Fans Reference Guide

Fans Reference Guide

4th edition, 2001

Fans 9501 Covers 1/8/01 10:12 AM Page 3

First Edition, September 1993Second Edition, October 1997Third Edition, August 1999Fourth Edition, January 2001

Coordinated by:

Scott Rouse, P.Eng., MBA.Ontario Hydro 1997

Revised by:

Richard Okrasa, P. Eng., MBA.Ontario Hydro

Written by:

Ralph G. Culham, P. Eng.Consulting Engineerfor Technology Services Department, Ontario Hydro, 1993

Neither Ontario Hydro, nor any person acting on its behalf, assumes anyliabilities with respect to the use of, or for damages resulting from the use of, anyinformation, equipment, product, method or process disclosed in this guide.

Printed in CanadaCopyright © 1993, 1997, 1999, 2001 Ontario Power Generation

Making Energy Savings Good Business


FANS

Reference Guide

3rd Edition, 1999

Fans (00-06) 1/8/01 10:35 AM Page c

i

TA B L E O F C O N T E N T S

INTRODUCTION .......................................................................1

DEFINITIONS ...........................................................................3Fans ...............................................................................................3Blowers .........................................................................................3Velocity Pressure...........................................................................4Static Pressure ...............................................................................4Total Pressure................................................................................4Fan Total Pressure Rise .................................................................4Fan Velocity Pressure ....................................................................4Fan Static Pressure ........................................................................5Fan Duty .......................................................................................5Fan Output Power ........................................................................5Fan Efficiency................................................................................5System Curve................................................................................5Performance Curve .......................................................................5Fan Static Efficiency......................................................................5

FAN TYPES ..............................................................................7Centrifugal Fans............................................................................7

Airfoil .............................................................................................9Backward-inclined ........................................................................11Radial ..........................................................................................11Forward-curved ............................................................................12

Axial Fans....................................................................................12Propeller.......................................................................................16Tubeaxial .....................................................................................16Vanaxial ......................................................................................20

Special designs ............................................................................21

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Bifurcated Fans.............................................................................21Centrifugal Inline Fans .................................................................21Centrifugal Roof Exhausters ..........................................................25Utility Fans ..................................................................................25

Fan Designation and Arrangements ...........................................25Class Limits for Fans...................................................................31

PRINCIPLES OF OPERATION .....................................................33Centrifugal Fans..........................................................................34Axial Fans....................................................................................36

FAN PERFORMANCE CURVES....................................................39

FAN LAWS .............................................................................43Limitations ..................................................................................43Compressibility Factor................................................................44

FAN FORMULAE .....................................................................47Density........................................................................................47Fan Flow Rate .............................................................................48Head and Pressure ......................................................................49Velocity Pressure.........................................................................49Total Pressure..............................................................................50Fan-System-Effect Factor ............................................................50Fan Power and Efficiency ...........................................................50Fan Motor Power........................................................................52Example 1....................................................................................53

AIR SYSTEMS.........................................................................59Example 2....................................................................................60

FAN AND SYSTEM INTERFACE ..................................................65System Effect Factors..................................................................65Fan Outlet Conditions................................................................66

ii


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Fan Inlet Conditions ...................................................................68

FAN SELECTION .....................................................................69Pressure Definitions....................................................................70Parallel Fan Selection ..................................................................72Series Vs Parallel Operation........................................................74

FAN NOISE ............................................................................77Fan Sound Power........................................................................78Example 3....................................................................................81

FAN DUTY CONTROL .............................................................85

VIBRATION ISOLATION............................................................89

ELECTRIC MOTOR FAN DRIVE.................................................91Flywheel Effect............................................................................91AC Motors ..................................................................................92DC Motors..................................................................................93

ENERGY CONSUMPTION ANALYSIS ...........................................95Constant-Volume Fans ...............................................................96Variable-Volume Fans .................................................................97Example 4..................................................................................102

APPENDICES ........................................................................109Appendix A – Density Calculations .........................................109Appendix B – Drive Loss Calculations.....................................115Appendix C – Fan Outlet Loss Coefficients.............................119

CONVERSION TABLES ...........................................................127

ABBREVIATIONS AND SYMBOLS..............................................131

BIBLIOGRAPHY.....................................................................135

GLOSSARY...........................................................................139

iii


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1. Operating Point .........................................................................62. General Configuration and Component Terms

for Centrifugal Fans ...................................................................83. Airfoil .........................................................................................94. Typical Characteristics of Airfoil Fans.....................................105. Typical Characteristics of Backward-inclined Fans.................136. Typical Characteristics of Radial Fans.....................................147. General Configuration and Component

Terms for Axial Fans ................................................................158. Typical Characteristics of Propeller Fans.................................179. Typical Characteristics of Tubeaxial Fans ...............................18

10. Typical Characteristics of Vaneaxial Fans................................1911. Configuration of Bifurcated Fans ............................................2212. Typical Characteristics of Centrifugal Inline Fans...................2313. Typical Characteristics of Centrifugal Roof Exhausters..........2414. Typical Characteristics of Forward-curved Utility Fans..........2615. Typical Characteristics of

Backward-inclined Utility Fans ...............................................2716. Drive Arrangements for Axial Fans

with or without Diffuser and Outlet Box ...............................2817. Drive Arrangements for Centrifugal Fans ...............................2918. Drive Arrangements for Centrifugal Fans ...............................3019. Outlet Velocity Vector Diagram for

Backward-inclined Blades........................................................3520. Outlet Velocity Vector Diagram for Radial Blades..................3521. Outlet Velocity Vector Diagram for

Forward-curved Blades ............................................................35

iv

L I S T O F F I G U R E S

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22. Velocity Vector Diagram for an Axial Fan without Inlet Guide or Diffusion Vanes near the Impeller Hub...........36

23. Velocity Vector Diagram for an Axial Fan without Inlet Guide or Diffusion Vanes at the Blade Tip .....................36

24. Fan Test-rig Setup ....................................................................4125. Compressibility Factor ............................................................4526. Operating Point and System Curve.........................................5327. Fan Static-pressure Design Curve at 1,475 rpm

Intersecting Design Point A and Fan Static-pressure Curve at 983 rpm Intersecting Point B ....................................56

28. Fan Static Pressure Curve Intersecting the Design Point A and the Maximum Design Point D .........................................63

29. Deficient Fan and System Performance ..................................6630. Fan-outlet Velocity Profiles......................................................6731. Design Operating Point Selection Range on a Typical

Centrifugal Fan Performance Curve ........................................7232. Pressure Flow Curves ..............................................................7333. Series Fan Operation................................................................7534. Outlet Damper Fan Control ....................................................8635. Throttle Control of a Fan with a Two-speed Motor...............8736. Inlet Vane Control of a Fan......................................................88

v


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C1. Plane Asymmetric Diffuser at Fan Outlet Without Ductwork ................................................................119

C2. Pyramidal Diffuser at Fan Outlet Without Ductwork ..........120C3. Plane Symmetric Diffuser at Fan Outlet With Ductwork.....121C4. Plane Asymmetric Diffuser at Fan Outlet With Ductwork...122C5. Plane Asymmetric Diffuser at Fan Outlet With Ductwork...123C6. Plane Asymmetric Diffuser at Fan Outlet With Ductwork...124C7. Pyramidal Diffuser at Fan Outlet With Ductwork................125

vi


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1. Fan Laws ..................................................................................462. Typical Manufacturer's Performance Data

for a 24-in. AFSW Centrifugal Fan at 70˚F andStandard Atmospheric Pressure...............................................54

3. Typical Number of Fan Blades ................................................794. Specific Sound Power Levels and Blade

Frequency Increments .............................................................805. Sound Correction Factors........................................................816. Summary for Example 3 ..........................................................837. Typical VAV-fan Constants....................................................1048. Motor Load Efficiencies.........................................................1059. The Solution to Example 4 ....................................................105

10. Summary of Example 4 .........................................................107

vii

L I S T O F TA B L E S

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A1. Standard Atmospheric Data for Altitudes to 3,000 m .............................................................114

A2. Density Calculations .............................................................114C1. Plane Asymmetric Diffuser at Fan

Outlet Without Ductwork.....................................................119C2. Pyramidal Diffuser at Fan Outlet

Without Ductwork ................................................................120C3. Plane Symmetric Diffuser at Fan

Outlet With Ductwork ..........................................................121C4. Plane Asymmetric Diffuser at Fan



Outlet With Ductwork ..........................................................124C7. Pyramidal Diffuser at Fan Outlet With Ductwork................125

viii

L I S T O F TA B L E S

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• Fans and blowers are the largest single type of user of electricityin industry. Applications in all industries include: boilercombustion air supply, dust and exhaust removal (pneumaticconveying), “bag” house, sewage aeration, drying, coolingindustrial processes, and ventilation. Issues such as indoor airquality and pollution control create a continuous demand forwell-designed, efficient and cost-effective ventilation andblower systems.

• Selecting the right size and type of fan and blower isfundamental to an energy-efficient system.

• The first step in any fan application is defining the needs of thesystem being supplied.

• Enhancing the performance of an existing air system with anew, energy-efficient electronic control system offers significantpotential for energy savings. In some cases, retrofitting with amore efficient fan or blower and interconnecting ductwork willbe the most appropriate way to reduce energy consumption.

Chapter 1: Introduction 1

C H A P T E R 1

INTRODUCTION

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• This guide contains the information required to select anindustrial or commercial fan and blower system. Supportinghandbooks and reference material are identified in theBibliography.

• Chapter 7 provides the formulae necessary to determine theenergy consumption of a heating and ventilating fan system,particularly variable-volume fans.

• Because the personal computer is a popular design tool, theformulae in this guide were designed to be used in a spread-sheet program. Hourly analysis programs determine energy usemore accurately, and some of these programs can be used forsystem design and selection.

• This guide demonstrates how to use the American Society ofHeating, Refrigerating and Air Conditioning Engineers(ASHRAE) Modified Bin Method on a spread-sheet program todetermine annual energy consumption when fan power is afunction of outdoor air temperature. This procedure isreasonably accurate relative to the time required to perform theanalysis.

• Once the annual energy consumption is determined, a life-cyclecosting analysis of a proposed system can be done. This guideexcludes life-cycle costing techniques as they are welldocumented in texts such as the ASHRAE Handbook, 1991HVAC Applications Volume.

2 Fans Reference Guide

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FAN

• Device that causes flow of a gaseous fluid by creating apressure difference by exchanging momentum from the fanblades to air/gas particles.

• The fan impeller converts rotational mechanical energy intoboth static and kinetic energy within the gaseous fluid.

• The proportion of static versus kinetic energy created and theinherent energy conversion efficiency depends on the type offan (blade design).

• The gaseous fluid transported by a fan is most often air and/ortoxic fumes, whereas blowers may transport a mixture ofparticulate and air.

BLOWER

• Similar to a fan, except it can produce a much higher staticpressure. Sometimes higher pressure is achieved by a multistageimpeller arrangement.

• Engineering practice distinguishes fans and blowers for lowpressure and centrifugal compressors for high pressure.

Chapter 2: Definitions 3

C H A P T E R 2

DEFINITIONS

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• The demarcation between blowers and compressors is set at a7% increase in the density of the air from blower inlet toblower outlet.

• Fan and blower definitions and formulae, assumingincompressibility, apply below this demarcation withinsignificant errors.

VELOCITY PRESSURE

• That pressure at a point in an airstream existing by virtue of theair density and its rate of motion.

STATIC PRESSURE

• That pressure at a point in an airstream existing by virtue of theair density and its degree of compression, and is independent ofthe rate of motion of the air.

TOTAL PRESSURE

• That pressure at a point in an airstream existing by virtue of theair density and the degree of compression and rate of motion ofthe air; hence it is the sum of the static and velocity pressure(also called stagnation pressure).

FAN TOTAL PRESSURE RISE

• The fan total pressure at outlet, minus the fan total pressure atinlet. Note: when moving air enters a closed area, it convertsvelocity pressure to static pressure.

FAN VELOCITY PRESSURE

• The pressure corresponding to the average velocity determinedfrom the volume flow rate and fan outlet area.


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FAN STATIC PRESSURE

• The fan total pressure rise diminished by the fan velocitypressure.

FAN DUTY

• The range of operating points, giving the fan inlet volume flowat a rated fan pressure.

FAN OUTPUT POWER

• The fan output power or the useful power, delivered by a fan toan incompressible fluid, is equal to the product of the fan flowrate and the fan total pressure divided by a constant dependingon the units.

FAN EFFICIENCY

• The fan total or mechanical efficiency is defined as the ratio offan air power to fan-shaft input power.

SYSTEM CURVE

• The set of operating points defined by the duct friction, bends,and other pressure losses that make up the connected systemthe fan must serve.

PERFORMANCE CURVE

• The set of operating points defined by a particular fan design,size, and speed. Where the system and performance curvesmeet is the fan’s operating point.

FAN STATIC EFFICIENCY

• This is not a true efficiency but has been used traditionally inthe fan industry. It is equal to the fan total efficiency times theratio of fan static to fan total pressures

Chapter 2: Definitions 5

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Figure 1: Operating Point


P

V

FanPerformance

CurveSystem Curve

Operating Point

Volume Flow

Fan TotalPressure

or Fan StaticPressure

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• The two general classifications of fans – centrifugal and axial –are established according to the direction of flow through theimpeller.

- Axial fans have high volume capability for large duct sizeventilation applications.

- Centrifugal fans have high pressure capability for applicationssuch as boilers, baghouses, conveyors, and sewage aerators.

• These general classifications are subdivided into groups withinherent performance characteristics to suit a specificapplication.

• All other fans fall under a special design classification, includingmixed-flow fans.

CENTRIFUGAL FANS

• Centrifugal fans are divided into four main subclassifcationsaccording to impeller type: airfoil, backward-inclined, radial andforward-curved.

Chapter 3: Fan Types 7

C H A P T E R 3

FAN TYPES

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Figure 2: General Configuration and Component Terms forCentrifugal Fans

Reprinted with permission from the Air Movement and Control Association from Publication 201–90

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Airfoil

• The most efficient centrifugal fan design, but the most expensive.

• Airfoil (AF) have an impeller with typically 10 to 16 blades ofairfoil contour (see Figure 2), curved away from the direction ofrotation.

• Air leaves the impeller wheel at a velocity of less than its tipspeed, and relatively deep blades allow for efficient airexpansion within the blade passages.

• For a given duty, these fans rotate at the highest speed.

• The fan is in a scroll-type housing designed to efficientlyconvert velocity pressure to static pressure.

• To achieve high static-pressure efficiency, a close tolerancebetween the wheel and the housing inlet cone must bemaintained.

• Due to the high operating speed, the airfoil blades and the closetolerances, an AF fan is the most expensive to construct andrepair.

• It is the most efficient centrifugal fan design at approximately90%.

Rotation

Figure 3: Airfoil

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Figure 4: Typical Characteristics of Airfoil Fans

00

2 4 6 8 1210

1

2

3

4

5

0

20

40

60

80

100

cfm x 1,000

24 in.

surge

AFSW Centrifugal1,170 rpm

stat

ic p

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ure

(sp)

/ ho

rsep

ower

(hp)

stat

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(se)

hp

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APPLICATION

• Usually commercial heating, ventilating and air-conditioning(HVAC) systems and clean-air industrial applications where thepower savings can be significant.

• Best suited to applications that require low-to-medium staticpressure and a large flow volume.

Backward-inclined

• Backward-inclined (BI) or backward curved fans have animpeller with typically 10 to 16 blades of uniform thicknessincllined or curved away from the direction of rotation.

• The fan is in the same scroll-type housing as an AF fan.

• BI fans are slightly less efficient than AF fans at approximately80%.

APPLICATION

• In systems that require low-to-high static pressure, specificallyin commercial HVAC systems with moderate flow volume.Also used in industrial systems that require some tolerance to acorrosive or erosive environment. They are being usedincreasingly in industrial process ventilation with wear liners.

Radial Fans

• Radial (R) fans have an impeller wheel of high mechanicalstrength with typically six to 10 blades of heavy gauge materialradiating out from the hub.

• The blades can be either straight radial or modified radial with aslight curve. They are often equipped with removeable wearplates to extend the useful life of the fan impeller.

• For a given duty, R fans operate at medium speed.


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• They are the least efficient fan at 50% to 60%, but they do notclog and are easily repaired.

APPLICATION

• Primarily in industrial systems in a corrosive or erosiveenvironment, such as material handling of airborne particulateor where high static pressure is required.

Forward-curved

• Forward-curved (FC) fans have an impeller wheel made of lightgauge material, with typically 24 to 64 shallow blades withboth the heel and the tip curved forward.

• Air leaves the blade at a velocity greater than the tip speed, andprimarily kinetic energy is transferred to the air.

• These fans are the smallest of the centrifugal type and, for agiven duty, rotate at the slowest speed.

• The fan housing is a scroll design similar to the other centrifugalfan housings, except the tolerance between the inlet cone andthe wheel is not as critical allowing lighter gauge material to beused.

• FC fans are less efficient than AF and BI fans at approximately70%.

APPLICATION

• Generally in packaged and built-up, commercial and residentialHVAC systems with low-to-medium static pressures and lowair volumes. (See Figure 14, p. 26.)

AXIAL FANS

• Divided into three subclassifications according to impeller type:propeller, tubeaxial and vaneaxial.


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Figure 5: Typical Characteristics of Backward-inclined Fans

00

2 4 6 8 1210

1

2

3

4

5

0

20

40

60

80

100

cfm x 1,000

24 in.

surge

BISW Centrifugal1,170 rpm

stat

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(sp)

/ ho

rsep

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(hp)

stat

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(se)

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Figure 6: Typical Characteristics of Radial Fans

00

2 4 6 8 1210

2

4

6

8

10

0

20

40

60

80

100

cfm x 1,000

22 in.

surge

Radial1,170 rpm

stat

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(sp)

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rsep

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(hp)

stat

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(se)

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Figure 7: General Configuration and Component Terms for Axial Fans

Reprinted with permission from the Air Movement and Control Association from Publication 201–90.

