Fan Application Guide (Technical Memoranda)

Fan application guide

CIBSE TM42: 2006

Engineering a sustainable built environment

The Fan Manufacturers Association

The rights of publication or translation are reserved. No part of this publication may be reproduced, stored in a retrieval system or transmitted in any form or by any means without the prior permission of the Institution. © 2006 CIBSE/FMA CIBSE is registered charity number 278104 ISBN-10: 1-903287-68-5 ISBN-13: 978-1-903287-68-2 This document is based on the best knowledge available at the time of publication. However no responsibility of any kind for any injury, death, loss, damage or delay however caused resulting from the use of these recommendations can be accepted by the Chartered Institution of Building Services Engineers, the authors or others involved in its publication. In adopting these recommendations for use each adopter by doing so agrees to accept full responsibility for any personal injury, death, loss, damage or delay arising out of or in connection with their use by or on behalf of such adopter irrespective of the cause or reason therefore and agrees to defend, indemnify and hold harmless the Chartered Institution of Building Services Engineers, the authors and others involved in their publication from any and all liability arising out of or in connection with such use as aforesaid and irrespective of any negligence on the part of those indemnified. Typeset by CIBSE Publications Printed in Great Britain by Latimer Trend & Co. Ltd., Plymouth PL6 7PY Cover illustration: Ron Mulholland (Howden Group Ltd.)

Note from the publisher This publication is primarily intended to provide guidance to those responsible for the design, installation, commissioning, operation and maintenance of air moving systems. It is not intended to be exhaustive or definitive and it will be necessary for users of the guidance given to exercise their own professional judgement when deciding whether to abide by or depart from it.

Foreword A person’s health can be seriously impaired by inadequate ventilation — the quality, temperature and movement of the air. Condensation and the growth of mould is a risk to buildings. Air freshness should be controlled with the removal of odours, pollutants and excessive moisture. The essence of the correct air motion is a matter of identifying the correct system, selecting the correct terminations, controls and duct routing and, perhaps most importantly, selecting the correct fan. It is a common perception that the fan is simply an air extraction device; a tool to move air between given points, e.g between the bathroom and outdoors. Fans are not only used to enhance personal comfort and well-being. Many uses of fans go unnoticed such as cooling a personal computer or the special fans used to extract hazardous fumes from industrial processes. This revised and updated Fan application guide considers the principles of air movement and the various fan types available in order to achieve the best results. The authors are all engineers with considerable experience in the fan industry. In the following pages, they discuss the principles and practice of air extract/supply system design, and offer guidance on fan selection to ensure that such systems perform their intended function efficiently. This publication has been produced not for the fan engineer but for the engineer who uses fans. It is aimed not only at specifiers, mechanical services designers and architects but also at those responsible for building services and plant maintenance. The contents are not exhaustive and it is strongly recommended that designers consult the fan manufacturers before finalising the design. If this were done more frequently there would be less need for this guide!

Acknowledgements The CIBSE and FMA gratefully acknowledge Ron Mulholland of Howden Group Ltd. for preparing the illustrations for this publication. The authors gratefully acknowledge the helpful comments provided by the CIBSE referees, John Armstrong, Mike Duggan and Brian Moss. Permission of the Carbon Trust to revise and reproduce material from GPG383 is gratefully acknowledged. Crown copyright material is reproduced with the permission of the Controller of HMSO and the Queen’s Printer for Scotland under licence number C02W0002935.

TM42 Working Group Geoff Lockwood (ebm-papst UK Ltd.) (chairman) Ian Andrews (Nuaire Ltd.) Colin Biggs (Nuaire Ltd.) Ken Butcher (CIBSE) Paul Cowell (ebm-papst UK Ltd.) Ian Davis (Vortice Ltd.) Mike Duggan (Federation of Environmental Trade Associations) Neil Jones (Fläkt Woods Ltd.) Ron Mulholland (Howden Group Ltd.)

Authors and contributors Paul Cowell (ebm-papst UK Ltd.) Geoff Lockwood (ebm-papst UK Ltd.) Ron Mulholland (Howden Group Ltd.) Ian Davis (Vortice Ltd.) Colin Biggs (Nuaire Ltd.) Ian Andrews (Nuaire Ltd.) Dan Hopkins (ebm-papst UK Ltd.)

Editor Ken Butcher

CIBSE Publishing Manager Jacqueline Balian

FETA Technical Manager Mike Duggan

Contents 1

How a fan works 1.1 Axial fans 1.2 Mixed flow fans 1.3 Centrifugal fans 1.4 Tangential flow fans 1.5 High pressure fans 1.6 Summary of fan types

2

Fan laws and system resistance 2.1 Fan laws 2.2 System resistance laws 2.3 Practical facts of the combined fan and system laws

9 9 10 10

3

Fan selection 3.1 Fan characteristic curves 3.2 Electrical supply 3.3 Efficiency 3.4 Noise level considerations 3.5 Air performance of different fan types

11 11 12 12 12 14

4

Installation and system effects

14

5

Fan control 5.1 The need for control 5.2 Savings in fan power 5.3 Sensors and controllers

19 19 19 22

6

Parallel and series operation 6.1 Parallel operation 6.2 Series operation

22 22 23

7

Acoustics 7.1 Acoustic terms 7.2 Noise level 7.3 Human perception of sound 7.4 Fan manufacturers’ data 7.5 Application and installation effects 7.6 Control of noise

24 24 25 25 26 26 26

8

Safety and maintenance 8.1 Installation 8.2 Commissioning 8.3 Operation 8.4 Maintenance

28 28 28 29 29

1 1 2 3 5 5 6

References

30

Appendix 1: Definitions and explanations

30

Appendix 2: Airflow and pressure measurement

32

Appendix 3: Information required for fan selection

38

Appendix 4: Electric motors

40

Index

44

1

Fan application guide

1

How a fan works

Fans can be divided into two main categories: axial and centrifugal. There are some types which are hybrids with a foot in both camps but in general terms they divide easily into these two main groups.

There is a limit to the performance of a flat plate. Enhanced lift can be obtained by introducing a camber or curve to the blade and ultimately by moving to an aerofoil shape. Aerofoils are significantly better due to their lower drag and their ability to make the air stick to the top surface at higher incidence, see Figure 1.2. In other words they generate lift more efficiently. Force on blade

1.1

Axial fans

1.1.1

Principles of operation

Axial fans work by generating aerodynamic lift within the rotating blades. This is not strictly true in that aerodynamic lift forces generated on the blades only act to apply a thrust load on the impeller; it is the equal and opposite force that is important, that is the force imparted by the blades on the air. To obtain a better understanding consider a simple plate in an airstream, see Figure 1.1. Here the flow passing over the plate wing generates a lift force which acts to force the plate upwards. This can only happen if the air is made to deflect from its normal flow path and move downwards. Newton’s Law states that for every action there exists an equal an opposite reaction. In this case the upward lift force on the blade is balanced by an equal and opposite force making the air flow downwards.

Forward motion Force on air Figure 1.2 Forces on an aerofoil in a moving airstream

Note that the blade lift force in itself does nothing to generate flow in a fan; it is the reaction that provides the ability of the blade to deflect the air stream. To get a better understanding, consider what would happen if an aircraft were flying level but only about 10 metres above the ground. As the aircraft flew overhead, a blast of air would be felt due to the downwash of the wing , see Figure 1.3. Forward motion

Aircraft wing Moving air

Force on blade Still air

Forward motion Force on air Figure 1.1 Forces on a flat plate in a moving airstream

An important phenomenon, which is crucial to the success of all such devices, is the Coanda effect. Although Figure 1.1 shows air being deflected by the bottom plate surface, and this will generate a reaction or lift force, it is what happens on the top surface that is important and which provides a major contribution. The Coanda effect, which affects all viscous fluids, acts to make a fluid literally stick to a surface even though there is a natural tendency for the fluid to strip away from that surface. Its effect is pronounced on the upper surface where the air sticks to the surface rather than breaking away. Its effect is not limited to the air local to the blade surface, but acts at significant distances from the top of the plate.

Figure 1.3 Downwash from aircraft wing

The downwash generated by an imaginary series of aircraft flying one behind the other illustrates the principle of a basic fan. Connect the blades to a central hub and rotate the impeller and the result is an axial fan, see Figure 1.4. Moving air

Forward motion

Moving air

Figure 1.4 Air movement for basic axial fan

Fan impeller

2

Fan application guide match any changing traffic direction and are fitted with close-coupled silencers.

Axial fans come in various shapes and sizes including: —

Low pressure high volume ducted propeller: normally used in cooling applications in conjunction with a close-coupled tube or fin type heat exchanger. Some of the blade shapes are quite complex and are designed to minimise noise. These machines generate small pressures and move large quantities of air.

Figure 1.5 Low pressure, high volume ducted propeller fans

—

Vane-axial: used in general heating, ventilating and air conditioning systems in low-, medium- and high-pressure applications where an in-line duct arrangement is an advantage. Downstream air velocity distribution is good due to the incorporation of flow straightening vanes on the output side of the impeller. Also used in industrial applications. More compact than a comparable centrifugal-type fan.

Figure 1.8 Jet fan

—

Variable pitch fans: axial fans are ideally suited for blade pitch adjustment. By so doing the output characteristic can be altered while still retaining high efficiency. Some designs only allow the blade pitch to be adjusted at rest, while others permit blade pitch adjustments to be performed with the fan running at full power.

Figure 1.6 Vane-axial fan

—

Tube-axial: low- to medium-pressure heating, ventilating and air conditioning applications where the air distribution downstream of the fan is not critical. Also used in some industrial applications such as drying ovens, paint spray booths and fume exhaust systems.

Figure 1.7 Tube-axial fans

—

Jet fans: used extensively for road tunnel ventilation when the tunnel length is not too great. The fan produces a high velocity jet which entrains surrounding air and generally ‘drives’ the air through the tunnel. Fans are normally bi-directional to

Figure 1.9 Variable pitch fans where the pitch may be adjusted with the machine at rest

1.1.2

Performance limitations

Axial machines are more suitable for high volume medium pressure duties while achieving very high efficiencies. Figure 1.10 shows a typical performance characteristic for a variable pitch machine and highlights not only the excellent peak efficiency but how the efficiency contour is elongated thus providing a high efficiency across a broad spectrum of duties. However, axial fans do suffer from a severe aerodynamic stall. The characteristics show that, for most blade angles, the pressure falls off rapidly just left of the peak pressure. This is the onset of stall. Inadvertent stall operation for prolonged periods with an axial fan is bad as the cantilever style blades can vibrate leading to fatigue failure. For this reason axial machines must not be operated near to the stall condition.

1.2

Mixed flow fans

Before considering centrifugal fans there are some hybrid types which combine aspects of both axial and centrifugal fans. They are similar to axial fans but with a radial flow

How a fan works

3 Figure 1.10 Performance characteristics for a typical variable pitch machine

Variable pitch axial fan aero performance Hub ratio: 1.73 ll

Sta

Pressure

% 87.5 84.5%

70 65

78.5%

25

30

35

40

50

45

60 55 Blade pitch angle

Flow

component. The impeller flow passage is conical in nature with a larger diameter at the outlet. The rotation imparts both aerodynamic lift from the blades, and a centrifugal component. The flow from the impeller discharges into an annular chamber where stationery vanes reduce the whirl component. Mixed flow fan impellers, which are shrouded, are more robust than axial impellers and have a less severe stall characteristic. Although most units are of the ‘compact’ type and generally small, larger machines have been successfully used up to an absorbed power of 7 MW.

1.3

Centrifugal fans

1.3.1

Principles of operation

Unlike axial fans, which rely on aerodynamic lift from carefully designed aerofoils, centrifugal fans rely on dragging the air round in circles and using centrifugal force to generate air flow. This may seem crude but a welldesigned centrifugal machine can move air efficiently, withstand more abuse and generate considerably more pressure than an axial fan. Impellers can be arranged backto-back to form double inlet machines matching most high volume axial duties. Figure 1.11 shows three different centrifugal fan blade arrangements (Only one blade is shown for clarity.) The

diagrams show a particle of air from when it enters the impeller to when it exits at the periphery at high speed. Centrifugal impellers add energy to the flow stream by accelerating the air, which enters radially, with little or no spin, travels through the impeller, and leaves with a significant spin velocity. (The spin, or tangential component of velocity at the impeller exit, is shown by the green arrow.) The radial bladed impeller appears as though it should, for a given size and speed, produce a greater tangential velocity at the exit than do other types of fan. In theory it should therefore, for the same size and speed, add more energy to the air and produce more pressure. However this is not the case.

In practice, the radial bladed is not as good as it appears because of how the air behaves within the blade passage. Consider what happens as the air enters the impeller inlet in a radial direction and is suddenly struck by the radial blade nose. This is indicated by the sudden change in direction of the red line depicting the particle movement. The air resists this treatment and objects by breaking away from the leading edge of the blade. The backward inclined bladed machines, by comparison, allow a much better flow distribution over the leading edge of the blade and gently introduce the air into the impeller blade passage before imparting spin. For this and other reasons, radial bladed machines are the least efficient, with backward inclined bladed types the most efficient.

Path of air particle Tangential velocity

Single blade rotation (a) Radial blade

(b) Backward inclined blade

(c) Backward curved blade

Figure 1.11 Path of air particle entering impeller of a centrifugal fan

4

Fan application guide Fan discharge casing

Fan inlet

Impeller

Most centrifugal fans are designed to operate within a casing, see Figure 1.12. This casing is important as it is used to recover a significant portion of the kinetic energy in the flow leaving the impeller, converting it to static pressure. However, some are designed for use without a scroll casing in compact applications. Examples of centrifugal machines are shown in Figures 1.13 and 1.14. Backward inclined bladed fans are usually used in large heating, ventilation and air conditioning systems, where the power savings can be significant. They can be used on low-to-medium and high-pressure systems. They are also used in large sizes, up to about 4 m diameter, for nonerosive industrial applications.

Figure 1.12 Cased centrifugal machine

Fan discharge

Centrifugal fans are commonly arranged with back-toback impellers producing what is known as a double inlet arrangement. This produces double the flow of a single inlet impeller at the same pressure. 1.3.2

Performance aspects

All centrifugal machines have a similar shape of pressure–volume characteristic, see Figure 1.15. They do exhibit stall but it is significantly less severe than that suffered by axial machines, which means that they will still deliver almost the same pressure/volume, even when the stall is encountered. The robust nature of the centrifugal fan impeller geometry makes them much more tolerant of inadvertent stall operation. However the fan should not be operated in stall as significant pressure pulsations can exist within the fan casing and duct system. Well designed centrifugal machines can achieve high efficiencies into the high eighties (%). Volume control can be achieved by a number of means. The output characteristic shown in Figure 1.15 is for a typical fixed speed fan fitted with inlet vane control at Stall

Design 86% 83% 80% 75% Pressure

Figure 1.13 Examples of small single inlet centrifugal machines

70%

80°

Figure 1.14 Double inlet centrifugal machine

Centrifugal fan Aerofoil blades Radial vane control Constant Speed

70°

60°

50°

40° 20°

0°

Flow Figure 1.15 Performance characteristics for a typical centrifugal machine

How a fan works

5 Air spinning in the same direction as the impeller at the impeller inlet Impeller rotation Inlet vane control

Air at entry being spun by the inlet vane control

Figure 1.16 Spin imparted to inlet air by inlet vanes

various angle settings. By spinning the inlet air in the same direction as the impeller, high efficiencies can still be maintained even at lower duties, see Figures 1.16 and 1.17.

