H = 0.64 hp

Qc + ∆Q

Qc - ∆Q Qc

QUOTED FAN PERFORMANCE CURVES

Q, FAN FLOW RATE

H, FAN POWER INPUT

Hc + ∆H Hc Hc - ∆H

Qc + ∆Q

Qc - ∆Q

Qc Q, FAN FLOW RATE

Figure T.1

280 | Field Performance Measurement

EXAMPLE 2: CALCULATION OF UNCERTAINTIES IN TEST RESULTS BASED ON MAXIMUM MEASUREMENT UNCERTAINTIES TEST VALUES Reference: Example 2C in Annex A SITE MEASUREMENTS

eQ = [ec2 + eA2 + (ef/2)2 + (eρ/2)2 + eN2]0.5 = [0.0752 + 0.0202 + (0.1136/2)2 + (0.02176/2)2 + 0.0102]0.5 = 0.0973 eH = [eh2 + eρ2 + (3eN)2]0.5 = [0.0702 + 0.021762 + (3 × 0.010)2]0.5 = 0.0792

td3 = 86.5°F tw3 = 75.5°F Ps4 = -1.57 in. wg Ps5 = 1.22 in. wg Pv2 = 0.61 in. wg

∆P = eP Psc = 0.0780 × 2.54 = 0.20 in. wg

CONVERTED RESULTS

Psc - ∆P = 2.54 - 0.20 = 2.34 in. wg

Psc + ∆P = 2.54 + 0.20 = 2.74 in. wg

Qc = 25964 cfm Psc = 2.54 in. wg Hc = 17.11 hp MEASUREMENT UNCERTAINTIES Reference: Maximum values per Section T.6 eb ed eW eN eh ec eA ef eg

= = = = = = = = =

0.007 0.020 10/(td3 - tw3) 0.010 0.070 0.075 0.020 0.1136 {0.0033 + [0.2 Pv/(Ps5 - Ps4)]2}0.5 CALCULATIONS

∆Q = eQQc = 0.0973 × 25964 = 2526 cfm Qc + ∆Q = 25964 + 2526 = 28490 cfm Qc - ∆Q = 25964 - 2526 = 23438 cfm ∆H = eHHc = 0.0792 × 17.11 = 1.36 hp Hc + ∆H = 17.11 + 1.36 = 18.47 hp Hc - ∆H = 17.11 - 1.36 = 15.75 hp

eg = {0.0033 + [0.2 Pv/(Ps5 - Ps4)]2}0.5 = {0.0033 + [(0.2 × 0.61)/(1.22 + 1.57)]2}0.5 = 0.07219 ev2 = [(0.00000725 tw - 0.0000542) ∆(td - tw)]2 = [(0.00000725 × 75.5 - 0.0000542) 10]2 = 0.0000243 eρ = (eb2 + ev2 + ed2)0.5 = (0.0072 + 0.0000243 + 0.0202)0.5 = 0.02176 eP = [eg2 + eρ2 + (2eN)2]0.5 = [0.072192 + 0.021762 + (2 × 0.010)2]0.5 = 0.0780

Field Performance Measurement | 281

GRAPHICAL PRESENTATION

TEST POINT MAXIMUM UNCERTAINTY RANGE Qc = 25964 cfm ∆Q = 2526 cfm Psc + ∆P Psc = 2.54 in. wg ∆P = 0.20 in. wg

Ps, FAN STATIC PRESSURE

Psc

Hc = 17.11 hp ∆H = 1.36 hp

Psc - ∆P

Qc - ∆Q

Qc + ∆Q

Qsc

H, FAN POWER INPUT

Q, FAN FLOW RATE

QUOTED FAN PERFORMANCE CURVES

Hc + ∆H Hc Hc - ∆H Qc - ∆Q

Qc + ∆Q Qsc

Q, FAN FLOW RATE

Figure T.2

282 | Field Performance Measurement

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The Air Movement and control Association International, Inc. is a not-for-profit international association of the world’s manufacturers of related air system equipment primarily, but limited to: fans, louvers, dampers, air curtains, airflow measurement stations, acoustic attenuators, and other air system components for the industrial, commercial and residential markets.