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Propeller

• Have an impeller with two or more BI blades that are usuallymade from single, light-gauge material attached to a smalldiameter hub.

• Because primarily kinetic energy is transferred to the air withlittle static energy, these fans are limited to low-pressureapplications.

• The efficiency of these fans is low.

• The fan housing can be a simple ring or circular guard, anorifice plate, or an inlet cone with close tolerance to the bladetips to create a venturi for optimum performance.

APPLICATION

• Low static-pressure, high volume, commercial and industrialsystems.

Tubeaxial

• Have an impeller with typically four to eight blades attached toa hub that is usually less than half the diameter of the wheel.

• The blades can be AF construction or single thickness.

• Because the greatest portion of the work transferred to the air isstatic energy, these fans can be used in applications where thereis resistance to flow, e.g., ductwork systems.

• Tubeaxial fans are more efficient than propeller fans.

• The housing is a cylindrical tube with a close tolerance to theimpeller blade tips; this results in higher performance thanpropeller fans.


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Figure 8: Typical Characteristics of Propeller Fans

00

2 4 6 8 1210

1

2

3

4

5

0

20

40

60

80

100

cfm x 1,000

24 in.

surge

Propeller870 rpm

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(sp)

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(hp)

stat

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(se)

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Figure 9: Typical Characteristics of Tubeaxial Fans

00

2 4 6 8 1210

1

2

3

4

5

0

20

40

60

80

100

cfm x 1,000

24 in.

surge

Tubeaxial1,770 rpm

stat

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(sp)

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rsep

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(hp)

stat

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(se)

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Figure 10: Typical Characteristics of Vaneaxial Fans

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APPLICATION

• Low and medium static-pressure, commercial, ducted systemswhere the axial arrangement saves space and the downstreamflow pattern is not critical.

• Industrial systems where airborne contaminants collect on theimpeller blades and require periodic cleaning.

Vaneaxial

• Usually have short AF blades radiating from a hub greater thanhalf the diameter of the impeller.

• The blades are either fixed, adjustable or controllable (variablepitch-in-motion).

• The discharge from the impeller has a rotative component,unless inlet guide vanes are used.

• Because stationary diffusion vanes downstream of the impellerconvert rotary energy produced by the blades into staticpressure (as in an axial blower or compressor), primarily staticenergy is transferred to the air.

• Vaneaxial fans are the most efficient axial fan.

• The housing is a cylindrical tube with a close tolerance to theimpeller blade tips.

• The housing may include a set of inlet guide vanes and/ordownstream diffusion vanes equal in number to the impellerblades and preferably of the AF type.

APPLICATION

• Low to high static-pressure, commercial HVAC systems, andindustrial ventilation systems where the axial arrangementsaves space, and the downstream flow patterns and efficiencyare important.


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SPECIAL DESIGNS

Bifurcated Fans

• Air flows around the motor mounted directly on the fan shaft(see fig 11).

• Essentially axial fans with a special casing that allow the drivingmotor to be removed from the airstream while maintaining adirect-drive arrangement.

• In corrosive environments, the casing may be plastic or coated.

• The mating flanges at each end of the casing are identical, butthe casing diameter is increased in barrel fashion to allowpassage of a similar cross section of air, concentric with themotor enclosure.

APPLICATION

• Generally used to extract sticky, corrosive or volatile fumes inindustrial applications where it is critical to protect the motorfrom the airstream.

Centrifugal Inline Fans

• Have a direct-drive or a belt-driven AF or BI impeller mountedperpendicular to a rectangular or tubular casing with ampleclearance around the blade tips.

• The air discharged radially from the blade tips must turnthrough 90 degrees to pass through the fan exit, which is in linewith the impeller inlet.

APPLICATION

• Commercial applications where high efficiency, low soundlevels and space are prime considerations.


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Figure 11: Configuration of Bifurcated Fans

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Figure 12: Typical Characteristics of Centrifugal Inline Fans

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2 4 6 8 1210

1

2

3

4

5

0

20

40

60

80

100

cfm x 1,000

24 in.

surge

Centrifugal Inline870 rpm

stat

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(sp)

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rsep

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(hp)

stat

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(se)

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Figure 13: Typical Characteristics of Centrifugal Roof Exhausters

00

2 4 6 8 1210

1

2

3

4

5

0

20

40

60

80

100

cfm x 1,000

24 in.

surge

Centrifugal Roof Exhauster870 rpm

stat

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ure

(sp)

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rsep

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(hp)

stat

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(se)

hpse

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Centrifugal Roof Exhausters

• Have a direct-drive or a belt-driven AF or BI impeller mountedin a multicomponent housing comprising of a curb cap with anintegral inlet venturi, a fan shroud with drive-mountingsupport, and a weatherproof motor hood.

• The impeller has an inlet cone that allows mixed flow throughthe impeller-blade passages, and air exits radially from the bladetips through a concentric discharge passage.

• The fan shroud redirects the air - either discharging it down orblasting it up.

APPLICATION

• The down-discharge configuration is used for exhaustingrelatively clean air, while the up-blast configuration is used forhot and/or contaminated air.

Utility Fans

• Utility fans are self-contained units consisting of either an FC orBI irnpeller, a motor, and a direct (or belt-driven) drive.

APPLICATION

• Commercial and industrial ventilation applications requiringlow-to-medium air volumes and pressures.

FAN DESIGNATION AND ARRANGEMENTS

• The Air Moving and Conditioning Association, Inc (AMCA) hasdevised standard designations for fan rotation, dischargeorientation, motor position for belt or chain drive, inlet boxposition, and drive arrangements for both centrifugal and axialfans.


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Figure 14: Typical Characteristics of Forward-curved Utility Fans

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2 4 6 8 1210

1

2

3

4

5

0

20

40

60

80

100

cfm x 1,000

22 in.

surge

FCSW Utility500 rpm

stat

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(sp)

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(hp)

stat

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(se)

hp

se

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Figure 15: Typical Characteristics of Backward-inclined Utility Fans

00

2 4 6 8 1210

1

2

3

4

5

0

20

40

60

80

100

cfm x 1,000

22 in.

surge

BISW Utility1,170 rpm

stat

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ure

(sp)

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rsep

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(hp)

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(se)

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Figure 16: Drive Arrangements for Axial Fans with or withoutDiffuser and Outlet Box

Reprinted with permission from the Air Movement and Control Association from Publication 201–90.

Optional on allarrangements

For belt drive or direct connection. Impeller overhung. Two bearings located either upstream or downstream of impeller

Arr. 1 two-stageArr. 1

Arr. 3For belt drive or direct connection. Impeller between bear-ings that are on internal supports. Drive through inlet.

Arr. 4 Arr. 4 two-stageFor direct connection. Impeller Impeller overhung on motorshaft. No bearings on fan. Motor on internal supports.

Arr. 7For belt drive or direct connection. Arr. 3 plus commonbase for prime mover.

Arr. 8 one-or two-stageFor belt drive or direct connection. Arr. 1 plus common basefor prime mover.

Arr. 9 motor on casing Arr. 9 motor on integral baseFor belt drive. Impeller overhung. Two bearings on internal supports. Motor on casing or on integral base. Drive through belt fairing.

Note: all fan orientations may be horizontal or vertical

Inlet box Diffuser

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Figure 17: Drive Arrangements for Centrifugal FansReprinted with permission from the Air Movement and Control Association from Publication 201–90.

SW- Single WidthSI- Single InletDW- Double WidthDI- Double Inlet

Arrangements 1,3,7and 8 are also availablewith bearings mountedon pedestals or baseset independant of thefan housing

Arr. 1 SWSI For belt drive or directconnection impeller overhung. Twobearings on base.

Arr. 3 SWSI For belt drive ordirect connection. One bearingon each side and supported byfan housing.

Arr. 3 DWDI For belt drive ordirect connection. One bearingon each side and supported byfan housing.

Arr. 4 SWSI For direct drive.Impeller overhung on primemover shaft. No bearings onfan. Prime mover base mountedor integrally directly connected.

Arr. 7 SWSI For belt drive ordirect connection. Arrangement3 plus base for prime mover.

Arr. 9 SWSI For belt drive. Impelleroverhung, two bearings, with primemover outside base.

Arr. 10 SWSI For belt drive.Impeller overhung, two bearings,with prime mover inside base.

Arr. 7 DWDI For belt drive ordirect connection. Arrangement3 plus base for prime mover.

Arr. 8 SWSI For belt drive ordirect connection. Arrangement1 plus extended base for primemover.

Arr. 2 SWSI For belt drive or direct con-nection. impeller overhung. Bearings inbracket supported by fan housing

Fans (00-06) 1/8/01 10:36 AM Page 29


Figure 18: Drive Arrangements for Centrifugal FansReprinted with permission from the Air Movement and Control Association from Publication 201–90.

SW- Single Width SI- Single Inlet DW- Double Width DI- Double Inlet

Arr. 1 SWSI with Inlet Box For belt driveor direct connection. Impeller overhung,two bearings on base. Inlet box may beself-supporting

Arr. 3 SWSI with Independent PedestalFor belt drive or direct, connection fan.Housing is self-supporting. One bearing oneach side supported by independantpedestals.

Arr. 3 SWSI with Inlet Box andIndependent Pedestals For belt drive ordirect connection fan. Housing is self-support-ing. One bearing on each side supported byindependent pedestals with shaft extendingthrough inlet box.

Arr. 3 DWDI with Independent PedestalFor belt drive or direct connection fan.Housing is self-supporting. One bearing oneach side supported by independantpedestals.

Arr. 3 DWDI with Inlet Box and IndependentPedestals For belt drive or direct connectionfan. Housing is self-supporting. One bearing oneach side supported by independent pedestalswith shaft extending through inlet box.

Arr. 3 SWSI with Inlet Box For belt drive ordirect connection. Impeller overhung, two bear-ings on base plus exended base for primemover. Inlet box may be self-supporting.

Fans (00-06) 1/8/01 10:36 AM Page 30

• Fan rotation is determined to be clockwise or counterclockwiseby viewing the fan from the drive side.

• The choice of fan arrangement depends on the application – theenvironment of the airstream being handled and the size of thefans are primary considerations.

• The discharge position and the drive arrangement must bedetermined to fit the fan system properly.

• Manufacturers identify the arrangements available for the fansin their product line.

CLASS LIMITS FOR FANS

• AMCA has adopted a standard that defines the operating limitsfor various classes of centrifugal fans used in general ventilationapplications.

• The standard uses limits based on mean brake horsepower persquare foot of outlet area, expressed in terms of outlet velocity andstatic pressure.

• There are three class limits for centrifugal fans - Class I is thelightest duty and Class III is the heaviest duty.

• When selecting a fan, it is important to ensure the duty pointdoes not exceed the performance range for the fan class.


Fans (00-06) 1/8/01 10:36 AM Page 31

• All fans produce total pressure, which represents the static andkinetic energy imparted to the air by the impeller.

• The rotating blades of the fan impeller convert mechanicalenergy into static and kinetic energy by changing the velocityvector of the incoming air.

• Centrifugal fans produce total pressure from the centrifugalforce of the air radiating out between the blade passages and bythe kinetic energy imparted to the air by virtue of its velocityleaving the impeller.

• The absolute velocity vector in the case of centrifugal fans is thesum of the tangential and radial velocity components.

• Axial fans produce total pressure from the change in velocitypassing through the impeller, with none being produced bycentrifugal force.

• The absolute velocity vector in the case of axial fans is the sumof the axial and tangential velocity components.

Chapter 4: Principles of Operation 33

C H A P T E R 4

PRINCIPLES OF OPERATION

Fans (00-06) 1/8/01 10:36 AM Page 33

CENTRIFUGAL FANS

• The operation of centrifugal fans can best be described byvelocity vector diagrams.

• The height of the diagram – indicated by the relative radialvelocity vector Vr – is based on the volume of air flowingthrough the fan.

• The air velocity relative to the blade – indicated by Vb is nearlytangential to the blade as some slip occurs due to boundarylayer effects.

• The tip speed component wr is perpendicular to the wheelradius, where w is the rotational speed of the impeller in radiansper second and r is the radius of the impeller at the blade tip.

• Because the speed of the wheel is the same for each case, thevector wr is constant.

• The absolute velocity indicated by Vs is the resultant of Vb andwr.

• The relative tangential velocity vector indicated by Vt isprojected from Vs in the direction of wr.

• If volume decreases, the vector Vr decreases and as the vector Vb

does not change for a given blade, Vt increases with BI blades,remains constant with R blades and decreases with FC blades.

• As the pressure of the fan depends on the product of Vt and wr,the pressure characteristic rises as volume decreases for the BIblade, is constant for the R blade and decreases for the FCblade.

These vector diagrams illustrate that, at a given speed, thesmallest fan selection will be a forward curved fan.Conversely, the largest will be an airfoil.


Fans (00-06) 1/8/01 10:36 AM Page 34

Figure 19: Outlet Velocity Vector Diagram for Backward-inclinedBlades

Figure 20: Outlet Velocity Vector Diagram for Radial Blades

Figure 21: Outlet Velocity Vector Diagram for Forward-curvedBlades


Fans (00-06) 1/8/01 10:36 AM Page 35

AXIAL FANS

• The principle of operation can be described by the use of avelocity vector diagram.

• Velocity diagrams for axial fans are drawn for a uniform axialvelocity indicated by Va. The axial velocity remains nearlyconstant from blade root to tip.

• The tip speed component wr is perpendicular to the axis and isshown as the blade section under consideration.

Figure 22: Velocity Vector Diagram for an Axial Fan without InletGuide or Diffusion Vanes near the Impeller Hub

Figure 23: Velocity Vector Diagram for an Axial Fan without InletGuide or Diffusion Vanes at the Blade Tip


Fans (00-06) 1/8/01 10:36 AM Page 36

• The air velocity relative to the blade indicated by Vb is nearlytangential to the blade as some slip occurs due to boundarylayer effects.

• The relative tangential velocity vector indicated by Vt isprojected from Va in the opposite direction of wr.

• The mean relative velocity drawn to bisect Vt is shown as Vm.

This is used in aerodynamic theory to calculate the circulationaround the airfoil.


Fans (00-06) 1/8/01 10:36 AM Page 37

• The manufacturer guarantees fan performance according tostandard air conditions. When selecting a fan, it is necessary toknow the actual air inlet conditions (temperature, pressure,density), and use the Fan Laws to correct the publishedperformance to actual conditions.

• Fan performance curves are developed from data obtained fromtests executed in accordance with AMCA and ASHRAEstandards.

• The most common procedure to develop a performance curveis to test the fan from shut-off conditions to nearly-free deliveryconditions.

• A fan is generally tested in a set-up that closely simulates howit will be installed in an air-moving system.

• Propeller fans are normally tested in the wall of a chamber, andpower roof exhausters are tested mounted on a curb to exhaustvertically from a chamber.

• Centrifugal, tubeaxial, and vaneaxial fans are usually testedwith an outlet duct with provision for restricting the flow at thedischarge.

Chapter 5: Fan Performance Curves 39

C H A P T E R 5

FAN PERFORMANCE CURVES

Fans (00-06) 1/8/01 10:36 AM Page 39

• A static- and velocity-pressure measuring station is locatedwithin the duct downstream of flow straighteners.

• At shut-off the duct is completely blanked off, and at free-delivery the duct outlet is wide open; test data is recorded whilemaintaining constant fan speed and air density.

• Under these conditions and at the same fan speed, the flow isgraduated to obtain sufficient data to define a correspondingperformance curve.

• For each test point, the pressures are measured and thecorresponding flow rate is determined. The measured pressuresare corrected back to fan inlet conditions.

• Fan performance curves are plotted with the inlet flow rate (incubic foot per minute or litres per second) on the abscissa. Totalpressure, static pressure, fan horsepower and fan efficiency areplotted on the ordinate axis.

• It is not practical to test a fan at every speed at which it canoperate or at every inlet density it may encounter.

• By using a series of equations referred to as the Fan Laws, it ispossible to accurately predict the fan's performance at otherspeeds and densities.

• Manufacturers usually publish fan performance curves at adensity of 0.075 Ib/ft3 and an inlet temperature of 70˚F.


Fans (00-06) 1/8/01 10:36 AM Page 40

Figure 24: Fan Test-rig SetupReprinted with permission from the Air Movement and Control Association from Publication 201–90.

Chapter 5: Fan Performance Curves 41

Fans (00-06) 1/8/01 10:36 AM Page 41

• The Fan Laws relate the performance variables for anydynamically similar series of fans at the same point of rating onthe performance curve.

• The variables are fan size, D; rotational speed, N; gas density, p;volume flow rate, Q; pressure, p;total efficiency Ntj and power(shaft), P.

• Fan Law No. 1 governs the effect of changing size, speed ordensity on volume flow, pressure and power level.

• Fan Law No. 2 governs the effect of changing size, pressure ordensity on volume flow rate, speed and power.

• Fan Law No. 3 governs the effect of changing size, volume flowor density on speed, pressure and power.

LIMITATIONS

• The Fan Laws may be applied to a particular fan to determinethe effect of speed change. However, caution should beexercised since the Laws apply only when all flow conditionsare similar.

Chapter 6: Fan Laws 43

C H A P T E R 6

FAN LAWS

Fans (00-06) 1/8/01 10:36 AM Page 43

• These Fan Laws do not include correction for compressible flow.

COMPRESSIBILITY FACTOR

• As air travels through a fan, it is compressed and the outletvolume will be less than at the inlet. The fan laws as presentedin this chapter do not account for this effect.

• A fan selected without using compressibility will be larger insize than required and the fan input power will be understated.

• The compressibility effect is quite small when fan pressure riseis below 10” Wg., and is customarily ignored below thisthreshold.

• For applications where the fan pressure rise is more than 10îWg., the chart on the following page may be used as follows:

1. Estimate the total efficiency of the fan that will be selected.

2. Obtain the compressibility factor, Kp from the chart for therequired fan static pressure rise.

3. For fan selection only, multiply the required pressure andflow by the compressibility factor, Kp. The fan input powerobtained using the fan laws for selection must be divided byKp.

