Fans Reference Guide

Fans Reference Guide 4th edition, 2001

First Edition, September 1993 Second Edition, October 1997 Third Edition, August 1999 Fourth 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 Engineer for Technology Services Department, Ontario Hydro, 1993

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

Making Energy Savings Good Business

Printed in Canada Copyright © 1993, 1997, 1999, 2001 Ontario Power Generation

First Edition, September 1993 Second Edition, October 1997 Third Edition, August 1999 Fourth 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 Engineer for Technology Services Department, Ontario Hydro, 1993

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

Making Energy Savings Good Business

Printed in Canada Copyright © 1993, 1997, 1999, 2001 Ontario Power Generation

FANS Reference Guide

3rd Edition, 1999

T A B L E

OF

CONTENTS

INTRODUCTION .......................................................................1 DEFINITIONS ...........................................................................3 Fans ...............................................................................................3 Blowers .........................................................................................3 Velocity Pressure...........................................................................4 Static Pressure ...............................................................................4 Total Pressure................................................................................4 Fan Total Pressure Rise .................................................................4 Fan Velocity Pressure ....................................................................4 Fan Static Pressure ........................................................................5 Fan Duty .......................................................................................5 Fan Output Power ........................................................................5 Fan Efficiency................................................................................5 System Curve................................................................................5 Performance Curve .......................................................................5 Fan Static Efficiency......................................................................5 FAN T YPES ..............................................................................7 Centrifugal Fans............................................................................7 Airfoil.............................................................................................9 Backward-inclined........................................................................11 Radial ..........................................................................................11 Forward-curved ............................................................................12 Axial Fans....................................................................................12 Propeller.......................................................................................16 Tubeaxial.....................................................................................16 Vanaxial ......................................................................................20 Special designs ............................................................................21 i

T A B L E

OF

CONTENTS

B ifurcated Fans.............................................................................21 Centrifugal Inline Fans .................................................................21 Centrifugal Roof Exhausters ..........................................................25 Utility Fans ..................................................................................25 Fan Designation and Arrangements ...........................................25 Class Limits for Fans...................................................................31

PRINCIPLES OF OPERATION .....................................................33 Centrifugal Fans ..........................................................................34 Axial Fans....................................................................................36 FAN PERFORMANCE CURVES ....................................................39 FAN LAWS .............................................................................43 Limitations ..................................................................................43 Compressibility Factor................................................................44 FAN FORMULAE .....................................................................47 Density........................................................................................47 Fan Flow Rate .............................................................................48 Head and Pressure ......................................................................49 Velocity Pressure.........................................................................49 Total Pressure..............................................................................50 Fan-System-Effect Factor ............................................................50 Fan Power and Efficiency ...........................................................50 Fan Motor Power ........................................................................52 Example 1....................................................................................53 AIR S YSTEMS.........................................................................59 Example 2....................................................................................60 FAN AND S YSTEM INTERFACE ..................................................65 System Effect Factors..................................................................65 Fan Outlet Conditions ................................................................66 ii

T A B L E

OF

CONTENTS

Fan Inlet Conditions ...................................................................68 FAN SELECTION .....................................................................69 Pressure Definitions ....................................................................70 Parallel Fan Selection ..................................................................72 Series Vs Parallel Operation........................................................74 FAN NOISE ............................................................................77 Fan Sound Power ........................................................................78 Example 3....................................................................................81 FAN DUTY CONTROL .............................................................85 V IBRATION ISOLATION............................................................89 ELECTRIC MOTOR FAN DRIVE .................................................91 Flywheel Effect............................................................................91 AC Motors ..................................................................................92 DC Motors..................................................................................93 ENERGY CONSUMPTION ANALYSIS ...........................................95 Constant-Volume Fans ...............................................................96 Variable-Volume Fans .................................................................97 Example 4..................................................................................102 APPENDICES ........................................................................109 Appendix A – Density Calculations.........................................109 Appendix B – Drive Loss Calculations.....................................115 Appendix C – Fan Outlet Loss Coefficients.............................119 CONVERSION TABLES ...........................................................127 ABBREVIATIONS AND S YMBOLS ..............................................131 BIBLIOGRAPHY .....................................................................135 GLOSSARY ...........................................................................139 iii

LIST

OF

FI GU RES

1. Operating Point .........................................................................6 2. General Configuration and Component Terms for Centrifugal Fans ...................................................................8 3. Airfoil .........................................................................................9 4. Typical Characteristics of Airfoil Fans.....................................10 5. Typical Characteristics of Backward-inclined Fans.................13 6. Typical Characteristics of Radial Fans .....................................14 7. General Configuration and Component Terms for Axial Fans ................................................................15 8. Typical Characteristics of Propeller Fans.................................17 9. Typical Characteristics of Tubeaxial Fans ...............................18 10. Typical Characteristics of Vaneaxial Fans................................19 11. Configuration of Bifurcated Fans ............................................22 12. Typical Characteristics of Centrifugal Inline Fans...................23 13. Typical Characteristics of Centrifugal Roof Exhausters..........24 14. Typical Characteristics of Forward-curved Utility Fans..........26 15. Typical Characteristics of Backward-inclined Utility Fans ...............................................27 16. Drive Arrangements for Axial Fans with or without Diffuser and Outlet Box ...............................28 17. Drive Arrangements for Centrifugal Fans ...............................29 18. Drive Arrangements for Centrifugal Fans ...............................30 19. Outlet Velocity Vector Diagram for Backward-inclined Blades........................................................35 20. Outlet Velocity Vector Diagram for Radial Blades..................35 21. Outlet Velocity Vector Diagram for Forward-curved Blades............................................................35 iv

LIST

FI GU RES

OF

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 ....................................................................41 25. Compressibility Factor ............................................................45 26. Operating Point and System Curve.........................................53 27. 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 ..................................66 30. Fan-outlet Velocity Profiles......................................................67 31. Design Operating Point Selection Range on a Typical Centrifugal Fan Performance Curve ........................................72 32. Pressure Flow Curves ..............................................................73 33. Series Fan Operation................................................................75 34. Outlet Damper Fan Control ....................................................86 35. Throttle Control of a Fan with a Two-speed Motor...............87 36. Inlet Vane Control of a Fan......................................................88

v

LIST

OF

FI GU RES

C1. Plane Asymmetric Diffuser at Fan Outlet Without Ductwork ................................................................119 C2. Pyramidal Diffuser at Fan Outlet Without Ductwork ..........120 C3. Plane Symmetric Diffuser at Fan Outlet With Ductwork.....121 C4. Plane Asymmetric Diffuser at Fan Outlet With Ductwork...122 C5. Plane Asymmetric Diffuser at Fan Outlet With Ductwork...123 C6. Plane Asymmetric Diffuser at Fan Outlet With Ductwork...124 C7. Pyramidal Diffuser at Fan Outlet With Ductwork................125

vi

LIST

OF

TA B L E S

1. Fan Laws ..................................................................................46 2. Typical Manufacturer's Performance Data for a 24-in. AFSW Centrifugal Fan at 70˚F and Standard Atmospheric Pressure...............................................54 3. Typical Number of Fan Blades ................................................79 4. Specific Sound Power Levels and Blade Frequency Increments .............................................................80 5. Sound Correction Factors........................................................81 6. Summary for Example 3 ..........................................................83 7. Typical VAV-fan Constants....................................................104 8. Motor Load Efficiencies.........................................................105 9. The Solution to Example 4 ....................................................105 10. Summary of Example 4 .........................................................107

vii

LIST

OF

TA B L E S

A1. Standard Atmospheric Data for Altitudes to 3,000 m .............................................................114 A2. Density Calculations .............................................................114 C1. Plane Asymmetric Diffuser at Fan Outlet Without Ductwork.....................................................119 C2. Pyramidal Diffuser at Fan Outlet Without Ductwork ................................................................120 C3. Plane Symmetric Diffuser at Fan Outlet With Ductwork ..........................................................121 C4. Plane Asymmetric Diffuser at Fan Outlet With Ductwork ..........................................................122 C5. Plane Asymmetric Diffuser at Fan Outlet With Ductwork ..........................................................123 C6. Plane Asymmetric Diffuser at Fan Outlet With Ductwork ..........................................................124 C7. Pyramidal Diffuser at Fan Outlet With Ductwork................125

viii

C H A P T E R 1

INTRODUCTION • Fans and blowers are the largest single type of user of electricity in industry. Applications in all industries include: boiler combustion air supply, dust and exhaust removal (pneumatic conveying), “bag” house, sewage aeration, drying, cooling industrial processes, and ventilation. Issues such as indoor air quality and pollution control create a continuous demand for well-designed, efficient and cost-effective ventilation and blower systems. • Selecting the right size and type of fan and blower is fundamental to an energy-efficient system. • The first step in any fan application is defining the needs of the system being supplied. • Enhancing the performance of an existing air system with a new, energy-efficient electronic control system offers significant potential for energy savings. In some cases, retrofitting with a more efficient fan or blower and interconnecting ductwork will be the most appropriate way to reduce energy consumption.

Chapter 1: Introduction

1

• This guide contains the information required to select an industrial or commercial fan and blower system. Supporting handbooks and reference material are identified in the Bibliography. • Chapter 7 provides the formulae necessary to determine the energy consumption of a heating and ventilating fan system, particularly variable-volume fans. • Because the personal computer is a popular design tool, the formulae in this guide were designed to be used in a spreadsheet program. Hourly analysis programs determine energy use more accurately, and some of these programs can be used for system design and selection. • This guide demonstrates how to use the American Society of Heating, Refrigerating and Air Conditioning Engineers (ASHRAE) Modified Bin Method on a spread-sheet program to determine annual energy consumption when fan power is a function of outdoor air temperature. This procedure is reasonably accurate relative to the time required to perform the analysis. •

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

2

Fans Reference Guide

C H A P T E R 2

DEFINITIONS F A N

• Device that causes flow of a gaseous fluid by creating a pressure difference by exchanging momentum from the fan blades to air/gas particles. • The fan impeller converts rotational mechanical energy into both static and kinetic energy within the gaseous fluid. • The proportion of static versus kinetic energy created and the inherent energy conversion efficiency depends on the type of fan (blade design). • The gaseous fluid transported by a fan is most often air and/or toxic fumes, whereas blowers may transport a mixture of particulate and air. B LOWER

• Similar to a fan, except it can produce a much higher static pressure. Sometimes higher pressure is achieved by a multistage impeller arrangement. • Engineering practice distinguishes fans and blowers for low pressure and centrifugal compressors for high pressure. Chapter 2: Definitions

3

• The demarcation between blowers and compressors is set at a 7% increase in the density of the air from blower inlet to blower outlet. • Fan and blower definitions and formulae, assuming incompressibility, apply below this demarcation with insignificant errors. V ELOCITY P RESSURE

• That pressure at a point in an airstream existing by virtue of the air density and its rate of motion. S TATIC P RESSURE

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

• That pressure at a point in an airstream existing by virtue of the air density and the degree of compression and rate of motion of the air; hence it is the sum of the static and velocity pressure (also called stagnation pressure). F A N T OTAL P RESSURE R ISE

• The fan total pressure at outlet, minus the fan total pressure at inlet. Note: when moving air enters a closed area, it converts velocity pressure to static pressure. F A N V ELOCITY P RESSURE

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

4


F A N S TATIC P RESSURE

• The fan total pressure rise diminished by the fan velocity pressure. F A N D U T Y

• The range of operating points, giving the fan inlet volume flow at a rated fan pressure. F A N O UTPUT P OWER

• The fan output power or the useful power, delivered by a fan to an incompressible fluid, is equal to the product of the fan flow rate and the fan total pressure divided by a constant depending on the units. F A N E FFICIENCY

• The fan total or mechanical efficiency is defined as the ratio of fan air power to fan-shaft input power. S YSTEM C URVE

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

• The set of operating points defined by a particular fan design, size, and speed. Where the system and performance curves meet is the fan’s operating point. F A N S TATIC E FFICIENCY

• This is not a true efficiency but has been used traditionally in the fan industry. It is equal to the fan total efficiency times the ratio of fan static to fan total pressures Chapter 2: Definitions

5

Fan Total Pressure or Fan Static Pressure

Fan Performance Curve System Curve

P

Operating Point

Volume Flow V

Figure 1: Operating Point

6 Fans Reference Guide

C H A P T E R 3

FAN T YPES • The two general classifications of fans – centrifugal and axial – are established according to the direction of flow through the impeller. - Axial fans have high volume capability for large duct size ventilation applications. - Centrifugal fans have high pressure capability for applications such as boilers, baghouses, conveyors, and sewage aerators. • These general classifications are subdivided into groups with inherent performance characteristics to suit a specific application. • All other fans fall under a special design classification, including mixed-flow fans. C ENTRIFUGAL F ANS

• Centrifugal fans are divided into four main subclassifcations according to impeller type: airfoil, backward-inclined, radial and forward-curved.