Improvements in casing design have brought this type into prominence for use in small domestic appliances where the long, narrow rectangular shape of inlets and outlets can be advantageous.

1.4

1.5

Tangential flow fans

Impellers similar to those of multi-vane forward curved centrifugal fans are used. The action of this type of fan is radically different to that of centrifugal fans. A vortex is formed and maintained by the blade forces with its axis parallel to the shaft and near to a point on the impeller circumference, see Figure 1.18.

High pressure fans

Commonly known as blowers these types are generally used for industrial applications where gases are being stored under pressure, or where air is required for aeration in, say, a water treatment plant. Discharge pressures can be quite high at around 100 kPa, however the flows are relatively small. In this type of application centrifugal impellers can be multi-staged to boost the pressure or high speed turbo blowers can be used.

Fan inlet Impeller Forced vortex Fan casing Impeller rotation

Rotational flow field Figure 1.17 Pre-rotation of inlet air by louvred dampers in a fan inlet box prior to entering impeller for part-load duty

Figure 1.18 Tangential flow fan

Fan discharge

6

Fan application guide

Turbo-blowers, see Figure 1.19, which look more like pumps than fans, run at speeds between 5000 and 15 000 rpm. Their impellers are small, usually unshrouded, are machined from solid and incorporate an axial inducer section on the inlet portion.

1.6

Summary of fan types

Table 1.1(1) summarises the characteristics performance of the common fan types.

and Figure 1.19 Typical high pressure fan (‘blower’) and detail of impeller

Table 1.1 Summary of fan types and applications (reproduced from GPG383(1) by permission of The Carbon Trust; Crown copyright)

Highest efficiency of all centrifugal fan designs. Air leaves the impeller at a velocity less than the tip speed and relatively deep blades provide for efficient expansion within the plate passages; 10–16 blades of aerofoil contour curved away from the direction of rotation. For a given duty, this is the highest efficiency machine.

Peak efficiency of up to 90% occurs at 50–65% of fullopen volume. The high efficiency zone also corresponds to a stable area of the operating curve. The absorbed power becomes lower towards a free delivery. The power curve is of a non-overloading type.

Efficiency is only slightly less than that of the aerofoil fans. Backward-inclined or backward curved blades are single thickness. 10 to 16 blades curved or flat inclined away from the direction of rotation. Efficient for the same reasons as the aerofoil fan given above.

The operating characteristics of this type of fan are similar to the aerofoil design mentioned above. Peak efficiency, which can be up around 85%, is only slightly lower than the aerofoil design. Power curves are normally of the non-overloading type.

Centrifugal backward inclined plate blade

General use for heating, ventilating and air conditioning (HVAC) systems. Usually applied only to large systems where the power savings can be significant. Can be used on low-to-medium and high pressure systems. Used in large sizes up to around 4 m diameter for non-erosive duties in most industrial applications.

Same HVAC applications as the aerofoil fan. Also extensively used in industrial applications. Can cope with medium erosive gases and in very onerous applications it can even be fitted with sacrificial erosion liners.

Fan characteristic*

Pressure 100

E

80 60

P

40

Efficiency

Centrifugal aerofoil

Applications

20 0

Volume flow rate

Pressure 100

E

P

80 60 40

Efficiency

Performance characteristic

Pressure and power

Efficiency

Pressure and power

Fan type

20 0

Volume flow rate

* E = efficiency; P = air power

Higher pressure characteristics than the above fans due to the higher velocities of the fluid leaving the impeller. The pressure volume curve may have a break just left of the peak pressure, but this is usually not sufficient to cause difficulty. Efficiencies are normally less than 70%. The power curve is of the overloading type so care must be taken when selecting the driving motor.

Used primarily for handling gases where particulate matter is present and has a tendency to adhere to the impeller blades. Has a selfcleaning advantage when used in handling gases with a particulate content. The flat bladed versions can be easily protected with sacrificial erosion liners. Not commonly found in hvac applications.

Pressure 100 80

E

60 40

P

Efficiency

Simplest of all the centrifugal fans and also least efficient. Has the highest mechanical strength and the impeller is easily repaired. For a given duty point this fan requires medium speed. 6 to 16 blades, shrouded or unshrouded.

Pressure and power

Centrifugal radial or radial-tipped

20 0

Volume flow rate

Table continues

How a fan works

7

Table 1.1 Summary of fan types and applications — continued Performance characteristic

Forward curved

Efficiency is less than an aerofoil or a backward curved plate bladed fan. The impeller is usually fabricated in lightweight and low-cost construction. The impellers normally have between 24 to 64 shallow blades with both the inlet and the outlet edges curved forward in the direction of impeller rotation. The velocity of the air leaving the impeller is normally higher than the peripheral speed of the impeller with the result that this design has the smallest impeller for a given duty. The impeller is limited in mechanical strength and is normally only used for lowspeed applications and for small sizes.

The pressure volume curve is less steep than that of a backward-curved impeller and there is a depression in the pressure to the left of the peak pressure. The power curve is of a constantly rising type and as such care must be taken when selecting the drive motor.

Axial entry mixed flow blades providing a centrifugal output flow, normally into a volute casing. Machined from steel or aluminium these impellers are normally unshrouded. Impellers are normally mounted on the high-speed output shaft of a gearbox and generally run at speed in excess of 3000 rpm.

High-pressure development can be combined with high efficiency. Normally fitted with either close-coupled inlet or discharge variable geometry flow control vanes. The power characteristic is of the overloading type so care must be taken when selecting the driving motor and in designing the drive train. The unstable region on the output characteristic, commonly referred to as surge, produces severe duct pulsation and should be avoided.

60

E

40

P

20 0

Volume flow rate

Surge

100

E

80 60

P 40 20 0

Volume flow rate

Although it is possible to reach efficiencies and noise levels comparable to those of a backward-curved centrifugal fan with a more compact in-line casing arrangement, these fans are not very popular due to the relatively high costs and the limited flexibility in output duty.

P ressure 100 80

E

60

P

40 20 0

Volume flow rate

High flow rate but very lowpressure capabilities with maximum efficiency of around 60% being achieved close to free delivery. The discharge from this type of fan has a large rotational or swirl component due to the action of the blades and the fact that there are no flow straightening vanes.

For low-pressure highvolume air moving applications such as circulation within a room space or ventilation through a wall without the use of ducting.

100 80

P ressure

E

20

* E = efficiency; P = air power

High flows rate, low pressure capability with a maximum efficiency of around 70%. Due to the lower rotational speeds the rotational flow-field after the impeller is minimised.

Normally used in all types of cooling applications in conjunction with a closecoupled tube or fin type heat exchanger.

Pressure and power

l

The efficiency of this type of fan is much improved by the incorporation of a close fitting shroud and inlet bell. In addition complex aerodynamic blade designs are used and commonly overlap on some arrangements. Blades are normally fabricated from plastic and/or reinforced glass fibre. Fans can be from 50 mm up to 15 m diameter, but at this large size they generally operate at low speeds, normally less than 100 rpm.

60 40

P

Plate-mounted axial flow/partition

Efficiency

Pressure and power

P ressure

Efficiency

A significant part of the pressure is developed by the centrifugal action and static pressure generating capacity is higher than an axial fan rotating at the same speed.

This design is an extension into the fan field of the high-strength, high tipspeed designs used for turbo-compressors, finds a place in the high-power applications of heavy industry. A typical application would be aeration of fluid ponds in the water-treatment industry.

Efficiency

Efficiency is low. Impellers are usually of inexpensive construction and limited to low-pressure applications. Impeller is usually of 2 or more blades, usually of single thickness and attached to a relatively small hub. Energy transfer is primarily in the form of velocity pressure.

80

0

fl

100 80

E

P ressure

60 40

P

Efficiency

Propeller

Similar to an axial fan but with a radial flow component. The impeller flow passage is conical in nature with a larger diameter at the outlet. The rotation imparts both aerodynamic lift from the blades, and a centrifugal component. The flow from the impeller discharges into an annular chamber where stationery vanes reduce the whirl component.

100

Pressure

Pressure and power

Mixed flow axial discharge

Used primarily in lowpressure HVAC applications such as domestic heating systems, and air conditioning units

Fan characteristic*

Pressure and power

Mixed flow centrifugal exit

Applications

Efficiency

Efficiency

Pressure and power

Fan type

20 0

Volume flow rate

Table continues

8

Fan application guide

Table 1.1 Summary of fan types and applications — continued Performance characteristic

Tube-axial

Somewhat more efficient than a propeller fan design and capable of developing a more useful static pressure rise. The blade number varies from around 4 to 8 and the hub is usually less than 50% of the blade tip diameter. Blades can be of aerofoil or single thickness cross-section.

High flow combined with medium pressure generating capability. The performance curve shows a dip in the pressure generating capacity left of the peak pressure and operation in this zone should be avoided. The discharge air pattern is annular and exhibits a distinct whirling motion as it leaves the impeller. There are no stationary downstream guide vanes to recover this rotational flow component.

Low to medium pressure duelled hvac applications where the air distribution down-stream of the fan is not critical. Also used in some industrial applications such as drying ovens, paint spray booths and fume exhaust systems.

Well-designed blades permit a medium pressure capability with high efficiency. The most efficient of these types of fans employ aerofoil blades. Blades can be fixed or adjustable at rest or even in motion using mechanical or hydraulic blade pitch adjustment systems. The hub diameter is normally around 50% of the blade tip diameter.

The performance curve includes a dip, caused by aerodynamic stall, to the left of the peak pressure which should be avoided. Special anti-stall chambers can be fitted close to the blade periphery where there is a requirement to operate the fan at lower flow. Downstream guide vanes are used to correct the circular motion imparted to the air by the impeller and improve the static pressure produced. Two impellers can be designed on the same shaft to form a two-stage unit, effectively doubling the static pressure capability.

Used in general hvac systems in low, medium and high-pressure applications where an in-line duct arrangement is of an advantage. Air velocity distribution on the downstream side is good. Also used in industrial applications. Relatively more compact than a comparable centrifugal-type fan.

Impellers similar to those of multi-vane forward curved centrifugal fans are used. The action of this type of fan is radically different. A vortex is formed and maintained by the blade forces and has its axis parallel to the shaft and near to a point on the impeller circumference.

Efficiency is low but the fans are quiet for their duty. To obtain a reasonable efficiency, an adequate outlet diffuser is necessary since most of the static pressure is derived from the conversion of the high velocity pressure leaving the impeller

Improvements in casing design have brought this type into prominence for the use in certain small domestic appliances. They present long, narrow rectangular shape of inlets and outlets.

P ressure 100 80

E

60

P 40 20 0

Volume flow rate

100 80

E 60 40

P

Efficiency

Pressure and power

P ressure

20 0

Volume flow rate

100 80 60

E 40

P ressure P

Efficiency

Cross flow

Fan characteristic*

Pressure and power

Vane-axial

Applications

Efficiency

Efficiency

Pressure and power

Fan type

20 0

Volume flow rate

Many models use aerofoil or backward inclined plate bladed impellers as described above. The impellers are designed to provide high volume flow at very low pressure. Mixed flow impellers, as described above, can also be used.

The fans are normally designed without connecting ductwork and therefore operate against a very low static pressure. Only static pressure and efficiency are shown for this type of fan.

Used for low static pressure exhaust systems such as general factory, kitchen, warehouse, and commercial installations where the low static pressure rise limitation can be tolerated.

100

P ressure

80 60

E

40

P

Efficiency

Used primarily for lowpressure return air systems in hvac applications. The ‘straight-through’ casing design offers some advantages.

Pressure and power

Performance is similar to backward-curved fans, except the pressure-volume characteristic is depressed as a result of the 90-degree change of direction at the impeller exit. The efficiency will be lower than an equivalent backward curved, conventional scrollcasing type of fan. Some designs may show a dip in the pressure curve similar to an axial unit.

20 0

Volume flow rate

100 80

P ressure 60

E P

40

Efficiency

Roof ventilator, centrifugal

This fan usually has an impeller similar to the aerofoil, backward- inclined or backward- curved plate bladed units as described above. Due to the annular casing constraint this type of fan has a lower efficiency than a conventional ‘scrollcasing’ type.

Pressure and power

Tubular centrifugal

20 0

Volume flow rate

* E = efficiency; P = air power

Table continues

Fan laws and system resistance

9

Table 1.1 Summary of fan types and applications — continued Performance characteristic

Applications

Roof ventilator, mixed flow

Multi-bladed shrouded mixed-flow impellers with an outlet flow offering a good match with the weather cowl.

The fans are normally designed without connecting ductwork and therefore operate against a very low static pressure. Only static pressure and efficiency are shown for this type of fan. The fan units generally have a lower dBA noise level compared with axial units and as such are quieter.

Used for low static pressure exhaust systems such as general factory, kitchen, warehouse, and commercial installations where only low static pressure rise is required.

Fan characteristic*

100 80

P ressure 60

E P

40

Efficiency

Efficiency

Pressure and power

Fan type

20 0

A large variety of propeller designs are used where the duty requires high volume flow at low static pressure.

The fans are normally designed without connecting ductwork and therefore operate against a very low static pressure. Only static pressure and efficiency are shown for this type of fan.

Used for low static pressure exhaust systems such as general factory, kitchen, warehouse, and commercial installations where the low static pressure rise limitation can be tolerated.

100 80 60

P ressure 40

E

Efficiency

Roof ventilator, axial

Pressure and power

Volume flow rate

20

P

0

Volume flow rate

These fans are unique in that they do not have a scroll casing. The discharge from the impeller is unconstrained and enters the room or the enclosure directly. Without a scroll casing a significant amount of the kinetic energy in the impeller discharge air stream is not converted to useful static pressure and is lost. The performance of the fan can be affected by its position within the room or enclosure.

Used mainly in air handling units 100 80

P ressure E

60 40

P

Efficiency

Many models use aerofoil or backward-inclined platebladed impellers as described above. The impellers are designed to provide high volume flow at medium pressures. Efficiencies are low.