4. If the actual efficiency is more than 5% different than whatwas estimated in step #1, return to step #1 using the newefficiency.

5. When using equations 7 and 9 in chapter 7, multiply theresulting power by Kp.


Fans (00-06) 1/8/01 10:36 AM Page 44

Figure 25: Compressibility Factor

Fans Reference Guide 45

1.000

.990

.980

.970

.960

.950

.9400 10 20 30 40 50 60 70

Fan Static Pressure Rise - inches W.G.

Approximate Kp 1%

Kp

0.50

0.55

0.600.650.70

Fans (00-06) 1/8/01 10:36 AM Page 45

Table 1: Fan Laws

Source: ASHRAE Handbook, 1988 Equipment Volume.


Law No. Formulae

1a Q1 = Q2 x (D1/D2)3 x (N1/N2)

1b p1 = p2 x (D1/D2)2 x (N1/N2)2 x r1/r2

1c P1 = P2 x (D1/D2)5 x (N1/N2)3 x r1/r2

2a Q1 = Q2 x (D1/D2)2 x (p1/p2)1/2 x (r2/r1)1/2

2b N1 = N2 x (D2/D1) x (p1/p2)1/2 x (r2/r1)1/2

2c P1 = P2 x (D1/D2)2 x (p1/p2)3/2 x (r2/r1)1/2

3a N1 = N2 x (D2/D1)3 x (Q1/Q2)

3b p1 = p2 x (D2/D1)4 x (Q1/Q2)2 x r1/r2

3c P1 = P2 x (D2/D1)4 x (Q1/Q2)3 x r1/r2

4 P = Qp / (6362 ht)

SELD

OM

USE

D

Fans (00-06) 1/8/01 10:36 AM Page 46

• The following fan formulae require certain constants andparameters specific to each application. The constants are givenin this guide or in the referenced material.

• The formulae in this section are valid for incompressible flow

DENSITY

• Fan performance data, unless otherwise identified, is based ondry air at the standard atmospheric pressure of 14.7 psi., 29.921in.Hg (101.325 kPa) and a temperature of 68˚F (20˚C). The airdensity at standard AMCA test conditions is 0.075 lbm/ft.3

(or 1.2 kg/m3 when SI units are used).

• In most applications, fans process moist air at temperatures andpressures other than standard conditions. Therefore, the airdensity must be corrected to obtain the actual fan performance.

• For fans processing moist air, the moisture content of anairstream is determined by measuring the wet-bulbtemperature, the dew-point temperature, or relative humidity.

Chapter 7: Fan Formulae 47

C H A P T E R 7

FAN FORMULAE

Fans (07-14) 1/8/01 10:50 AM Page 47

• Wet-bulb and dry-bulb temperatures are most often determinedat fan inlet conditions, using a sling thermometer.When theairstream exceeds 180˚F (82˚C), the dew-point temperature ismore reliable to determine moisture content.

Density, when the dry-bulb temperature falls between 42˚F andl00˚F (5˚C and 38˚C), may be determined by using thepsychrometric density chart in AMCA Publication 203-90, FieldPerformance Measurement of Fan Systems, Appendix N. Thenumerical method in Chapter 16, Appendix A of this guide maybe used with a scientific calculator or PC spread-sheet program.

• EQUATION 1: When the gas density (P) at one plane isdetermined, the density at any point in the fan system may bedetermined.

“Boyle’s Law”: P = constant or px = p1

Tr Tx rx T1 r1

whererx = density at plane x, lb./ft.3 (kg/m3)r1 = density at plane 1, lb./ft.3 (kg/m3)Tx = absolute temperature at plane x, ˚Fabs (K)TI = absolute temperature at plane 1, ˚Fabs (K)px = absolute pressure at plane x, in.Hg (kPa)p1 = absolute pressure at plane l, in.Hg (kPa)

FAN FLOW RATE

• EQUATION 2: The flow rate at a reference plane. The fan flowrate is the primary performance parameter.

Q = VAwhere

Q = flow rate, ft.3/min. (L/s)V = average velocity at reference plane, ft./min.(m/s)A = area of reference plane, ft.2 (m2)


Fans (07-14) 1/8/01 10:50 AM Page 48

HEAD AND PRESSURE

• The common unit is inches of water, “water gauge”.

• Head is the height of a fluid column of water supported by gasflow, while pressure is the normal force per unit area. With gasor air, it is convention to measure pressure on a column ofliquid, as pressure measured in terms of unit area is notpractical.

• The term (V2/2g) refers to velocity head, and the term (rV2/2gc)refers to velocity pressure.

• Velocity head is independent of fluid density.

VELOCITY PRESSURE

• EQUATION 3: Velocity pressure is not independent of density.pv = r(V/cf)2

wherepv = velocity pressure, in.Wg (Pa)r = density, lbm/ft.3 (kg/m3)V = mean fluid velocity, ft./min. (m/s)cf = conversion factor, 1097 ( 1.414)

• EQUATION 4: For a standard air density of 0.075 lbm/ft.3

(1.20 kg/m3)1 Equation 3 becomes the following:Pv = (V/cf)2

whereV = mean fluid velocity, ft./min. (m/s) cf = conversion factor, 4005 ( 1.29)

1 The SI standard density of 1.20 kg/m3 is not an exact equivalent of the imperial standarddensity. Source: Jorgensen, R. (ed.) Fan Engineering 8th ed. Buffalo: NY, Buffalo ForgeCompany, 1983. (Ref. A). The SI density derived directly from the imperial equivalentwould be a value of 1.2014 kg/m3. Source: Metric Conversion Handbook for MechanicalEngineers in the Building Industry 2nd ed. Public Works Canada, 1983. (Ref. B)


Fans (07-14) 1/8/01 10:50 AM Page 49

TOTAL PRESSURE

• EQUATION 5: The sum of the static pressure and the velocitypressure is total pressure.

Pt = Ps + Pv

wherept = total pressure, in.Wg (Pa) ps = static pressure, in.Wg (Pa)pv = velocity pressure, in.Wg (Pa)

FAN-SYSTEM-EFFECT FACTOR (AT INLET)

• EQUATlON 6: The fan-system-effect pressure drop.SEF = Co r (Vo /cf)2

whereSEF = fan-system-effect pressure loss, in.Wg (Pa)Co = fan-system-effect loss coefficient, dimensionlessr = density, lbm/ft.3 (kg/m3)cf = conversion factor, 1097 (1.414)

and wherefor centrifugal fans:Vo = inlet velocity based on area at the inlet collar, or

outlet velocity based on outlet area, fpm (m/s)

for axial fans:Vo = inlet or outlet velocity based on area calculated

from fan diameter, fpm (m/s)

FAN POWER AND EFFICIENCY

• EQUATION 7: The air horsepower, or fan output power, PFo, isdetermined from product of the flow and total pressure rise.

PFo = Qpt

cfwhere:

PFo = output power, hp (W)Q = flow, ft.3/min. (L/s)


Fans (07-14) 1/8/01 10:50 AM Page 50

pt = total pressure rise, in.Wg (Pa)cf = conversion factor, 6349.6 (1,000)2

• The fan input power, PFi, is the measured power delivered tothe fan shaft.

2 The conversion factor 6349.6 was derived from converting the metric form of the equationto the imperial equivalent utilizing the conversion factor in Ref. B (see p.47). When theformula is applied directly to imperial units, the conversion factor to use is 6354 from Ref. A (see p.47). In strict SI terms, flow would be in m3/s and the conversion factor wouldbe 1.0.

• EQUATION 8: The fan mechanical or total efficiency is theratio of the output power to the input power.

ht = PFo/PFi

whereht = mechanical (total) efficiency, dimensionlessPFo = output power, hp (W)PFi = input power, hp (W)

• EQUATION 9: The fan input power.

PFi = Q pt

ht cfwhere

PFi = input power, hp (W)Q = flow, ft3/min. (L/s)pt = total pressure rise, in.Wg (pa)ht = total efficiency, dimensionless ratiocf = conversion factor, 6349.6 ( 1,000)

• EQUATION 10: The fan static efficiency is the product ofmechanical efficiency and the ratio of static pressure to totalefficiency.

hs = (ps/pt) ht

wherehs = static effciency, dimensionless ratio


Fans (07-14) 1/8/01 10:50 AM Page 51

ht = total efficiency, dimensionless ratiops = static pressure, in.Wg (pa)pt = total pressure, in. Wg (pa)

• EQUATION 11: By substitution, the fan input power is also:

PFi = Q pshs cf

wherePFi = input power, hp (W)Q = flow, ft.3/min. (L/s)hs = static efficiency, dimensionless ratiops = static pressure, in.Wg (Pa)cf = conversion factor, 6349.6 ( 1,000)

FAN MOTOR POWER

• EQUATION 12 AND EQUATION 13: The fan motor outputpower.

PMo = Q pt

ht hD cf

orPMo = PFi

hD

wherePFi = input power, hp (W)PMo = motor output power, hp (W)Q = flow, ft.3/min. (L/s)ht = fan total efficiency, dimensionless ratiohD = fan drive efficiency, dimensionless ratiocf = conversion factor, 6349.6 (1,000)


Fans (07-14) 1/8/01 10:50 AM Page 52

• EQUATION 14: The fan motor input power.

PMi = PMo

hM cf

wherePMi = motor input power, kWPMo = motor output power, hp (W)hM = motor efficiency, dimensionless ratiocf = conversion factor, 1.3410 ( 1,000)

EXAMPLE 1

• A large cafeteria at a manufacturing plant requires an exhaustfan to ensure proper indoor air quality for the patrons. Positiveexhaust is provided by a belt-driven, 24-in., airfoil, single-width(AFSW) centrifugal fan with an inlet and outlet area of 4.11sq.ft. The fan is to be equipped with a two-speed 1,800/1,200-rpm motor and, on the high-speed setting, is required to deliver10,000 cfm of air at 70˚F at a static pressure of 2.5 in.Wg. This isidentified as operating point A (see Figure 25). Standardatmospheric conditions are assumed for this example.

Figure 26. Operating Point and System Curve


Fans (07-14) 1/8/01 10:50 AM Page 53

Example 1.1

• Using formulae and manufacturers' catalogue data, determinespeed, the high-speed fan horsepower, and the correspondingfan total and static efficiencies. Assuming drive losses of 5.6%,calculate the motor size required.

Table 2. Typical ManufacturersÕ Performance Data for a 24-in.,AFSW Centrifugal Fan at 70ûF and Standard Atmospheric

Pressure

• The velocity pressure of the fan must be established todetermine the fan static pressure and the total efficiency of the fan.

• EQUATION 15: Outlet velocity (rearrange Equation 2, p. 46).

V = Q/A= 10,000/4.11= 2,433 fpm


Fan Static Pressure

1Ó 1 - 1/4Ó 2 - 1/2Õ 3Ó 3 - 1/2Ó 4 - 1/2Ó

rpm bhp rpm bhp rpm bhp rpm bhp rpm bhp rpm bhp

6600 1605 963 1.8 1003 2.1 1191 3.5 1260 4.2 1329 4.2 1465 6.3

7000 1703 1000 2.0 1040 2.3 1222 3.8 1288 4.4 1353 5.1 1482 6.6

9000 2189 1188 3.2 1223 3.6 1387 5.4 1446 6.2 1503 7.0 1610 8.6

9400 2287 1226 3.5 1260 3.9 1420 5.8 1479 6.6 1534 7.5 1640 9.1

9800 2384 1267 3.8 1301 4.2 1456 62 1512 7.1 1567 8.0 1671 9.6

10200 2481 1310 4.2 1339 4.6 1494 6.7 1546 7.5 1601 8.4 1703 10.2

cfm

, std

air

outle

tve

locit

y

Fans (07-14) 1/8/01 10:50 AM Page 54

• Velocity pressure at the inlet & outlet (Equation 4, p. 47).pv = (V/4,005)2

= 0.369 in.Wg

• Fan static pressure is ps = (p2 - p1)- pv1

= 2.5” - .369” = 2.131”

• By interpolation, the fan power is 5.83 hp and from cataloguedata the corresponding impeller speed is 1,429 rpm.

• Total pressure (Equation 5, p. 47)pt = ps + pv

= 2.131” + .369”= 2.5” in.Wg

• EQUATION 16: Fan total efficiency (rearrange Equation 9, p. 49)

ht = (Q pt)/(cf PFi)= (10,000 x 2.5)/(6,349.6 x 5.83) x 100= 67.53%

• Fan static efficiency (Equation 10, p. 49).hs = (2.131/2.5) x 67.53

= 57.56%

• Motor output power including drive losses (Equation 13, p. 50).

PMo = PFi/hD

= 5.83/(1 - 0.056)= 6.176 hp

• Therefore a 7.5-hp motor is required.


Fans (07-14) 1/8/01 10:50 AM Page 55

Figure 27: Fan Static-pressure Design Curve at 1,475 rpmIntersecting Design Point A and Fan Static-pressure Curve at

983 rpm Intersecting Point BExample 1.2

• Using fan laws and assuming constant density, determine thelow-speed flow and total pressure to give operating point B (seeFigure 26). Using formulae and the manufacturers’ data in Table2, p. 52, determine the corresponding fan power speed, andtotal and static efficiencies.

• EQUATION 17: Air volume delivered by the fan on the lowspeed setting (rearrange Fan Law No. la).

Q2 = Q1 (N2/N2)=10,000 x (1,200/1,800)= 6,666 cfm

• EQUATION 18: The corresponding impeller speed.

N2 = N1 (rpm2/rpml)= 1,429 x 1,200/1,800= 953 rpm


Fans (07-14) 1/8/01 10:50 AM Page 56

• EQUATION 19: The corresponding static pressure delivered bythe fan (rearrange Fan Law No. 3b).

ps2 = psl (Q2/Q1)2

= 2.131 x (6,666/10,000)2

= 0.947 in.Wg

• By interpolation, the fan power at point B is 1.73 hp from Table2, and the corresponding impeller speed is 953 rpm.Alternately, using the fan laws,

Pfi = 5.83hp x (1,200/1,800)3

= 1.73hp

• Outlet velocity (Equation 14, p. 50).

V = Q/A= 6,666/4.11= 1,622 fpm

• Velocity pressure (Equation 4, p. 47).

pv = (1,622/4,005)2

= 0.164 in.Wg

• Total pressure (Equation 5, p. 47).

pt = 0.947 + 0.16= 1.11 in.Wg

• Fan total efficiency (Equation 12, p. 50).

ht = (Q pt)/cf Pfi)= (6,666 x 1.11)/(6,349.6 x 1.73) x 100= 67.35%

• Fan static efficiency (Equation 10, p. 49).

hs = (0.947/1.11) x 67.35= 57.46%


Fans (07-14) 1/8/01 10:50 AM Page 57

• A fan provides the energy to overcome the resistance to flowthrough air-system components. A fan's performance isinterdependent with the system elements.

• Components that contribute to system resistance includestraight ductwork, elbows, fittings, filters, humidifierdistributors, heat-transfer coils, dampers, acoustic silencers, birdscreens, registers, grilles and diffusers.

• Most air systems operate in the turbulent-flow regime ratherthan laminar-flow conditions.

• Pressure losses in system elements are therefore mainly relatedto turbulence and flow separation, the kinetic energy beingdissipated by viscous shear in the air.

• The pressure loss of each of the air system's elements may becalculated with manufacturers' data and the procedures in theASHRAE Handbook, 1989 Fundamentals Volume.

• A given rate of airflow through a system requires a specific totalpressure generated by the system fan.

Chapter 8: Air Systems 59

C H A P T E R 8

AIR SYSTEMS

Fans (07-14) 1/8/01 10:50 AM Page 59

• EQUATION 20: If the flow rate is changed, the resulting totalpressure required for turbulent-flow systems.

(Æp2/Æp1) = (Q2/Q1)2

• Figure 25 (p.51) shows the characteristic system curve plottedin a parabolic fashion according to the relationship establishedin Equation 20.

• Example 2 shows the effect of the relationship in Equation 20.

EXAMPLE 2

• The exhaust system in Example 1 has a filter bank to protect theheat-recovery coil. The duct-system and coil static losses are 2.0 in.Wg and clean-filter losses are 0.5 in.Wg at the designflow rate of 10,000 cfm.

• The required static pressure of the fan is 2.5 in.Wg at the designflow rate. This identifies the design operating point A fromwhich the design curve A is plotted. It is assumed that the fan isplenum mounted and hence the inlet velocity pressure is ~ 0” Wg, and fan static pressure = pressure rise of 2.5” Wg.

• When the fan is set at low speed, from the Fan Laws shown inExample 1, the flow rate is 6,666 cfm.

• This second design point is point B on curve in Figure 27.

• The filter specifies a maximum dirty pressure loss of 1.5 in.Wgat the design flow rate, which means in the dirty condition, thetotal system static losses are 3.5 in.Wg.

• This gives a new design point C, from which the dirtymaximum design operating system curve C is plotted (seeFigure 27).


Fans (07-14) 1/8/01 10:50 AM Page 60

Example 2.1

• Using formulae and manufacturers' catalogue data (table 2, p. 52), determine for operating point C, the speed, the high-speed fan power, and the corresponding fan total and staticefficiencies. Assuming the same drive losses of 5.6%, calculatethe motor service factor .

• By interpolation, the fan power at point C is 8.2 hp from Table2, and the corresponding impeller speed is 1,584 rpm.

• Total pressure required (Equation 5, p. 47).Pt = 3.5 + 0.37

= 3.87 in.Wg

• Fan total efficiency (Equation 16, p. 53),ht = (10,000 x 3.87)/(6,349.6 x 8.20) x 100

= 74.3%

• Fan static efficiency (Equation 10, p. 49).hs = (3.50/3.87) x 74.3

= 67.2%

• EQUATION 21: Motor service factor.

SF = PFi/(hD x PMo)= 8.20/(0.944 x 7.5)=1.16

• Fan power exceeds standard motor service factor of 1.15.