Chapter 3: Fan Types

7

Figure 2: General Configuration and Component Terms for Centrifugal Fans Reprinted with permission from the Air Movement and Control Association from Publication 201–90

8


Airfoil

• The most efficient centrifugal fan design, but the most expensive. • Airfoil (AF) have an impeller with typically 10 to 16 blades of airfoil contour (see Figure 2), curved away from the direction of rotation. • Air leaves the impeller wheel at a velocity of less than its tip speed, and relatively deep blades allow for efficient air expansion 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 efficiently convert velocity pressure to static pressure. • To achieve high static-pressure efficiency, a close tolerance between the wheel and the housing inlet cone must be maintained. • Due to the high operating speed, the airfoil blades and the close tolerances, an AF fan is the most expensive to construct and repair. • It is the most efficient centrifugal fan design at approximately 90%.

Rotation

Figure 3: Airfoil Chapter 3: Fan Types

9

24 in. ) p h ( r e w o p e s r o h / ) p s ( e r u s s e r p c i t a t s

AFSW Centrifugal 1,170 rpm 5

100 surge

4

80

3

60 hp

2

40 se

1

) e s ( y c n e i c i f f e c i t a t s

20

sp 0

0 0

2

4

6

8

10

12

cfm x 1,000

Figure 4: Typical Characteristics of Airfoil Fans

10


APPLICATION

• Usually commercial heating, ventilating and air-conditioning (HVAC) systems and clean-air industrial applications where the power savings can be significant. • Best suited to applications that require low-to-medium static pressure and a large flow volume. Backward-inclined

• Backward-inclined (BI) or backward curved fans have an impeller with typically 10 to 16 blades of uniform thickness incllined 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 approximately 80%. APPLICATION

• In systems that require low-to-high static pressure, specifically in commercial HVAC systems with moderate flow volume. Also used in industrial systems that require some tolerance to a corrosive or erosive environment. They are being used increasingly in industrial process ventilation with wear liners. Radial Fans

• Radial (R) fans have an impeller wheel of high mechanical strength with typically six to 10 blades of heavy gauge material radiating out from the hub. • The blades can be either straight radial or modified radial with a slight curve. They are often equipped with removeable wear plates to extend the useful life of the fan impeller. • For a given duty, R fans operate at medium speed. Chapter 3: Fan Types

11

• They are the least efficient fan at 50% to 60%, but they do not clog and are easily repaired. APPLICATION

• Primarily in industrial systems in a corrosive or erosive environment, such as material handling of airborne particulate or where high static pressure is required. Forward-curved

• Forward-curved (FC) fans have an impeller wheel made of light gauge material, with typically 24 to 64 shallow blades with both the heel and the tip curved forward. • Air leaves the blade at a velocity greater than the tip speed, and primarily kinetic energy is transferred to the air. • These fans are the smallest of the centrifugal type and, for a given duty, rotate at the slowest speed. • The fan housing is a scroll design similar to the other centrifugal fan housings, except the tolerance between the inlet cone and the wheel is not as critical allowing lighter gauge material to be used. • FC fans are less efficient than AF and BI fans at approximately 70%. APPLICATION

• Generally in packaged and built-up, commercial and residential HVAC systems with low-to-medium static pressures and low air volumes. (See Figure 14, p. 26.) A XIAL F ANS

• Divided into three subclassifications according to impeller type: propeller, tubeaxial and vaneaxial. 12



BISW Centrifugal 1,170 rpm 5

100 surge

4

80 hp

3

60

2

40

1


20

se sp

0

0 0

2

4

6

8

10

12

cfm x 1,000

Figure 5: Typical Characteristics of Backward-inclined Fans


13


Radial 1,170 rpm hp

10

100

8

80

6

60 sp

4

40

se 2


20 surge

0 0

2

0 4

6

8

10

12

cfm x 1,000

Figure 6: Typical Characteristics of Radial Fans

14


Figure 7: General Configuration and Component Terms for Axial Fans Reprinted with permission from the Air Movement and Control Association from Publication 201–90.


15

Propeller

• Have an impeller with two or more BI blades that are usually made from single, light-gauge material attached to a small diameter hub. • Because primarily kinetic energy is transferred to the air with little static energy, these fans are limited to low-pressure applications. • The efficiency of these fans is low. • The fan housing can be a simple ring or circular guard, an orifice plate, or an inlet cone with close tolerance to the blade tips to create a venturi for optimum performance. APPLICATION

• Low static-pressure, high volume, commercial and industrial systems. Tubeaxial

• Have an impeller with typically four to eight blades attached to a 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 is static energy, these fans can be used in applications where there is 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 the impeller blade tips; this results in higher performance than propeller fans.



Propeller 870 rpm 5

100

4

80 surge

3

60

2

40 se

1


20 hp sp

0 0

0 2

4

6

8

10

12

cfm x 1,000

Figure 8: Typical Characteristics of Propeller Fans Chapter 3: Fan Types

17


Tubeaxial 1,770 rpm 5

100

4

80

3

60 hp

2

40 surge se

1


20

sp 0

0 0

2

4

6

8

10

12

cfm x 1,000

Figure 9: Typical Characteristics of Tubeaxial Fans

18


Figure 10: Typical Characteristics of Vaneaxial Fans


19

APPLICATION

• Low and medium static-pressure, commercial, ducted systems where the axial arrangement saves space and the downstream flow pattern is not critical. • Industrial systems where airborne contaminants collect on the impeller blades and require periodic cleaning. Vaneaxial

• Usually have short AF blades radiating from a hub greater than half the diameter of the impeller. • The blades are either fixed, adjustable or controllable (variable pitch-in-motion). • The discharge from the impeller has a rotative component, unless inlet guide vanes are used. • Because stationary diffusion vanes downstream of the impeller convert rotary energy produced by the blades into static pressure (as in an axial blower or compressor), primarily static energy is transferred to the air. • Vaneaxial fans are the most efficient axial fan. • The housing is a cylindrical tube with a close tolerance to the impeller blade tips. • The housing may include a set of inlet guide vanes and/or downstream diffusion vanes equal in number to the impeller blades and preferably of the AF type. APPLICATION

• Low to high static-pressure, commercial HVAC systems, and industrial ventilation systems where the axial arrangement saves space, and the downstream flow patterns and efficiency are important. 20


S PECIAL D ESIGNS

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 driving motor to be removed from the airstream while maintaining a direct-drive arrangement. • In corrosive environments, the casing may be plastic or coated. • The mating flanges at each end of the casing are identical, but the casing diameter is increased in barrel fashion to allow passage of a similar cross section of air, concentric with the motor enclosure. APPLICATION

• Generally used to extract sticky, corrosive or volatile fumes in industrial applications where it is critical to protect the motor from the airstream. Centrifugal Inline Fans

• Have a direct-drive or a belt-driven AF or BI impeller mounted perpendicular to a rectangular or tubular casing with ample clearance around the blade tips. • The air discharged radially from the blade tips must turn through 90 degrees to pass through the fan exit, which is in line with the impeller inlet. APPLICATION

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


21

Figure 11: Configuration of Bifurcated Fans

22



Centrifugal Inline 870 rpm 5

100

4

80 surge

3

60 hp

2 1

40

y c n e i c i f f e c i t a t s

20

se

sp

) e s (

0

0 0

2

4

6

8

10

12

cfm x 1,000

Figure 12: Typical Characteristics of Centrifugal Inline Fans


23


Centrifugal Roof Exhauster 870 rpm 5

100

surge

4

80

3

60

2

40 hp se

1

0

2

4

6

y c n e i c i f f e c i t a t s

20

sp

0

) e s (

0 8

10

12

cfm x 1,000

Figure 13: Typical Characteristics of Centrifugal Roof Exhausters

24


Centrifugal Roof Exhausters

• Have a direct-drive or a belt-driven AF or BI impeller mounted in a multicomponent housing comprising of a curb cap with an integral inlet venturi, a fan shroud with drive-mounting support, and a weatherproof motor hood. • The impeller has an inlet cone that allows mixed flow through the impeller-blade passages, and air exits radially from the blade tips through a concentric discharge passage. • The fan shroud redirects the air - either discharging it down or blasting it up. APPLICATION

• The down-discharge configuration is used for exhausting relatively clean air, while the up-blast configuration is used for hot and/or contaminated air. Utility Fans

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

• Commercial and industrial ventilation applications requiring low-to-medium air volumes and pressures. F A N D ESIGNATION

AND

A RRANGEMENTS

• The Air Moving and Conditioning Association, Inc (AMCA) has devised standard designations for fan rotation, discharge orientation, motor position for belt or chain drive, inlet box position, and drive arrangements for both centrifugal and axial fans.


25


FCSW Utility 500 rpm 5

100 hp

surge

4

80

3

60

2

40 se

1


20

sp 0

0 0

2

4

6

8

10

12

cfm x 1,000

Figure 14: Typical Characteristics of Forward-curved Utility Fans



BISW Utility 1,170 rpm 5

100

4

80

surge

se

3

60 hp

2

40 sp

1

20

0

0 0

2

4

6

8

10


12

cfm x 1,000

Figure 15: Typical Characteristics of Backward-inclined Utility Fans


27

Optional on all arrangements Inlet box

Diffuser

Arr. 1 two-stage

Arr. 1

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

Arr. 3

Arr. 4

For belt drive or direct connection. Impeller between bearings that are on internal supports. Drive through inlet.

Arr. 7

Arr. 4 two-stage

For direct connection. Impeller Impeller overhung on motor shaft. No bearings on fan. Motor on internal supports.

Arr. 8 one-or two-stage

For belt drive or direct connection. Arr. 3 plus common base for prime mover.

Arr. 9 motor on casing

For belt drive or direct connection. Arr. 1 plus common base for prime mover.

Arr. 9 motor on integral base

For 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

Figure 16: Drive Arrangements for Axial Fans with or without Diffuser and Outlet Box Reprinted with permission from the Air Movement and Control Association from Publication 201–90.

28


SW- Single Width SI- Single Inlet DW- Double Width DI- Double Inlet Arrangements 1,3,7 and 8 are also available with bearings mounted on pedestals or base set independant of the fan housing Arr. 1 SWSI For belt drive or direct connection impeller overhung. Two bearings on base.

Arr. 2 SWSI For belt drive or direct connection. impeller overhung. Bearings in bracket supported by fan housing

Arr. 3 SWSI For belt drive or direct connection. One bearing on each side and supported by fan housing.

Arr. 3 DWDI For belt drive or direct connection. One bearing on each side and supported by fan housing.

Arr. 4 SWSI For direct drive. Impeller overhung on prime mover shaft. No bearings on fan. Prime mover base mounted or integrally directly connected.

Arr. 7 SWSI For belt drive or direct connection. Arrangement 3 plus base for prime mover.

Arr. 7 DWDI For belt drive or direct connection. Arrangement 3 plus base for prime mover.

Arr. 8 SWSI For belt drive or direct connection. Arrangement 1 plus extended base for prime mover.

Arr. 9 SWSI For belt drive. Impeller overhung, two bearings, with prime mover outside base.

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

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


29

SW- Single Width

SI- Single Inlet

DW- Double Width

DI- Double Inlet

Arr. 1 SWSI with Inlet Box For belt drive or direct connection. Impeller overhung, two bearings on base. Inlet box may be self-supporting

Arr. 3 SWSI with Independent Pedestal For belt drive or direct, connection fan. Housing is self-supporting. One bearing on each side supported by independant pedestals.