Pressure and power

Plenum fan, centrifugal

20 0

Volume flow rate

* E = efficiency; P = air power

2

Fan laws and system resistance

Change in fan pressure: pf 2 = pf 1 × (n2 /n1)2 × (D2 /D1)2 × (ρ2 / ρ1)

2.1

Fan laws

Change in fan power:

Using the fan laws, the performance of geometrically similar fans of different sizes or speeds can be predicted accurately enough for practical purposes. Total accuracy would require that the effects of, for example, surface roughness of the fan, the viscosity of the gas and scale effect be taken into account, but for the vast majority of fan calculations this is not necessary. It is important to note, however, that the laws apply to the same point of operation on the fan characteristic. They cannot be used to predict other points on the fan’s curve. The fan laws are most often used to calculate changes in flow rate, pressure, and power of a fan when the size, rotational speed or gas density is changed. In the following laws the suffix ‘1’ has been used for initial known values and the suffix ‘2’ for the resulting calculated value. Change in volume flow: qv2 = qv1 × (n2 /n1) × (D2 /D1)3

(2.2)

(2.1)

Pr2 = Pr1 × (n2 /n1)3 × (D2 /D1)5 × (ρ2 / ρ1)

(2.3)

where qv is the volume flow rate, pf is the fan pressure (total, static, or dynamic), ρ is the gas density, n is the fan rotational speed, D is the impeller diameter, Pr is the mechanical power input to the impeller. (All values in any consistent system of units). When a significant change of density occurs between the fan inlet and discharge, the laws apply to the arithmetic mean of the density and volume. However, for fans operating at pressures below 2 kPa (200 mm w.g.) the above laws may be taken to apply to inlet volume and inlet density. These laws are simplified when a variable is unchanged. For example, when the gas density is constant, the ratio (ρ2 / ρ1) equals 1, and can be omitted from the equation. Similarly, if the diameter is also constant as with an existing fan and (D2 / D1) is 1, this too can be omitted and only the speed variation laws apply. These are as follows:

10

Fan application guide

2.3

Change in volume flow: qv2 = qv1 × (n2 /n1)

(2.4)

p2 = p1 × (n2 /n1)2

(2.5)

Change in fan power: Pr2 = Pr1 × (n2 /n1)3

(2.6)

System resistance laws

As air moves through a ducted system, the energy (pressure) given to the air by the fan is progressively lost by friction of the air against the duct walls, by turbulence at bends, dampers and changes of duct section and by the losses through heaters, filters or other items of equipment in the system. The loss of pressure due to all of these sources, known as the system resistance, is for practical purposes proportional to the square of the velocity at the point of loss. For a fixed system, it may be said that the pressure required to pass a given volume of air through the system will vary as the square of the volume flow rate, i.e. pf ⬀ qv2. Therefore, to double the air flow, a pressure four times greater is required from the fan. This is true only for a constant system and a constant air density. Should the system be altered, e.g. by closure of a damper, then the above laws do not apply.

Fan shaft power, Pr

Similarly, the pressure loss or system resistance will vary directly with air density.

Fan power characteristic

PrB PrA PrC

As noted above, the pressure loss through a fixed air system increases or decreases as the square of the rate of flow. When a fan is connected to the system the flow will stabilise at a rate at which the pressure rise of the fan is exactly equal to the system loss. This is represented by point A in Figure 2.1, where the system resistance (pressure loss) curve is seen to cross the fan pressure– volume curve. This is called the operating point. A change in the system resistance curve such as the dashed curves (caused, for example, by changes of damper position or miscalculation of overall system resistance) will result in different operating points B and C. Although the pressure difference is the same for points A and C, the flow is much less for C. An operating point widely different from the design may lead to reduced or excessive fan power. In a typical fan system a percentage change in fan speed will result in an equal percentage change in air volume handled. This arises because the fan laws for speed change exactly match the system laws for volume flow rate change and the operating point on the fan characteristic remains proportionally the same. It must be remembered, however, that the power taken by the fan will vary as the cube of the speed. Thus, if it is desired to increase the volume flow rate by 10% this can be done by increasing the fan speed by 10% (where facilities for speed change exist) and the power taken by the fan impeller will increase by 33%. The full and broken curves in Figure 2.2 illustrate such a change.

Fan shaft power, Pr

Change in fan pressure:

2.2

Practical facts of the combined fan and system laws

PrA1 PrA

ptA1

C

A

ptB

Fan output characteristic

B

Fan total pressure, pt

Fan total pressure, pt

}

Fan power characteristic at (N+10)% Fan power characteristic at N %

Fan design operating point ptA ptC

33% power increase

ptA

Fan output characteristic at (N+10)%

A1

21% pressure increase

Fan output characteristic at N %

A

System resistance curve for operating point A

System resistance curve for operating point A

10% flow increase qvA qvA1 Volume flowrate, qv

Volume flowrate, qv Figure 2.1 Change in system resistance curve

Figure 2.2 Change in fan speed

PrA1

Fan power characteristic of larger fan

PrA

Fan power characteristic

Fan total pressure, pt

ptA1

Fan shaft power, Pr

11

Fan power characteristic at standard air density PrA PrA1

Fan power characteristic at lower air density

A1

ptA

A

Fan output characteristic with a 10% larger fan

Fan total pressure, pt

Fan shaft power, Pr

Fan selection

ptA

A

ptA1

A1

System resistance curve for operating point A

Fan output characteristic at standard air density

Fan output characteristic at lower air density

System resistance curve for reduced air density qvA qvA1

Fan output characteristic

qvA qvA1 Volume flowrate, qv

Volume flowrate, qv

Figure 2.3 Change in fan size

Figure 2.4 Change in air density

On the other hand, if an increase in flow rate is achieved by replacing the fan by a larger one of the same type it is no longer possible to calculate the precise performance by reference to the fan and system laws only. This is because the intersection of the fan characteristic with the system resistance curve will be a different working point and new determination of both the working point and the power taken will have to be made, see Figure 2.3.

the engineer to make the right choice. When specifying a product the following information will be required: —

the volume flow requirement (and any possible increase that may occur due to design changes)

—

the pressure required to overcome the system resistance

—

the air path (i.e. is the air required to travel in line with the fan (axial) or would a fan type that alters the direction of airflow (centrifugal or mixed flow) be more suitable?)

—

the required electrical supply (voltage, frequency and phase)

—

the efficiency of the fan (if required to be specified)

—

any size restrictions that may exist

—

the maximum acceptable noise level

—

the required control options (speed control, pressure or temperature sensors?)

—

the ambient temperature

Although the volume flow is unchanged the mass flow is changed (being proportional to density). It is important to consider this in heat exchange calculations and for fans at high altitudes. In accordance with the fan laws, the fan power will vary directly as the air density.

—

any special applications (e.g. smoke extract, potentially explosive gases, corrosive atmospheres etc.).

3

3.1

A change in air density alters both the fan performance curve and the system pressure loss curve in accordance with the laws previously stated, viz: —

For a given volume flow rate, the fan pressure is proportional to the air density.

—

System pressure loss is proportional to the air density.

A change of air density from standard to 20% below standard is illustrated in Figure 2.4 where both the fan pressure and system resistance are reduced by 20% the volume flow rate remaining constant. This might occur in high temperature operation or at high altitudes.

Fan selection

In selecting a fan for any given application it is often found that several different fan types and sizes can provide the required performance; however, some prove to be ‘better’ selections than others. It is the responsibility of

Appendix 3 gives a detailed list of the information required to assist in the selection of a suitable fan.

Fan characteristic curves

The graph shown in Figure 3.1 shows a typical fan performance characteristic for an axial flow fan, however, many of the features shown will also apply to other fan types. Fan curves are published with the volume flow rate

12

Fan application guide where f is the electrical supply frequency (Hz) and N is the number of pairs of poles in the motor. For example, for a 4-pole motor running from a 50 Hz electrical supply the synchronous speed is:

Stall region

50 × 60 Synchronous speed = ——— = 1500 rpm 2

Static pressure / Pa

Peak efficiency

(i.e. 2 pairs of poles in a 4-pole motor).

Optimum range

If the same motor is used on a 60 Hz supply the synchronous speed will increase to 1800 rpm. The fan performance will therefore increase as will the absorbed power and noise level. Due to the extra power required to run the fan at a higher speed the maximum permissible ambient temperature is often reduced when running the same product from a 60 Hz supply 0 0

3.3

Volume flowrate / (m3/s)

Figure 3.1 Typical fan performance characteristic

across the x-axis and pressure development on the y-axis; the units for either property can vary but, in the UK, m3/s and Pa (pascals) respectively are the most common. The performance curves published from some manufactures may only show part of this curve in order to influence the fan selection within the best design parameters. It is always important to select the fan as near to its peak efficiency as possible. However, it is also important to note the stall point on the characteristic; when the fan enters the stall range a dip in performance is often seen as the impeller creates turbulence. This is also accompanied with a rise in the noise level and significant drop in efficiency. The turbulent air will also lead to fluctuating stresses in the impeller, which, for some fan designs, can ultimately lead to impeller failure, it is therefore important to avoid this part of the performance characteristic.

3.2

Electrical supply

Many smaller fans are designed to operate from a single phase electrical supply. However, as the required power increases (usually at approximately 1 kW) it becomes necessary to use a 3-phase power supply. 3-phase systems require additional wiring but also offer speed reduction options and soft-starting facilities by configuring the wiring in either a star or delta configuration. (Note: this depends on the motor design and is not possible with all 3-phase motors.) The motor will also be designed to operate at a nominal voltage (with a +/– tolerance), which may vary depending on the intended electrical supply. The frequency of the electrical supply will affect the running speed of the motor. Equation 3.1 gives the synchronous speed in rpm of a motor depending on the number of magnetic poles and supply frequency. f × 60 Fan synchronous speed = ——– N

(3.1)

Efficiency

The efficiency of a fan varies greatly across its operating range, while it is possible to design fans for ‘high efficiency’, it is also vitally important to ensure that the selection of the fan is correct and operating at its highest efficiency. Part L of the Building Regulations(2) specifies the fan efficiency in terms of specific fan power (SFP). The system design has the greatest effect on this quantity as a reduced system resistance will provide a lower SFP. While ensuring a reasonable level of efficiency for the system, the SFP does not necessarily ensure the most efficient fan is being applied to the application.

3.4

Noise level considerations

When comparing noise levels of fans it is essential to check that the same test methods are used in order to ensure that the figures are measured on a like-for-like basis. The main considerations are given below; for further guidance refer to section 7. —

Is the noise level a sound power or sound pressure level?

—

If sound pressure levels are given, at what distance from the source?

—

Has the level been ‘A-weighted’? (‘A-weighting’ adjusts the individual frequency levels to match the sensitivity of the human ear).

—

Are the levels in-duct or non-ducted?

—

Were the measurements taken on the inlet or outlet, or averaged over a sphere or hemisphere?

Comparing the noise spectra for different types of fans, see Figure 3.2, it can be seen that centrifugal designs produce most of their noise at low frequencies whereas axial flow designs generate higher frequency noise. Most people will accept higher levels of low frequency noise and this is one of the reasons why centrifugal fans are generally used when noise is an important consideration. However, higher frequency noise can be more easily suppressed using simple attenuation devices whereas reducing the lower frequency noise requires larger, more expensive attenuators. It is sometimes less expensive to install a high

Installation and application effects

13

0

Fan static pressure

–20

Noise

–10

Static pressure

Octave band sound power level

Region of maximum efficiency Axial fans

Noise

–30 Centrifugal fans –40 62.5

125

250

500

1000

2000

4000

0 0

8000

Mid-frequency of band / Hz

Air flow / (m3/h)

Figure 3.3 Noise compared to peak efficiency

Figure 3.2 General comparison of noise

speed, small diameter axial flow fan fitted with an attenuator, than a slower speed centrifugal fan generating similar noise levels and giving the same aerodynamic performance.

close to the fan to prevent unwanted noise passing along the system. Attenuation can be provided by lining the ducts with absorptive material or by inserting proprietary attenuator units. Duct lining, especially at bends, is satisfactory providing sufficient length of duct is available and thick enough lining is used.

Figure 3.3 shows how fan generated noise varies as the fan duty and emphasises that fans are at their quietest when operating near peak efficiency, and noisiest when running at, or near, the stalled condition.

Where length is limited and the noise to be absorbed is in the high- and mid-frequency range (axial flow fans) then a simple in-line attenuator unit is ideal. A splitter attenuator

If the sound power of a fan is unacceptably high, and no other selection is possible, attenuation must be introduced

5.0

Radial bladed

centrifugal

4.5

4.0

Total pressure / kPa

3.5

3.0 te Pla

al ug trif n e dc de a l b

2.5 be Tu

ax

ial

al

ug

rif nt

e

2.0

l

oi

of

r Ae

ad bl

d ar rw

ug

f tri

ed rv

1.5

al

ce

n ce

cu

Fo

Ax ial

1.0

0.5 Figure 3.4 Comparative fan designs at an equal absorbed power of 10 kW (reproduced from Practical guide to noise control by permission of Fläkt Woods Ltd.)

Ducted propeller

5

10

15

20

Volume flow / (m3/s)

25

30

35

14

Fan application guide

or, more expensive, a packaged bend splitter attenuator, must be used if low frequency noise (centrifugal fans) is to be prevented from entering the system.

The perfect fan installation is one that has non-turbulent air entering the fan inlet and has no restrictions near the outlet. The following are examples of good and bad designs and installations.

3.5

Figures 4.1 showing how poorly fitted flexible ducting will create turbulent air at the fan inlet, increasing pressure loss and reducing the performance of the fan.

Air performance of different fan types

Figure 3.4 shows typical air performance characteristics for the most common fan types and helps to illustrate which fans are more suitable for various system resistances. There are always other combinations available, such as multi-stage axial that can offer significantly higher-pressure development and within the centrifugal fan range that are many different impeller types.

4

Installation and system effects

Figure 4.2 illustrates a number of effects. The sharp bend at the fan entry (Figure 4.2a) creates an area of high turbulence on one portion of the blade reducing its effective working area leading to a loss of performance. The increased turbulence over the blades will result in increased noise levels. A bend upstream in the duct (Figure 4.2b) causes an area of turbulence after the bend and the curve in the duct adds swirl at the fan inlet. An inlet cone can help turn the air into the inlet (Figure 4.2c) and a dividing plane will minimise the swirl in the air at the fan inlet (Figure 4.2d).

Fan performance curves are measured with the fan installed in perfect conditions. The air performance test rig ensures clean non-turbulent air enters the fan inlet and that the air can exit without hindrance. Air turbulence adversely affects the impeller performance. It reduces pressure generation and increases the turbulence across the blades amplifying the noise generation. Fans installed in less than perfect conditions will perform less well than expected and, more than likely, be noisier than expected. Fans move air from one point to another and generate pressure to overcome resistance to flow. Resistance to flow of air is due to the friction of air particles over surfaces and the friction of the air particles moving past each other. This resistance increases with increased air velocity. High velocity and turbulent air has a higher resistance to flow than slow non-turbulent air. Energy is required to overcome this resistance.