Example 2.2

• Using Fan Laws, determine the actual air volume, fan power,and fan total and static efficiency on the high-speed settingunder dirty filter conditions identified as point D on the systemcurve (see Figure 27).


Fans (07-14) 1/8/01 10:50 AM Page 61

• EQUATION 22: Actual air volume delivered by the fan at aconstant speed of 1,475 rpm (rearrange Fan Law la).

Q2 = Q1/N1/N2)= 10,000/(1,584/1,475)= 9,312 cfm

• EQUATION 23: Static pressure.

ps2 = ps1/Q1/Q2)2

= 3.5/ ( 10,000/9,312)2

= 3.03 in.Wg

• By interpolation, the fan power at point D is 6.56 hp fromcatalogue data.

• Outlet velocity (Equation 15, p. 53).V = 9,312/4.11

= 2,266 fpm

• Velocity pressure (Equation 4, p.47).pv = ( 8,266/4,005)2

= 0.32 in H20

• Total pressure (Equation 5, p.47).pt = 3.03+0.32

= 3.35 in H20

• Fan total efficiency (Equation 16, p. 53).ht = ( 9,312 x 3.35)/( 6,349.6 x 6.56) x 100

= 74.9%

• Fan static efficiency (Equation 10, p.49)hs = ( 3.03/3.35) x 74.9

= 67.7%


Fans (07-14) 1/8/01 10:50 AM Page 62

Figure 28: Fan Static Pressure Curve Intersecting the DesignPoint A and the Maximum Design Point D


Fans (07-14) 1/8/01 10:50 AM Page 63

SYSTEM EFFECT FACTORS

• A fan is normally tested with open inlets and straight ductattached to the outlet. This results in uniform airflow into thefan and efficient static-pressure recovery at the fan outlet.

• If these conditions are not matched in the actual installation,the performance of the fan degrades. This must be allowed forwhen selecting the fan.

• Figure 28 illustrates deficient fan and system performance withthe calculated system-design pressure and flow shown as Point 1.

• Since no allowance was made for system effect, the actualoperating condition is Point 4 – at the intersection of the fanpressure-volume curve and the actual system curve.

• The difference between Point 1 and Point 4 projected on theabscissa is the deficiency in flow.

• To compensate for the deficiency, a system effect factor equalto the pressure difference between Points 1 and 2 must beadded to the calculated system-pressure losses, with the fanselected to operate at Point 2.

Chapter 9: Fan and System Interface 65

C H A P T E R 9

FAN AND SYSTEM INTERFACE

Fans (07-14) 1/8/01 10:50 AM Page 65

• The fan system-effect factor is the product of Co times thevelocity pressure and is calculated using Equation 6, p. 48.

• For centrifugal fans, velocity is based on the area of the inletcollar and the outlet area; for axial fans, it is based on the fandiameter.

• Appendix C explains how to determine Co (the system effectfactors For SWSI centrifugal fans). Other system-effect factorsare beyond the scope of this guide, but are covered in AMCAPublication 201-90.

Figure 29: Deficient Fan and System Performance

FAN OUTLET CONDITIONS

• System-effect factors must be calculated and added to thesystem resistance losses whenever 100% recovery at the outletof a fan cannot be achieved.


Fans (07-14) 1/8/01 10:50 AM Page 66

• Complete recovery can be achieved if the outlet effective ductlength is 2.5 diameters or more for a velocity of 2,500 fpm (13 m/s) or less. Add one duct diameter for each additional1,000 fpm (5 m/s).

Figure 30. Fan-outlet Velocity ProfilesReprinted with permission from the Air Movement and Control Association from

Publication 201-90

• In some cases, fan outlets are connected directly to a larger ductor plenum without a transition. This causes a pressure loss ofup to one velocity head (V2/2g), based on the highest velocityof the fan outlet.

Chapter 9: Fan and System Interface 67

Fans (07-14) 1/8/01 10:50 AM Page 67

• The highest velocity is in the blast area of a centrifugal fan (thearea between the cutoff and the scroll), and in the swept area ofan axial fan.

FAN INLET CONDITIONS

• To achieve rated fan performance, air must enter a fanuniformly over the inlet area in an axial direction withoutprerotation.

• The ideal inlet condition allows air to enter axially anduniformly without spin.

• A spin, in the same direction as the impeller rotation, reducesfan flow and pressure, whereas a counter-rotating vortex at theinlet slightly increases fan flow and pressure and substantiallyincreases fan power.

• The most common cause of reduced fan performance isnonuniform flow into the inlet of a fan.

• Turbulence and uneven flow into the fan impeller is typicallycaused by an elbow at the inlet of a fan.

• System-effect Co factors for inlet conditions that cause spin arenot available because of the multitude of variations.

• The system-effect factor can be eliminated by including anappropriate length of straight duct between the elbow and thefan inlet.

• For fans installed in cabinets or adjacent to walls, adequatedistance must be maintained in front of the inlet to allow forunobstructed flow.


Fans (07-14) 1/8/01 10:50 AM Page 68

• Fan selection involves consideration of the following:• volume flow rate and variation• fan total or static pressure and system effects• air density• air temperature• environment (corrosion, erosion, flammability)• permissible noise levels• attitude of fan and space available• type of fan required• type of drive and accessories• speed capability of motor driver.

• The fan can be selected once the system-pressure-loss curve isknown.

• The system-pressure-loss curve is defined by accounting for thesystem flow, resistances, and system-effect factors (according toASHRAE methods).

• The fan size, speed power, and noise spectrum is determinedusing one of the many methods available from fanmanufacturers, e.g., tables or PC-based computer programs.

Chapter 10: Fan Selection 69

C H A P T E R 1 0

FAN SELECTION

Fans (07-14) 1/8/01 10:50 AM Page 69

• Computer selection programs allow a designer to evaluate thevarious fan options for optimizing system efficiency andperformance, obtain noise-level data, and plot the fanperformance curve quickly.

• The performance data in fan tables is based on arbitraryincrements of flow rate and static pressure, and shading may beused to indicate an AMCA fan-class demarcation.

• Fan Laws cannot be used to obtain adjacent data points becauseeach data point represents a different point of operation on thefan performance curve.

• Intermediate points of operation can be determined byinterpolation, as the listed data points are close enough forreasonable accuracy for fan selection.

• Using a performance curve in conjunction with a computerprogram or tables is very important, particularly in VAVsystems which have more than one point of operation.

• Using the performance curve optimizes fan selection to avoidoperation close to, or in, the stall region and maximizeefficiency throughout the operating range.

PRESSURE DEFINITIONS

• When using performance tables or charts, it is important tounderstand what definition of pressure has been used by thefan manufacturer. There are three possible ways to state thefan's pressure requirement: Fan Total Pressure, Fan StaticPressure and Fan Static Pressure Rise.


Fans (07-14) 1/8/01 10:50 AM Page 70

• These definitions arise from Bernoullí’s equation which is usedto calculate system friction losses. The relationship betweenany two points (1 and 2) in a system is given below:

The Friction Losses in this equation must be overcome by the fan.

• Fan Total Pressure is defined as

• Fan Static Pressure is defined as

• Fan Static Pressure Rise is defined as

Dps = Ps2 - Ps1

• The inlet and outlet velocities (V1 and V2) in these equations aretaken to be at the terminals of the fan manufacture’s supply,which may include silencers, inlet boxes, outlet diffusers, etc.The velocities may not be identical to those in the adjacentducts.

The most common definition in North America is Fan StaticPressure for centrifugal fan and Fan Total Pressure for axial flowfans. The Europeans use Fan Total Pressure almost exclusively forall fans.

Chapter 10: Fan Selection 71

Ps1 + r1 = Ps2 + r2 + Friction LossV1

2

1096

V2 2

1096

Fanpt = Ps2 + r2 - Ps1 + r1

V2 2

1096

V1 2

1096

Fanps = Ps2 - Ps1 + r1

V1 2

1096

Fans (07-14) 1/8/01 10:50 AM Page 71

PARALLEL FAN SELECTION

• Selecting parallel fans with a characteristic pressure reductionleft of the peak pressure point typical of FC fans requires carefulconsideration.

• When these fans are operated in parallel, a fluctuating loadcondition may result if one of the fans operates to the left of thepeak static point on its performance curve.

• Figure 31 shows the pressure flow curves of a single fan (curveA-A) and the same fan operating in parallel with an identical fan(curve B-B).

Figure 31: Design Operating Point Selection Range on a TypicalCentrifugal Fan Performance Curve


Fans (07-14) 1/8/01 10:50 AM Page 72

Figure 32: Pressure Flow Curves

• The figure-eight curve plots possible combinations of volumeflow at each pressure value for the individual fans.

• Points to the right of B-C are the result of the fan operating tothe right of its peak rating point; for all systems stable operationoccurs with a system resistance curve below C-C.

• For points of operation to the left of B-C it is possible to satisfysystem requirements with one fan operating at one point ofrating while the other fan operates at another.

• Figure 31 shows point BD1 – 4,700 cfm at a static pressure of2.1 in.Wg – can be satisfied with each fan operating at 2,350 cfm at 2.1 in static pressure.


Fans (07-14) 1/8/01 10:50 AM Page 73

• The system curve can also be satisfied at point BD2 by one fanoperating at 1,350 cfm at a static pressure of 1.9 in.Wg and thesecond fan operating at 3,150 cfm at the same static pressure.

• This is because the system curve D-D passes through thecombined performance curve at two points, but operation canbe unstable under such conditions.

• With fan selection at point BD2, one fan is under loaded whilethe other fan is heavily loaded, and surge can occur in thesystem.

SERIES VS PARALLEL OPERATION

• In any 2 stage arrangement, the same mass flow per unit timemust be handled by each stage (if there is no leakage). Thedensity of the gas passing through each stage will be different,so it follows that the volume handled by each stage will bedifferent.

• Figure 33 shows the static pressure and horsepower curves for asingle fan. Also shown, are the static pressure curves for two ofthese fans (a) connected in series, and for comparison (b)connected in parallel. Strictly speaking, the static pressure curvefor the two fans in series will be slightly higher than that shownsince there is only one velocity head to be deducted from thecombined total pressure from 2 stages in order to compute thecombined static pressure available. For simplicity of discussion,the static pressure is doubled for any given volume for twoidentical fans in series.

• In Figure 33, system B passes through P, the intersection of thecombined series and combined parallel curves, i.e. eithercombination will give the same volume on this system.However, in series, each fan will consume the horsepower atpoint D, whereas in parallel each fan will take the power shownat point E.


Fans (07-14) 1/8/01 10:50 AM Page 74

Figure 33: Series Fan Operation

• On any system to the left of point P. The two fans in seriescombination will always produce more volume than theparallel configuration. On any system to the right of point P.Two fans in parallel will always produce more volume thanthey will connected in series. Whenever a second fan is to beadded to one existing on a given system to increase flow, it isadvisable to plot pressure volume curves for both series andparallel connection if maximum possible flow is desired. Thepower absorbed by each fan should also be carefully noted.


Volume Flow

Stat

ic P

ress

ure

System ASystem B

System CD

P

Static Pressure�Two Fans in Series

Static�Pressure�One Fan

Horsepower�One Fan

Static Pressure Two�Fans in Parallel

Fans (07-14) 1/8/01 10:50 AM Page 75

• Fan noise is an important criteria for the proper selection of fantype and size for an application.

• The noise from a fan is predominantly from aerodynamicsources and includes factors such as lift, rotation, vortexshedding, and wake.

• The noise generated by a fan depends on the fan design, thevolume flow rate, total pressure, and efficiency. This noise isproportional to the product of the pressure squared and theflow.

• Low outlet velocity does not necessarily relate to lower soundpower, and fan selection should not be based solely on fan tipspeed.

• The only valid basis for comparison is the actual sound powerlevels generated by the different fan types when they areoperating at the required system flow and pressure.

• For constant-volume systems, the recommended practice for aselected fan type is that the fan size and speed be selected sooperation falls at or near the peak efficiency point of the fanperformance curve.

Chapter 11: Fan Noise 77

C H A P T E R 1 1

FAN NOISE

Fans (07-14) 1/8/01 10:50 AM Page 77

• A fan is normally quietest when selected within the mostefficient operating range, which is also advantageous for energyconservation.

• Fan selection criteria for VAV systems include two other factors:the efficiency and stability of the fan through the entire range ofmodulation, and the acoustic impact of the modulation system.

• Fan selection for VAV systems is a compromise between fansurge and fan inefficiency, and the narrower the range ofmodulation the more acceptable the compromise will be.

• Variable inlet vanes may generate significant low frequencynoise as the vanes modulate to the close position and requireadditional attenuation with a corresponding increase in systempressure drop. Maximum sound levels occur at approximately75% open VIVs.

• The other modulation systems – variable-speed motors anddrives, and variable-pitch fan blades – generate less noise as thefan modulates to the no-flow operating point.

• The sound power generation of a specific fan at its operatingpoint should be obtained from manufacturers' AMCA test dataor from manufacturers' computer fan-selection programs.

• The data is presented as sound power levels in eight octavebands and as weighted overall sound level.

• If test data is not available, the octave-band sound power levelscan be estimated using the following procedure.

FAN SOUND POWER

• Fans generate a tone at the blade passage frequency, and thenumber of decimals to be added is the blade frequencyincrement (BFI). The octave band to which the BFI is addeddepends on the type of fan and the impeller speed.


Fans (07-14) 1/8/01 10:50 AM Page 78

• EQUATION 24 : Blade frequency

Bf = N x no. of bladescf

whereBf = blade frequency, HzN = impeller speed, rpm (r/s)cf = conversion factor 60 ( 1 )

Table 3: Typical Number of Fan Blades

• Blade frequency can be estimated using data from Table 3.

• EQUATION 25 : Estimating sound power levels at actualoperating conditions.

Lw = Kw + 10 log Q/cf1 + 20 log p/cf2 + Cwhere

Lw = estimated sound power level (dB re 1 pW)Kw = specific sound power levelQ = flow rate, cfm (L/s)


Impeller No. ofFan Type Size/Drive Blades

CENTRIFUGAL

Airfoil and Backward-inclined 24 in. and over 10Under 24 in. 12

Forward-curved 52

Radial 6

AXIAL

Vaneaxial 12

Tubeaxial Belt drive 6Direct drive 4

Propeller 6

Fans (07-14) 1/8/01 10:50 AM Page 79

p = fan pressure rise, in.Wg (Pa)C = correction factor for point of operation, dBcf1 = conversion factor, 1 (0.472)cf2 = conversion factor, 1 (249)

• Estimated sound power level is calculated for all seven bandswith KW selected from Table 4. The BFI is added to the octaveband in which the blade passage frequency falls. Soundcorrection factor is selected from Table 5.

Table 4: Specific Sound Power Levels and Blade FrequencyIncrements


Sound Power Level, KW (dB re 1 pW)

Octave-band Centre Frequency, Hz

Fan Type Impeller Size 63 125 250 500 1,000 2,000 4,000 BFI

CENTRIFUGAL

Airfoil and 36 in. and over 32 32 31 29 28 23 15 3backward-inclined Under 36 in. 36 38 36 34 33 28 20

Forward-inclined All 47 43 39 33 28 25 23 2

Radial blade and 40 in. and over 45 39 42 39 37 32 30 8Pressure blower 20 in. to 40 in. 55 48 48 45 45 40 38

Under 20 in. 63 57 58 50 44 39 39

Axial

Vanaxial 40 in. and over 39 36 38 39 37 34 32Under 40 in. 37 39 43 43 43 41 28 6

Tubeaxial 40 in. and over 41 39 43 41 39 37 34Under 40 in. 40 41 47 46 44 43 37 5

PropellerCooling tower All 48 51 58 56 55 52 46 5

Note: These values are the specific power levels radiated from either the inlet or the outlet ofthe fan. If the total sound power level being radiated is desired, add 3 db to each of theabove values.

Fans (07-14) 1/8/01 10:50 AM Page 80

Table 5: Sound Correction Factors

EXAMPLE 3

• The fan in Example 1, p. 51, is operating at its design conditionof 10,000 cfm, at static pressure of 2.5 in.Wg, static efficiency of60.9% and impeller speed of 1,475 rpm.

• Determine the sound power level in seven octave bands byassuming the number of impeller blades to be 10 and byestimating the off-peak, static-efficiency correction factor.

• A simple method to calculate the off-peak, static-efficiencycorrection factor with reasonable accuracy, is to determine thestatic efficiency of the fan operating at the same impeller speed,but at a static pressure and flow of about 55% WOcfm.

• From manufacturers' data, the operating point close to the55%-WOcfm line is 7,000 cfm at 4.5 in.Wg static pressure and6.6-bhp power.


Correction Factor, C, for Off-peak Operation

Static Efficiency % of peak Correction Factor dB

90 to 100 085 to 89 375 to 84 665 to 74 955 to 64 1250 to 54 15

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• EQUATION 26 : Peak static efficiency given by rearrangingEquation 11, p. 49.

h2 = QPs

cf PFi

= 7,000 x 4.56,349 x 6.6

= 0.752

• Static efficiency as a percentage of peak is as follows: 60.9/75.2 x 100 = 81%

• From Table 6, the off-peak correction factor C is 6 dB.

• The additional sound power levels due to the volume flow rateand pressure are given by Equation 25, p. 73.

Lw = Kw + 10 log 10,000 + 20 log 2.5 + C= Kw + 40 + 6= Kw + 54 dB

• The specific sound power levels (KW) for Table 6 are obtainedfrom the second line of Table 4.

• BFI is given by Equation 24, p. 73.

Bf =1,475 x 10/60= 246

• The closest octave band is 250 Hz.

• The magnitude of the off-peak, efficiency correction factorsuggests that a more efficient fan should be selected for thegiven duty.


Fans (07-14) 1/8/01 10:50 AM Page 82

• A larger fan would allow for operation closer to the surge line ata slower impeller speed, lower power and lower noise levels,but with less ability to cope with higher system static pressuresat the design flow.