Arr. 3 SWSI with Inlet Box and Independent Pedestals For belt drive or direct connection fan. Housing is self-supporting. One bearing on each side supported by independent pedestals with shaft extending through inlet box.

Arr. 3 DWDI with Independent Pedestal For belt drive or direct connection fan. Housing is self-supporting. One bearing on each side supported by independant pedestals.

Arr. 3 DWDI with Inlet Box and Independent Pedestals For belt drive or direct connection fan. Housing is self-supporting. One bearing on each side supported by independent pedestals with shaft extending through inlet box.

Arr. 3 SWSI with Inlet Box For belt drive or direct connection. Impeller overhung, two bearings on base plus exended base for prime mover. Inlet box may be self-supporting.

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

30


• Fan rotation is determined to be clockwise or counterclockwise by viewing the fan from the drive side. • The choice of fan arrangement depends on the application – the environment of the airstream being handled and the size of the fans are primary considerations. • The discharge position and the drive arrangement must be determined to fit the fan system properly. • Manufacturers identify the arrangements available for the fans in their product line. C LASS L I M I TS

FOR

F ANS

• AMCA has adopted a standard that defines the operating limits for various classes of centrifugal fans used in general ventilation applications. • The standard uses limits based on mean brake horsepower per square foot of outlet area , expressed in terms of outlet velocity and static pressure. • There are three class limits for centrifugal fans - Class I is the lightest duty and Class III is the heaviest duty. • When selecting a fan, it is important to ensure the duty point does not exceed the performance range for the fan class.


31

C H A P T E R 4

PRINCIPLES OF OPERATION • All fans produce total pressure, which represents the static and kinetic energy imparted to the air by the impeller. • The rotating blades of the fan impeller convert mechanical energy into static and kinetic energy by changing the velocity vector of the incoming air. • Centrifugal fans produce total pressure from the centrifugal force of the air radiating out between the blade passages and by the kinetic energy imparted to the air by virtue of its velocity leaving the impeller. • The absolute velocity vector in the case of centrifugal fans is the sum of the tangential and radial velocity components. • Axial fans produce total pressure from the change in velocity passing through the impeller, with none being produced by centrifugal force. • The absolute velocity vector in the case of axial fans is the sum of the axial and tangential velocity components.

Chapter 4: Principles of Operation

33

C ENTRIFUGAL F ANS

• The operation of centrifugal fans can best be described by velocity vector diagrams. • The height of the diagram – indicated by the relative radial velocity vector V r – is based on the volume of air flowing through the fan. • The air velocity relative to the blade – indicated by V b is nearly tangential to the blade as some slip occurs due to boundary layer effects. • The tip speed component wr is perpendicular to the wheel radius, where w is the rotational speed of the impeller in radians per 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, the vector wr is constant. • The absolute velocity indicated by V s is the resultant of V b and wr. • The relative tangential velocity vector indicated by V t is projected from V s in the direction of wr. • If volume decreases, the vector V r decreases and as the vector V b does not change for a given blade, V t 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 V t and wr, the pressure characteristic rises as volume decreases for the BI blade, is constant for the R blade and decreases for the FC blade. These vector diagrams illustrate that, at a given speed, the smallest fan selection will be a forward curved fan. Conversely, the largest will be an airfoil. 34


Figure 19: Outlet Velocity Vector Diagram for Backward-inclined Blades

Figure 20: Outlet Velocity Vector Diagram for Radial Blades

Figure 21: Outlet Velocity Vector Diagram for Forward-curved Blades


35

A XIAL F ANS

• The principle of operation can be described by the use of a velocity vector diagram. • Velocity diagrams for axial fans are drawn for a uniform axial velocity indicated by V a. The axial velocity remains nearly constant from blade root to tip. • The tip speed component wr is perpendicular to the axis and is shown as the blade section under consideration.

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

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


• The air velocity relative to the blade indicated by V b is nearly tangential to the blade as some slip occurs due to boundary layer effects. • The relative tangential velocity vector indicated by V t is projected from V a in the opposite direction of wr. • The mean relative velocity drawn to bisect V t is shown as V m. This is used in aerodynamic theory to calculate the circulation around the airfoil.


37

C H A P T E R 5

FAN PERFORMANCE CURVES • The manufacturer guarantees fan performance according to standard air conditions. When selecting a fan, it is necessary to know the actual air inlet conditions (temperature, pressure, density), and use the Fan Laws to correct the published performance to actual conditions. • Fan performance curves are developed from data obtained from tests executed in accordance with AMCA and ASHRAE standards. • The most common procedure to develop a performance curve is to test the fan from shut-off conditions to nearly-free delivery conditions. • A fan is generally tested in a set-up that closely simulates how it will be installed in an air-moving system. • Propeller fans are normally tested in the wall of a chamber, and power roof exhausters are tested mounted on a curb to exhaust vertically from a chamber. • Centrifugal, tubeaxial, and vaneaxial fans are usually tested with an outlet duct with provision for restricting the flow at the discharge. Chapter 5: Fan Performance Curves

39

• A static- and velocity-pressure measuring station is located within the duct downstream of flow straighteners. • At shut-off the duct is completely blanked off, and at freedelivery the duct outlet is wide open; test data is recorded while maintaining constant fan speed and air density. • Under these conditions and at the same fan speed, the flow is graduated to obtain sufficient data to define a corresponding performance curve. • For each test point, the pressures are measured and the corresponding flow rate is determined. The measured pressures are corrected back to fan inlet conditions. • Fan performance curves are plotted with the inlet flow rate (in cubic foot per minute or litres per second) on the abscissa. Total pressure, static pressure, fan horsepower and fan efficiency are plotted on the ordinate axis. • It is not practical to test a fan at every speed at which it can operate or at every inlet density it may encounter. • By using a series of equations referred to as the Fan Laws, it is possible to accurately predict the fan's performance at other speeds and densities. • Manufacturers usually publish fan performance curves at a density of 0.075 Ib/ft 3 and an inlet temperature of 70˚F.

40


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

Chapter 5: Fan Performance Curves

41

C H A P T E R 6

FAN LAWS • The Fan Laws relate the performance variables for any dynamically similar series of fans at the same point of rating on the 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 or density on volume flow, pressure and power level. • Fan Law No. 2 governs the effect of changing size, pressure or density on volume flow rate, speed and power. • Fan Law No. 3 governs the effect of changing size, volume flow or density on speed, pressure and power. L IMITATIONS

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

Chapter 6: Fan Laws

43

• These Fan Laws do not include correction for compressible flow. C OMPRESSIBILITY F ACTOR

• As air travels through a fan, it is compressed and the outlet volume will be less than at the inlet. The fan laws as presented in this chapter do not account for this effect. • A fan selected without using compressibility will be larger in size than required and the fan input power will be understated. • The compressibility effect is quite small when fan pressure rise is below 10” Wg., and is customarily ignored below this threshold. • For applications where the fan pressure rise is more than 10î Wg., 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, K p from the chart for the required fan static pressure rise. 3. For fan selection only, multiply the required pressure and flow by the compressibility factor, K p. The fan input power obtained using the fan laws for selection must be divided by K p. 4. If the actual efficiency is more than 5% different than what was estimated in step #1, return to step #1 using the new efficiency. 5. When using equations 7 and 9 in chapter 7, multiply the resulting power by K p.

44


Approximate Kp 1% 1.000

.990

.980

Kp .970

0.50

.960

0.55 0.60 0.65 0.70

.950

.940 0

10

20

30

40

50

60

70

Fan Static Pressure Rise - inches W.G.

Figure 25: Compressibility Factor


45

Table 1: Fan Laws

D E S

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)

U M O D L E

S

Source: ASHRAE Handbook, 1988 Equipment Volume.


C H A P T E R 7

FAN FORMULAE • The following fan formulae require certain constants and parameters specific to each application. The constants are given in this guide or in the referenced material. • The formulae in this section are valid for incompressible flow D ENSITY

• Fan performance data, unless otherwise identified, is based on dry air at the standard atmospheric pressure of 14.7 psi., 29.921 in.Hg (101.325 kPa) and a temperature of 68˚F (20˚C). The air density 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 and pressures other than standard conditions. Therefore, the air density must be corrected to obtain the actual fan performance. • For fans processing moist air, the moisture content of an airstream is determined by measuring the wet-bulb temperature, the dew-point temperature, or relative humidity.

Chapter 7: Fan Formulae

47

• Wet-bulb and dry-bulb temperatures are most often determined at fan inlet conditions, using a sling thermometer.When the airstream exceeds 180˚F (82˚C), the dew-point temperature is more reliable to determine moisture content. Density, when the dry-bulb temperature falls between 42˚F and l00˚F (5˚C and 38˚C), may be determined by using the psychrometric density chart in AMCA Publication 203-90, Field Performance Measurement of Fan Systems, Appendix N. The numerical method in Chapter 16, Appendix A of this guide may be used with a scientific calculator or PC spread-sheet program. • EQUATION 1: When the gas density (P) at one plane is determined, the density at any point in the fan system may be determined. “Boyle’s Law”: P = constant or px = p1 T r T x rx T 1 r1 where rx = density at plane x, lb./ft.3 (kg/m3) r1 = density at plane 1, lb./ft.3 (kg/m3) T x = absolute temperature at plane x, ˚Fabs (K) T I = 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) F AN F LOW R ATE

• EQUATION 2: The flow rate at a reference plane. The fan flow rate is the primary performance parameter. Q = VA where 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)

48


H EAD

AND

P RESSURE

• The common unit is inches of water, “water gauge”. • Head is the height of a fluid column of water supported by gas flow, while pressure is the normal force per unit area. With gas or air, it is convention to measure pressure on a column of liquid, as pressure measured in terms of unit area is not practical. • The term (V2 /2g) refers to velocity head, and the term ( r V 2 /2gc) refers to velocity pressure. • Velocity head is independent of fluid density. V ELOCITY P RESSURE

• EQUATION 3: Velocity pressure is not independent of density. pv = r(V/cf)2 where pv = 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 where V = 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 standard density. Source: Jorgensen, R. (ed.) Fan Engineering 8th ed. Buffalo: NY, Buffalo Forge Company, 1983. (Ref. A). The SI density derived directly from the imperial equivalent would be a value of 1.2014 kg/m3. Source: Metric Conversion Handbook for Mechanical Engineers in the Building Industry 2nd ed. Public Works Canada, 1983. (Ref. B)


49

T OTAL P RESSURE

• EQUATION 5: The sum of the static pressure and the velocity pressure is total pressure. Pt = Ps + Pv where pt = total pressure, in.Wg (Pa) ps = static pressure, in.Wg (Pa) pv = velocity pressure, in.Wg (Pa) F AN -S YSTEM -E FFECT F ACTOR ( AT

INLET )

• EQUATlON 6: The fan-system-effect pressure drop. SEF = Co r (V o /cf)2 where SEF = fan-system-effect pressure loss, in.Wg (Pa) Co = fan-system-effect loss coefficient, dimensionless r = density, lbm /ft.3 (kg/m3) cf = conversion factor, 1097 (1.414) and where for centrifugal fans: V o = inlet velocity based on area at the inlet collar, or outlet velocity based on outlet area, fpm (m/s) for axial fans: V o = inlet or outlet velocity based on area calculated from fan diameter, fpm (m/s) F AN P OWER

AND

E FFICIENCY

• EQUATION 7: The air horsepower, or fan output power, PFo, is determined from product of the flow and total pressure rise. PFo = Qpt cf where: PFo = output power, hp (W) Q = flow, ft.3 /min. (L/s) 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 to the fan shaft. 2

The conversion factor 6349.6 was derived from converting the metric form of the equation to the imperial equivalent utilizing the conversion factor in Ref. B (see p.47). When the formula 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 would be 1.0.