Swirl can either be with or against the fan impeller rotation

Noise is caused by pressure fluctuations that act upon the human ear. Turbulent air generates pressure fluctuations and is, therefore, responsible for causing noise. (a)

Bad

(b)

Good

Flare fitted to improve entry

Bad Good Figure 4.1 Good and bad practice: flexible ducting

(c)

Central splitter fitted to eliminate swirl

Figure 4.2 Problems associated with connection to ductwork

(d)

Installation and application effects

15

Figure 4.3 demonstrates that inlet conditions to centrifugal fans are also critical. A sharp turn at fan entry generates areas of high turbulence in the corners leading to higher inlet velocities; the high turbulence will buffet the impeller creating noise and vibration. Use of turning vanes avoids turbulence and provides clean smooth air into the fan inlet. Good outlet conditions are just as important as inlet conditions. Figure 4.4 shows how a sharp outlet produces high turbulence, increasing the system resistance and noise, whereas turning vanes reduce turbulence, save energy and reduce noise.

Cascade turning vanes fitted

(a)

(b) Cascade turning vanes fitted

Optimising a fan installation is often a matter of common sense. Figures 4.4 and 4.5 illustrate the effect of a sharp change in air direction. Significant investments in developing efficient aerodynamic impellers are negated by buffeting, turbulent air generating noise, and vibration. Inlet cones and long tapers reduce turbulence at the inlet, enabling the impellers to operate more efficiently. It is not always possible to use fans in applications or installations providing perfect air flow conditions. The following guidance suggests how ancillary items should be used and how clearances affect fan performance.

(c)

(d)

Figure 4.3 Effect of turning vanes on inlet

If an axial fan is not fitted in a short or long tube casing, its performance can be improved by employing a full bell mouth inlet incorporated in an wall plate. Figure 4.6 illustrates the improved performance of this arrangement compared to a simple aperture. The position of the fan in the wall plate will also affect the fan performance. The fan curve in Figure 4.7 shows how the characteristic is changed with position. The use of a guard at the fan inlet will also reduce the fan performance. The guard adds a restriction to flow and adds turbulence to the air, see Figure 4.8 Obstructions at the fan inlet and outlet will reduce an axial fan’s performance. Figure 4.9 shows how the fan performance will be reduced as the obstruction gets closer. The size and position of inlet cones will improve or adversely affect the fan performance of centrifugal fans. Figure 4.10 (page 18) shows how the fan performance and efficiency is affected by the inlet ring gap dimensions.

Splitters added to bends

(a)

In practice, installed fans will invariably be affected by more than one of the above conditions. In addition there are losses in the system to the passage of air flow, such as the resistance to flow in ducts, resistance to flow through filters and across dampers etc. The combined effect gives a total system loss. Figure 4.11 (page 18) shows static and dynamic losses and gains through a notional system. Turbulent air entering a fan impairs its aerodynamic efficiency. Turbulence also increases the system resistance. As well as causing a loss of fan volume and pressure development, turbulence causes increased noise, vibration and structural stress. All these effects are minimised by reducing turbulence at the inlet and outlet of the fan. (b)

Further information on ductwork installation may be found in the Fan and Ductwork Installation Guide(3).

Figure 4.4 Effect of turning vanes on outlet

16

Fan application guide Figure 4.5 Use of inlet cones and tapers

Inlet cone improves the air-flow

Smooth taper improves the air-flow

1

Wall ring

2

Figure 4.6 Effect on fan performance of not using a wall ring

Aperture

Efficiency

Static pressure

1

2

1 2

0

2

x / D = 7% x

D

Aligned on the inlet side

3

x / D = 0% x

D

Air flow / (m3/h)

Immersed on the inlet side

Figure 4.7 Effect on fan performance of position of a wall ring

2

x / D = –7% x

D

1

3

Efficiency

Projecting on the inlet side

Static pressure

1

0

1

3

0 0

Air flow / (m3/h)

2

Fan control

1

17

Without guard grille

2

With guard grille

Figure 4.8 Effect on fan performance of a guard grille

1

0

Efficiency

Static pressure

2

1

2

0

Air flow / (m3/h)

Figure 4.9 Effect on fan performance of obstructions at inlet and outlet

Obstruction on the inlet side

1 2

D

3 4 5

∞

x/D= x / D = 35% x / D = 18% x / D = 9% x / D = 5%

Static pressure

1

2

3 4 5

0 x

0

Air flow / (m3/h)

1 2 3

D

4 5

∞

x/D= x / D = 35% x / D = 18% x / D = 9% x / D = 5%

Static pressure

Obstruction on the outlet side

2 5

x

0

0

4

3 3

Air flow / (m /h)

1

18

Fan application guide Figure 4.10 Effect on fan performance of inlet ring dimensions

2 3

S / D = 0.4% S / D = 1.0% S / D = 1.4%

2 3

Efficiency

1

D

Static pressure

S

1

2

3

0

0

1

Air flow / (m3/h)

1 2

X

2

D

3

X / D = 0.6% X / D = 0.0% X / D = –0.8%

1 3

Efficiency

1

Static pressure

3

2

0

Inlet louvres

Heater

Air flow / (m3/h)

0

Change shape (contraction)

Fan outlet diffuser

Balance or control damper

Delivery grille

Filter Air in

Ventilated space

Air accelerating

Air decelerating Fan

Ducting, bends, etc

Pressure

Loss in diffuser

Loss across grille

Louvre loss

Pressure increase in the fan

Loss in ducts, bends etc. Exit velocity loss

Filter loss Heater loss

Loss across damper

Loss in change of duct shape

Figure 4.11 Static and dynamic losses through a notional system

Fan control

5

19

Fan control

For every litre/second of air that is unnecessarily moved through a building, around two watts of power are wasted. This is a generalisation, but one that is reflected in the UK Building Regulations(2) where maximum permissible specific fan powers are specified for various ventilation systems in non-residential buildings. For occupied buildings during the heating season, or where air conditioning is used, the energy waste is greater, since the air that is moved will have been heated or cooled to bring it to the building internal temperature before being discharged. This ventilation load, roughly equivalent to 16 W per litre/s, can account for a very significant proportion of the overall conditioning energy requirement for the building. Process fans can be similarly wasteful if their operation is not regulated and controlled effectively, leading to increased overall plant costs. For these reasons it is considered prudent, if not essential, to provide duty control for fans and fan systems.

5.1

5.1.2

Buildings and, to a lesser extent, industrial processes are rarely static systems. There may be significant alterations in the function or capacity required of a fan system during its life cycle. If such changes are planned or anticipated, sufficient flexibility can be allowed for in the system design. 5.1.3

Variation of ventilation rate with occupancy patterns

Many ventilated spaces can be characterised according to the way in which they are used, and the output of the fan system varied accordingly. Typical examples are offices where a constant level of daytime occupancy occurs. In this situation, fans used for ventilation only can be operated at constant rate during working hours and shut down at other times. In meeting or conference rooms and larger public spaces, a more complex situation exists. Occupancy levels may vary widely, often in an unpredictable way. In this type of space, the preferred option is for some degree of demandcontrolled ventilation, where the ventilation rate is varied in response to occupancy levels directly, or to the air quality level in the space.

The need for control 5.1.4

There are many methods of control, and the selection of the appropriate technique depends on both the range of duty required and the frequency with which the changes are to be made. Importantly, the reduction in flow rate should be accompanied by a reduction in power requirement, noise level and component wear rates. Cost, both initial and for ongoing maintenance is also crucial. These issues are discussed below. 5.1.1

Alteration of requirements

System uncertainties

The set-up and proving of the fan and system (system commissioning) is carried out after initial installation. Commonly, the system designer will have made allowance in the pressure drop calculations for system balancing, deviation from design in the installed system due to constructional limitations, and the often unknown effects of interaction between adjacent system components. Additional factors for duct leakage may also be applied.

Variation of ventilation rate with ambient conditions

A building’s fresh air input rate must be maintained at a minimum level relative to occupancy (typically 10 litre/s per person), but internal comfort may be enhanced using different overall air movement rates, e.g. greater velocities in summer or lower levels at night. In most systems it is advisable to provide a suitable degree of local control to account for the variability in what is perceived as comfortable. 5.1.5

Air as a thermal transport medium

Air may be conveniently used to transfer heat between source and sink, in a wide variety of building and process applications. Similar analysis of degree of variation and frequency of change to those discussed above may be applied to such systems.

The fan selected will often have been the next largest ‘standard’ model in the manufacturer’s range. These allowances may combine to produce an overall system flow rate well in excess of that required. Commissioning procedures commonly result in duty reductions of up to 25%.

5.2

Conversely, systems with inadequate design or poor installation practices may have resistances far higher than the selected fan can tolerate, sometimes necessitating costly upgrading or replacement.

It is important to understand that fan and system design may be optimised for a number of different parameters including for example, energy use and whole life cost. The required target must be carefully specified since the resultant fan/system design may vary.

System characteristic estimation is not yet an exact science even if software packages make it seem so. There remains a need to adjust system performance.

Savings in fan power

The power supplied to the fan is used to provide a motive force to the air and to overcome the frictional and turbulence-related energy losses encountered in the distribution and air treatment systems.

Along with other system elements, the methods of reduction of fan duty used must be considered in terms of

20

Fan application guide

the impact that they have on the power requirement of the system overall. In essence, if the load/speed of a fan never alters and is perfectly matched to its required output once installed, then the addition of a control can only reduce the system efficiency. However, these conditions are rarely, if ever, encountered. The methods of duty control commonly employed are considered in the following sections. 5.2.1

Alteration of the fan rotational speed

See Figure 5.1. For a fan installed in a typical building ventilation system, the power absorbed is approximately proportional to the cube of fan speed. Note that other (typically process) load types may have different characteristics.

Where infrequent duty changes are needed, the use of belt driven fans, with high efficiency drive systems is often used for decreasing fan speed and operating at speeds other than those provide by standard motors (note that any increase in speed generally requires reference to the fan manufacturer). Belt/pulley changing requires a degree of skill, and some considerable downtime, these factors being the principal disadvantages together with the power loss inherent in the drive. More frequent speed changes with fixed ratios (usually of 1:2 or 1:1.5) may be achieved at relatively low cost with multi-speed AC motors and basic electrical switchgear under automatic (e.g. time based) or manual control.

In practice, the varying efficiency of motor and drive systems with speed may modify this ‘perfect’ control characteristic.

For smaller induction motors (up to around 2.2 kW output) the possibility exists for variable speed operation using a fixed frequency, variable voltage speed controller, of either a transformer or solid state (triac) based type. The latter tend to be smaller and lower cost but may cause motor noise due to their non-sinusoidal output waveform.

Although larger fans may be driven by a wide variety of prime movers, the majority of fans are driven by electric motors.

A degree of two-speed operation can be achieved in this output range by using a star/delta wiring changeover, which produces an effective voltage control.

Fan shaft power, Pr

In outputs above 200 W, these are typically AC induction motors where the rotational speed is dependant on supply frequency and internal construction and having efficiencies that vary from 60 to 90%+, over the range to 75 kW. In smaller sizes, where AC motor efficiency may be as low as 20%, recent developments in DC motor technology have seen these used with the benefits of lower losses and more stable speed controllability.

Pr1

Pr2 Pr3

50% full speed

Design operating point

Full speed Fan static pressure, ps

Fan power characteristic at full speed

75% full speed

Fan output characteristics

ps1

1 75% full speed

ps2

3

qv3

qv2

Volume flowrate, qv Figure 5.1 Flow control by speed regulation

The motor and/or control may fail if they are not correctly matched. Typically a peak in motor current occurs at around 65% of maximum speed and must be allowed for. Specifiers are strongly advised to seek the approval of the fan manufacturer for any speed control device used, and to ensure that all products comply with the latest legislative requirements. Undoubtedly, one of the most significant technological advances in the past 30 years has been the availability of cost effective variable frequency speed control drives (often known as ‘inverters’), in output ratings from 150 W to several thousand kilowatts. By varying the effective supply frequency to the motor, speed reductions down to 20% may be achieved without significant de-rating. This technique is applicable to a widely available range of suitable motors. These products have made practical a wide range of sophisticated control possibilities and the ability to interface easily with building management systems. The power conversion efficiency of such drives is typically above 96%, although the imperfect output waveform may reduce peak motor efficiency by 1 or 2 %. Motor efficiency may itself reduce significantly at speeds below 75%.

2 50% full speed

ps3

The reason for the stated limitation in motor output is the temperature increase in the motor due both to increased losses and the reduced ability to reject heat as speed decreases.

qv1

There are many manufacturers of these devices, offering many advanced features. A limitation on their use has been the lack of experience of HVAC installers with the technology, particularly with regard to electromagnetic emissions and immunity.

Fan shaft power, Pr

Fan control

21 arrangement that is capable of full automation. Such systems can be controlled down to zero or even negative flow, and can maintain operational efficiency over a wide duty range. The mechanical complexity of these systems tends to restrict their practical range of application to fans of over 1000 mm in diameter.

Pr1 Pr2 Pr3 Pr4 Pr5

ps1

1 2

ps2

3

ps3 ps4 ps5

Fan output characteristics at different blade angles

ing

4

eas ecr

le d

5

de

Bla

ch pit

ang

The disadvantage of these systems is that the maximum capacity of the fan may be limited, and base noise levels can be increased. Fan shaft power, Pr

Fan static pressure, ps

Variable pitch inlet vanes may be used, typically with centrifugal fans, to pre-rotate the air at the fan inlet in the same direction as the impeller. This has the effect of reducing the ability of the fan to impart energy to the air, hence reducing fan output, see Figure 5.3.

Pr1 Pr2

Pr3

Pr4

Fan power characteristic with inlet vanes fully open

Pr5

qv5 qv4 qv3 qv2 qv1

Design operating point

Volume flowrate, qv

Specifiers are advised to check with the manufacturers of inverters regarding exact capabilities, limitations and compatibility (with the motor) issues that may influence selection. Alongside the power reductions achieved by speed control, a large reduction in overall noise level, approximately proportional to the fifth power of fan speed, may be observed. No other control method can match this result. 5.2.2

Fan static pressure, ps

Figure 5.2 Flow control by blade angle regulation (axial fan)

This interaction can be changed by alteration of the fan blade geometry or by altering the direction of the airflow relative to the impeller as it enters or leaves.

Other fan types (mixed flow and centrifugal) have been adapted to incorporate similar adjustment methods.

2

ps3

3

ps4

4 5

s ne

g

sin

clo

Va

qv5 qv4 qv3 qv2 qv1 Volume flowrate, qv Figure 5.3 Flow control by inlet swirl vanes (backward bladed centrifugal fan at constant speed)

5.2.3 Such an approach is exemplified by the variable blade angle adjustment often used on the aerofoil axial flow fan, see Figure 5.2. Within the operational limits defined by flow separation from the blades, these products may provide arguably the most efficient means of duty control across all types of system characteristic, whilst significant noise reductions may also be achieved.