Table 6. Summary for Example 3


Sound Power Level, KW (dB re 1 pW)

Octave-band Centre Frequency, Hz

63 125 250 500 1,000 2,000 4,000

KW 36 38 36 34 33 28 20

Equation 11 54 54 54 54 54 54 54

BFI 3

Total dB 90 93 93 89 86 82 74

Fans (07-14) 1/8/01 10:50 AM Page 83

• Fans are required to perform over a range of flows andpressures called, “The Duty Cycle”. How the fan is controlledto achieve the required range, can have a significant energycost.

• The type of fan control should be selected on the basis of cost,the precision of control required and the frequency andmagnitude of system flow changes.

• For HVAC systems with infrequent changes in flow rate and/orwhere control is a secondary consideration, a two- or three-speed motor is a low-cost solution.

• Where a system requires continuously varying flow rates overspeed ranges of 2.6:1 and input power of less than 15 kW, anadjustable-speed pulley drive is satisfactory and can improvesystem efficiency.

• Magnetic or hydraulic slip couplings can be used in systemswith power greater than 15 kW. However these couplings havean inherent power loss since the torque from the motor istransmitted unchanged to the impeller, in spite of the differencein rotational speed.

Chapter 12: Fan Duty Control 85

C H A P T E R 1 2

FAN DUTY CONTROL

Fans (07-14) 1/8/01 10:50 AM Page 85

• The most efficient method of speed control, with the potentialfor precise control, is the electronic adjustable speed drive (seefigure 32).

• Mechanical methods of volume control are often used incommercial and industrial HVAC systems.

• The simplest, most inefficient method of control, is a fandischarge damper; the damper artificially increases systemresistance and the fan works along its system curve.

Figure 34: Outlet Damper Fan Control

• Single fan performance curve

• Multiple system resistance curves


Volume Flow

Out

let P

ress

ure

unstable�region Higher system �

resistance�curve as outlet �damper closes

Fans (07-14) 1/8/01 10:50 AM Page 86

• To provide a broader range of flow control at lower energypenalties, the discharge damper can be used with multi-speedmotors, or with adjustable speed drives.

• Figure 33 shows discharge-damper control sequenced withmotor-speed control.

• There is no advantage other than initial cost to using an outletdamper. Its use should be avoided.

Figure 35: Throttle Control of a Fan witha Two-speed Motor

• Inlet-vane control can provide precise flow control, down toabout 40% of the full flow rate.

• This control device rotates the inlet airflow the same as theimpeller and so reduces the work done by the impeller.

Chapter 12: Fan Duty Control 87

Fans (07-14) 1/8/01 10:50 AM Page 87

• Another inlet-control device is a cone that varies the effectivefan-inlet area as a Function of the axial distance the cone ispositioned from the fan inlet. Used in HVAC only.

• Figure 31 shows the effect of inlet-vane control on BI,centrifugal-fan performance.

• The flow rate from variable pitch-in-motion vaneaxial fans canbe dynamically regulated by varying the attack angle of theimpeller blades; this maintains high efficiencies over a wideblade-pitch range.

Figure 36: Inlet Vane Control of a Fan


Fans (07-14) 1/8/01 10:50 AM Page 88

• All rotating machinery have critical speeds called resonantfrequencies where excessive vibration can cause damage. It isnecessary to have each fan and foundation installation checkedto avoid these speeds, or correct the fan balance.

• Most fans are shipped statically and dynamically balanced butcorrosion, erosion, dust and airborne contaminants collectingon the impeller may cause imbalance over time.

• Therefore, consider fan isolation when designing theinstallation.

• The transmission of vibration to a building structure involvesvibratory force, frequency of vibratory force (disturbingfrequency), natural frequency of isolator and floor, and stiffnessof isolator and floor.

• It is important to select vibration isolators to compensate forfloor deflection; and to avoid resonance, the natural frequencyof the isolator should be different to the disturbing frequency.

• The degree of fan isolation and balancing depends on the floorspan, and the fan type, size, speed and power.

Chapter 13: Vibration Isolation / Fan Balancing 89

C H A P T E R 1 3

VIBRATION ISOLATION/FAN BALANCING

Fans (07-14) 1/8/01 10:50 AM Page 89

• Use the "Vibration Isolator Selection Guide" in the ASHRAEHVAC Applications Volume Handbook to determine the base andisolator type and minimum deflection.

• Install flexible duct connectors at the inlet and at the dischargeto reduce transmission to the duct work.

• Consider resilient, structural steel or concrete, inertial bases forfans in critical areas or in long buildings of light construction.

Note: The best vibration isolation design will not compensate for aresonant frequency problem. Isolation will only reduce thetransmission of forces due to imbalance.


Fans (07-14) 1/8/01 10:50 AM Page 90

• The AC electric motor is the main type of prime mover used todrive fans and there are many types.

• The selection of a high-efficiency motor is important, but thestarting motor current and torque are more important.

• Excessive starting time raises the temperature of the motorwindings beyond acceptable levels.

FLYWHEEL EFFECT

• The time to accelerate a fan to operating speed depends on thefan/impeller inertia (flywheel effect) and the startingcharacteristics of the electric motor.

• The fan/impeller inertia is given as WR2 in the industry. Thismust be corrected to represent the apparent inertia as seen bythe motor when the fan operates at a different speed from themotor. R is called the radius of gyration.

• Additional resistance to starting will be air power consumed bythe fan. Therefore it is advisable to start centrifugal fans withdampers closed. (Axial fans should have dampers open.)

Chapter 14: Electric Motor Fan Drive 91

C H A P T E R 1 4

ELECTRIC MOTOR

FAN DRIVE

Fans (07-14) 1/8/01 10:50 AM Page 91

• The characteristic starting-torque curves of the fan motor andthe maximum allowable time for acceleration (usually about 10seconds) are available from manufacturers.

• The available acceleration torque is the difference between themotor torque and the fan torque.

AC MOTORS

• Polyphase (usually three-phase) AC motors are almost alwaysused in fan applications that require more than 2 hp.

• The AC induction motor, usually with a squirrel-cage rotor andno external connections, is the most suitable for three- phase-power fan drive as it is inexpensive and reliable.

• It is a constant-speed motor with a flat torque characteristic inrelation to motor speed.

• However, its starting current is high – as much as seven or eighttimes higher than the running current.

• The extra starting current for fans with low inertia and largemotors can cause problems with electrical supply and demand.Special consideration must be given to reducing the startingcurrent.

• It is usual to reduce the starting voltage at start-up and step upthe voltage until the fan reaches running speed.

• Another solution is to use a wound-rotor induction motor inwhich the polyphase windings of the motor are connected toan external resistor via slip rings.

• Starting torque and starting current can be controlled byadjusting the external resistance.

• Small fans requiring less than 2 hp usually use single-phasepower supply.


Fans (07-14) 1/8/01 10:50 AM Page 92

• Single-phase induction motors are non-self starting; to start themotor, a second starting phase is created by connecting an extrastator winding through a capacitor.

• The starting capacitor displaces the phase of the current and isdisconnected after running speed is obtained, and a differentsize capacitor is used for running.

• Shaded pole motors are generally unsuitable for fan drive,because of their inherent poor starting torque.

DC MOTORS

• Sometimes DC motors are preferable for fan drive, particularlyin applications requiring speed modulation.

• The series motor is most suitable, because it has moderatestarting current and self-regulating, stable operatingcharacteristics.

Chapter 14: Electric Motor Fan Drive 93

Fans (07-14) 1/8/01 10:50 AM Page 93

EXAMPLE: Building Ventilation

• Note: This procedure involves many calculations that are easilydone using a spreadsheet computer program.

• It is often necessary to estimate the energy consumption of afan, particularly for life-cycle costing. The energy costs areusually determined for a period of a year.

• EQUATION 27: Estimating the energy consumption of a faninvolves integrating the fan shaft input power divided bysystem efficiencies over time.

where:E = energy consumption, kWhPFi t(n) = fan shaft input power for time period, hp (W)hDt(n) = drive efficiency for time period,

dimensionless ratiohMt(n) = motor efficiency for time period,

dimensionless ratiohVt(n) = variable-speed drive efficiency for time period,

dimensionless ratioChapter 15: Energy Consumption Analysis 95

C H A P T E R 1 5

ENERGY CONSUMPTION

ANALYSIS

PFi t(n)

hDt(n) x hMt(n) x hVt(n) x cf

nE=∑

1x t(n)

Fans (15-21) 1/8/01 10:54 AM Page 95

cf = conversion factor, 1.3410 ( 1,000)t(n) = time at fan motor power, hours

• The fan shaft input power is calculated by Equation 11 (p. 49).Appendix B outlines the procedures for estimating driveefficiencies. Motor efficiencies are obtained frommanufacturers' data. The procedure for using Equation 27depends on the fan being analyzed.

CONSTANT-VOLUME FANS

• For a constant-volume fan processing a gas at a constanttemperature rise, the procedure is straight forward as the energyconsumption is integrated over one time period.

• For a constant-volume fan processing outdoor-air, the fanpower varies as a function of the outdoor-air temperature.

• When energy use is a function of outdoor-air temperature, theannual energy consumption can be determined using acomputerized, hourly analysis program.

• In the absence of such a program, a reasonable assessment canbe made by using ASHRAE bin weather data. When time-of-use energy rates apply, then the weather data can be separatedinto on-peak and off-peak periods.

• EQUATION 28: Total energy consumption:

where:E = energy consumption, kWhPFi b(n) = fan shaft input power at bin temperature, kWhDb (n) = drive efficiency at bin temperature,

dimensionless ratio


PFi b(n) X t b(n)

hDb (n) x hMb (n) x cf

nE=∑

1

Fans (15-21) 1/8/01 10:54 AM Page 96

hMb (n) = motor efficiency at bin temperature,dimensionless ratio

t b(n) = time at temperature bin, hourscf = conversion factor, 1.3410 ( 1,000)

• EQUATION 29: The fan shaft input power at each temperaturebin when a constant-volume fan processes air of varyingtemperature:

PFi b(n) = PFi s x rbrs

where:PFi b(n) = fan-shaft input power at bin

temperature, hp (W)PFi s = fan-shaft input power at standard

conditions, hp (W)rb = actual bin density, lbm/ft.3 (kg/m3)rs = standard air density, lbm/ft.3 (kg/m3)

VARIABLE-VOLUME FANS

• For an AF or BI centrifugal fan with variable-inlet vanes (VIVs),the load profile of the fan system and the design point of thefan (with the VIVs completely open) must be determined orestimated.

• From the full-load design point, the fan flow as a percentage ofwide-open cubic feet per minute (WOcfm) must be determinedfrom the fan curve or a computer program. This is necessary todetermine which constants are used in the following equations,which estimate the part-load conditions when the system curveoriginates at the apex of zero flow and zero static pressure.

Chapter 15: Energy Consumption Analysis 97

Fans (15-21) 1/8/01 10:54 AM Page 97

• EQUATION 30: Load profile expressed in terms of WOcfm.

%Q(n) = %Qd x Lf(n)

100where:

%Q(n) = percentage fan load for load point, %WOcfm%Qd = percentage design fan load, %WOcfmLf(n) = load factor for load point, %

• EQUATION 31: Percentage design fan load.

%Qd = Qd x 100QWOcfm

where%Qd = percentage design fan load, %WOcfmQd = flow at design, cfm (L/s)QWOcfm = flow at WOcfm, cfm (L/s)

• EQUATION 32: Percentage fan-shaft input power at each loadpoint, including the design point.

%hp(n) = c + a x exp (%Q(n) x b)

where:%hp(n) = percentage fan power for load point, %hp%Q(n) = percentage fan load for load point, %WOcfma,b,c = constants determined from table closest to

design WOcfm

• EQUATION 33: The fan-shaft input power in horsepower atthe design point is determined from manufacturers' data andcorrected for VIV losses.

PFi c = PFi d x fP

where:PFi c = corrected fan-shaft input power at design, hpPFi d = fan input power at design, hpfP = fan power correction factor


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• EQUATION 34: Typical fan power correction factor.

fP = 1 + 1.933 994 x exp (0.026 075 x %Qd)100

wherefP = fan power correction factor, dimensionless ratio%Qd = %WOcfm at design

• EQUATION 35: The fan input power at each load point.

PFi (n) = PFi c x %hp(n) x rn x cf%hpd rd

where:PFi (n) = fan shaft input power for time period, hp (W)PFi c = corrected fan input power at design, hp%hp(n) = percentage fan power for time period, %hp%hpd = percentage fan power at design, %hprn = actual density at load point, lbm/ft.3 (kg/m3)rd = design air density used to determine hpd,

lbm/ft.3 (kg/m3)cf = conversion factor, 1.0 (745.70)

• EQUATION 36: To determine the fan motor input power, themotor efficiency at each load point must be determined as afunction of the load factor.

LfM = PMo(n)

PM

whereLfM = motor load factor, dimensionless ratioPMo(n) = fan motor output power at load point, hpPM = nominal nameplate rating of motor, hp


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• EQUATION 37: Fan motor output power (derivative ofEquation 13, p.50).

PMo(n) = PFi(n)

hD(n)

where:PMo(n) = fan motor output power, hpPFi(n) = fan-shaft input power at load point, hphD(n) = drive efficiency, dimensionless ratio

• EQUATION 38: The fan motor input power at each load pointconsidering the part load efficiencies by combining Equation 13(p. 50) and Equation 14 (p. 50).

PMi(n) = PFi(n)

hD(n) x hM(n) x cf

where:PMi(n) = fan motor input power at load point, kWPFi(n) = fan-shaft input power at load point, hp (W)hD(n) = drive efficiency at load point,

dimensionless ratiohM(n) = motor efficiency at load point,

dimensionless ratiocf = conversion factor,1.3410 (1,000)

• The total energy consumption in kilowatt-hours is thendetermined by adding the product of the time in hours and thefan motor input power in kilowatts at each load point.

• EQUATION 39: Total energy consumption.

(PMi(n) x t(n))

where:E = energy consumption, kWh


nE=∑

1

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PMi(n) = fan motor input power for load point, kWt(n) = time at load point, hours

• EQUATION 40: Percentage static pressure at each load pointincluding the design point.

%sp(n) = a x %Q (n)b

where%sp(n) = percentage fan static pressure at load point, %sp%Q (n) = percentage fan flow at load point, %WOcfma,b = constants determined from table closest to

design WOcfm

• EQUATION 41: The fan static pressure at each load point.

Ps(n) = Psd x %sp(n) x rn x cf%spd rs

where:Ps(n) = static pressure at load point, in.Wg (Pa)Psd = static pressure at design, in.Wg (Pa)%sp(n) = percentage static pressure at load point, %sp%spd = percentage static pressure at design, %sprn = actual density at load point, lbm/ft.3 (kg/m3)rs = standard density used to determine psd, lbm/ft.3

(kg/m3 )cf = conversion factor, 1.0 (248.84)

• EQUATION 42: The fan static efficiency at each load point(rearrange Equation 11, p. 49).

hs(n) = Q(n) x Ps(n)

PFi (n) x cf

where:hs(n) = static efficiency at load point,

dimensionless ratio


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Q(n) = flow at load point, ft.3/min. (L/s)Ps(n) = static pressure at load point, in.Wg (Pa)PFi (n) = fan-shaft input power at load point, hp (W)cf = conversion factor, 6349.6 ( 1,000)

• EQUATION 43: Fan speed at the design point determined frommanufacturers' data and corrected for VIV losses.

NFc = NFd x Nf

where:NFc = corrected fan speed with VIVs, rpmNFd = fan speed at design without VIVs, rpmNf = fan speed correction factor

• EQUATION 44: Typical fan speed correction.

Nf = 1 + -41.329 73 + 11.168 903 x ln %Qd

100

where:Nf = fan-speed correction factor, dimensionless ratio%Qd = %WOcfm at design

EXAMPLE 4

• Uses the energy-analysis formulae in this chapter and thedensity-calculation formulae in Appendix A.

• A ventilation system is required for a new welding shop at amanufacturing plant. The shop is located in Toronto, Ontario,at 176m above sea level, and is to operate three shifts a day yearround. Analysis has determined the welding booth requires aminimum of 40,000 cfm of exhaust air for contaminate control.To remove excess heat in warm weather, doubling the exhaustvolume to 80,000 cfm has been considered.

• One proposed system is two exhaust fans of 40,000 cfm eachand a roof-mounted, VAV, makeup-air system with a blow-


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through, 54-in.-diameter, DWDI centrifugal fan with VIVs and aglycol heating coil. The makeup-air unit fan delivers amaximum of 80,000 cfm at a static pressure of 4.5 in.Wg, at0.075 lb/ft3.

• Also included is a roof-mounted gravity relief damper toprevent overpressurization of the zone. Whenever the systemmust operate, the control sequence would be as follows:

• One exhaust fan and the makeup-air unit start and runcontinuously. The VIVs in the unit modulate and providemakeup airflow as determined by a calibrated velocitypressure sensor in the supply duct. The supply air should bereset from the minimum volume of 40,000 cfm at an outdoortemperature of l3˚C to the maximum volume of 80,000 cfm atan outdoor temperature of l9˚C. When the VIVs are fullyopen, the second exhaust fan starts and runs continuously. Anormally-open control valve on the heating coil is modulatedto maintain a minimum supply-air temperature of 12.8˚C.

• To determine the annual energy consumption of the make-up-air fan, considering density flow variation, use ASHRAE MetricBin Weather Data and a PC spreadsheet program using themodified bin method to simplify calculations.

• The fan requires a 100-hp motor, and motor-efficiency data is inTable 8.