• EQUATION 8: The fan mechanical or total efficiency is the ratio of the output power to the input power. ht = PFo /PFi

where ht = mechanical (total) efficiency, dimensionless PFo = output power, hp (W) PFi = input power, hp (W) • EQUATION 9: The fan input power. PFi = Q pt ht cf where PFi = input power, hp (W) Q = flow, ft 3 /min. (L/s) pt = total pressure rise, in.Wg (pa) ht = total efficiency, dimensionless ratio cf = conversion factor, 6349.6 ( 1,000) • EQUATION 10: The fan static efficiency is the product of mechanical efficiency and the ratio of static pressure to total efficiency. hs = (ps /pt ) ht

where hs = static effciency, dimensionless ratio Chapter 7: Fan Formulae

51

ht = total efficiency, dimensionless ratio

ps = static pressure, in.Wg (pa) pt = total pressure, in. Wg (pa) • EQUATION 11: By substitution, the fan input power is also: PFi = Q ps hs cf where PFi = input power, hp (W) Q = flow, ft.3 /min. (L/s) hs = static efficiency, dimensionless ratio ps = static pressure, in.Wg (Pa) cf = conversion factor, 6349.6 ( 1,000) F AN M OTOR P OWER

• EQUATION 12 AND EQUATION 13: The fan motor output power. PMo = Q pt ht hD cf or PMo = PFi hD

where PFi = input power, hp (W) PMo = motor output power, hp (W) Q = flow, ft.3 /min. (L/s) ht = fan total efficiency, dimensionless ratio hD = fan drive efficiency, dimensionless ratio cf = conversion factor, 6349.6 (1,000)

52


• EQUATION 14: The fan motor input power. PMi = PMo hM cf where PMi = motor input power, kW PMo = motor output power, hp (W) hM = motor efficiency, dimensionless ratio cf = conversion factor, 1.3410 ( 1,000) E XAMPLE 1

• A large cafeteria at a manufacturing plant requires an exhaust fan to ensure proper indoor air quality for the patrons. Positive exhaust is provided by a belt-driven, 24-in., airfoil, single-width (AFSW) centrifugal fan with an inlet and outlet area of 4.11 sq.ft. The fan is to be equipped with a two-speed 1,800/1,200rpm motor and, on the high-speed setting, is required to deliver 10,000 cfm of air at 70˚F at a static pressure of 2.5 in.Wg. This is identified as operating point A (see Figure 25). Standard atmospheric conditions are assumed for this example.

Figure 26. Operating Point and System Curve Chapter 7: Fan Formulae

53

Example 1.1 • Using formulae and manufacturers' catalogue data, determine speed, the high-speed fan horsepower, and the corresponding fan 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 Fan Static Pressure r i a d t s , m f c

y t t i c e l t o l u e o v

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

• The velocity pressure of the fan must be established to determine 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

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 catalogue data 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.


55

Figure 27: Fan Static-pressure Design Curve at 1,475 rpm Intersecting Design Point A and Fan Static-pressure Curve at 983 rpm Intersecting Point B

Example 1.2 • Using fan laws and assuming constant density, determine the low-speed flow and total pressure to give operating point B (see Figure 26). Using formulae and the manufacturers’ data in Table 2, p. 52, determine the corresponding fan power speed, and total and static efficiencies. • EQUATION 17: Air volume delivered by the fan on the low speed 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


• EQUATION 19: The corresponding static pressure delivered by the 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 Table 2, 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% Chapter 7: Fan Formulae

57

C H A P T E R 8

AIR S YSTEMS • A fan provides the energy to overcome the resistance to flow through air-system components. A fan's performance is interdependent with the system elements. • Components that contribute to system resistance include straight ductwork, elbows, fittings, filters, humidifier distributors, heat-transfer coils, dampers, acoustic silencers, bird screens, registers, grilles and diffusers. • Most air systems operate in the turbulent-flow regime rather than laminar-flow conditions. • Pressure losses in system elements are therefore mainly related to turbulence and flow separation, the kinetic energy being dissipated by viscous shear in the air. • The pressure loss of each of the air system's elements may be calculated with manufacturers' data and the procedures in the ASHRAE Handbook, 1989 Fundamentals Volume . • A given rate of airflow through a system requires a specific total pressure generated by the system fan.

Chapter 8: Air Systems

59

• EQUATION 20: If the flow rate is changed, the resulting total pressure required for turbulent-flow systems. (Æp2 /Æp1) = (Q2 /Q1)2 • Figure 25 (p.51) shows the characteristic system curve plotted in a parabolic fashion according to the relationship established in Equation 20. • Example 2 shows the effect of the relationship in Equation 20. E XAMPLE 2

• The exhaust system in Example 1 has a filter bank to protect the heat-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 design flow rate of 10,000 cfm. • The required static pressure of the fan is 2.5 in.Wg at the design flow rate. This identifies the design operating point A from which the design curve A is plotted. It is assumed that the fan is plenum 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 in Example 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.Wg at the design flow rate, which means in the dirty condition, the total system static losses are 3.5 in.Wg. • This gives a new design point C, from which the dirty maximum design operating system curve C is plotted (see Figure 27).

60


Example 2.1 • Using formulae and manufacturers' catalogue data (table 2, p. 52), determine for operating point C, the speed, the highspeed fan power, and the corresponding fan total and static efficiencies. Assuming the same drive losses of 5.6%, calculate the motor service factor . • By interpolation, the fan power at point C is 8.2 hp from Table 2, 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 setting under dirty filter conditions identified as point D on the system curve (see Figure 27).


61

• EQUATION 22: Actual air volume delivered by the fan at a constant 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 from catalogue 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%

62


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


63

C H A P T E R 9

FAN AND S YSTEM INTERFACE S YSTEM E FFECT F ACTORS

• A fan is normally tested with open inlets and straight duct attached to the outlet. This results in uniform airflow into the fan 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 for when selecting the fan. • Figure 28 illustrates deficient fan and system performance with the calculated system-design pressure and flow shown as Point 1. • Since no allowance was made for system effect, the actual operating condition is Point 4 – at the intersection of the fan pressure-volume curve and the actual system curve. • The difference between Point 1 and Point 4 projected on the abscissa is the deficiency in flow. • To compensate for the deficiency, a system effect factor equal to the pressure difference between Points 1 and 2 must be added to the calculated system-pressure losses, with the fan selected to operate at Point 2. Chapter 9: Fan and System Interface

65

• The fan system-effect factor is the product of Co times the velocity pressure and is calculated using Equation 6, p. 48. • For centrifugal fans, velocity is based on the area of the inlet collar and the outlet area; for axial fans, it is based on the fan diameter. • Appendix C explains how to determine Co (the system effect factors For SWSI centrifugal fans). Other system-effect factors are beyond the scope of this guide, but are covered in AMCA Publication 201-90.

Figure 29: Deficient Fan and System Performance F AN O UTLET C ONDITIONS

• System-effect factors must be calculated and added to the system resistance losses whenever 100% recovery at the outlet of a fan cannot be achieved. 66 Fans Reference Guide

• Complete recovery can be achieved if the outlet effective duct length is 2.5 diameters or more for a velocity of 2,500 fpm (13 m/s) or less. Add one duct diameter for each additional 1,000 fpm (5 m/s).

Figure 30. Fan-outlet Velocity Profiles

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

• In some cases, fan outlets are connected directly to a larger duct or plenum without a transition. This causes a pressure loss of up to one velocity head (V 2 /2g), based on the highest velocity of the fan outlet. Chapter 9: Fan and System Interface

67

• The highest velocity is in the blast area of a centrifugal fan (the area between the cutoff and the scroll), and in the swept area of an axial fan. F AN I NLET C ONDITIONS

• To achieve rated fan performance, air must enter a fan uniformly over the inlet area in an axial direction without prerotation. • The ideal inlet condition allows air to enter axially and uniformly without spin. • A spin, in the same direction as the impeller rotation, reduces fan flow and pressure, whereas a counter-rotating vortex at the inlet slightly increases fan flow and pressure and substantially increases fan power. • The most common cause of reduced fan performance is nonuniform flow into the inlet of a fan. • Turbulence and uneven flow into the fan impeller is typically caused by an elbow at the inlet of a fan. • System-effect Co factors for inlet conditions that cause spin are not available because of the multitude of variations. • The system-effect factor can be eliminated by including an appropriate length of straight duct between the elbow and the fan inlet. • For fans installed in cabinets or adjacent to walls, adequate distance must be maintained in front of the inlet to allow for unobstructed flow.

68


C H A P T E R 1 0

FAN SELECTION • 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 is known. • The system-pressure-loss curve is defined by accounting for the system flow, resistances, and system-effect factors (according to ASHRAE methods). • The fan size, speed power, and noise spectrum is determined using one of the many methods available from fan manufacturers, e.g., tables or PC-based computer programs.

Chapter 10: Fan Selection

69

• Comp Comput uter er selec selecti tion on prog progra rams ms allow allow a desi design gner er to eval evalua uate te the the various fan options for optimizing system efficiency and performance, obtain noise-level data, and plot the fan performance curve quickly. quickly. • The The perf perfor orma manc ncee data data in in fan fan table tabless is bas based ed on on arbit arbitra rary ry increments of flow rate and static pressure, and shading may be used to indicate an AMCA fan-class demarcation. • Fan Fan Laws Laws cann cannot ot be be used used to obt obtain ain adj adjace acent nt dat dataa point pointss beca because use each data point represents a different point of operation on the fan performance curve. • Inte Interm rmed ediat iatee poin points ts of of opera operati tion on can can be be dete determ rmine ined d by interpolation, as the listed data points are close enough for reasonable accuracy for fan selection. • Using Using a perfo perform rmanc ancee curv curvee in con conju junc nctio tion n with with a com compu pute terr program or tables is very important, particularly particul arly in VAV VAV systems which have more than one point of operation. • Using Using the the perf perfor orma manc ncee curve curve opti optimi mize zess fan sel select ection ion to to avoid avoid operation close to, or in, the stall region and maximize efficiency throughout the operating range. P RESSURE D EFINITIONS

• When When usin usingg perf perfor orma manc ncee table tabless or cha chart rts, s, it it is imp impor orta tant nt to to understand what definition of pressure has been used by the fan manufacturer. manufacturer. There are three possible ways to state the fan's pressure requirement: Fan Total Pressure, Fan Static Pressure and Fan Static Pressure Rise.

70


• These These defin definitio itions ns arise arise from from Bern Bernoull oullí’ í’ss equat equation ion which which is used used to calculate system friction losses. The relationship relationship between any two points (1 and 2) in a system is given below: Ps1 +

V 1

2

1096

2

V 2

r1 = Ps2 +

1096

r2 + Friction Loss

The Friction Losses in this equation must be overcome by the fan. • Fan Total Pressure is defined as Fanpt =

Ps2 +

V 2 1096

2

r2

-

V 1

Ps1 +

1096

2

r1

• Fan Static Pressure is defined as Fanps =

Ps2 -

Ps1 +

V 1 1096

2

r1

• Fan Static Pressure Rise is defined as Dps =

Ps2 - Ps1

• The The in inle lett an and d ou outl tlet et velo veloci citi ties es (V 1 and V 2) in these equations are taken to be at the terminals of the fan manufacture’s supply, supply, which may include silencers, inlet boxes, outlet diffusers, etc. The velocities may not be identical to those in the adjacent ducts. The most common definition in North America is Fan Static Pressure for centrifugal fan and Fan Total Pressure for axial flow fans. The Europeans use Fan Total Total Pressure almost exclusively for all fans. Chapter 10: Fan Selection Selection

71

A N S ELECTION P ARALLEL F AN

• Select Selecting ing paralle parallell fans fans with with a charact characteris eristic tic pressu pressure re reduct reduction ion left of the peak pressure point typical of FC fans requires careful consideration. • When When the these se fans fans are are ope operat rated ed in par parall allel, el, a fluc fluctu tuat atin ingg load load condition may result if one of the fans operates to the left of the peak static point on its performance curve. • Figure Figure 31 shows shows the pressu pressure re flow curves curves of a singl singlee fan fan (curve (curve A-A) and the same fan operating in parallel with an identical fan (curve B-B).

Figure 31: Design Operating Point P oint Selection Range on a Typical Typical Centrifugal Fan Performance Curve

72


Figure 32: Pressure Flow Curves

• The figure-eight curve plots possible combinations of volume flow at each pressure value for the individual fans. • Points to the right of B-C are the result of the fan operating to the right of its peak rating point; for all systems stable operation occurs with a system resistance curve below C-C . • For points of operation to the left of B-C it is possible to satisfy system requirements with one fan operating at one point of rating while the other fan operates at another. • Figure 31 shows point BD1 – 4,700 cfm at a static pressure of 2.1 in.Wg – can be satisfied with each fan operating at 2,350 cfm at 2.1 in static pressure.