1

ps2

ps5

Alteration of the fan/air geometry

The operational capacity of all fans is determined by the interaction of the fan blades with the air.

ps1

Fan output characteristics at different inlet valve positions

Alteration of the system geometry

Altering the system characteristic as well as, or instead of, the fan may vary the output of a system. Typically, this means using volume control dampers to throttle the airflow essentially by increasing the system resistance, see Figure 5.4

Alteration of the blade angle may be a manual process, similar in complexity to changing a pulley ratio (although great care must be taken not to affect the impeller balance), and may be attempted in situ.

This method of control, whilst it may be low cost and reliable, may not achieve worthwhile energy savings and may increase noise and vibration levels. Some fans however, (multivane forward curved centrifugal units) that display a power characteristic that falls as operating resistance increases, may achieve a reasonable level of energy saving using damper control, see Figure 5.5

Fans for which the blade pitch can be varied whilst the impeller is in motion provide a more sophisticated

Bypass control of fans, essentially recirculating the system airflow through the fan, can be applied where the required

Fan shaft power, Pr

22

Fan application guide

Pr2

5.3

Pr1 Pr3 Pr4

Fan output characteristic

Da mp Fan static pressure, ps

ps1

4

ps2

ers

The increasing use of electronic integration has meant that a large variety of compact and low cost control devices are available to the system designer. Whereas process systems may require specialised devices with well defined accuracy and stability levels, building environmental control systems need to reflect the needs of occupants, most of whom will be operating with different calibration curves!

clo

sin

3

ps3

g

Fan output characteristic 2

ps4

1

Additional pressure loss provided by damper

qv4

Sensors and controllers

qv3

Several manufacturers offer complete fan specific control systems and components including system enablers such as time clocks and occupancy detectors. Variable output sensors for temperature, relative humidity and air quality or carbon dioxide (used as an analogue of occupancy levels) are available for demand controlled ventilation solutions, as are ancillary controls for dampers, heating coils etc. Typically these fan controls operate individually or in zones, and most have the ability to interface with larger building management systems, with most of the individual processing occurring at the fan and simply reporting status to the central system.

qv2 qv1

Volume flowrate, qv Figure 5.4 Flow control by system damper regulation (backward bladed centrifugal fan at constant speed)

As well as sensor accuracy, long term stability is of great importance and ideally the control systems should offer straight-forward means of system set-up, commissioning and calibration.

reduction in the flow handled by the fan would otherwise result in unstable operation. Careful consideration of the fan characteristic is required.

6

Parallel and series operation

6.1

Parallel operation

5.2.4

Multi-stage fans

Typically employed in larger systems, a successful control method is to use multiple fans in switched series/parallel combinations and possibly in conjunction with multispeed motors.

Where two or more fans each receive air from, and deliver air into a common system, they are said to be operating in parallel.

It is essential that the fan manufacturer is made aware of this sort of design requirement, since conditions of unstable operation of some or all of the fans may occur as combinations change.

It would seem that two identical fans designed to run in parallel within a system have a capability of handling twice the amount of air handled by a single fan at the same pressure. However the effect of adding a second fan in

Recirculation zones

Figure 5.5 Air flow through parallel and opposed dampers

Significantly distorted flow profile

Even flow distribution across the duct

Parallel and series operation

23

This effect is shown in Figure 6.1 where the air volume qvA handled by a single fan operating at point A is shown to increase to qvA1 when a second identical fan is introduced parallel to it. The operating point moves up the system resistance line from A to A1. Relative increase in flow is governed by the point of working on the characteristic as shown by reference to two other typical system resistance lines B to B1 and C to C1. Furthermore, if the actual working point of a single fan had been D there would be no increase in flow at all if a second fan were added in parallel. In fact, there would be a serious risk of instability due to hunting in such a case. Note that certain types of fans have characteristics which may, under certain conditions, make them unsuitable for parallel operation. This applies particularly to the zone of operation already referred to above at or near to point D in Figure 6.1 or to any fans operating at or near the peak pressure point of their characteristics. The phenomenon known as hunting occurs when each of two fans in parallel alternately takes a greater or lesser share of the total air volume with a consequent fluctuation in power sharing. If one fan is switched off the other may absorb considerably more power. With several independently driven fans working in parallel, some control over the total flow can be obtained by switching off a suitable number of fans. When this form of control is used, it is essential to provide isolating dampers or non-return valves on the inlet or outlet of all fans to prevent short-circuiting of air through the stationary fans. Dampers are also required when the parallel arrangement is used to allow a stand-by fan to be brought into action quickly when in the event of a failure of one fan, the remaining fans continue to supply into the system. Fans operated in parallel should be of the same type, size and speed, otherwise undesirable performance complications may result. It is strongly advised that the advice of the fan manufacturer be sought when considering the use of fans in parallel.

Operation in this zone not recommended

Fan static pressure, ps

D

psA

Series operation

Two or more fans may be connected together so that the flow passes through each fan in turn. In this arrangement the flow is constant but the pressure is increased by each successive fan. The fans may be either separate machines or a number of impellers on a common shaft. Where separate machines are installed, these can have all impellers rotating in the same direction or successively rotating in opposite directions. Where they are axial type and rotate in the same direction, inter-stage guide vanes must be used. This also applies to more than one impeller, either of the axial or centrifugal type on the same shaft. The guide vanes ensure that each impeller receives its flow with little or no pre-swirl, and that each impeller therefore absorbs approximately the same power. In practice, this arrangement results in each fan producing roughly the same static pressure rise. Thus the static pressure rise for the complete machine approaches the sum of the static pressure rises of each stage, less any inter-stage losses in the connecting channels or ductwork. In the case where the impellers rotate in opposite directions, as is quite common on axial flow fans, the static pressure rise for a pair of impellers can be as much as 2.5 times that for the single impeller running without guide vanes. The effect of adding a second stage to an existing fan is therefore given by moving up the system resistance line from A to A1 in Figure 6.2. The relative increase in flow depends on the point of operation on the characteristic as shown by other typical system resistance lines B to B1 and

Fan 1

B1

Fan 2

B1

Working point for two fans in parallel

2 psA psA1

A1

Working point for one fan psA

B A

C1 C

Fan 1

psA1

Fan 2 A1

Working point for one fan

6.2

Fan static pressure, ps

parallel is not to double the air flow because of the increase in resistance within the system.

Working point for two fans in parallel

Volume flowrate, qv Figure 6.2 Characteristic for two fans in series

C1

B

A C

Volume flowrate, qv Figure 6.1 Characteristic for two fans in parallel

Figure 6.3 Example of contra-rotating axial fans in series

24

Fan application guide

C to C1. Series operation can be used as a method of controlling the flow through a system by shutting down fans as appropriate, but the resistance to flow of those fans not being driven should be allowed for in the calculations and reference should be made to the fan manufacturers.

7

Acoustics

Sound power Sound power is a measure of the strength of a sound source. For example, the sound power of a starting pistol is always the same. How loud someone hears the sound (sound pressure) depends on how close they are to the pistol and if it was fired indoors or outdoors, see Figure 7.1. Sound power is denoted Lw and also expressed in decibels (dB) expressed above a different reference point of 1 pW (picowatt).

Acoustics is a complicated subject. The following gives an overview. For a more detailed treatment refer to Woods of Colchester’s Practical Guide to Noise Control(4) and chapter 5 of CIBSE Guide B(5). All fans propagate sound as a consequence of the work they do. When the sound becomes annoying it is commonly referred to as noise. An understanding of acoustic terms, the source of the sound and how the noise occurs will give an appreciation of the differences between fan types and similar fans.

7.1

Acoustic terms

Sound level Sound level is an ambiguous term and should be avoided. It gives no indication as to whether it is sound pressure, sound power, A-, B- or C-weighting nor the distance from the sound source. Nor does it give any indication of the environmental conditions around the fan, i.e. whether it is in a fully reverberant, semi-reverberant or free field.

ld

sure fie

ing pres

Expand

Figure 7.1 Strength of a sound power source

The sound spectrum defines the sound intensity at various frequencies across a range of frequencies, typically 44 to 22 500 Hz. The actual measurement is divided into octave or 1/3 -octave segments, see Figures 7.2 and 7.3. These bandwidths are described by their band centre frequency. A narrow band spectrum, see Figure 7.4, measures a value for individual frequencies, normally in 1 Hz steps. The data are not published in manufacturers’ catalogues and are not measured for each fan made. It is a technique used 90

Sound pressure

70 60 Level

Sound pressure is what we feel (if the sound is loud enough). It is also what can be measured. However, it is affected by what is around the sound source and the distance to the source. Sound pressure is denoted Lp and is expressed in decibels (dB) above a reference level of 20 µPa. Table 7.1 illustrates the sound pressure for various noise environments.

80

50 40 30 20

Table 7.1 Examples describing the loudness of sound pressure values Sound pressure / dB

10

Typical environment

140

30 m from military aircraft at take off

130

Pneumatic chipping and rivetting (at operators position)

120

Boiler shop (maximum levels)

110

Automatic punch press (operator position)

100

Automatic lathe shop

0

500

1000

2000

4000

8000

16000

80 70 60

70

Loud radio (in average domestic room)

40

60

Restaurant

30

50

Conversational speech at 1 m

20

40

Whispered conversation at 2 m

10

30

—

Level

Kerbside of a busy street

50

0

80 10 0 12 5 16 0 20 0 25 0 31 5 40 0 50 0 63 0 80 10 0 0 12 0 5 16 0 0 20 0 0 25 0 0 31 0 5 40 0 0 50 0 0 63 0 0 80 0 10 00 00 12 0 5 16 00 0 20 00 00 0

Background in television and recording studios

250

90

80

Threshold of hearing

125

Figure 7.2 Typical sound spectrum measured in octave bandwidths

Construction site — pneumatic hammer

0

63

Octave band / Hz

90

20

31.5

1/3rd-octave

band / Hz

Figure 7.3 Typical sound spectrum measured in 1/3 -octave bandwidths

Acoustics

25

100

The resulting figure at (A) will be less than at (B), because sound decays with distance. The only common approach is to consider the total sound power level emanating from the fan (C). This figure is not associated with any distance and is a measure of the total sound energy leaving the fan unit.

dB(A)

80 60 40 20 0

0

2

4

6

8

10

12

14

16

18

20

If distance is brought into consideration, fan manufacturers generally have no choice other than to quote values assuming free field conditions, as they will not have details of its eventual installation.

Frequency / kHz Figure 7.4 Typical narrow band spectrum

to pinpoint any peaks in the spectrum that could be the cause of noise. The frequency and intensity of the peak can give clues to the root cause. A-, B- and C-weighting is a method to describe a spectrum with a single number. The weightings approximate how the ear hears sound. The weighting number is added to the octaves and then the octaves are logarithmically added together. The main weightings are: —

A-weighting: approximate hearing at 55 dB

—

B-weighting: approximate hearing at 70 dB.

—

C-weighting: approximate hearing at 85 dB.

Once true installation effects are taken into account, actual levels will always exceed free field values due to reflections from adjacent surfaces such as floors and walls. Actual installations will be somewhere between the free field and reverberant environments. As an example, consider a 1-metre diameter fan with a quoted sound power level of 100 dB. This fan will give the following sound pressure levels: (a)

80 dB at a distance of 3 impeller diameters (i.e. 3 m in this case)

(b)

89 dB at a distance of 1 metre

(c)

102 dB within a 1-metre diameter duct

(d)

96.5 dB within a reverberant room of approx. 100 m2 surface area and an absorption coefficient of 0.1.

An A-weighting indicates how annoying a noise might be and C-weighting indicates the potential for hearing damage. In practice A-weighting is invariably used. Noise criteria (NC) and noise rating (NR) curves are curves applied to the sound spectra octave measurements. Spectra that fall under a particular curve are said to meet that curve. They cannot be applied to fan data as the ambient conditions, such as reflections are not considered

7.2

Noise level

Fan manufacturers provide sound ratings in various ways such as average sound pressure level in free field at a distance of 3 impeller diameters (A) from the fan inlet or discharge, or average sound pressure level in free field at a distance of 1 metre (B) from the fan inlet or discharge, see Figure 7.5.

These differences vary with the size of the fan as (a) and (c) especially are size related.

7.3

Human perception of sound

Human speech is mainly in the 1–4 kHz range, so human hearing is most sensitive in this range at lower sound levels. At higher sound levels, the ear has approximately the same sensitivity over the whole frequency range, i.e. a ‘flat’ response. The ear can hear over a very large range of sound pressure levels, from a quiet whisper to the roar of a jet engine. In order to accommodate this range, and partially mimic how sound is experienced, a logarithmic scale is used. The type of log scale is called a decibel scale: 10 log10. This gives: Lw = 10 log10 (Ps1 / Ps0) Lp = 10 log10 (ps12 / ps02) = 20 log10 (ps1 / ps0)

A

B

C

Figure 7.5 Sound figures at different points within an fan system

(7.1)

(7.2)

where Ps1 is the sound power being measured (W) and Ps0 is the reference sound power (10–12 W), ps1 is the sound pressure being measured (Pa) and ps0 is the reference sound pressure (2 × 10–5 Pa). Pure tones are seen as more damaging and more unacceptable, so there is sometimes a weighting added to them. Typically 5 dB would be added to any octave band level that has a pure tone before adding the weighting and summing the octave bands.

26

Fan application guide

7.4

Fan manufacturers’ data

Fan manufacturers may quote sound power or sound pressure (at a distance of 1 m or 3 m from the inlet or the outlet), a linear figure or a weighted figure, a free field measurement or one taken in an exhaust duct. Every method is different but they all quote a dB figure. It is important to understand which one is being quoted and understand that often there is no simple correlation between the different methods. Therefore they cannot always be compared.

7.5

Application and installation effects

—

Reverberant conditions: conditions occurring in a room with hard reflecting surfaces, e.g. a bathroom, result in the sound reflecting, combing and mixing and taking a longer time to decay. The result is a higher noise level.

—

Mechanical vibration: sound transfers as mechanical vibration into housings, support members and structures. Added to this is the vibration due to the rotating imbalances of the impeller and motor. Mounting arrangements will transmit this to other parts of the system or building and could later return as sound into the environment.

—

Breakout noise: sound from the fan will enter the duct or application. Some of the sound can break out of the duct or application. The breakout noise is a function of the internal sound power, internal area, wall material, thickness and length of the duct or application.

—

Peak operating conditions: the fan will have a peak operating condition where the efficiency is at its highest and noise is at its lowest level. Operation away from this point will lead to increased noise.