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Table 7: Typical VAV-fan Constants


Dependent Independent%WOcfm

Variable Variable Constant 50 55 60 65

%hp %Q a 1.015668 1.251667 1.847918 1.915869

b 0.078326 0.068488 0.055539 0.052399

c 42.50 42.25 41.50 41.50

%sp %Q a 0.036628 0.028756 0.022564 0.017582

b 2.000501 2.000751 2.000808 1.999086

Dependent Independent%WOcfm

Variable Variable Constant 70 75 80 85

%hp %Q a 2.012019 2.462608 3.751074 2.274535

b 0.048141 0.042180 0.033279 0.037477

c 41.50 41.00 39.50 41.50

%sp %Q a 0.013297 0.009782 0.007228 0.004959

b 2.000341 2.001993 1.997487 1.996321

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Table 8: Motor Load Efficiencies

Table 9: The Solution to Example 4


Motor Load Factor Percentage

Size

25 50 75 100 115 125

100 hp 89.05 91.86 93.50 93.05 92.83 92.70

GIVEN PARAMETERS VALUE REMARKS

Elevation, m 176

Standard Density (r), Ib/ft3 0.075

Flow (Q), cfm 80000

Static Pressure (ps), in.Wg 4.5

FAN SELECTION PARAMETERS

Power (PFi d) at Std Cond, bhp 74

Speed (NFd) at Std Cond, rpm 693

Max Flow (Qmax) 130300 From performance curve

at zero pressure, cfm at 693 rpm

%WOcfm (%Q), % 61.40 Q/Qmax

Maximum Static Pressure (Ps max) 5.5 From performance curve

in.Wg at 693 rpm

Motor Power (PM), bhp 100 Nominal motor size

Motor Efficiency (hM). % 93.5 Peak efficiency at 75%

load

contÕd

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Table 9: The Solution to Example 4 (contÕd)

Summary

• In reviewing the output from the spreadsheet program, it isapparent that the fan operates at a static efficiency of only 16%for over 6,000 hours a year. Therefore, the proposed controlscheme and system arrangement is not very efficient andanother scheme should be considered. Also, the ductworkwould have to be doubled in size for this arrangement, whichwould add to the installation cost.


VAV FAN CONSTANTS

%WOcfm 60 Value nearest Q/Qmax

Variables Constants

%hp %Q a 1.847918 Data from Table 7

b 0.055539

c 41.50000

%sp %Q a 0.022564

b 2.000808

CALCULATED PARAMETERS

Station pressure 99.261 Interpolated

(Pb), kPa

Drive efficiency (hD). % 4.00 Equation B1, p. 109

Percent power at design using 93.249 Equation 32, p. 92

(%hpd) 60%WOcfm

Percent pressure at design using 81.500 Equation 40, p. 95

(%spd) 60%WOcfm

VIV correction factor (fp) 1.09245 Equation 34. p. 93

VIV correction factor (Nf) 1.04400 Equation 44, p. 96

Corrected power (PFI C) bhp 80.841 Equation 33. p. 92

Corrected speed (NFc), rpm 723 Equation 44, p. 96

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Table 10. Summary of Example 4

Column Description Column Description

1 Bin dry-bulb temperature, ûC. 11 Fan static pressure, in.Wg; Equation 41.2 Bin mean coincident wet-bulb temperature, ûC. 12 Percent fan power; Equation 32.3 Bin hours of occurrence, h. 13 Fan-shaft input power, bhp; Equation 35.4 Saturation pressure, Pa.; Equation A6 and Equation A7, ûC. 14 Fan static efficiency, %o; Equation 42.5 Saturation humidity ratio, dimensionless; Equation A9. 15 Fan motor output power, bhp; Equation 37.6 Humidity ratio, dimensionless; Equation A11. 16 Motor load factor; dimensionless ratio; Equation 36.7 Density, Ib/ft.3; Equation A14 and Equation A15 converted to IP units. 17 Motor efficiency; Equation derived from a regression analysis of 8 Load factor derived from example control sequence. Table 8 data (alternatively linear interpolation can be used).9 Percent fan load; Equation 30. 18 Fan motor input power, kW; Equation 38.

10 Percent fan static pressure; Equation 40. 19 Fan energy consumption, kWh; Equation 27 (product of columns 3 and 18).

Tdb Twb Hr PWS Ws W Rho Lfn %Qn %spn psn %hpn PFin Seff PMon Lf M eff PMin En

1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19

34 25 8 3169.211 0.02051 0.01672 0.0684 100.00 60 81.50 4.11 93.25 73.78 70.13 76.85 0.768 0.935 61.31 49131 23 42 2810.437 0.01812 0.01477 0.0693 100.00 60 81.50 4.16 93.25 74.73 70.13 77.84 0.778 0.934 62.12 260928 21 186 2487.663 0.01599 0.01307 0.0702 100.00 60 81.50 4.21 93.25 75.68 70.13 78.83 0.788 0.934 62.92 1170425 19 352 2197.793 0.01408 0.01159 0.0711 100.00 60 81.50 4.26 93.25 76.62 70.13 79.81 0.798 0.934 63.72 2242822 18 742 2064.288 0.01321 0.01155 0.0718 100.00 60 81.50 4.31 93.25 77.40 70.13 80.63 0.806 0.934 64.38 4777019 15 630 1705.445 0.01087 0.00922 0.0728 100.00 60 81.50 4.37 93.25 78.48 70.13 81.75 0.818 0.934 65.30 4113616 13 644 1497.808 0.00953 0.00830 0.0737 75.00 45 45.83 2.49 64.00 54.50 43.10 56.77 0.568 0.924 45.83 2951513 11 969 1312.737 0.00834 0.00752 0.0745 50.00 30 20.36 1.12 51.28 44.18 15.93 46.b2 0.460 0.915 37.50 3633710 8 588 1072.839 0.00680 0.00598 0.0755 50.00 30 20.36 1.13 51.28 44.76 15.93 46.63 0.466 0.916 37.97 22325

7 5 811 872.485 0.00552 0.00471 0.0765 50.00 30 20.36 1.15 51.28 45.33 15.93 47.22 0.472 0.916 38.43 311674 3 621 758.030 0.00479 0.00438 0.0773 50.00 30 20.36 1.16 51.28 45.85 15.93 47.76 0.478 0.917 38.85 241241 0 1731 611.212 0.00385 0.00345 0.0783 50.00 30 20.36 1.17 51.28 46.42 15.93 48.35 0.484 0.917 39.31 68045

-2 -5 488 401.763 0.00253 0.00133 0.0794 50.00 30 20.36 1.19 51.28 47.09 15.93 49.05 0.491 0.918 39.85 19449-5 -7 461 338.193 0.00213 0.00133 0.0803 50.00 30 20.36 1.20 51.28 47.62 15.93 49.60 0.496 0.918 40.28 18569-8 -10 274 259.902 0.00163 0.00084 0.0813 50.00 30 20.36 1.22 51.28 48.19 15.93 50.20 0.502 0.919 40.75 11165

-11 -13 138 198.518 0.00125 0.00046 0.0823 50.00 30 20.36 1.23 51.28 48.78 15.93 50.81 0.508 0.919 41.22 5688-14 -16 56 150.676 0.00095 0.00016 0.0833 50.00 30 20.36 1.25 51.28 49.36 15.93 51.42 0.514 0.920 41.69 2335-17 -19 13 113.618 0.00071 -0.00007 0.0843 50.00 30 20.36 1.26 51.28 49.96 15.93 52.04 0.520 0.920 42.17 548-20 -22 5 85.096 0.00053 -0.00025 0.0853 50.00 30 20.36 1.28 51.28 50.57 15.93 52.67 0.527 0.921 42.66 213

TOTAL ENERGY CONSUMPTION (kWh) 395619

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APPENDIX A - DENSITY CALCULATIONS

Moist Air Parameters for Density Determination

• Moist air is defined as a binary mixture of dry air and watervapour. The maximum amount of water vapour at saturation inmoist air depends on the temperature and pressure.

• EQUATION A1: Moist air obeys the perfect gas equation.

p V= n R T

where:total pressure (p) is the sum of the partial pressure of dryair (pa) and the partial pressure of water vapour (pw), V isthe total mixture volume, the total moles (n) is the sum ofthe number of moles of dry air (na) and the number ofmoles of water vapour (nw), T is the absolute temperatureand R is the universal gas constant[1545.32 ft-lbf /lb-mol-F(abs) or 8.31441 J/(g-mol)-K].

• EQUATION A2: The density (p) of moist air is the ratio of thetotal mass to the total volume.

p = (ma + mw)/V

Chapter 16: Appendices 109

C H A P T E R 1 6

APPENDICES

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where:ma = mass of dry airmw = mass of water vapour

• EQUATION A3: The humidity ratio (W) of a given moist airsample is defmed as the ratio of the mass of water vapour tothe mass of dry air.

W = mw/ma

• The saturation humidity ratio (Ws) is the humidity ratio ofmoist air saturated with water (or ice) at the same temperatureand pressure.

• EQUATION A4: Relative humidity (f) is the ratio of the molefraction of water vapour in a given moist air sample to the molefraction in a saturated air sample at the same temperature andpressure.

f = xw/xws | t,p

• EQUATION A5: The dew point temperature (td) is thetemperature of moist air saturated at the same pressure (p) andwith the same humidity ratio (W) as that of the given sample ofair. It is defined as the solution td(p, W) to:

Ws(p,td) = W

• The thermodynamic wet-bulb temperature (twb) is thetemperature at which water (liquid or solid), by evaporatinginto moist air at a given dry-bulb temperature (t) and humidityratio (W), can bring air into saturation adiabatically at the sametemperature (twb), while the pressure (p) is constant.

Density Calculations

• The numerical method for calculating the density of moist air isa multistep process and initially involves determining the water-vapour saturation pressure. The saturation pressure in SI units


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at the wet-bulb temperature is calculated by the followingHyland and Wexler formulae published in the ASHRAEHandbook, 1989 Fundamentals Volume.

• EQUATION A6: The saturation pressure over ice for thetemperature range of -l00˚C to O˚C.

In(pws) = C1/T + C2 + C3T + C4T2 + C5T3 + C6T4 + C7In(T)

where:C1 = -5.674 535 9 E-3C2 = 6.392 524 7C3 = -9.677 843 E-3C4 = 6.221 157 Ol E-7C5 = 2.074 782 5 E-9C6 = -9.484 024 E-13C7 = 4.163 501 9

• EQUATION A7: The saturation pressure over water for thetemperature range of O˚C to 200˚C.

In(pws) = C8/T + C9 + Cl0T + C11T2 + C12T3 + C13In(T)

where:C8 = -5.800 220 6 E-3C9 = 1.391 499 3C10 = -4.864 023 9 E-2C11 = 4.176 476 8 E-5C12 = -1.445 209 3 E-8C13 = 6.545 967 3

and where for both equations:pws = saturation pressure, PaT = absolute temperature, K

• EQUATION A8: Saturation pressure in IP units:

pws(IP) = 0.020855 pws(SI)


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where:pws(IP) = saturation pressure, lb/ft2

pws(SI) = saturation pressure, Pa

• EQUATION A9: Saturation humidity ratio.

Ws = 0.62198 pws/(p - pws)

where:Ws = saturation humidity ratio, dimensionlessp = absolute pressure, lbf/ft2 (Pa)pws = saturation pressure, lbf/ft2 (Pa)

• EQUATION A10: The humidity ratio using the wet-bulbtemperature.

W = ( 1,093 - 0.556 twb) Ws(wb) -0.240 (tdb - twb)1,093 + 0.444 (tdb - twb)

• EQUATION A11: or in SI units:

W = (2,501- 2.381 twb) Ws(wb) - (tdb - twb)2,501 + 1.805tdb - 4.186twb

where for both Equation A10 and Equation All:W = humidity ratio, dimensionlessWs = saturation humidity ratio, dimensionlesstdb = dry-bulb temperature, ˚F (˚C)twb = wet-bulb temperature, ˚F (˚C)

• EQUATION A12: The humidity ratio.

W = 0.62198 pw/(p - pw)

where:W = humidity ratio, dimensionlessp = total pressure, lbf/ft.2 (Pa)pw = partial pressure of water vapour, lbf/ft.2 (Pa)


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• EQUATION A13: Relative humidity using the perfect gasrelationships:

f = Pw/Pws | t,p

where:f = relative humidity, dimensionlesspw = partial pressure of water vapour, lbf/ft.2 (Pa)pws = saturation pressure, lbf/ft.2 (Pa)

• EQUATION A14: Volume.

v = RaT ( 1 + 1 .6078 W)p

where:v = volume, ft.3/lbm, (m3/kg)Ra = gas constant for air, 53.352 ft.lbf/lbm.F(abs)

(287.055 J/kg.K)T = absolute temperature, ˚F(abs) (K)p = total pressure, lbf/ft2 (Pa)W = humidity ratio, dimensionless

• EQUATION A15: Density is determined by the inverse of thevolume.

r = 1/v

where:r = density, lbm/ft.3 (kg/m3)

• The equation to use to determine density depends on themeasured parameters. Table A2 outlines which equations mustbe used with the appropriate parameters.


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Table A1. Standard Atmospheric Data for Altitudes to 3,000 m

Source: ASHRAE Handbook, 1989 Fundamentals Volume

Table A2. Density Calculations


Altitude Temperature Pressure

m ft. ûC ûF kPa in.Hg

0 0 15.0 59.0 101.325 29.921

500 1,640 11.8 53.2 95.461 28.19

1,000 3,281 8.5 47.3 89.874 26.54

2,000 6,562 2.0 35.6 79.495 23.47

3,000 6,562 -4.5 23.9 70.108 20.70

GivenParameters To Obtain Use Comments

tdb, twb, p pws | twb Equation A6 or Using twb

Equation A7

Ws | twb Equation A9 Using pws | twb

W Equation A10 or Using ps | twb

Equation A11

v Equation A14 Using W

r Equation A15 Using v

tdb, td, p pw = pws | td Equation A6 or Using tdEquation A7

W Equation A12 Using Pws | td for Pw



tdb, f, p Pws | tdb Equation A6 or Using tdb

Equation A7

Pw Equation A13 Using pws

W Equation A12 Using pw



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APPENDIX B - DRIVE LOSS CALCULATIONS

• Power transmission losses must be considered in energyconsumption analysis whenever a direct-drive system is notused. The types of drive systems with losses include hydraulicand gear drives, belt drives (including V-belts and rubber chain),and variable-speed drives including eddy current clutches andelectronic and mechanical variable-speed devices.

• Determining precise drive losses involves laboratory testingprocedures. However, using the methods in this appendix,losses for the common drive systems can be estimated withsuitable accuracy.

V-belt Drives

• Expressed as a percentage of motor output, these lossesdiminish logarithmically as the motor size increases. Inaddition, there is a range for each motor size where typicallythe losses increase as speed increases. The value of the drive-belt loss can be determined by the graph in AMCA Publication203-90, or the mean drive loss can be determined by thefollowing equations:

• EQUATION Bl: Fractional horsepower motors:LD =9.4-4.651 27 In PM

• EQUATION B2: Motors from 1 to 10 horsepower:LD = 9.4 -1.867 47 In PM

• EQUATION B3: Motors from 10 to 100 horsepower:LD = 6.2 - 0.477 724 In PM

• EQUATION B4: Motors over 100 horsepower:LD = 4.0

where for all applicable equations:LD = drive loss in percent of motor output, %PM = nominal rated motor output power, hp


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• EQUATION B5: Actual motor output power:PMo = PFi/cf(1.0 - LD/l00)

where:PMo = motor output power, bhp (kW)PFi = fan shaft input power, bhpLD = drive loss in percent of rated motor output

power, %cf = conversion factor, 1.0 (0.745 70)

Rubber Chain Drives

• Relatively new drive method that is more efficient than V-beltdrive systems. The drive pulleys are ribbed and the belt istoothed to prevent slippage. Drive losses occur due to thebending forces as the belt rotates around the pulleys. Ask themanufacturer for the drive losses, or assume that LD = 2.0.

• Fan motor output power is determined using Equation B5.

Electronic Variable-speed Drives

• Manufacturers publish the part-load efficiencies and powerfactor of their electronic variable-speed drives as a function ofoutput speed at constant load and, in the case of fan systems,as a function of output speed with load reducing with the cubeof the speed change. If the manufacturer does not supply thisdata, the part-load drive efficiency can be determined bymultiplying the full-load design drive efficiency by the part-loadcorrection factor.

• EQUATION B6:

cfv =1.0 + (0.203 176 x 1n Nfv)

where:cfv = variable-torque drive correction factor,

dimensionless ratio


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Nfv = speed fraction of variable-speed drive,dimensionless ratio

• EQUATION B7: Variable-speed drive efficiency.

hV = hvd x cfv100

where:hV = part-load drive efficiency, dimensionless ratiohvd = drive efficiency at full-load design, %

• The part-load efficiency of the motor must be considered todetermine the input power to the variable-speed drive. Data onmotor part-load efficiency at reduced speed may be difficult toobtain. In the absence of this data use the part-load efficiency atfull rated speed.

• EQUATION B8: Variable-speed drive input power.

PVi = PFi /hV hD hM

where:Pvi = variable-speed drive input power, hp (kW)PFi = fan-shaft input power, hp (kW)hV = variable-speed drive efficiency,

dimensionless ratiohD = drive efficiency, dimensionless ratiohM = motor efficiency, dimensionless ratio

• For installed systems where the input load and power factorcan be measured, the output of the variable-speed drive can bedetermined by;

• EQUATION B9: For single - phase power:

PVo = (E I pf hD)/cf


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• EQUATION B10: For three - phase power:

PVo = (√3 E I pf hD)/cf

where for both equations:PVo = variable-speed drive output power, hp (kW)E = average of the measured phase voltsI = average of the measured phase amps

= conversion factor, 745.70 ( 1,000)pf = power factor, dimensionless ratiohD = drive efficiency, dimensionless ratio

Other Drive Systems

• For systems such as eddy current clutches and hydraulic couplings,ask the manufacturer for the drive efficiencies. If this data is notavailable, then the actual fan input power must be determinedby laboratory tests.