73

• The system curve can also be satisfied at point BD2 by one fan operating at 1,350 cfm at a static pressure of 1.9 in.Wg and the second fan operating at 3,150 cfm at the same static pressure. • This is because the system curve D-D passes through the combined performance curve at two points, but operation can be unstable under such conditions. • With fan selection at point BD2, one fan is under loaded while the other fan is heavily loaded, and surge can occur in the system. S ERIES V S P ARALLEL O PERATION

• In any 2 stage arrangement, the same mass flow per unit time must be handled by each stage (if there is no leakage). The density of the gas passing through each stage will be different, so it follows that the volume handled by each stage will be different. • Figure 33 shows the static pressure and horsepower curves for a single fan. Also shown, are the static pressure curves for two of these fans (a) connected in series, and for comparison (b) connected in parallel. Strictly speaking, the static pressure curve for the two fans in series will be slightly higher than that shown since there is only one velocity head to be deducted from the combined total pressure from 2 stages in order to compute the combined static pressure available. For simplicity of discussion, the static pressure is doubled for any given volume for two identical fans in series. • In Figure 33, system B passes through P, the intersection of the combined series and combined parallel curves, i.e. either combination will give the same volume on this system. However, in series, each fan will consume the horsepower at point D, whereas in parallel each fan will take the power shown at point E. 74


System A Static Pressure Two Fans in Series

e r u s s e r P c i t a t S

System B System C

D

Horsepower One Fan P Static Pressure One Fan

Static Pressure Two Fans in Parallel

Volume Flow

Figure 33: Series Fan Operation

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


75

C H A P T E R 1 1

FAN NOISE • Fan noise is an important criteria for the proper selection of fan type and size for an application. • The noise from a fan is predominantly from aerodynamic sources and includes factors such as lift, rotation, vortex shedding, and wake. • The noise generated by a fan depends on the fan design, the volume flow rate, total pressure, and efficiency. This noise is proportional to the product of the pressure squared and the flow. • Low outlet velocity does not necessarily relate to lower sound power, and fan selection should not be based solely on fan tip speed. • The only valid basis for comparison is the actual sound power levels generated by the different fan types when they are operating at the required system flow and pressure. • For constant-volume systems, the recommended practice for a selected fan type is that the fan size and speed be selected so operation falls at or near the peak efficiency point of the fan performance curve. Chapter 11: Fan Noise

77

• A fan is normally quietest when selected within the most efficient operating range, which is also advantageous for energy conservation. • Fan selection criteria for VAV systems include two other factors: the efficiency and stability of the fan through the entire range of modulation, and the acoustic impact of the modulation system. • Fan selection for VAV systems is a compromise between fan surge and fan inefficiency, and the narrower the range of modulation the more acceptable the compromise will be. • Variable inlet vanes may generate significant low frequency noise as the vanes modulate to the close position and require additional attenuation with a corresponding increase in system pressure drop. Maximum sound levels occur at approximately 75% open VIVs. • The other modulation systems – variable-speed motors and drives, and variable-pitch fan blades – generate less noise as the fan modulates to the no-flow operating point. • The sound power generation of a specific fan at its operating point should be obtained from manufacturers' AMCA test data or from manufacturers' computer fan-selection programs. • The data is presented as sound power levels in eight octave bands and as weighted overall sound level. • If test data is not available, the octave-band sound power levels can be estimated using the following procedure. F AN S OUND P OWER

• Fans generate a tone at the blade passage frequency, and the number of decimals to be added is the blade frequency increment (BFI). The octave band to which the BFI is added depends on the type of fan and the impeller speed. 78


• EQUATION 24 : Blade frequency Bf = N x no. of blades cf where Bf N cf

= blade frequency, Hz = impeller speed, rpm (r/s) = conversion factor 60 ( 1 ) Table 3: Typical Number of Fan Blades Impeller Size/Drive

Fan Type

No. of Blades

CENTRIFUGAL Airfoil and Backward-inclined

24 in. and over Under 24 in.

10 12

Forward-curved

52

Radial

6

AXIAL Vaneaxial Tubeaxial

12 Belt drive Direct drive

6 4

Propeller

6

• Blade frequency can be estimated using data from Table 3. • EQUATION 25 : Estimating sound power levels at actual operating conditions. Lw where Lw K w Q

= K w + 10 log Q/cf 1 + 20 log p/cf 2 + C = estimated sound power level (dB re 1 pW) = specific sound power level = flow rate, cfm (L/s) Chapter 11: Fan Noise

79

p C cf 1 cf 2

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

• Estimated sound power level is calculated for all seven bands with KW selected from Table 4. The BFI is added to the octave band in which the blade passage frequency falls. Sound correction factor is selected from Table 5. Table 4: Specific Sound Power Levels and Blade Frequency Increments 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

Airfoil and backward-inclined


32 36

32 38

31 36

29 34

28 33

23 28

15 20

3

Forward-inclined

All

47

43

39

33

28

25

23

2

Radial blade and Pressure blower

40 in. and over 20 in. to 40 in. Under 20 in.

45 55 63

39 48 57

42 48 58

39 45 50

37 45 44

32 40 39

30 38 39

8


39 37

36 39

38 43

39 43

37 43

34 41

32 28

6


41 40

39 41

43 47

41 46

39 44

37 43

34 37

5

All

48

51

58

56

55

52

46

5

CENTRIFUGAL

Axial Vanaxial Tubeaxial Propeller Cooling tower Note:

80

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


Table 5: Sound Correction Factors

Correction Factor, C, for Off-peak Operation Static Efficiency % of peak

Correction Factor dB

90 to 100 85 to 89 75 to 84 65 to 74 55 to 64 50 to 54

0 3 6 9 12 15

E XAMPLE 3

• The fan in Example 1, p. 51, is operating at its design condition of 10,000 cfm, at static pressure of 2.5 in.Wg, static efficiency of 60.9% and impeller speed of 1,475 rpm. • Determine the sound power level in seven octave bands by assuming the number of impeller blades to be 10 and by estimating the off-peak, static-efficiency correction factor. • A simple method to calculate the off-peak, static-efficiency correction factor with reasonable accuracy, is to determine the static 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 the 55%-WOcfm line is 7,000 cfm at 4.5 in.Wg static pressure and 6.6-bhp power.

Chapter 11: Fan Noise

81

• EQUATION 26 : Peak static efficiency given by rearranging Equation 11, p. 49. h2

= QPs cf PFi = 7,000 x 4.5 6,349 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 rate and pressure are given by Equation 25, p. 73. Lw = Kw + 10 log 10,000 + 20 log 2.5 + C = K w + 40 + 6 = K w + 54 dB • The specific sound power levels (K W) for Table 6 are obtained from 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 factor suggests that a more efficient fan should be selected for the given duty.

82


• A larger fan would allow for operation closer to the surge line at a slower impeller speed, lower power and lower noise levels, but with less ability to cope with higher system static pressures at 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

89

86

82

74

BFI Total dB

3 90

93

93

Chapter 11: Fan Noise

83

C H A P T E R 1 2

FAN DUTY CONTROL • Fans are required to perform over a range of flows and pressures called, “The Duty Cycle”. How the fan is controlled to achieve the required range, can have a significant energy cost. • The type of fan control should be selected on the basis of cost, the precision of control required and the frequency and magnitude of system flow changes. • For HVAC systems with infrequent changes in flow rate and/or where control is a secondary consideration, a two- or threespeed motor is a low-cost solution. • Where a system requires continuously varying flow rates over speed ranges of 2.6:1 and input power of less than 15 kW, an adjustable-speed pulley drive is satisfactory and can improve system efficiency. • Magnetic or hydraulic slip couplings can be used in systems with power greater than 15 kW. However these couplings have an inherent power loss since the torque from the motor is transmitted unchanged to the impeller, in spite of the difference in rotational speed.

Chapter 12: Fan Duty Control

85

• The most efficient method of speed control, with the potential for precise control, is the electronic adjustable speed drive (see figure 32). • Mechanical methods of volume control are often used in commercial and industrial HVAC systems. • The simplest, most inefficient method of control, is a fan discharge damper; the damper artificially increases system resistance and the fan works along its system curve.

e r u s s e r P t e l t u O

unstable region

Higher system  resistance curve as outlet  damper closes

Volume Flow

Figure 34: Outlet Damper Fan Control

• Single fan performance curve • Multiple system resistance curves


• To provide a broader range of flow control at lower energy penalties, the discharge damper can be used with multi-speed motors, or with adjustable speed drives. • Figure 33 shows discharge-damper control sequenced with motor-speed control. • There is no advantage other than initial cost to using an outlet damper. Its use should be avoided.

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

• Inlet-vane control can provide precise flow control, down to about 40% of the full flow rate. • This control device rotates the inlet airflow the same as the impeller and so reduces the work done by the impeller. Chapter 12: Fan Duty Control

87

• Another inlet-control device is a cone that varies the effective fan-inlet area as a Function of the axial distance the cone is positioned 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 can be dynamically regulated by varying the attack angle of the impeller blades; this maintains high efficiencies over a wide blade-pitch range.

Figure 36: Inlet Vane Control of a Fan

88


C H A P T E R 1 3

V IBRATION ISOLATION/ FAN BALANCING • All rotating machinery have critical speeds called resonant frequencies where excessive vibration can cause damage. It is necessary to have each fan and foundation installation checked to avoid these speeds, or correct the fan balance. • Most fans are shipped statically and dynamically balanced but corrosion, erosion, dust and airborne contaminants collecting on the impeller may cause imbalance over time. • Therefore, consider fan isolation when designing the installation. • The transmission of vibration to a building structure involves vibratory force, frequency of vibratory force (disturbing frequency), natural frequency of isolator and floor, and stiffness of isolator and floor. • It is important to select vibration isolators to compensate for floor deflection; and to avoid resonance, the natural frequency of the isolator should be different to the disturbing frequency. • The degree of fan isolation and balancing depends on the floor span, and the fan type, size, speed and power.

Chapter 13: Vibration Isolation / Fan Balancing

89

• Use the "Vibration Isolator Selection Guide" in the ASHRAE HVAC Applications Volume Handbook to determine the base and isolator type and minimum deflection. • Install flexible duct connectors at the inlet and at the discharge to reduce transmission to the duct work. • Consider resilient, structural steel or concrete, inertial bases for fans in critical areas or in long buildings of light construction. Note: The best vibration isolation design will not compensate for a

resonant frequency problem. Isolation will only reduce the transmission of forces due to imbalance.

90


C H A P T E R 1 4

ELECTRIC MOTOR FAN DRIVE • The AC electric motor is the main type of prime mover used to drive fans and there are many types. • The selection of a high-efficiency motor is important, but the starting motor current and torque are more important. • Excessive starting time raises the temperature of the motor windings beyond acceptable levels. F LYWHEEL E FFECT

• The time to accelerate a fan to operating speed depends on the fan/impeller inertia (flywheel effect) and the starting characteristics of the electric motor. • The fan/impeller inertia is given as WR 2 in the industry. This must be corrected to represent the apparent inertia as seen by the motor when the fan operates at a different speed from the motor. R is called the radius of gyration. • Additional resistance to starting will be air power consumed by the fan. Therefore it is advisable to start centrifugal fans with dampers closed. (Axial fans should have dampers open.)