—

Poor inlet and outlet conditions: fan manufacturers measure the sound with clean and free inlet and outlet conditions. Inlet obstructions cause turbulence that increases the broadband noise. Inlet obstructions can also lead to blade passing problems, where a relatively few blades pass a fixed object causing fluctuating pressure waves that produce an annoying tonal effect. This also applies to outlet conditions.

—

Poor ductwork: noise is also generated by the airflow as it is obstructed or turned along ducts or within applications. Sharp bends, sudden changes in area, inlet/outlet grilles, louvres and filters, and high velocity air can all add to the overall noise level

These include the following: —

—

—

Free field conditions: fan manufacturers normally measure the acoustic performance of their fans in free field conditions, i.e. those occurring in free space, in a room with sound absorbent surfaces (anechoic or semi-anechoic room), where the sound is effectively deadened by its rapid dispersion and absorption, see Figure 7.6. The same sound levels will not be produced by fans installed in other acoustic environments. They will, therefore, differ from the manufacturer’s data. Sound transmission: fan noise will be transmitted by the following mechanisms: (a)

radiation from the fan inlet

(b)

radiation from the fan outlet

(c)

mechanical vibration in fan housing, mounting and connected structures, which is then transmitted as noise

(d)

mechanical vibration in fan housing, mounting and connected structures, which is transmitted as sound

(e)

break-out of sound through inlet and outlet ducts.

Increased sound levels: the above may increase the sound level measured in free field conditions. The following will also increase sound levels: (a)

not using the fan at its peak operating efficiency

(b)

poor fan inlet conditions

(c)

poor fan outlet conditions

(d)

poor design of duct work and associated equipment.

The above installation deficiencies can result in a noisy system, even though the fan is ‘quiet’. Consider a ventilation system such as that shown in Figure 7.7. Noise will emit from the fan inlet and outlet of the duct (3). It will also transfer through walls windows and ceilings (1). Vibrations from the fan can transfer though ceilings and roofs and emit into the room (2). Noise will also break out through duct walls (4). All these will combine to produce an increased noise level, with reflections off walls (2), further adding to the problem.

7.6

Control of noise

There are a number of ways to control unwanted sound that utilise acoustic enclosures and attenuators. These expensive solutions may not be necessary if the source of the sound is considered prior to installation.

Figure 7.6 Free field and reverberant room

Noise is unwanted sound and sound is pressure waves propagating from the source. Sound is the result of fluctuating forces, variations in pressure such as turbulent airflow within a duct. A fan is a moving machine which generates a pressure difference from inlet to outlet and in doing so causes many fluctuating forces due to the turbulent flow at the inlet, through the fan and at the outlet.

Acoustics

27 Figure 7.7 Noise aspects of a typical ventilation system 3

1

2 1

4 2

3

3

3

3

3

2

One way of reducing the sound source is therefore to reduce the turbulence produced by the fan. This accounts for the difference in noise between different types and makes of fan. Another source of noise is high turbulent flow. This noise source is not generated by the fan but rather by the velocity of the air within the application or within the ducts. There are other effects and components within the fan assembly that can create noise issues. These typically cause peaks in the sound spectrum that the human ear can register as an annoying noise. Most fans cause noise related to the blade passage frequency. This is a function of the number of blades in the impeller and its rotational speed. Sometimes it is caused by fixed objects being placed too close to the rotating impeller at its inlet or discharge. It can also be caused by the disturbed flow coming off the preceding blade, as it rotates. The resulting discrete noise peak can easily be noticed above the otherwise general broadband levels. If the fan manufacturer is supplying complete fan units, then there is less likelihood that this will occur, however every endeavour should be made to present the fan impeller with uniform, turbulent-free conditions within the installation. Mechanical vibration of elements of the fan can also breakout as noise. The electric motor generates noise which could be higher than the sound level produced by the fan. There are also noise effects resulting from poor application of the motor. If the motor is used significantly away from its peak efficiency then electromagnetic

imbalances result in mechanical vibrations at certain frequencies. This is particularly true for single phase motors. Electric motors and fans cannot be perfectly balanced, there is always a residual imbalance. Depending on the degree of balance selected by the manufacturer, this can create a vibration at a particular frequency which can break out and peak above the broadband noise in the noise spectrum. Various methods of speed control can cause undue noise. Mechanical dampers used to restrict the amount of airflow can also create turbulence resulting in noise. Variable voltage motor controllers generally work by chopping the current wave form into the motor. This can create mechanical vibrations in the motor. Variable frequency drives provide a variable supply frequency to the motor by pulse width modulation (PWM). This PWM waveform is typically at a high frequency and can cause motor and fan vibrations. Methods of control should therefore consider the possible affects discussed above. The following strategies for minimising noise should be considered: —

Select a fan to operate at its peak efficiency.

—

Reduce the speed of the fan as much as possible to reduce the sound generated by the impeller.

—

Keep air velocities as low as possible to minimise turbulence.

—

Ensure smooth transitions and generous bends in duct work to minimise turbulence.

28

Fan application guide

—

Avoid obstructions in front of and after the fan.

—

Ensure the motor has been selected to operate at its peak efficiency.

—

Consider the mechanical imbalance of the electric motor and impeller and if necessary isolate the motor and impeller from the installation.

—

Use output filters on speed control devices or consider isolating motor and fan from the installation.

Noise attenuation can be achieved in a number of ways. It could be an acoustic enclosure around the fan and this enclosure could even be a plant room. Sound attenuators can be used on the inlet and outlet of the fan and sound

8

Safety and maintenance

Installation, commissioning, operation and maintenance must be undertaken by suitably qualified personnel and local regulations applied.

8.1

Installation

Reference should be made to the instructions supplied with the fan to ensure a correct and safe installation. Ancillary items such as isolation mounts and guards should be installed according to the instructions so that the fan is securely mounted and that contact protection is ensured.

Discharge silencer

Noise propagation into the plant room

Open inlet (a)

Discharge silencer

attenuation cladding can be used on the application or duct work, see Figure 7.8. The selection of attenuators and attenuation material depends on the frequency of the sound that needs to be attenuated and the amount of attenuation required.

Noise propagation into plant room from duct and casing ‘breakout’

Refer to section 4 of this guide for advice regarding fitting of ducting, and inlet rings or wall plate if supplied as separate items, to optimise performance. Large fans often have access points. If access points are added during installation ensure the correct type are used; cam or hinged type access doors are suitable for low and medium pressure and screwed-down type should be used in high pressure installations.

8.2

Commissioning

The fan will also be supplied with information to assist commissioning and these should be followed to complete a safe installation. The following are key points to check during commissioning:

Inlet silencer

—

Electrical wiring must conform to the relevant regulations(6).

—

Details on the fan data plate must agree with the supply voltage and frequency.

—

Motor overload protection devices such as thermal contacts and thermisters must be connected to a suitable motor overload protection device.

—

Ensure any loose material is removed from the fan casing before switching on.

—

The tightness of all key fasteners, including bearing and casing location bolts prior to the start of the fan for the first time, or after any maintenance work.

—

Ensure the fan rotates in the correct direction. Most fans will overload their motor if they are rotating in the wrong direction.

—

On larger fans, and where there is a possibility that the fan rotor could ‘turbine’ under the action of ventilation flow, a permanent rotor brake should be provided.

(b)

Fan positioned to minimise exposed duct area for minimising noise propagation into plant room Discharge silencer

Minimal noise propagation into plant room from casing ‘breakout’

Inlet silencer

(c) Figure 7.8 Applications of noise attenuation; (a) attenuation applied at discharge, (b) attenuation applied at both discharge and inlet, (c) fan repositioned to minimise breakout noise from ductwork

Commissioning of air distribution systems is dealt with in CIBSE Commissioning Code C(7).

Fan maintenance and cleaning

8.3

Operation

To guarantee a long and safe operating life of the fan check that it is working at or near to its best operating point. Operation away from this point will increase energy consumption, noise and often increase the operating temperature of the motor. Operation at the extremes of the fan performance may be dangerous. In particular check the following: —

Ensure the fan is not operated in a stalled condition as this can lead to material fatigue and the impeller exploding.

—

The fans should only be operated for short periods of time under extreme low flow to avoid mechanical damage to the impeller, casing, etc.

—

8.4

Do not start and stop the fan at high frequency unless the fan has been suitably designed to do so.

Maintenance

It is essential that any machinery is properly maintained. Fan efficiency can reduce by more than 50% if maintenance is inadequate. The fan will have been supplied with maintenance instructions and these should be easily accessible to assist maintenance. The instructions will give advice as to the frequency of maintenance and the actions required. The following generalised advice is also recommended. —

The equipment must be isolated. The isolator should ideally be locked in position.

—

It is also important to note that even with the removal of power, some fan impellers can have a tendency to rotate, turbine under the action of ventilation. It is recommended that impellers are locked with integral brakes applied, if fitted.

—

Vibration can sometimes loosen the fan mounting brackets often causing them to become misaligned.

—

Fan blades have a tendency to attract dust and such a build-up on can reduce fan performance by up to 30%.

—

Impellers should be regularly inspected for erosion, corrosion, cracking, etc.

—

Allowing dust to build up on grilles also restricts the fan's performance.

—

If dirt builds up on a motor it can act as insulation which causes the motor to operate at elevated temperatures, resulting in shortened motor life can be shortened and motors commonly burn-out of motors.

—

An excessive build up of dust or damage to the fan blade can add imbalance causing excessive vibration. This can lead to fatigue and dangerous failure. Check for excessive imbalance.

Dirt can be removed by cleaning with suitable equipment. Alternatively, fan blades and louvres can be cleaned with a brush and detergent. The fan motor casing must be totally enclosed to prevent water from damaging motor windings. If the motor is not totally enclosed, then it must be removed and cleaned separately before the housing is cleaned. If using a brush or high-pressure washer care must be taken not to damage the impeller blades.

29 Distorted blades will cause an imbalance and as a consequence, the fan life and performance will be significantly reduced. One of the biggest maintenance problems with belt-driven fans is poor belt adjustment. It is essential that belts are adjusted regularly if the system’s maximum efficiency and air delivery are to be achieved. Drive belts should therefore be easy to adjust. When a new belt (or a new fan) has been installed, the belt should be re-tensioned after two weeks in order to take up any slackness resulting from stretching. Subsequently the belt should be checked, and adjusted if necessary, on a monthly basis. Where applicable, all set-screws in bearings must be periodically checked and tightened. The majority of fans have sealed ball race bearings and do not require lubrication. Where greasing of bearings is required, the manufacturer’s recommendations must be adhered to. Ensure that the correct oils and/or greases and quantities are used. It is particularly important that lubricants of the correct working range are used. Oil attracts dust and as a consequence any excess should be wiped clear (e.g. fan blades, see above). Other electrical and electronic equipment that makes up the ventilation system must also be maintained and must not be neglected simply because it is out of sight. Check the electrical and thermal overload protection for the motor, and if the motor has a resettable overload protection check that it is not stuck on (or off!). Motors with automatic reset thermal overload can start without warning therefore the power to the fan must be turned off before reaching into the blades or belt. Fuses should be a tight fit, and it is recommended that circuit breakers be manually activated each month so that any internal corrosion is broken away. Examine wiring and other electrical system components for signs of deterioration, loose connections and damage. Any poor connections or corrosion will result in increased electrical resistance in that component; this can in turn lead to overheating and the increased potential of fire. Control gear should be protected within air-tight enclosures or by locating them in a separate room and utilising remote sensors, however care must be taken to maintain the sensors; dirt can insulate a sensor and affect its accuracy. It is wise to visually inspect such sensors and re-calibrate them periodically. Fan housings can be cleaned and re-painted with a suitable corrosion-resistant paint to prolong their life. It is also imperative that guards be checked for damage or incorrect fitting. Guards should be painted a suitable colour, so that they stand out. A service schedule for all components should be maintained and all service activities recorded. If the system incorporates two fans, one for regular use and the second for back-up use if the primary unit should fail, it is essential that both units be used and that similar levels of maintenance are applied to both units. After cleaning and servicing the fan, a label should be placed on the fan housing with details of the current service date and any relevant contact information. The label should also indicate the date at which the next routine maintenance inspection is due. Most importantly of all, keep a supply of spares — the equipment manufacturer will be able to advise on what spares should be carried.

30

Fan application guide

References

4

Sharland I Woods practical guide to noise control (Colchester: Woods of Colchester) (1972)

1

Energy savings in fans and fan systems Good Practice Guide GPG383 (The Carbon Trust) (2004) (www.thecarbontrust.co.uk)

5

Noise and vibration control for HVAC ch. 5 in CIBSE Guide B: Reference data (London: Chartered Institution of Building Services Engineers) (2001–2)

2

Conservation of fuel and power in buildings other than dwellings Building Regulations 2000 Approved Document L1/2 (London: NBS/RIBA Enterprises) (2006)

6

BS 7671: 2001: Requirements for electrical installations. IEE Wiring Regulations. Sixteenth edition (London: British Standards Institution) (2001)

3

Fan and ductwork installation guide (Reading: HEVAC Association) (1993)

7

Air distribution systems CIBSE Commissioning Code A (London: Chartered Institution of Building Services Engineers) (1996)

Appendix 1: Definitions and explanations A1.1

Pressure

Pressure can be expressed as atmospheric, static, velocity and total. The units of measure are typically in pascals (Pa) or kilopascals (kPa). Atmospheric pressure ( pa )

Total pressure ( pt )

Atmospheric air experiences a pressure caused by the weight of the air above. This is the atmospheric (or barometric) pressure (pa ), and is quite substantial, being typically 100 kilopascals (kPa). If air is blown into a balloon, it can be described as being ‘under pressure’; the ‘skin’ of the balloon applies the pressure. The air inside the balloon experiences a greater pressure than the atmospheric air outside, though the difference will be relatively small, say 105 kPa compared with atmospheric pressure of 100 kPa. The term ‘pressure’ describes both that inside and outside the balloon, assuming the air to be still in both cases. The term absolute pressure is used for clarity. Thus, in the example the absolute pressure inside the balloon is 105 kPa and outside the balloon is 100 kPa. Static pressure ( ps ) The difference between the absolute pressure at the point under consideration and the atmospheric pressure is what is important in fan engineering. This is termed the static pressure and corresponds to the potential energy of the air stream. In the example of the balloon above, the static pressure is 5 kPa (i.e. 105 – 100). Static pressure is shown as positive when the absolute pressure is greater than atmospheric pressure, and negative when it is less than atmospheric pressure. Static pressure can also be described as the pressure exerted against the side of the duct measured at right angles to the direction of flow. Velocity pressure ( pd2 ) Velocity pressure is an important quantity to which all the pressure and drag effects of a moving air stream can be related. What counts is the velocity of the body relative to the undisturbed air. Velocity pressure corresponds to the kinetic energy of the air stream: pd2 = 1/2 ρ v2

The movement of air or gas exerts a force on an object in its path. This is mainly because the pressure on the windward side is greater than that on the leeward side. The air or gas is not stopped by the object, but flows round it. The air is brought to rest at one point on the surface of the object.