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APPENDIX C - FAN OUTLET LOSS COEFFICIENTS (REF.ASHR)

Table C1. Plane Asymmetric Diffuser at Fan Outlet WithoutDuctwork

Figure C1. Plane Asymmetric Diffuser at Fan Outlet WithoutDuctwork


C0

q A1 / A2

degree 1.5 2.0 2.5 3.0 3.5 4.0

10 0.51 0.34 0.25 0.21 0.18 0.17

15 0.54 0.36 0.27 0.24 0.22 0.20

20 0.55 0.38 0.31 0.27 0.25 0.24

25 0.59 0.43 0.37 0.35 0.33 0.33

30 0.63 0.50 0.46 0.44 0.43 0.42

35 0.65 0.56 0.53 0.52 0.51 0.50

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Table C2. Pyramidal Diffuser at Fan Outlet Without Ductwork

Figure C2. Pyramidal Diffuser at Fan Outlet Without Ductwork


C0

q A1 / A2

degree 1.5 2.0 2.5 3.0 3.5 4.0

10 0.54 0.42 0.37 0.34 0.32 0.31

15 0.67 0.58 0.53 0.51 0.50 0.51

20 0.75 0.67 0.65 0.64 0.64 0.65

25 0.80 0.74 0.72 0.70 0.70 0.72

30 0.85 0.78 0.76 0.75 0.75 0.76

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Table C3. Plane Symmetric Diffuser at Fan Outlet With Ductwork

Figure C3. Plane Symmetric Diffuser at Fan Outlet With Ductwork


C0

q A1 / A2

degree 1.5 2.0 2.5 3.0 3.5 4.0

10 0.05 0.07 0.09 0.10 0.11 0.11

15 0.06 0.09 0.11 0.13 0.13 0.14

20 0.07 0.10 0.13 0.15 0.16 0.16

25 0.08 0.13 0.16 0.19 0.21 0.23

30 0.16 0.24 0.29 0.32 0.34 0.35

35 0.24 0.34 0.39 0.44 0.48 0.50

Fans (15-21) 1/8/01 10:54 AM Page 121

Table C4. Plane Asymmetric Diffuser at Fan Outlet With Ductwork

Figure C4. Plane Asymmetric Diffuser at Fan Outlet With Ductwork


C0

q A1 / A2

degree 1.5 2.0 2.5 3.0 3.5 4.0

10 0.08 0.09 0.10 0.10 0.11 0.11

15 0.10 0.11 0.12 0.13 0.14 0.15

20 0.12 0.14 0.15 0.16 0.17 0.18

25 0.15 0.18 0.21 0.23 0.23 0.26

30 0.18 0.25 0.30 0.33 0.35 0.35

35 0.21 0.31 0.38 0.41 0.43 0.44

Fans (15-21) 1/8/01 10:54 AM Page 122




C0

q A1 / A2

degree 1.5 2.0 2.5 3.0 3.5 4.0

10 0.05 0.08 0.11 0.13 0.13 0.14

15 0.06 0.10 0.12 0.14 0.15 0.15

20 0.07 0.11 0.14 0.15 0.16 0.16

25 0.09 0.14 0.18 0.20 0.21 0.22

30 0.13 0.18 0.23 0.26 0.28 0.29

35 0.15 0.23 0.28 0.33 0.35 0.36

Fans (15-21) 1/8/01 10:54 AM Page 123




C0

q A1 / A2

degree 1.5 2.0 2.5 3.0 3.5 4.0

10 0.11 0.13 0.14 0.14 0.14 0.14

15 0.13 0.15 0.16 0.17 0.18 0.18

20 0.19 0.22 0.24 0.26 0.28 0.30

25 0.29 0.32 0.35 0.37 0.39 0.40

30 0.36 0.34 0.46 0.49 0.51 0.51

35 0.44 0.54 0.61 0.64 0.66 0.66

Fans (15-21) 1/8/01 10:54 AM Page 124

Table C7. Pyramidal Diffuser at Fan Outlet With Ductwork

Figure C7. Pyramidal Diffuser at Fan Outlet With Ductwork


C0

q A1 / A2

degree 1.5 2.0 2.5 3.0 3.5 4.0

10 0.10 0.18 0.21 0.23 0.24 0.25

15 0.23 0.33 0.38 0.40 0.42 0.44

20 0.31 0.43 0.48 0.53 0.56 0.58

25 0.36 0.49 0.55 0.58 0.62 0.64

30 0.42 0.53 0.59 0.64 0.67 0.69

Fans (15-21) 1/8/01 10:54 AM Page 125

Chapter 17: Conversion Tables 127

C H A P T E R 1 7

CONVERSION TABLES

From To Multiply by

ûF ûC TûC = (tûF Ð 32)/1.8

ûF ûR or F(abs) T(ûR) = tûF + 459.67

ûF K T(K) = (tûF + 459.67) /1.8

ûR or F(abs) K T(K) = TûR /1.8

ûC K T(K) = tûC + 273.15

K ûC t (ûC) = TK - 273.15

Fans (15-21) 1/8/01 10:54 AM Page 127


Inch/Pound (IP) to Metric (SI) SI� to lP

length

1 in. 25.400 mm 1 mm 0.039 37 in.

1 ft. 0.304 80 m 1 m 3.2808 ft.

area

1 in.2 645.16 mm2 1 mm2 0.00155 in.2

1 ft.2 0.092 903 m2 1 m2 10.764 ft.2

mass

1 Ibm 0.453 59 kg 1 kg 2.2046 Ibm

volume

1 ft.3 0.028 317 m3 1 m3 35.315 ft.3

1 ft.3 28.317 I 1 0.035 315 ft.2

1 gal. Imp. 4.546 1 I 1 I 0.219 97 gal. Imp.

1 gal. US 3.785 4 I 1 I 0.264 17 gal. US

density

1 Ibm/ft.3 16.018 kg/m3 1 kg/m3 0.062 430 Ibm/ft.3

specific v

ft3/Ibm 0.062 43 m3/kg 1 m3/kg 16.018 ft3/Ib.

velocity

1 fps 0.304 80 m/s 1 m/s 3.280 8 fps

1 fpm 0.005 0800 m/s 1 m/s 196.85 fpm

force

1 Ibf 4.448 2 N 1 N 0.224 81 Ibf

torque

1 Ibf.ft. 1.355 8 N.m 1 N.m 0.737 56 Ibf.ft.

Fans (15-21) 1/8/01 10:54 AM Page 128

Note: 1 Water and mercury at 20˚C (68˚F)2 M = 103 in Mbh

Chapter 17: Conversion Tables 129

Inch/Pound (IP) to Metric (SI) SI to IP

flow rate

1 cfs 28.317 m3/s 1 m3/s 35.315 cfs

1 cfm 0.471 95 m3/s 1 m3/s 2.118 9 cfm

1 gpm (Imp.) 0.075 77 L/s 1 L/s 13.198 gpm (Imp.)

1 gpm (US) 0.063 09 L/s 1 L/s 15.850 gpm (US)

pressure / head

1 psi 6.894 8 kPa 1 kPa 0.145 03 psi

1 psf 0.047 88 kPa 1 kPa 20.885 psf

1 ft.Wg(1) 2.986 1 kPa 1 kPa 0.334 88 ft.Wg� (1)

1 in.Wg(1) 248.84 Pa = 0.036 psi 1 kPa 4.018 6 in.Wg� (1)

1 in.Hg(1) 3.376 9 kPa 1 kPa 0.296 12 in.Hg� (1)

1 psi = 27.8 in Wg

energy, work

1 Btu 1.055 1 kJ 1 kJ 0.947 85 Btu

1 kWh 3600.0 kJ 1 MJ 0.277 78 kWh

1 ft.lbf 1.355 8 J 1 J 0.737 56 ft.lbf

power

1 Btu/h 0.293 07 W 1 kW 3.412 2 MBh (2)�

1 hp 746.00 W 1 kW 1.340 5 hp (electric)

1 hp = 550 ft-lb/sec

Fans (15-21) 1/8/01 10:54 AM Page 129

• Atmospheric pressure (standard)14.7 psia, 101 kPapsia = pounds/sq.in. “absolute”

i.e., includes atmospheric pressure

• p sig.= pounds/sq.in. “gauge”= pressure measured above local atmospheric pressure,

( i.e., not including atmospheric pressure )


Fans (15-21) 1/8/01 10:54 AM Page 130

ABBREVIATIONS

A = fan outlet area, ft.2 (m2)

Bf = blade frequency

C = constant

Co = system effect coefficient, dimensionless

cf = conversion factor

cfm = cubic feet per minute

cfs = cubic feet per second

D = fan size or impeller diameter

E = energy consumption, kWh

fp = fan power correction factor, dimensionless ratio

I = amperage

In = natural logarithm

K = value for calculating system effect factors

Kw = specific sound power level, dB re 1 pW

LD = drive loss, %

lf = load factor

Chapter 18: Abbreviations and Symbols 131

C H A P T E R 1 8

ABBREVIATIONS AND

SYMBOLS

Fans (15-21) 1/8/01 10:54 AM Page 131

Lw = sound power level, dB re lpW

m = mass, lbm (g)

N = rotational speed, rpm (r/s)

Nf = fan speed correction factor, dimensionless ratio

P = power, hp (kW)

p = pressure, in.Wg (Pa) or psi (kPa)

pB = barometric pressure, in.Hg (kPa)

pf = power factor, dimensionless ratio

PFi = shaft power input to the fan, hp (kW)

PFo = air power output of the fan, hp (kW)

PMi = motor input power, kW

psia = pounds per square inch, atmospheric

psig = pounds per square inch, gauge

pt = fan total pressure rise, in.Wg (Pa)

Pvi = variable-speed input power, hp (kW)

PMo = motor power output to the fan drive, hp (kW)

ps = fan static pressure rise, in.Wg (Pa)

pv = fan velocity pressure, in.Wg (Pa)

Pvo = variable-speed output power, hp (kW)

pws = saturation pressure, lb/ft.2, Pa

Q = volume flow rate at inlet conditions, cfm (L/s)

R = universal gas constant, Ft-lbf/lb-mol.F(abs)(J/g-mol.K)

r = radius, ft., in. (m, cm)

rpm = revolutions per minute

SF = service factor

SEF = system effect factor


Fans (15-21) 1/8/01 10:54 AM Page 132

T = thermodynamic temperature, ˚R (K)

t = customary temperature, ˚F (˚C)

tb = time at temperature bin, hours

td = temperature differential

tdb = dry-bulb temperature, ˚F (˚C)

twb = wet-bulb temperature, ˚F (˚C)

V = velocity, ft./min. (m/s)

Va = axial velocity component

Vb = velocity relative to blade

Vm = mean velocity component

Vr = radial velocity component

Vs = absolute velocity

Vt = tangential velocity component

W = humidity ratio, dimensionless

Wg = water gauge

Ws = saturation humidity ratio

WOcfm = wide open cubic feet per minute, %

%hp = percent fan power for load point

%sp = percent static pressure for load point

SYMBOLS

˚C = degree Celsius

dB = decibel

˚F = degree Fahrenheit

g = gram

hp = horsepower

Chapter 18: Abbreviations and Symbols 133

Fans (15-21) 1/8/01 10:54 AM Page 133

in.Hg = inch of mercury

K = kelvin

kPa = kilopascal (103 x pascal)

L/s = litre per second

m/s = metre per second

Pa = pascal

pW = picowatt

˚R = degree Rankine

rad/s = radian per second

W = watt

∆t = temperature difference

hM = motor efficiency, dimensionless ratio

hD = drive efficiency, dimensionless ratio

hV = variable-speed drive efficiency, dimensionless ratio

hs = static efficiency of fan, dimensionless ratio

ht = total efficiency of fan, dimensionless ratio

q = plane angle

v = volume, ft.3/lbm (m3/kg)

r = density, lbm/ft3 (kg/m3)

∑ = summation of

f = relative humidity

w = rotational speed, rad/s


Fans (15-21) 1/8/01 10:54 AM Page 134

• Air Movement and Control Association (AMCA). Fans andSystems. Publication 201-90. 30 West University Drive,Arlington Heights, Ill. 60004-1893, (708)394-0150, 1990.

• AMCA. Field Performance Measurement of Fan Systems. Publication203-90. Arlington Heights, Ill., 1990.

• American Society of Heating, Refrigerating and AirConditioning Engineers (ASHRAE). Handbook, l988 EquipmentVolume. Atlanta, 1988.

• ASHRAE. Handbook, 1989 Fundamentals Volume. Atlanta, 1989.

• ASHRAE. Handbook, 1991 HVAC Applications Volume. Atlanta,1991.

• ASHRAE. Metric Bin Weather Data, Toronto International Airport.Atlanta, n.d.

• ASHRAE. Simplified Energy Analysis using the Modified Bin Method.Report TC 4. I . Atlanta, 1983.

• E.A. Avallone and T. Baumeister. MARKS' Standard Handbook forMechanical Engineers. 9th ed. New York: McGraw-Hill, 1987.

Chapter 19: Bibliography 135

C H A P T E R 1 9

BIBLIOGRAPHY

Fans (15-21) 1/8/01 10:54 AM Page 135

• Buffalo. High Pressure Type HL Industrial Exhausters. Bulletin FI-115A. Kitchener, Ont.: Canada Blower/Canada Pumps, 1987.

• Buffalo. Type BL Centrifugal Fans. Bulletin F107-B, Kitchener,Ont.: Canada Blower/Canada Pumps, 1987.

• Greenheck Fan Corp. CAPS- Computer Selection Program.Schofield, Wis., 1986.

• Greenheck Fan Corp. Centrifugal Fans - Backward Inclined andAirfoil Single and Double Width. Catalogue Cent. Fab (BI/AF) R. Schofield, Wis., 1990.

• Greenheck Fan Corp. Centrifugal Roof Exhausters - Models G andGB. Catalogue G/GB R. Schofield, Wis., 1989.

• Greenheck Fan Corp. Industrial Fans – Open Radial & Radial Tip.Catalogue IF 1-86 M. Schofield, Wis., 1986.

• Greenheck Fan Corp. Inline Fans – Models SQ and BSQ.Catalogue DSQ/BSQ R. Schofield, Wis., 1989.

• Greenheck Fan Corp. Sidewall Propeller Fans – Belt Drive.Catalogue SPF-M. Schofield, Wis., 1989.

• Greenheck Fan Corp. Sidewall Propeller Fans – Direct Drive.Catalogue SD-APR. 89-M. Schofield, Wis., 1989.

• Greenheck Fan Corp. Tubeaxial Fans – Direct and Belt Drive.Catalogue TAB/TAD 2 R. Schofield, Wis., 1989.

• Greenheck Fan Corp. Utility Fans – Forward Curved and BackwardInclined. Catalogue SFD/SFB 3-86 M. Schofield, Wis., 1986.

• Greenheck Fan Corp. Vane Axial – Response Control and PresetPitch Fans. Catalogue VR, VP R M. Schofield, Wis., 1989.

• Jorgensen, R. (ed.) Fan Engineering. 8th ed. Buffalo, NY, BuffaloForge Company, 1983.


Fans (15-21) 1/8/01 10:54 AM Page 136

• Public Works Canada. Metric Conversion Handbook for MechanicalEngineers in the Building Industry. 2nd ed. 1983.

• The Trane Company. CDS – Customer Direct Service ComputerSelection Program. Vol 10.1. La Crosse, Wis., n.d.

• The Trane Company. Centrifugal Fans – Sizes 12-89 Single andDouble Width. catalogue PL-AH-FAN-000-DS-6-1083. La Crosse,Wis., 1983.

Chapter 19: Bibliography 137

Fans (15-21) 1/8/01 10:54 AM Page 137

abscissa• horizontal coordinate of a point in a plane Cartesian coordinate

system obtained by measuring parallel to the x-axis. (Compareordinate.)

absolute humidity• in a mixture of water vapour and dry air, the mass of water

vapour in a specific volume of the mixture. Compare relativehumidity.

absolute (thermodynamic) temperature• temperature as measured above absolute zero.

absolute (dynamic) viscosity• force per unit area required to produce unit relative velocity

between two parallel areas of fluid unit distance apart, alsocalled coefficient of viscosity.

absolute zero temperature• zero point on an absolute temperature scale.

Chapter 20: Glossary 139

C H A P T E R 2 0

GLOSSARY

Fans (15-21) 1/8/01 10:54 AM Page 139

adiabatic exponent• exponent k in the equation pvk = constant, representing an

adiabatic change (k is the ratio of the specific heat at a constantpressure to the specific heat at constant volume).

adiabatic process• thermodynamic process during which no heat is extracted from

or added to the system.

aerodynamic excitation• time varying loads acting on the blades of a fan due to

nonconformities of the air flow. Note: Spatial nonuniformities ofairflow that are steady in time give rise to harmonic excitationat frequencies that are integer multiples of the rotation rate ofthe fan. TIme excitations of the airflow give rise to randomexcitation.

air• ambient local atmospherical air supply at fan intake.

air change• introduction of new, cleansed, or recirculated air to a space.

air-conditioning system• assembly of equipment for air treatment to control

simultaneously its temperature, humidity, cleanliness anddistribution to meet the requirements of a conditioned space.

airflow resistance• deterrent (due to friction, change of direction, etc.) to the

passage of air within a system of airways or an apparatus.

air power (operational)• power required to move air at a given rate of flow against a

given resistance. The ratio of air power to input power of a fanor blower is termed efficiency.


Fans (15-21) 1/8/01 10:54 AM Page 140

air power (theoretical)• power required to drive a fan or blower as though there were

no losses in the fan or blower (100% efficiency).

algorithm• prescribed set of well defined rules, or process, for the solution

of a problem in a finite number of steps, e.g., a full statement ofan arithmetical procedure for evaluating sine X to a statedprecision.

ambient air• surrounding air (usually outdoor air or the air in an enclosure

under study).

ANSI• American National Standards Institute

apparent power• product of the volts and amperes of a circuit. This product

generally is divided by 1,000 and designated in kilovolt-amperes (kVA). It comprises both real and reactive power.

ARI• Air-Conditioning and Refrigeration Institute.

ASHRAE• American Society of Heating, Refrigerating, and Air

Conditioning Engineers.