Chapter 14: Electric Motor Fan Drive

91

• The charact characteris eristic tic starti startingng-tor torque que curves curves of the fan motor motor and the maximum allowable time for acceleration (usually about 10 seconds) are available from manufacturers. • The availabl availablee acce accelera leratio tion n torq torque ue is is the the differ differenc encee betw between een the motor torque and the fan torque. AC MOTORS

• Poly Polyph phase ase (us (usual ually ly thre threee-ph phase ase)) AC moto motors rs are are almost almost alw alway ayss used in fan applications that require more than 2 hp. • The AC indu inducti ction on m moto otorr, usuall usually y with with a squir squirrelrel-cag cagee rotor rotor and no external connections, is the most suitable for three- phasepower fan drive as it is inexpensive and reliable. • It is a consta constantnt-spee speed d moto motorr with with a flat flat torq torque ue chara characte cterist ristic ic in in relation to motor speed. • Howeve Howeverr, its its start starting ing curren currentt is high – as much much as seven seven or eight eight times higher than the running current. • The extra extra starting starting curren currentt for for fans fans with with low inertia inertia and large large motors can cause problems with electrical supply and demand. Special consideration must be given to reducing the starting current. • It is is usual usual to reduc reducee the the star startin tingg volta voltage ge at at start start-up -up and step up the voltage until the fan reaches running speed. • Anot Anothe herr solut solutio ion n is to to use use a woun woundd-ro roto torr induc inducti tion on mot motor or in in which the polyphase windings of the motor are connected to an external resistor via slip rings. • Star Starti ting ng tor torqu quee and and star starti ting ng cur curre rent nt can can be be cont contro rolle lled d by adjusting the external resistance. • Smal Smalll fans fans requ requir iring ing less less than than 2 hp usu usuall ally y use use single single-p -pha hase se power supply. 92


• SingleSingle-pha phase se indu inductio ction n moto motors rs are are non non-sel -selff start starting ing;; to to start start the motor, a second starting phase is created by connecting an extra stator winding through a capacitor. • The starti starting ng capacit capacitor or displace displacess the the phase phase of of the the current current and is disconnected after running speed is obtained, and a different size capacitor is used for running. • Shad Shaded ed pole pole mot motor orss are are gener general ally ly unsu unsuit itab able le for for fan fan drive drive,, because of their inherent poor starting torque. D C M OTORS

• Sometim Sometimes es DC motors motors are prefer preferable able for fan dri drive, ve, partic particular ularly ly in applications requiring speed modulation. • The The serie seriess moto motorr is mos mostt suit suitab able, le, bec becau ause se it has has mod moder erat atee starting current and self-regulating, stable operating characteristics.

Chapter 14: Electric Motor Fan Drive

93

C H A P T E R 1 5

ENERGY CONSUMPTION ANALYSIS EXAMPLE: Building Ventilation • Note: This procedure involves many calculations that are easily done using a spreadsheet computer program. • It is oft often en nec necess essary ary to est estim imat atee the the energ energy y cons consum umpt ptio ion n of a fan, particularly for life-cycle costing. The energy costs are usually determined for a period of a year. • EQUA EQUATIO TION N 27: Est Estima imatin tingg the ene energy rgy con consum sumptio ption n of a fan involves integrating the fan shaft input power divided by system efficiencies over time. n

PFi t(n)

1

hDt(n) x hMt(n) x h Vt(n) x cf

E=∑ E= ∑ where: E PFi t(n) hDt(n) hMt(n) h Vt(n)

x t (n) (n)

= energy consumption, kWh = fan shaft input power for time period, hp (W) = drive efficiency for time period, dimensionless ratio = motor efficiency for time period, dimensionless ratio = variable-speed drive efficiency for time period, dimensionless ratio Chapter 15: Energy Consumption Analysis

95

cf t (n)

= conversion factor, 1.3410 ( 1,000) = 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 drive efficiencies. Motor efficiencies are obtained from manufacturers' data. The procedure for using Equation 27 depends on the fan being analyzed. C ONSTANT- VOLUME F ANS

• For a constant-volume fan processing a gas at a constant temperature rise, the procedure is straight forward as the energy consumption is integrated over one time period. • For a constant-volume fan processing outdoor-air, the fan power varies as a function of the outdoor-air temperature. • When energy use is a function of outdoor-air temperature, the annual energy consumption can be determined using a computerized, hourly analysis program. • In the absence of such a program, a reasonable assessment can be made by using ASHRAE bin weather data. When time-ofuse energy rates apply, then the weather data can be separated into on-peak and off-peak periods. • EQUATION 28: Total energy consumption: n

PFi b(n) X t b(n)

1

hDb (n) x hMb (n) x cf

E=∑ where: E PFi b(n) hDb (n)

96

= energy consumption, kWh = fan shaft input power at bin temperature, kW = drive efficiency at bin temperature, dimensionless ratio


hMb (n)

t b(n) cf

= motor efficiency at bin temperature, dimensionless ratio = time at temperature bin, hours = conversion factor, 1.3410 ( 1,000)

• EQUATION 29: The fan shaft input power at each temperature bin when a constant-volume fan processes air of varying temperature: PFi b(n)

= PFi s x rb rs

PFi b(n)

= fan-shaft input power at bin temperature, hp (W) = fan-shaft input power at standard conditions, hp (W) = actual bin density, lbm /ft.3 (kg/m3) = standard air density, lbm /ft.3 (kg/m3)

where: PFi s rb rs V ARIABLE - VOLUME F ANS

• For an AF or BI centrifugal fan with variable-inlet vanes (VIVs), the load profile of the fan system and the design point of the fan (with the VIVs completely open) must be determined or estimated. • From the full-load design point, the fan flow as a percentage of wide-open cubic feet per minute (WOcfm) must be determined from the fan curve or a computer program. This is necessary to determine which constants are used in the following equations, which estimate the part-load conditions when the system curve originates at the apex of zero flow and zero static pressure.

Chapter 15: Energy Consumption Analysis

97

• EQUATION 30: Load profile expressed in terms of WOcfm. %Q(n) where: %Q(n) %Qd Lf (n)

= %Qd x Lf (n) 100 = percentage fan load for load point, %WOcfm = percentage design fan load, %WOcfm = load factor for load point, %

• EQUATION 31: Percentage design fan load. %Qd

=

Qd

x 100

Q WOcfm where %Qd = percentage design fan load, %WOcfm Qd = flow at design, cfm (L/s) Q WOcfm = flow at WOcfm, cfm (L/s) • EQUATION 32: Percentage fan-shaft input power at each load point, 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, %WOcfm a,b,c = constants determined from table closest to design WOcfm • EQUATION 33: The fan-shaft input power in horsepower at the design point is determined from manufacturers' data and corrected for VIV losses. PFi c

= PFi d x f P

PFi c PFi d f P

= corrected fan-shaft input power at design, hp = fan input power at design, hp = fan power correction factor

where:

98


• EQUATION 34: Typical fan power correction factor. f P where f P %Qd

=1+

1.933 994 x exp (0.026 075 x %Q d) 100

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

• EQUATION 35: The fan input power at each load point. PFi (n) where: PFi (n) PFi c %hp(n) %hpd rn rd cf

= PFi c x %hp(n) x rn x cf %hpd rd = fan shaft input power for time period, hp (W) = corrected fan input power at design, hp = percentage fan power for time period, %hp = percentage fan power at design, %hp = actual density at load point, lbm /ft.3 (kg/m3) = design air density used to determine hpd, lbm /ft.3 (kg/m3) = conversion factor, 1.0 (745.70)

• EQUATION 36: To determine the fan motor input power, the motor efficiency at each load point must be determined as a function of the load factor. Lf M = PMo(n) PM where Lf M PMo(n) PM

= motor load factor, dimensionless ratio = fan motor output power at load point, hp = nominal nameplate rating of motor, hp


99

• EQUATION 37: Fan motor output power (derivative of Equation 13, p.50). PMo(n)

= PFi(n) hD(n)

where: PMo(n) PFi(n) hD(n)

= fan motor output power, hp = fan-shaft input power at load point, hp = drive efficiency, dimensionless ratio

• EQUATION 38: The fan motor input power at each load point considering the part load efficiencies by combining Equation 13 (p. 50) and Equation 14 (p. 50). PMi(n) where: PMi(n) PFi(n) hD(n) hM(n)

cf

=

PFi(n) hD(n) x hM(n) x cf

= fan motor input power at load point, kW = fan-shaft input power at load point, hp (W) = drive efficiency at load point, dimensionless ratio = motor efficiency at load point, dimensionless ratio = conversion factor,1.3410 (1,000)

• The total energy consumption in kilowatt-hours is then determined by adding the product of the time in hours and the fan motor input power in kilowatts at each load point. • EQUATION 39: Total energy consumption. n

E=∑ (PMi(n) x t (n)) where: E 100

1

= energy consumption, kWh


PMi(n) t (n)

= fan motor input power for load point, kW = time at load point, hours

• EQUATION 40: Percentage static pressure at each load point including 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, %WOcfm a,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) Psd %sp(n) %spd rn rs

cf

= static pressure at load point, in.Wg (Pa) = static pressure at design, in.Wg (Pa) = percentage static pressure at load point, %sp = percentage static pressure at design, %sp = actual density at load point, lbm /ft.3 (kg/m3) = standard density used to determine psd, lbm /ft.3 (kg/m3 ) = 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 Chapter 15: Energy Consumption Analysis

101

Q(n) Ps(n) PFi (n) cf

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

• EQUATION 43: Fan speed at the design point determined from manufacturers' data and corrected for VIV losses. NFc where: NFc NFd Nf

= NFd x Nf = corrected fan speed with VIVs, rpm = fan speed at design without VIVs, rpm = fan speed correction factor

• EQUATION 44: Typical fan speed correction. Nf where: Nf %Qd

= 1 + -41.329 73 + 11.168 903 x ln %Q d 100 = fan-speed correction factor, dimensionless ratio = %WOcfm at design

E XAMPLE 4

• Uses the energy-analysis formulae in this chapter and the density-calculation formulae in Appendix A. • A ventilation system is required for a new welding shop at a manufacturing plant. The shop is located in Toronto, Ontario, at 176m above sea level, and is to operate three shifts a day year round. Analysis has determined the welding booth requires a minimum of 40,000 cfm of exhaust air for contaminate control. To remove excess heat in warm weather, doubling the exhaust volume to 80,000 cfm has been considered. • One proposed system is two exhaust fans of 40,000 cfm each and a roof-mounted, VAV, makeup-air system with a blow102


through, 54-in.-diameter, DWDI centrifugal fan with VIVs and a glycol heating coil. The makeup-air unit fan delivers a maximum of 80,000 cfm at a static pressure of 4.5 in.Wg, at 0.075 lb/ft 3. • Also included is a roof-mounted gravity relief damper to prevent overpressurization of the zone. Whenever the system must operate, the control sequence would be as follows: • One exhaust fan and the makeup-air unit start and run continuously. The VIVs in the unit modulate and provide makeup airflow as determined by a calibrated velocity pressure sensor in the supply duct. The supply air should be reset from the minimum volume of 40,000 cfm at an outdoor temperature of l3˚C to the maximum volume of 80,000 cfm at an outdoor temperature of l9˚C. When the VIVs are fully open, the second exhaust fan starts and runs continuously. A normally-open control valve on the heating coil is modulated to maintain a minimum supply-air temperature of 12.8˚C. • To determine the annual energy consumption of the make-upair fan, considering density flow variation, use ASHRAE Metric Bin Weather Data and a PC spreadsheet program using the modified bin method to simplify calculations. •

The fan requires a 100-hp motor, and motor-efficiency data is in Table 8.