(A1.1)

where pd2 is the velocity pressure (Pa), ρ is the density of air (or gas) (kg/m3) and v is the velocity of air (or gas) (m/s).

The sum of the static pressure and the velocity pressure at any point in the air is called the total pressure: pt = ps + pd2 = ps + 1/2 ρ v2

(A1.2)

where pt is the total pressure (Pa), ps is the static pressure (Pa) and pd2 is the velocity pressure (Pa). Volume flow (qv) Volume flow (qv) describes the quantity of air moved by the fan. Units of measure used are: —

cubic metres per second (m3/s)

—

cubic metres per hour (m3/h)

—

litres per second (litre/s).

A1.2

Power

Input power The term input power could refer to the shaft power (Pa) required to drive the impeller, typical for a large fan, or electrical input power (Pe ) consumed by the motor for direct driven integrated fan and motor units. Both use watts (W) or kilowatts (kW) as units of measure. In the case of shaft power, the wattage is derived from the torque required to rotate the impeller and the rotational speed. In the case of an integrated fan and motor, the wattage is a function of the applied voltage and electrical current consumed. It is important to understand what the input power refers to if shown on the fan characteristic. Output power The output power (Pu ) is the energy embodied within the air stream, and is a function of total pressure and volumetric flowrate:

where Pu is the output power (W), V is the volumetric flow rate (m3/s) and pt is the total pressure (Pa).

A1.3

Efficiency

As there are a number of terms and methods that can be used it is important to determine the basis of the efficiency stated in the fan characteristic. Impeller efficiency ( ηr ) Often referred to as ‘fan efficiency’ but is usually derived from the power input to the impeller, and is the shaft power compared to the air power output. Pu V × pt ηr = —– = ——– Pa Pa

(A1.4)

where ηr is the impeller efficiency (%), Pu is the output power (W), Pa is the shaft power (W), V is the volumetric flow rate (m3/s) and pt is the total pressure (Pa)

250 200 150 100 50 0 100

5.0 Stall region

Efficiency

4.0

80

60

3.0

2.0

40 Fan characteristic

1.0

Fan efficiency / %

(A1.3)

Static pressure rise across fan / kPa

Pu = V (m3/s) × pt (Pa)

31 Fan shaft power / kW

Appendix 2: Airflow and pressure measurement

20

0 0

20

40

60

80

0

Volume flow rate / (m3/s) Figure A1.1 Characteristics of a backward inclined aerofoil section fan

Overall efficiency ( ηe ) The overall efficiency considers the impeller efficiency along with the motor and drive efficiency, where the drive is the mechanical connection between motor output shaft and impeller shaft, e.g. belts and pulleys. It is the electrical power input compared to the air power output:

ηe = Pu / Pe = impeller × motor × drive efficiencies (A1.5) where ηe is the overall efficiency (%), Pu is the output power (W) and Pe is the electrical input power (W). Specific fan power ( SFP) The specific fan power is the power consumed at a specific duty and is expressed as watts per litre per second (W/(L/s)). This is an alternative to fan efficiency and would not be quoted by a fan manufacturer as the value varies at each point on the fan curve. It is a term referred to in Part L of the Building Regulations and is a method of limiting the power consumed in building ventilation systems by stating the maximum SFP allowed for a ventilation system. The SFP is the sum of the electrical power consumed by all the fans in a system divided by the volume flow of the system. It also includes the losses in all ancillary units such as variable speed drives.

SFP

A1.4

Σ Pe (W) = ———– qv (L/s)

(A1.6)

Fan curves

Although each fan manufacturer may express the fan characteristic using different units of measure they all will use a fan curve with volume flow on the x-axis and pressure on the y-axis. It is important to determine whether the pressure is static or total. If there are additional data regarding input power and efficiency, it is essential to determine whether the power refers to impeller, shaft or electric input power and whether the efficiency is for the impeller only or refers to the overall efficiency. Figure A1.1 is the fan curve for a typical large fan with additional data for shaft power. Note that the efficiency would in this instance be the fan efficiency described without drive losses above. The stall region is an area of aerodynamic instability caused by major flow separation from areas of the impeller.

32

Fan application guide

Appendix 2: Airflow and pressure measurement Fan manufacturers usually test their products using standardised airways and then calculate and publish the volume flow and pressure for ‘standard air’ (i.e. density of 1.2 kg/m3 and temperature of 20 °C). The test standard, BS 848: Part 1:1997 (ISO 5801: 1997)(A2.1) specifies four basic airway configurations designed to replicate different installation methods. Therefore, when selecting the fan it is important to consider the installation type that most closely resembles the installed condition. The four installation types are given in Table A2.1. Table A2.1 Installation types

With the use of a pitot tube, see below (Figure A2.4), it is possible to measure both the total pressure and static pressures in a duct, the dynamic/velocity pressure can therefore be established and, through calculation, the air velocity can be determined.

A2.2

Measurement of airflow rate

The quantity of air handled by a fan may be measured either at the entrance to or exit from the system or somewhere in the system itself, provided that all the air handled by the fan passes through the chosen section. In order to achieve an accurate measurement it is important that the area chosen should, as far as practicably possible, be free of swirl and in steady flow (avoid bends, dampers, positions close to the fan outlet etc.).

Installation type

Description

Example

Type A

Free inlet, free outlet

Plate mounted condenser fan

Type B

Free inlet, ducted outlet

Supply fan at the start of a duct system

Type C

Ducted inlet, free outlet

Exhaust fan at the end of a duct system

A2.2.1

Type D

Ducted inlet, ducted outlet

Supply or extract fan in a fully ducted system

Measurement by calibrated inlet device

Each of the types listed will provide a slightly different performance characteristic. In general, types A and C are similar, as are B and D. Site testing of products will always introduce inaccuracies in the procedure, often due to ductwork configurations etc. An assessment of the possible errors should therefore be carried out and consideration given to their likely effects on the performance results.

A2.1

Measurement of pressure

Flow measurement at system intake

The most common of these are the conical inlet to BS 848: Part 1:1997 (ISO 5801:1997), see Figure A2.2, and various forms of inlet nozzle with well established entry coefficients. Unless one of these is used it is difficult to make an accurate measurement of rate of flow at this point because of the rapidly contracting pattern of the flow. A traverse by an anemometer over an inlet opening, whether fitted with an inlet grille or not, will give a reading which may be useful for purposes of comparison with another similar intake. However it cannot be relied upon to give an accurate measurement of rate of air flow unless corrected by a factor previously determined for that specific type of inlet by a reliable calibration method.

Static pressure in a stream of flowing air is determined by measuring it in a manner such that the velocity of the air has no influence on the measurement. This is done by measuring it through a small hole or series of holes arranged at right angles to the flow in a surface lying parallel with the lines of flow, see Figure A2.1. The surface must not cause any disturbance to the flow apart from friction. Holes arranged in this manner are referred to as static holes or, when provided with short connecting pipes at the opposite side of the boundary wall, static tappings. A measurement of static pressure is made by connecting a tube from a static tapping to one side of a manometer, the other side being open to the atmosphere.

Figure A2.2 Conical inlet pressure tappings

Measurement by static suction in the entry part of the system If a somewhat lower level of accuracy is acceptable, a meaningful measurement of air flow within about plus or minus 10%, and sometimes much closer, can often be made utilising the static pressure existing just inside the duct entry of a system exhausting from a zone with an area much larger than the exhaust duct.

Figure A2.1 Static pressure tapping

This method makes use of the drop in static pressure which takes place in an air stream when air accelerates.

Appendix 2: Airflow and pressure measurement According to Bernoulli’s Theorem this drop in static pressure is exactly equal to the rise in dynamic pressure, see Appendix 1, which takes place at the same time provided there is no friction in the process. If there is friction the drop in static pressure will be numerically greater than the change in dynamic pressure by the additional loss of pressure due to friction. The pressure in the open area is reduced on entry to the duct to the value of the static suction pressure, which is easily measured. In its simplest form the technique is to drill one or more holes in the wall of the pipe a short distance from the entry, and to measure the pressure by holding PVC or rubber tubing as a snug fit around the hole, and taking the reading with a manometer. Alternatively, if space permits a standard pitot-static tube could be inserted in the duct, lined up into the air stream, and the static section consisting of the small holes in the outer tube connected to the measuring gauge and the reading taken. In both cases the other limb of the manometer is connected to the open area upstream of the duct area (in most cases the ambient atmosphere). Fortunately, a considerable amount of reliable information exists on what is known as the entry coefficient (α) for many types of entry, and using these the airflow rate may be calculated using the static suction pressure previously measured.

33 For air at standard density of 1.2 kg/m3:

–

qv = 1.291 A α √ps

(A2.1)

where qv is the volume flow rate (m3/s), A is the cross sectional area in m2, α is the coefficient of entry and ps is the static pressure (Pa). The entry coefficient α can be defined as the factor by – which √ ps must be multiplied to obtain the equivalent – value of √ pd (i.e. the dynamic pressure) in the flow formula given below in section 2.2.3. Therefore:

√–pd α = —— √–p

(A2.2)

s

A number of typical entry coefficients appropriate to a variety of system intake shapes are given in Table A2.2 from which a direct calculation of air volume flow may be made using equation A2.1. These are not fully comprehensive but give a fair indication of the situations where the method can be applied, particularly when no other more suitable sections for flow measurement are accessible or when the flow pattern at other sections is too disturbed to permit any accurate flow measurement by conventional pitot static tube traverse methods.

Table A2.2 Typical entry coefficients Entry type

Coefficient

Plain end of pipe (rectangular or circular)

Entry type

Coefficient

Exhaust booth

α = 0.72 α = 0.82 Flanged end of pipe (rectangular or circular)

α = 0.82

Sharp edged orifice Plain end of pipe plus small radius elbow

α = 0.62

α = 0.60

Flanged end of pipe plus small radius elbow

α = 0.74 Intake with hood H/D Tapered cone or rectangular-to-round

α ≈ 0.90

0.2 0.4 0.6 0.8 1.0

H

Coeff. α 0.60 0.65 0.67 0.69 0.69

D

Table continues

34

Fan application guide

Table A2.2 Typical entry coefficients — continued Entry type

Coefficient

Entry type

Rectangular duct (mounted on single wall)

Coefficient

Hoods

face area For: ———— ≥ 2 pipe area

α = 0.78

Coeff. α

Included angle 10°–100° 100°–140° 180° Rectangular duct (mounted in corner)

Circ.

Rect.

0.95 0.90 0.82

0.95 0.85 0.82

face area For: ———— = 1.2 to 2 pipe area

α = 0.75 Short taper < 1/2 face diam:

α = 0.95 Long taper > 1/2 face diam:

α = 0.85 Rectangular duct (mounted between three walls)

Wire mesh screen at entry Coeff. α

Free area ratio*

α = 0.73

0.3 0.4 0.5 0.6 0.7 0.8

0.35 0.45 0.52 0.58 0.62 0.66

Inlet louvres Free area ratio*

Coeff. α

0.4 0.5 0.6 0.7 0.8

Free area ratio*

0.30 0.38 0.47 0.55 0.62

Perforated sheet at entry

0.3 0.4 0.5 0.6 0.7 0.8

Coeff. α

0.4 0.5 0.6 0.7 0.8

0.37 0.47 0.56 0.65 0.72

Coeff. α

Free area ratio*

0.20 0.29 0.38 0.47 0.58 0.60

Side entry Ratio*

* Free area ratio = projected free area between louvres divided by gross duct area

B

A

Bench exhaust grille with taper connection W

H

α = 0.82 Number of inlets = n

0.4 0.6 0.8 1.0 1.2 1.4 1.6 1.8

Coeff. α for stated louvre blade angle None

30°

45°

0.26 0.34 0.43 0.50 0.58 0.63 0.68 0.68

0.22 0.27 0.34 0.43 0.47 0.50 0.54 0.54

— 0.21 0.26 0.32 0.39 0.43 0.45 0.45

n×A×B * Ratio = ———— W×H

Appendix 2: Airflow and pressure measurement

2.7 kPa

UNITS

VIEW

ON/OFF

29.7 Pa

29.7 Pa

ZERO

UNITS

STORE

LIGHT

35

ENTER MODE

UNITS

ZERO

VIEW

STORE

ON/OFF

LIGHT

VIEW

ON/OFF

ENTER

ZERO

20.0 Pa

STORE

LIGHT

ENTER MODE

MODE

UNITS

VIEW

ON/OFF

ZERO

STORE

LIGHT

ENTER MODE

(b)

(a)

(c)

Figure A2.3 Measurement of differential pressure; (a) positive pressure (static), (b) dynamic pressure, (c) dynamic and static pressure

For cases where no α-value is published, but an entry loss coefficient (known as a K-factor) is available (and these are widely published in system pressure loss data), the corresponding α-value may be readily calculated. The Kfactor is simply the total pressure loss coefficient in terms of a multiple of the dynamic pressure (corresponding to the mean velocity at the section considered) irretrievably lost at the inlet. The corresponding α-value is given by the following formula:

α=

1

(A2.3)

K +1

Example 1 What is the entry coefficient for a tapered exhaust hood with a loss of 1 dynamic head? From equation A2.3:

α=

1 K +1

1

=

1+1

= 0.71

Example 2 What is the entry coefficient for an inlet louvre with a Kfactor of 1.4 and a free area ratio of 70%. Again, from equation A2.3:

α=

1 1.4 + 1

=

1 2.4

Total pressure is measured by connecting a manometer to a tube with its open end facing directly into the flow. Again, for this measurement, the other side of the manometer is open to ambient atmosphere. Dynamic pressure cannot conveniently be measured directly but can very easily be measured as the difference between total pressure and static pressure by joining the total pressure connection to one side of a manometer and the static pressure connection to the other. This is sometimes referred to as a differential pressure and it is nearly always measured, in fan applications, by a digital manometer, see Figure A2.3. The pitot-static tube, often referred to simply as a pitot tube, is a convenient form of probe for inserting into a duct to provide, in a single instrument, both static pressure holes and a forward facing (total pressure) tube so that static pressure and dynamic pressure may be measured simultaneously on two separate manometers suitably connected. Figure A2.4 shows a section of the modified ellipsoidal nose form which is the only type recommended in BS 1042: Part 2.1: 1983 (ISO 3966: 1977)(A2.2). Static pressure in a duct may also be measured by providing a static tapping in the wall of the duct in a position where the flow is not disturbed by adjacent duct irregularities. It is important, however, that the hole used for measuring static pressure should be small (about 2 mm diameter) and free from burrs inside. This can be ensured by drawing a small fine round file through the hole

= 0.65

If the inlet duct dimensions behind the louvres are 0.5 m × 0.8 m and the static suction pressure is 220 Pa, what is the air flow rate in cubic metres per second?