ASTM• American Society for Testing and Materials.

baghouse fan• an exhaust fan for conveying smoke, dust, etc., into filters for

pollution control.


Fans (15-21) 1/8/01 10:54 AM Page 141

balance pressure• pressure in a system or container equal to that outside.

bin method• energy calculation method, usually used for prediction, in

which the annual (or monthly) energy use of a building iscalculated as the sum of the energy used for all the outdoortemperature bins. It allows heat pump (or other heater orcooler) performance, which is different for each bin, to beaccounted for.

boundary layer• region of retarded fluid-flow near the surface of a body moving

through the fluid, or past which the fluid moves.

brake horsepower (BHP)• actual power delivered by or to a shaft (from the use of a brake

to measure power).

British thermal unit (Btu)• the mechanical equivalent energy of a Btu is approximately

778.169 262 ft. lb. The heat energy of a Btu is approximatelythat required to raise the temperature of a pound of water from59˚F to 60˚F.

capacity• maximum load for which a machine, apparatus, device or

system is designed or constructed.

cell (in a cooling tower)• smallest tower subdivision that can function as an independent

heat exchange unit. It is bounded by exterior walls or partitions.Each cell may have one or more fans or stacks and one or moredistribution systems.


Fans (15-21) 1/8/01 10:54 AM Page 142

central fan system• mechanical indirect system of heating, ventilating or air-

conditioning, in which the air is treated or handled byequipment located outside the rooms served, usually at acontrol location, and is conveyed to and from the rooms bymeans of a fan and a system of distributing ducts.

cleansed air• air that has been treated to remove pollutants, particulates and

odours.

coil• cooling or heating element made of pipe or tube that may or

may not be finned, formed into helical or serpentine shape.

compressibility• ease with which a fluid may be reduced in volume by the

application of pressure.

compressor• device for mechanically increasing the pressure of a gas.

conditioned air• air treated to control its temperature, relative humidity, purity,

pressure and movement.

control/controller• manual or automatic device for regulating a system or

component in normal operation.

cooling tower• heat-transfer device in which atmospheric air cools warm water,

generally by direct contact (evaporation).

• mechanical-draft, water-cooling tower; tower throughwhich air movement is effected by one or more fans.


Fans (15-21) 1/8/01 10:54 AM Page 143

counterflow• in heat exchange between two fluids, the opposite direction of

flow; i.e., the coldest portion of one fluid meeting the coldestportion of the other.

critical speed• The speed at which a fan, duct, or other component will vibrate

in resonance.

damper• device used to vary the volume of air passing through an outlet,

inlet or duct, or generally through a confined cross section byvarying the cross-sectional area.

decibel• unit of air sound pressure and sound power.

design airflow• required airflow when the system is operating under assumed

maximum conditions, including diversity.

design conditions• specified environmental conditions, e.g., temperature and

humidity, required to be produced and maintained by a system.

design working pressure• in the U.S., the maximum working pressure for which an

apparatus has been designed. In some countries, the designpressure is greater than the maximum working pressure.

dew point• temperature at which water vapour has reached saturation

point (100% relative humidity).


Fans (15-21) 1/8/01 10:54 AM Page 144

dry air (definition for HVAC applications)• 1. air without entrained water vapour. 2. air unmixed with or

containing no water. Note: Composition of dry air is defined inISO 2533-1975, without contaminants or pollution.

dry-bulb temperature• temperature of air indicated by an ordinary thermometer.

duct system• series of ducts, elbows and connectors to convey air or other

gases from one location to another.

dynamic pressure• additional pressure exerted by a fluid due to motion, if that

motion were converted to a static pressure, as in a fluid jetimpinging on a surface.

equivalent length• resistance of fittings or appurtenances in a conduit through

which the fluid flows, expressed in length of straight conduit ofthe same diameter or shape that would have the sameresistance; also expressed in length/diameter units.

evasé• a diffuser duct section on fan outlet to regain static pressure.

As the diffuser, in fact, adds a loss, the fan total efficiency is reduced.

external vibration isolation• in an air-handling unit, isolation of its vibration by devices

external to the unit.

fan• device for moving air by two or more blades or vanes attached

to a rotating shaft.


Fans (15-21) 1/8/01 10:54 AM Page 145

fan air density• density of air corresponding to the absolute pressure and

absolute temperature at the fan inlet when the fan is operating.

fan appurtenances• accessories added to a fan for control, isolation, safety, static

pressure regain, wear, etc. (inlet boxes, inlet box dampers,variable inlet vanes, outlet dampers, vibration isolation bases,inlet screens, belt guards, diffusers, sound attenuators, wearprotection, turning gears).

fan blast area• fan scroll outlet area less the area of the cutoff.

fan boundary (inlet and outlet)• interface between the fan and the remainder of the system, at a

plane perpendicular to the airstream where it enters or leavesthe fan.

fan casing (volute, scroll)• the part of the casing of a centrifugal fan or compressor that

receives fluid forced outward from the impeller or diffuser andleads it to the discharge. (Compare fan shroud.)

fan coil (convector) unit• fan and a heat exchanger for heating and/or cooling assembled

within a common casing.

fan curve• diagram giving the pressure/volume characteristics of a fan, and

the power it requires.

fan free-discharge area• area where the fan chamber meets the discharge scroll. Used in

fan system-effect calculations, (the outlet boundary).


Fans (15-21) 1/8/01 10:54 AM Page 146

fan inlet (outlet) area• area of the fan or fan equipment for connection to attached

ductwork.

fan nodal line• The point of zero displacement on any component vibrating at

its natural frequency.

fan (constant speed) performance curve• graphical representation of static or total pressure and power

input over a range of air volume flow rate at a stated inletdensity and fan speed. It may include static and mechanicalefficiency curves.

fan power• power input at the fan shaft, or the total of the power input to

the fan shaft and the power loss attributable to the powertransmission device.

fan pressurization test• test for determining the air leakage of a building using a fan-

induced pressure difference.

fan propeller• propeller or disc-type wheel within a mounting ring or plate,

and including driving mechanism supports for either belt-driveor direct connection. (Compare impeller.)

fan shroud• protective housing that surrounds the fan and that may also

direct the flow of air. (Compare fan casing.)

fan sound power• sound power radiated into a duct, or through the housing.


Fans (15-21) 1/8/01 10:54 AM Page 147

fan static pressure• difference between fan total pressure and fan discharge velocity

pressure.

fan torsional excitation• type of excitation in which external force is applied to the fan

shaft in the form of torque pulsations.

fan total pressure• arithmetic difference between fan-outlet total pressure and fan-

inlet total pressure.

fan wheel• revolving part of a fan or blower.

fan wheel cone• inlet ring, impeller shroud, impeller rim annular plate, or conical

ring on the air inlet side of a centrifugal fan to which theimpeller blades are fixed.

filter mixing box• in air-handling units, a combination filter section outside-/

return-air mixing plenum, including control dampers.

flow nozzle• tube specially shaped to increase the discharge velocity of the

fluid, to minimize contraction losses.

flow velocity• velocity (local or average) of a fluid in a pipe, duct or canal, or

from an orifice.


Fans (15-21) 1/8/01 10:54 AM Page 148

frequency response• normalized motion response of a fan to a known excitation,

expressed as a function of the frequency of the excitation. Isusually given graphically by curves showing the relationship ofthe response to the excitation (and, where applicable, phaseshift or phase angle) as a function of frequency.

full-load amperes• current that a rotating machine will draw from the power line

when the machine is operating at rated voltage, speed andtorque.

gauge pressure• pressure above atmospheric pressure.

head• energy per unit mass of fluid divided by gravitational

acceleration. In fluid statics and dynamics, a vertical linearmeasure. Note: The terms head and pressure are oftenmistakenly used interchangeably.

head pressure• operating pressure measured in the discharge line at a pump,

fan or compressor outlet; i.e., at the head.

horsepower• work done at the rate of 550 ft- lb/sec. (745.7 W). (See also brake

horsepower.)

HVAC systems• provide either collectively or individually the processes of

comfort heating, ventilating and/or air conditioning within, orassociated with, a building.


Fans (15-21) 1/8/01 10:54 AM Page 149

hydrostatic pressure• pressure exerted by a fluid at rest.

impeller (rotor; wheel)• rotating part of a device (fan, blower, compressor or pump) that

moves fluid. (See also fan.)

impeller reaction• ratio of the variation of the fluid pressure in the impeller to the

total variation of pressure in the device.

impeller running noise frequency• in a turbomachine, the noise frequency resulting from the

rotational speed of the impeller times the number of blades.

intermediate pressure (interstage pressure)• pressure between stages of multistage compression.

internal vibration isolation• in an air-handling unit, spring isolation of all moving parts

within the unit that support the fan sled.

IP units (inch-pound units)• units using inches, pound and other designations; as opposed

to SI units in the metric system. Examples are foot, Btu,horsepower, gallon.

iterative procedure• process which repeatedly executes a series of operations until

some prescribed condition is satisfied.

joule (J)• 1. (electric work) work done by one ampere flowing through a

resistance of one ohm for one second. J = W·sec. (watt second).2. (heat or mechanical work) work done by a force of onenewton acting over one metre. J = N·m.


Fans (15-21) 1/8/01 10:54 AM Page 150

kelvin temperature• SI absolute temperature scale (K), on which the triple point of

water is 273.16K and the boiling point is approximately373.15K ( 1 K = 1˚C). Kelvin is 1/273.16 of the temperature ofthe thermodynamic triple point of water.

kinematic viscosity• ratio of absolute viscosity to density of a fluid.

laminar flow (streamline)• fluid flow in which all the particles move in substantially

parallel paths, occurs at low Reynolds numbers.

mixing box• compartment in which two air supplies are mixed together

before being discharged.

modulate• 1. adjust by small increments and decrements. 2. vary a voltage

or other variable with a signal.

noise (NC) criteria curves• curves that define the limits that the octave-band spectrum of a

noise source must not exceed if a certain level of occupantacceptance is to be achieved.

noise reduction (NR)• difference between the average sound pressure levels, or sound

intensity levels of two spaces – usually two adjacent roomscalled the source room and the receiving room respectively.

operating load point• actual system operating capacity at the time of taking an

instrument reading.


Fans (15-21) 1/8/01 10:54 AM Page 151

ordinate• the Cartesian coordinate obtained by measuring parallel to the

y-axis. (Compare abscissa.)

outdoor air• air outside a building, or air taken from outdoors and not

previously circulated through the system.

outlet area• gross overall discharge area of a given component in an air

distribution system.

output• capacity, duty, performance, net refrigeration produced by a

system.

phase• 1. in thermodynamics, one of the three states of matter, solid,

liquid, or gas. 2. position in a cycle.

pitot tube• small bore tube inserted perpendicular to a flowing stream with

its orifice facing the stream to measure total pressure.

polytropic process• one in which heat is being exchanged with the surroundings,

represented by the equation pvn = constant (n is the polytropicexponent). Describes the process in a fan.

pressure• thermodynamically, the normal force exerted by a

homogeneous liquid or gas, per unit of area, on the wall of thecontainer.

prime mover• engine, turbine, water wheel or similar machine that drives an

electric generator.


Fans (15-21) 1/8/01 10:54 AM Page 152

psychrometer• instrument for measuring relative humidities with wet- and dry-

bulb thermometers.

pump• machine for imparting energy to a fluid causing it to do work,

drawing a fluid into itself through an entrance port, and forcingthe fluid out through an exhaust port. Main types are air lift,centrifugal, diaphragm, positive displacement, reciprocating androtary.

Rankine temperature• absolute temperature scale conventionally defined by the

temperature of the triple point of water equal to 491.68˚R, with180 divisions between the melting point of ice and the boilingpoint of water under standard atmospheric pressure (l˚R= 11˚F).

rating standard• standard that sets forth a method of interpreting the results of

tests of individual units, at specified conditions, in relation to aproduct manufactured in quantity.

reactive power• portion of apparent power that does no work. It is measured

commercially in kilovars. Reactive power must be supplied tomost types of magnetic equipment, such as motors. It issupplied by generators or by electrostatic equipment, such ascapacitors.

real power• energy- or work-producing part of apparent power. It is

measured commercially in kilowatts. The product of real powerand length of time is energy, measured by watt-hour metersand expressed in kilowatt-hours (kWh).


Fans (15-21) 1/8/01 10:54 AM Page 153

reheat• application of sensible heat to supply air that has been

previously cooled below the temperature desired formaintaining the temperature of the conditioned space.

relative humidity• ratio of the partial pressure or density of water vapour to the

saturation pressure or density respectively, at the same dry-bulbtemperature, and barometric pressure of the ambient air.

saturation pressure• for a pure substance at a given temperature, the pressure at

which vapour and liquid, or vapour and solids, can exist inequilibrium.

sensor• device or instrument designed to detect and measure a variable.

specification• precise statement of a set of requirements to be satisfied by a

material, product, system or service that indicates theprocedures for determining whether each of the requirements issatisfied.

stall region• performance zone where unstable operation occurs,

characterized by aerodynamic blockage or the breakaway of theflow from certain passages between the blades.

standard air (IP)• dry air at 70˚F and 14.696 psia. Under these conditions, dry air

has a mass density oF 0.075 lb/ft3.

standard air (SI)• dry air at 20˚C and 101.325 kPa absolute. Under these

conditions, dry air has a mass density of 1.204 kg/m3.


Fans (15-21) 1/8/01 10:54 AM Page 154

stratified airflow• layers of air, usually at different temperatures or different

velocities, flowing through a duct or plenum.

stratified fluid flow• form of low velocity, two-phase flow in horizontal pipes, so

that the free surface of the liquid remains level between agaseous and liquid phase above and below it respectively.

system effects• usually conditions in a distribution system that affect fan and

pump performance and related testing, adjusting, and balancingwork. Can also affect the performance of other components(such as filters).

temperature profile• graph representing the distribution of temperatures in a plane

section of a body or a space, or over a period of time.

testing standard• standard that sets forth methods of measuring capacity, or other

aspects of operation, of a specific unit or system of a given classof equipment, together with a specification of instrumentation,procedure and calculations.

thermal transfer fluid• fluid circulated through closed circuits to transfer heat from one

location to another.

thermal watt• heat power expressed in watts.

throttling• 1. of a fluid, an irreversible adiabatic process which consists of

lowering pressure by an expansion without work. 2. reductionin fluid or current flow by adding resistance.


Fans (15-21) 1/8/01 10:54 AM Page 155

ton (of refrigeration)• time-rate of cooling equal to 12,000 Btu/h (approximately

3,517 W).

total pressure• in fluid flow, the sum of static pressure and velocity pressure.

turbine• fluid-energized acceleration machine for generating rotary

mechanical power from the energy in a fluid stream.

turbulent (eddy) flow• fluid flow in which the velocity varies in magnitude and

direction in an irregular manner throughout the mass.

turning vane (air splitter)• curved strip of short radius placed in a sharp bend or elbow in a

duct to direct air around the bend.

two-phase flow• simultaneous flow of two phases of a fluid, usually gas-liquid

flows.

valve• device to regulate or stop the flow of fluid in a pipe or a duct by

throttling.

variable air volume (VAV)• use of varying airflow to control the condition of air, in contrast

to constant flow with varying temperature.

variable flow• throttling control of water during a cooling or heating process.

velocity head• height of fluid corresponding to the kinetic energy per unit

mass of fluid divided by gravitational acceleration.


Fans (15-21) 1/8/01 10:54 AM Page 156

velocity pressure• in a moving fluid, the pressure due to the velocity and density

of the fluid, expressed by the velocity squared times the fluiddensity, divided by two (rv2/2).

velocity profile• graph that represents, in a plane section, the velocity

distribution in a flowing fluid.

vena contracta• smallest cross-sectional area of a fluid stream leaving an orifice.

venturi• contraction in a pipeline or duct that increases the fluid velocity

to lower its static pressure, followed by a gradual expansion toallow recovery of static pressure. Used for metering and otherpurposes that involve change in pressure.

viscosity• 1. property of semifluids, fluids and gases by which they resist

an instantaneous change of shape or arrangements of parts. Itcauses fluid friction whenever adjacent layers of fluid move inrelation to each other. 2. property of a fluid to resist flow orchange of shape.

viscous flow• 1. laminar flow or streamline flow. 2. type of gas flow in which

the average free path of gas molecules is much smaller than thesmallest cross-sectional dimension of the pipe conveying thesubstance.

voltampere (VA)• basic unit of apparent power. The practical unit of apparent

power is kilovolt-ampere (kVA), 1,000 voltamperes.


Fans (15-21) 1/8/01 10:54 AM Page 157

water column (wc)• tubular column located at the steam and water space of a boiler

to which protective devices, such as gauge cocks, water gaugeand level alarms are attached.

water gauge (Wg)• 1. gauge glass with attached fittings that indicates water level

within a vessel. 2. designation that water is the fluid in amanometer.

watt (power) (W)• 1. energy flow at the rate of one joule per second. 2. the work

done or energy generated by one ampere induced by an emf ofone volt. P = EI = I2R.

wet-bulb temperature• temperature indicated by a psychrometer when the bulb of one

thermometer is covered with a water-saturated wick over whichair is caused to flow to reach an equilibrium temperature ofwater evaporating into air, when the heat of vaporization issupplied by the air.

• wet-bulb temperature is lower than dry-bulb temperature. Thisdifference indicates the amount of humidity in the air. If W.B.temp. = D.B. temp., then you have 100% humidity.

• tables of W.B. and D.B. difference are available to show %relative humidity (% of maximum).


Fans (15-21) 1/8/01 10:54 AM Page 158

OTHER IN-HOUSE REFERENCE GUIDES:

• Adjustable Speed Drives• Energy Monitoring & Control Systems• Lighting• Motors• Power Quality• Power Quality Mitigation• Pumps

COMMENTS:

For any changes, additions and/or comments call or write to:

Scott RouseProject ManagerOntario Power Generation700 University Avenue, H15-A6Toronto, OntarioM5G 1X6Telephone (416) 592-8044Fax (416) 592-4841E-Mail [email protected]


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Fans Reference Guide

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