103

Table 7: Typical VAV-fan Constants %WOcfm Dependent Variable

Independent Variable

Constant

%hp

%Q

a

1.015668 1.251667 1.847918 1.915869

b

0.078326 0.068488 0.055539 0.052399

c

42.50

a

0.036628 0.028756 0.022564 0.017582

b

2.000501 2.000751 2.000808 1.999086

%sp

%Q

50

55

42.25

60

41.50

65

41.50

%WOcfm Dependent Variable

Independent Variable

Constant

%hp

%Q

a

2.012019 2.462608 3.751074 2.274535

b

0.048141 0.042180 0.033279 0.037477

c

41.50

a

0.013297 0.009782 0.007228 0.004959

b

2.000341 2.001993 1.997487 1.996321

%sp

104

%Q


70

75

41.00

80

39.50

85

41.50

Table 8: Motor Load Efficiencies Motor

Load Factor Percentage

Size

100 hp

25

50

75

100

115

125

89.05

91.86

93.50

93.05

92.83

92.70

Table 9: The Solution to Example 4 GIVEN PARAMETERS

VALUE

Elevation, m

176

Standard Density (r), Ib/ft3

0.075

Flow (Q), cfm

80000

Static Pressure (p s), in.Wg

4.5

REMARKS

FAN SELECTION PARAMETERS

Power (PFi d) at Std Cond, bhp

74

Speed (NFd) at Std Cond, rpm

693

Max Flow (Qmax) at zero pressure, cfm

130300

From performance curve at 693 rpm

%WOcfm (%Q), %

61.40

Q/Qmax

Maximum Static Pressure (Ps max) in.Wg

5.5

From performance curve at 693 rpm

Motor Power (PM), bhp

100

Nominal motor size

Motor Efficiency ( hM). %

93.5

Peak efficiency at 75% load contÕd


105

Table 9: The Solution to Example 4 (contÕd) VAV FAN CONSTANTS

%WOcfm

60

Value nearest Q/Qmax Data from Table 7

Variables

Constants

%hp %Q

a

1.847918

b

0.055539

c

41.50000

a

0.022564

b

2.000808

%sp %Q

CALCULATED CALCULA TED PARAMETERS

Station pressure (Pb), kPa

99.261

Interpolated

Drive efficiency ( hD). %

4.00

Equation B1, p. 109

Percent power at design using (%hpd)

93.249

Equation 32, p. 92 60%WOcfm

Percent pressure at design using (%spd)

81.500

Equation 40, p. 95 60%WOcfm

VIV correction factor (fp)

1.09245

Equation 34. p. 93

VIV correction factor (N f)

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

Summary • In revi review ewing ing the the out outpu putt from from the the spr spread eadsh sheet eet pr prog ogra ram, m, it it is apparent that the fan operates at a static efficiency of only 16% for over 6,000 hours a year year.. Therefore, the proposed control scheme and system arrangement is not very efficient and another scheme should be considered. Also, the ductwork would have to be doubled in size for this arrangement, which would add to the installation cost. 106


C H A P T E R 1 6

APPENDICES

A PPENDIX A - D ENSITY C ALCULATIONS

Moist Air Parameters for Density Determination

• Moist air is defined as a binary mixture of dry air and water vapour. The maximum amount of water vapour at saturation in

C H A P T E R 1 6

APPENDICES

A PPENDIX A - D ENSITY C ALCULATIONS

Moist Air Parameters for Density Determination

• Moist air is defined as a binary mixture of dry air and water vapour. The maximum amount of water vapour at saturation in moist 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 dry air (pa) and the partial pressure of water vapour (pw), V is the total mixture volume, the total moles (n) is the sum of the number of moles of dry air (na) and the number of moles of water vapour (nw), T is the absolute temperature and 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 the total mass to the total volume. p = (ma + mw)/V Chapter 16: Appendices

109

where: ma mw

= mass of dry air = mass of water vapour

• EQUATION A3: The humidity ratio (W) of a given moist air sample is defmed as the ratio of the mass of water vapour to the mass of dry air. W = mw /ma • The saturation humidity ratio (Ws) is the humidity ratio of moist air saturated with water (or ice) at the same temperature and pressure. • EQUATION A4: Relative humidity (f) is the ratio of the mole fraction of water vapour in a given moist air sample to the mole fraction in a saturated air sample at the same temperature and pressure. f = xw /xws | t,p

• EQUATION A5: The dew point temperature (t d) is the temperature of moist air saturated at the same pressure (p) and with the same humidity ratio (W) as that of the given sample of air. It is defined as the solution t d(p, W) to: Ws(p,t d) = W • The thermodynamic wet-bulb temperature (t wb) is the temperature at which water (liquid or solid), by evaporating into moist air at a given dry-bulb temperature (t) and humidity ratio (W), can bring air into saturation adiabatically at the same temperature (t wb), while the pressure (p) is constant. Density Calculations

• The numerical method for calculating the density of moist air is a multistep process and initially involves determining the watervapour saturation pressure. The saturation pressure in SI units 110


at the wet-bulb temperature is calculated by the following Hyland and Wexler formulae published in the ASHRAE Handbook, 1989 Fundamentals Volume . • EQUATION A6: The saturation pressure over ice for the temperature range of -l00˚C to O˚C. In(pws) = C1 /T + C2 + C3 T + C4 T 2 + C5 T 3 + C6 T 4 + C7In(T) where: C1 C2 C3 C4 C5 C6 C7

= -5.674 535 9 E-3 = 6.392 524 7 = -9.677 843 E-3 = 6.221 157 Ol E-7 = 2.074 782 5 E-9 = -9.484 024 E-13 = 4.163 501 9

• EQUATION A7: The saturation pressure over water for the temperature range of O˚C to 200˚C. In(pws) = C8 /T + C9 + Cl0 T + C11 T 2 + C12 T 3 + C13In(T) where: C8 = -5.800 220 6 E-3 C9 = 1.391 499 3 C10 = -4.864 023 9 E-2 C11 = 4.176 476 8 E-5 C12 = -1.445 209 3 E-8 C13 = 6.545 967 3 and where for both equations: pws = saturation pressure, Pa T = absolute temperature, K • EQUATION A8: Saturation pressure in IP units: pws(IP) = 0.020855 p ws(SI)

Chapter 16: Appendices

111

where: pws(IP) = saturation pressure, lb/ft 2 pws(SI) = saturation pressure, Pa • EQUATION A9: Saturation humidity ratio. Ws = 0.62198 p ws /(p - pws) where: Ws p pws

= saturation humidity ratio, dimensionless 2 (Pa) = absolute pressure, lb /ft f 2 (Pa) = saturation pressure, lb /ft f

• EQUATION A10: The humidity ratio using the wet-bulb temperature. W = ( 1,093 - 0.556 t wb) Ws(wb) -0.240 (t db - t wb) 1,093 + 0.444 (t db - t wb) • EQUATION A11: or in SI units: W = (2,501- 2.381 t wb) Ws(wb) - (t db - t wb) 2,501 + 1.805t db - 4.186t wb where for both Equation A10 and Equation All: W = humidity ratio, dimensionless Ws = saturation humidity ratio, dimensionless t db = dry-bulb temperature, ˚F (˚C) t wb = wet-bulb temperature, ˚F (˚C) • EQUATION A12: The humidity ratio. W = 0.62198 p w /(p - pw) where: W p pw

112

= humidity ratio, dimensionless 2 (Pa) = total pressure, lb /ft. f 2 (Pa) = partial pressure of water vapour, lb /ft. f


• EQUATION A13: Relative humidity using the perfect gas relationships: f = Pw /Pws | t,p

where: f

pw pws

= relative humidity, dimensionless 2 (Pa) = partial pressure of water vapour, lb /ft. f 2 (Pa) = saturation pressure, lb /ft. f

• EQUATION A14: Volume. v = R a T ( 1 + 1 .6078 W) p where: v R a T p W

= volume, ft.3 /lbm, (m3 /kg) = gas constant for air, 53.352 ft.lb /lb f m.F(abs) (287.055 J/kg.K) = absolute temperature, ˚F (abs) (K) = total pressure, lbf/ft2 (Pa) = humidity ratio, dimensionless

• EQUATION A15: Density is determined by the inverse of the volume. r = 1/v

where: r

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

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


113

Table A1. Standard Atmospheric Data for Altitudes to 3,000 m Altitude m 0 500 1,000 2,000 3,000

ft. 0 1,640 3,281 6,562 6,562

Temperature ûC ûF 15.0 59.0 11.8 53.2 8.5 47.3 2.0 35.6 -4.5 23.9

Source: ASHRAE

Handbook, 1989 Fundamentals Volume

kPa 101.325 95.461 89.874 79.495 70.108

Pressure in.Hg 29.921 28.19 26.54 23.47 20.70

Table A2. Density Calculations Given Parameters

tdb, twb, p

tdb, td, p

tdb, f, p

114

To Obtain

Use

Comments

pws | twb

Equation A6 or Equation A7

Using twb

Ws | twb

Equation A9

Using pws | twb

W


Using ps | twb

v

Equation A14

Using W

r

Equation A15

Using v

pw = pws | td


Using td

W

Equation A12

Using Pws | td for Pw

v

Equation A14

Using W

r

Equation A15

Using v

Pws | tdb


Using tdb

Pw

Equation A13

Using pws

W

Equation A12

Using pw

v

Equation A14

Using W

r

Equation A15

Using v


A PPENDIX B - D RIVE L OSS C ALCULATIONS

• Power transmission losses must be considered in energy consumption analysis whenever a direct-drive system is not used. The types of drive systems with losses include hydraulic and gear drives, belt drives (including V-belts and rubber chain), and variable-speed drives including eddy current clutches and electronic and mechanical variable-speed devices. • Determining precise drive losses involves laboratory testing procedures. However, using the methods in this appendix, losses for the common drive systems can be estimated with suitable accuracy. V-belt Drives

• Expressed as a percentage of motor output, these losses diminish logarithmically as the motor size increases. In addition, there is a range for each motor size where typically the losses increase as speed increases. The value of the drivebelt loss can be determined by the graph in AMCA Publication 203-90, or the mean drive loss can be determined by the following equations: • EQUATION Bl: Fractional horsepower motors: LD =9.4-4.651 27 In P M • EQUATION B2: Motors from 1 to 10 horsepower: LD = 9.4 -1.867 47 In P M • EQUATION B3: Motors from 10 to 100 horsepower: LD = 6.2 - 0.477 724 In P M • 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 Chapter 16: Appendices

115

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

= motor output power, bhp (kW) = fan shaft input power, bhp = drive loss in percent of rated motor output power, % = conversion factor, 1.0 (0.745 70)

Rubber Chain Drives

• Relatively new drive method that is more efficient than V-belt drive systems. The drive pulleys are ribbed and the belt is toothed to prevent slippage. Drive losses occur due to the bending forces as the belt rotates around the pulleys. Ask the manufacturer 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 power factor of their electronic variable-speed drives as a function of output speed at constant load and, in the case of fan systems, as a function of output speed with load reducing with the cube of the speed change. If the manufacturer does not supply this data, the part-load drive efficiency can be determined by multiplying the full-load design drive efficiency by the part-load correction factor. • EQUATION B6: cfv =1.0 + (0.203 176 x 1 n Nfv) where: cfv

116

= variable-torque drive correction factor, dimensionless ratio


Nfv

= speed fraction of variable-speed drive, dimensionless ratio

• EQUATION B7: Variable-speed drive efficiency. h V = hvd x cfv

100 where: h V hvd

= part-load drive efficiency, dimensionless ratio = drive efficiency at full-load design, %

• The part-load efficiency of the motor must be considered to determine the input power to the variable-speed drive. Data on motor part-load efficiency at reduced speed may be difficult to obtain. In the absence of this data use the part-load efficiency at full rated speed. • EQUATION B8: Variable-speed drive input power. P Vi = PFi /h V hD hM where: Pvi PFi h V hD hM

= variable-speed drive input power, hp (kW) = fan-shaft input power, hp (kW) = variable-speed drive efficiency, dimensionless ratio = drive efficiency, dimensionless ratio = motor efficiency, dimensionless ratio

• For installed systems where the input load and power factor can be measured, the output of the variable-speed drive can be determined by; • EQUATION B9: For single - phase power: P Vo = (E I pf hD)/cf


117

• EQUATION B10: For three - phase power: P Vo = (√3 E I pf hD)/cf where for both equations: P Vo = variable-speed drive output power, hp (kW) E = average of the measured phase volts I = average of the measured phase amps = conversion factor, 745.70 ( 1,000) pf = power factor, dimensionless ratio hD = 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 not available, then the actual fan input power must be determined by laboratory tests.

118


A PPENDIX C - F AN O UTLET L OSS C OEFFICIENTS ( REF.ASHR) Table C1. Plane Asymmetric Diffuser at Fan Outlet Without Ductwork

C0 A1 / A2

q

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

Figure C1. Plane Asymmetric Diffuser at Fan Outlet Without Ductwork


119

Table C2. Pyramidal Diffuser at Fan Outlet Without Ductwork

C0 A1 / A2

q

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

Figure C2. Pyramidal Diffuser at Fan Outlet Without Ductwork

120


Table C3. Plane Symmetric Diffuser at Fan Outlet With Ductwork

C0 A1 / A2

q

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

Figure C3. Plane Symmetric Diffuser at Fan Outlet With Ductwork


121

Table C4. Plane Asymmetric Diffuser at Fan Outlet With Ductwork

C0 A1 / A2

q

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

Figure C4. Plane Asymmetric Diffuser at Fan Outlet With Ductwork

122



C0 A1 / A2

q

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



123


C0 A1 / A2

q

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


124


Table C7. Pyramidal Diffuser at Fan Outlet With Ductwork

C0 A1 / A2

q

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

Figure C7. Pyramidal Diffuser at Fan Outlet With Ductwork


125

C H A P T E R 1 7

CONVERSION TABLES

From

To

Multiply by

ûF

ûC

TûC = (t F 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

û

Chapter 17: Conversion Tables

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

1m

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

1 kg

2.2046 Ibm

mass

1 Ibm

0.453 59 kg

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

1I

0.219 97 gal. Imp.

1 gal. US

3.785 4 I

1I

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

1N

0.224 81 Ibf

1 N.m

0.737 56 Ibf.ft.

force

1 Ibf

4.448 2 N torque

1 Ibf.ft. 128

1.355 8 N.m


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 in.Wg(1)

248.84 Pa = 0.036 psi

1 kPa

4.018 6 in.Wg

1 in.Hg(1)

3.376 9 kPa

1 kPa

0.296 12 in.Hg

(1) (1) (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

1J

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

Note:

1 Water

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

Chapter 17: Conversion Tables

129

• Atmospheric pressure (standard) 14.7 psia, 101 kPa psia = 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 )

130


C H A P T E R 1 8

ABBREVIATIONS

AND

S YMBOLS A BBREVIATIONS

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

K w

=

specific sound power level, dB re 1 pW

LD

=

drive loss, %

lf

=

load factor Chapter 18: Abbreviations and Symbols

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 pow ower er in inp put to the fan fan,, hp (kW)

PFo

=

air power output of the fan an,, hp (kW)

PMi

=

motor input power, kW

psia

=

pounds per square inch, atmospheric

psig

=

pounds per square inch, gauge

pt

=

fan total pr pres esssure ris isee, in.Wg (P (Paa)

Pvi

=

variab abllee-sspeed in inp put pow power er,, hp hp (k (kW)

PMo

=

mot otor or po pow wer ou outp tput ut to the the fa fan n dri drive ve,, hp hp (kW (kW))

ps

=

fan st stat atic ic pres esssure ris isee, in.Wg (Pa)

pv

=

fan vel elo ocit ity y pressur ure, e, in in..Wg (Pa)

Pvo

=

variab abllee-sspeed output po power er,, hp hp (k (kW)

pws

=

saturation pressure, lb/ft.2, Pa

Q

=

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

R

=

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

r

=

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

rpm

=

revolutions per minute

SF

=

service factor

SEF

=

system effect factor

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)

V a

=

axial velocity component

V b

=

velocity relative to blade

V m

=

mean velocity component

V r

=

radial velocity component

V s

=

absolute velocity

V t t

=

tangential ve 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

%s p

=

percent static pressure for load point

˚C

=

degree Celsius

dB

=

decibel

˚F

=

d e g r e e F a h r e n he i t

g

=

g ra m

hp

=

ho r s e po w e r

S YMBOLS

Chapter 18: Abbreviations and Symbols

133

in.Hg

=

inch of mercury

K

=

k e lv i n

kPa

=

kilopascal (103 x pascal)

L/s

=

litre per second

m/s

=

metre per second

Pa

=

pascal

pW

=

picowatt

˚R

=

d e g r e e Ra n k i n e

rad/s

=

radian per second

W

=

watt

∆t

=

temperature difference

hM

=

mot otor or ef effi ficcie ienc ncy y, di dim men ensi sion onle less ss ra rattio

hD

=

driv dr ivee ef effi fici cien ency cy,, di dime mens nsio ionl nles esss ra rati tio o

h V

=

vari va riab able le-s -spe peed ed dri drive ve eff effic icie ienc ncy y, dim dimen ensi sion onle less ss rat ratio io

hs

=

stat st atic ic ef efffic icie ienc ncy y of of fan fan,, dim dimen ensi sion onle less ss ra rattio

ht

=

tot otal al ef efffic icie ienc ncy y of of fan fan,, dim dimen ensi sion onle less ss ra rati tio o

q

=

p la ne a n g l e

v

=

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

r

=

density, lbm /ft 3 (kg/m3)

∑

=

summation of

f

=

relative humidity

w

=

rotational speed, rad/s

134


C H A P T E R 1 9

BIBLIOGRAPHY

• Air Movement and Control Association (AMCA). Fans and Systems. Publication 201-90. 30 West University Drive, Arlington Heights, Ill. 60004-1893, (708)394-0150, 1990. • AMCA. Field Performance Measurement of Fan Systems. Publication 203-90. Arlington Heights, Ill., 1990. • American Society of Heating, Refrigerating and Air Conditioning Engineers (ASHRAE). Handbook, l988 Equipment Volume. 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 for Mechanical Engineers . 9th ed. New York: McGraw-Hill, 1987.

Chapter 19: Bibliography

135

• Buffalo. High Pressure Type HL Industrial Exhausters. Bulletin FI115A. 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 and Airfoil Single and Double Width . Catalogue Cent. Fab (BI/AF) R. Schofield, Wis., 1990. • Greenheck Fan Corp. Centrifugal Roof Exhausters - Models G and GB. 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. T ubeaxial Fans – Direct and Belt Driv e. Catalogue TAB/TAD 2 R. Schofield, Wis., 1989. • Greenheck Fan Corp. Utility Fans – Forward Curved and Backward Inclined . Catalogue SFD/SFB 3-86 M. Schofield, Wis., 1986. • Greenheck Fan Corp. Vane Axial – Response Control and Preset Pitch Fans . Catalogue VR, VP R M. Schofield, Wis., 1989. • Jorgensen, R. (ed.) Fan Engineering . 8th ed. Buffalo, NY, Buffalo Forge Company, 1983.

136


• Public Works Canada. Metric Conversion Handbook for Mechanical Engineers in the Building Industry . 2nd ed. 1983. • The Trane Company. CDS – Customer Direct Service Computer Selection Program. Vol 10.1. La Crosse, Wis., n.d. • The Trane Company. Centrifugal Fans – Sizes 12-89 Single and Double Width . catalogue PL-AH-FAN-000-DS-6-1083. La Crosse, Wis., 1983.

Chapter 19: Bibliography

137

C H A P T E R 2 0

GLOSSARY

abscissa

• horizontal coordinate of a point in a plane Cartesian coordinate system obtained by measuring parallel to the x-axis. (Compare ordinate.) absolute humidity

• in a mixture of water vapour and dry air, the mass of water vapour in a specific volume of the mixture. Compare relative humidity. 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, also called coefficient of viscosity. absolute zero temperature

• zero point on an absolute temperature scale.

Chapter 20: Glossary

139

adiabatic exponent

• exponent k in the equation pvk = constant, representing an adiabatic change (k is the ratio of the specific heat at a constant pressure 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 of airflow that are steady in time give rise to harmonic excitation at frequencies that are integer multiples of the rotation rate of the fan. TIme excitations of the airflow give rise to random excitation. 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 and distribution 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 fan or blower is termed efficiency. 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 of an arithmetical procedure for evaluating sine X to a stated precision. 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 kilovoltamperes (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.


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 is calculated as the sum of the energy used for all the outdoor temperature bins. It allows heat pump (or other heater or cooler) performance, which is different for each bin, to be accounted 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 approximately that required to raise the temperature of a pound of water from 59˚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 more distribution systems.

142


central fan system

• mechanical indirect system of heating, ventilating or airconditioning, in which the air is treated or handled by equipment located outside the rooms served, usually at a control location, and is conveyed to and from the rooms by means 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 through which air movement is effected by one or more fans. Chapter 20: Glossary

143

counterflow

• in heat exchange between two fluids, the opposite direction of flow; i.e., the coldest portion of one fluid meeting the coldest portion 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 by varying 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 design pressure is greater than the maximum working pressure. dew point

• temperature at which water vapour has reached saturation point (100% relative humidity).

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 in ISO 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 jet impinging on a surface. equivalent length

• resistance of fittings or appurtenances in a conduit through which the fluid flows, expressed in length of straight conduit of the same diameter or shape that would have the same resistance; 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.


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, wear protection, 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 leaves the 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 and leads 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).

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) per formance curve

• graphical representation of static or total pressure and power input over a range of air volume flow rate at a stated inlet density and fan speed. It may include static and mechanical efficiency 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 power transmission device. fan pressurization test

• test for determining the air leakage of a building using a faninduced pressure difference. fan propeller

• propeller or disc-type wheel within a mounting ring or plate, and including driving mechanism supports for either belt-drive or 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.


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 faninlet 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 the impeller 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.

148


frequency response

• normalized motion response of a fan to a known excitation, expressed as a function of the frequency of the excitation. Is usually given graphically by curves showing the relationship of the response to the excitation (and, where applicable, phase shift 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 and torque. gauge pressure

• pressure above atmospheric pressure. head

• energy per unit mass of fluid divided by gravitational acceleration. In fluid statics and dynamics, a vertical linear measure. Note: The terms head and pressure are often mistakenly 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, or associated with, a building.


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 one newton acting over one metre. J = N·m. 150


kelvin temperature

• SI absolute temperature scale (K), on which the triple point of water is 273.16K and the boiling point is approximately 373.15K ( 1 K = 1˚C). Kelvin is 1/273.16 of the temperature of the 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 occupant acceptance is to be achieved. noise reduction (NR)

• difference between the average sound pressure levels, or sound intensity levels of two spaces – usually two adjacent rooms called the source room and the receiving room respectively. operating load point

• actual system operating capacity at the time of taking an instrument reading.


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 polytropic exponent). 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 the container. prime mover

• engine, turbine, water wheel or similar machine that drives an electric generator. 152


psychrometer

• instrument for measuring relative humidities with wet- and drybulb thermometers. pump

• machine for imparting energy to a fluid causing it to do work, drawing a fluid into itself through an entrance port, and forcing the fluid out through an exhaust port. Main types are air lift, centrifugal, diaphragm, positive displacement, reciprocating and rotary. Rankine temperature

• absolute temperature scale conventionally defined by the temperature of the triple point of water equal to 491.68˚R, with 180 divisions between the melting point of ice and the boiling point 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 a product manufactured in quantity. reactive power

• portion of apparent power that does no work. It is measured commercially in kilovars. Reactive power must be supplied to most types of magnetic equipment, such as motors. It is supplied by generators or by electrostatic equipment, such as capacitors. real power

• energy- or work-producing part of apparent power. It is measured commercially in kilowatts. The product of real power and length of time is energy, measured by watt-hour meters and expressed in kilowatt-hours (kWh).


153

reheat

• application of sensible heat to supply air that has been previously cooled below the temperature desired for maintaining 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-bulb temperature, 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 in equilibrium. 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 the procedures for determining whether each of the requirements is satisfied. stall region

• performance zone where unstable operation occurs, characterized by aerodynamic blockage or the breakaway of the flow 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/ft 3. 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. 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 a gaseous 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 balancing work. 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 class of 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. reduction in fluid or current flow by adding resistance. Chapter 20: Glossary

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. 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 fluid density, 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 to allow recovery of static pressure. Used for metering and other purposes 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. It causes fluid friction whenever adjacent layers of fluid move in relation to each other. 2. property of a fluid to resist flow or change 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 the smallest cross-sectional dimension of the pipe conveying the substance. voltampere (VA)

• basic unit of apparent power. The practical unit of apparent power is kilovolt-ampere (kVA), 1,000 voltamperes.


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 gauge and 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 a manometer. 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 of one volt. P = EI = I 2 R. wet-bulb temperature

• temperature indicated by a psychrometer when the bulb of one thermometer is covered with a water-saturated wick over which air is caused to flow to reach an equilibrium temperature of water evaporating into air, when the heat of vaporization is supplied by the air. • wet-bulb temperature is lower than dry-bulb temperature. This difference 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).

158


OTHER IN-HOUSE REFERENCE GUIDES: • • • • • • •

Adjustable Speed Drives Energy Monitoring & Control Systems Lighting Motors Power Quality Power Quality Mitigation Pumps

COMMENTS:

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Fans Reference Guide

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