Facing (total pressure) hole

Static pressure holes

Pointer

From equation A2.1:

–

qv = 1.291 A α √ps

––– = 1.291 × (0.5 × 0.8) × 0.65 √220 = 4.98 m3/s

Total pressure tapping Figure A2.4 Schematic of a pitot tube

Static pressure tapping

36

Fan application guide

backwards after drilling as clean a hole as possible. It is usually satisfactory to wet the square-cut end of a thick walled rubber or PVC tube about 5 to 6 mm bore and hold it hard against the outside smooth surface of the duct over the static hole until the manometer connected to the other end of the tube reaches a steady reading. The size and position of the hole is not so important in a plenum chamber or section of duct where the dynamic pressure is likely to be negligible compared with the static pressure.

A2.2.3

A2.2.2

It is important to carry out this traverse in a straight parallel section preferably not less than six duct diameters or duct widths downstream of any bend, obstruction or abrupt change of section. For this purpose the static pressure and total pressure connections of the pitot-static tube are connected to opposite sides of the manometer which then gives a reading of dynamic pressure (being the difference between static pressure and total pressure). The air velocity at the point of measurement can easily be calculated from the dynamic pressure reading according to the following formula which is for ‘standard’ air of 1.2 kg/m3 density (corresponding to 16 °C and 1000 mb atmospheric pressure and 55% relative humidity):

Flow measurement at system outlet

Measuring by means of a pitot-static tube or an anemometer at the air discharge point is relatively easy if the flow is reasonably uniform and straight but if an outlet grille is fitted it is very difficult to obtain a reliable reading unless a temporary extension of the airway after the grille is fitted to give the flow a chance to steady-up and re-form into a single flowing mass.

0.765 D 0.939 D (b)

where v is the air velocity (m/s). Figure A2.6 shows the relationship between dynamic pressure and air velocity for ‘standard’ air.

D

0.979 D

0.816 D

0.883 D

0.345 D

0.655 D

10 5 pa

×

T 289

×

pa pa + ps

× pd

(A2.5)

where v is the air velocity (m/s), pd is the dynamic pressure (Pa), pa is the atmospheric pressure (Pa), ps is the static pressure (Pa), T is the (absolute) temperature (K).

0.926 D

Position relative to inner wall for stated no. of points or traverse lines 7 6 5 0.074 0.061 0.053 0.288 0.235 0.203 0.437 0.366 0.5 0.712 0.563 0.5 0.926 0.765 0.634 0.939 0.797 0.947

Figure A2.5 Pitot tube transverse patterns recommended for (a) circular and (b) rectangular pattern ducts

1000

Dynamic pressure / Pa

0.288 D

0.074 D

0.712 D

0.5 D

0.563 D

(A2.4)

The expressions (105 / pa), (T / 289) and (pa / (pa + ps)) are corrections for atmospheric pressure, air temperature and duct pressure (all usually quite small) to bring the measured value of pd to the equivalent for standard air. The last of these can be ignored if the duct static pressure (ps) is not more than about 1000 Pa above or below atmospheric pressure. Similarly, on a site test as distinct from an acceptance test, the other two corrections can be

0.061 D

0.437 D

v = 1.291 √– pd

v = 1.291

(a)

0.235 D

In the case of a ducted air system to which there is reasonable access, it is invariably most accurate and most convenient to measure the airflow by making a traverse by pitot-static tube, see Figure A2.3, in conjunction with a digital or inclined manometer. Figure A2.5 shows traverse patterns recommended for circular and rectangular ducts based on the Log Tchebycheff Rule.

For non-standard air conditions the formula becomes: 0.117 D

1°

0.184 D

°± 60

0.021 D

Provided the airflow approach to the grille is reasonably straight and uniform and the grille itself does not give a pronounced directional change to the air flow, an extension piece having a length equal to twice the narrowest dimension of the whole opening will be adequate for this purpose. The method of calculating the volume flow rate is the same as for a traverse in a duct, see section A2.2.3 below.

Flow measurement in a duct

800

600

400

200

0 0

10

20 Air velocity / (m/s)

30

40

Figure A2.6 Relationship between dynamic pressure and air velocity for ‘standard’ air

Appendix 2: Airflow and pressure measurement

37

ignored if they are not likely to affect the equivalent pd value by more than 2 or 3 per cent. When averaging the readings taken on a duct traverse it is strictly correct to average the air velocities at the points (which is equivalent to averaging the square roots of the dynamic pressures). However, where no single dynamic pressure reading is greater than twice any other, the result will not be affected by more than about 1–2% if the average of dynamic pressures is used for calculating the mean velocity. This is then multiplied by the duct area to give the volume flow rate. For certain conditions it is sometimes convenient to carry out a traverse in the same pattern using a rotating vane anemometer. If the airway is so large that the operator needs to stand inside it, the operator always be at least 1.5 m downstream of the plane of measurement and well out of line with the instrument. Such a traverse should be carried out by taking a separate velocity reading at each point and then taking the average of all these readings as the mean velocity for the whole area. For this purpose, an electronic anemometer giving an instantaneous reading of the air velocity is convenient though a mechanical or other type of counting anemometer, requiring a greater dwell time at each point, may be used with similar accuracy.

Figure A2.7 Flow measurement using rotating vane anemometer hood

The point-by-point method is recommended in preference to a continuous traverse over the whole area, as used to be common, because a rotating vane anemometer responds many times faster to an increase in velocity than it does to a decrease in velocity (due to the inertia of the impeller). Thus it tends to read high if moved into and out of areas of high and low velocity. It is important to remember that anemometers require recalibration (usually annually). There are various other air velocity measuring instruments available such as those employing the cooling effect of air movement on a hot wire or other element, or where the air velocity deflects a small vane against spring pressure but all of these are liable to need special calibration which may depend not only upon the actual air velocity but also upon the particular situation (e.g. proximity to duct wall etc.) in which they are to be used. In general, they are unlikely to give as good a result as that obtained by using a pitot-static tube or a rotating vane anemometer. Most are provided with rather short probes and their main use is for measurement of relatively low air velocity in free space or point velocity at outlet grilles. A2.2.4

Flow measurement using a flow measuring hood

Obtaining an accurate measurement of the volume flow rate at an inlet or exhaust grill can be problematic due to the increase in velocity caused by the grill. One solution is to use a flow measuring or capture hood, this is a device that incorporates a fabric hood that fits over the entire grill and channels all of the air through a measuring section, see Figure A2.7. The measuring section can vary depending on the design and may use a rotating vane anemometer, thermo-anemometer or differential pressure meter to determine the airflow rate. Whichever method is used, a simple digital readout is usually provided making

Figure A2.8 Volume flow measurement using static pressure differential across inlet ring

capture hoods a very quick and easy method of establishing volume flow rates. With a centrifugal fan (usually backward curved designs) it is possible to establish the volume flow rate by measuring the static pressure rise into the fan, see Figure A2.8. This can be achieved by using a static pressure tapping in the upstream ductwork and usually a set of four pressure tappings (in order to average the result and reduce errors and fluctuations) on the fan inlet ring. The fan manufacturer should provide a tapped inlet ring and advise the flow coefficient; the following expression can then be used to establish the performance with a reasonable accuracy. The volume flow rate is given by:

—–

qv = k √ Δ p

(A2.6)

where k is the inlet ring coefficient and Δ p is the pressure rise (Pa).

38

Fan application guide

A2.2.5

Fan power and efficiency

required to drive the system rather than impeller input power.

Because of the variety of fan driving arrangements which can be employed (e.g. direct drive with or without coupling, belt drive etc.) it is normal to give fan performance data in terms of power input to the impeller and due allowance must then be made for bearing and indirect drive losses, as appropriate, in determining the required motor rating. An additional margin may also be necessary to accommodate possible variation in working conditions.

For low pressure fans where the fan static or total pressure does not exceed 2500 Pa, kp may be taken as unity. In these conditions equations A2.7 and A2.8 can therefore be simplified to:

The general expression for fan efficiency is given by the formula: qv p t ηt = kp —— Pr

( )

(A2.7)

qv pt ηt = —— Pe

(A2.9)

qv p s ηs = —— Pe

(A2.10)

where Pe is the electrical input power (W).

or: qv p s ηs = kp —— Pr

( )

(A2.8)

where ηt is the fan total efficiency (%), ηs is the fan static efficiency (%), kp is the compressibility coefficient, qv is the inlet volume flow rate (m3/s), pt is the fan total pressure (Pa), pt is the fan static pressure (Pa) and Pr is the power input to the impeller (W) With the introduction of the many different control methods used with fans and the requirements for increased efficiency, it is often most useful for comparison purposes to consider the ‘overall fan/control system efficiency’. This would be defined in a similar way to the above expression but using the electrical input power

Equation A2.9 gives the ‘overall’ total efficiency (ηt) and equation A2.10 gives the ‘overall’ static efficiency (ηs). Guidance for testing fans and calculating performance is given in detail in BS 848: Part 1: 1997 (ISO 5801: 1997): Fans for general purposes. Performance testing using standardized airways(A2.1).

References A2.1

BS 848: Part 1: 1997 (ISO 5801: 1997): Fans for general purposes. Performance testing using standardized airways (London: British Standards Institution) (1997)

A2.2

BS 1042: Part 2.1: 1983 (ISO 3966: 1977): Measurement of fluid flow in closed conduits. Velocity area methods. Method using Pitot static tubes (London: British Standards Institution) (1997)

Appendix 3: Information required for fan selection Certain essential information is required before a fan manufacturer is able to supply equipment that exactly matches the function for which it is intended. In addition, further information, though not essential, may prevent an unsuitable machine being supplied or will ensure that the best selection from a number of alternatives is made to give the cheapest, quietest or most efficient fan. It is, therefore, clearly in the interest of the fan user to provide the maximum amount of the information set out below. Flow rate The desired flow rate in combination with the pressure required to overcome the system resistance is know as the duty point. The required volume flow rate is normally dictated by the application (a mass of air to dissipate a heat load, a mass of air required for combustion in a furnace or a number of air changes per hour to provide adequate ventilation for an office, etc.). If the required flow rate is not known the requirements should be dis-

cussed with the fan manufacturer, who can often calculate the flow rate or suggest where further information or advice can be found. System pressure In order to deliver the required volume flow rate, a fan must overcome the resistance of the system in which it is installed. Resistances of ducted systems can be calculated but this is often not the case where a fan is applied to a ‘black box’. Although fluid dynamic software tools exist which can calculate system resistances of enclosures etc., it is advisable first to discuss the problem with the fan manufacturer, who may be able to help define the resistance from experience or measure it on an airflow test rig. Inlet gas density For most applications and installations this is not required as it is assumed that standard air is being transported. However if another gas, or dust laden air is being transposed it is important to define the gas/air density as this affects the power required to drive the impeller.

Appendix 4: Electric motors

39

Altitude of working site

Operation duty

Fan performance is stated at standard temperature and pressure. At altitude the density of the air is reduced affecting the fan performance. Advice on this change is given in chapter 2.

Whether the fan is for continuous or intermittent use can affect the selection. The number of starts per day and the flow conditions existing when the fan is started should be stated. Frequent starts per hour can add a high thermal load to the motor.

Ambient temperature The fan impeller will have a maximum operating temperature. However it is more likely that the drive motor maximum allowable temperature will be lower. The ambient temperature needs to be stated.

Duty control Is a change in performance required, permanently or periodically, and the proposed means to achieve this? Installation requirements

Space envelope How much space is available? The fan manufacturer can chose from a number of solutions and if space is tight compromises in other areas can be made.

Are any special features such as lagging cleats, cleaning door, anti-vibration mountings required? What is the intended mounting arrangement or foundation? Fan type

Application Brief details of the application or purpose for which the fan is required will often help the fan manufacturer select the right fan, or suggest alternatives.

Advise details of the type of fan preferred, its configuration, including handing and discharge, and the size of inlet and/or outlet ducts to which it will be connected, see Figure A3.1.

Ambient conditions

Drive

Ambient conditions such as humidity, and atmospheric contaminants, e.g. salt spray, acidic vapour etc., may mean that additional protection may be required.

Particulars of the drive arrangement, including whether shaft horizontal or vertical (up or down), see Figure A3.1, and prime mover preferred, electrical supply etc. Fans can be supplied with integral electric motor, motors mounted on a common frame with direct or indirect drive via belts and pulleys, or supplied without a drive.

135°

Standards Are there particular Standards or specifications applicable to the fan, the area where it is situated or the equipment of which it may form part?

LG 90

LG 135

LG 180

LG 225

LG 270

LG 315

RD 0

RD 45

RD 90

RD 135 Vertical axis

U

RD 180

RD 225

RD 270

RD 315 D

Figure A3.1 Designation of outlet positions and motor positions

A

B

Upward discharge

LG 45

AU

BU

Downward discharge

LG 0

B Motor downstream

Horizontal axis

A Motor upstream

AD

BD

40

Fan application guide

Special requirements

Reference

Will the fan be used to move dust laden air or be of an explosive nature? Movement of dust laden air will require specific impeller blades. Dust will build up on the impeller causing imbalance and stresses. Dust and other elements can be of an explosive nature. There are strict requirements defined by the European ‘ATEX’ Directive(A3.1) placing requirements on the end user to ensure correct selection and safe operation of fans. The fan manufacturer can give guidance.

A3.1

Directive 94/9/EC of the European Parliament and Council of 23 March 1994 on the approximation of the laws of the Member States concerning equipment and protective systems intended for use in potentially explosive atmospheres Official J. of the European Communities 19.04.1994 L100 (Brussels: Commission of the European Communities (1994)

Appendix 4: Electric motors The primary type of motor used in fan applications is the AC induction (squirrel cage) motor suitable for either single phase supply for small fans or 3-phase supply for medium to large fans. The motors are rated according to their output power. Small fans for domestic applications and in electronic cooling systems can be as small as one watt output, medium size motors to drive fans in building service and process applications can be from hundreds of watts to kilowatts and very large motors to drive forced draught fans in large coal fired boilers and be in the order of megawatts output. Large and very large motors used in fan applications are normally AC induction motors. In the small and medium range there are a number of technologies used from shaded pole design, split phase capacitor, direct-current (DC), electronic commutation DC (EC) as well as 3-phase induction motor. An important parameter to consider is the motor run-up time and on-off cycling time with respect to the motor and fan inertia. Systems with high inertia can prolong run-up times leading to high current and over heating of the motor.

3-phase motor

EC

motor

The motor characteristics are expressed in a torque speed curve, speed in the x-axis and torque on the y-axis. Typical characteristics are shown in the Figure A4.1.

A4.1

Alternating current (AC) motors

A4.1.1

3-phase: squirrel cage

These are the commonest types used from medium to large size fans in building services and industrial applications up to large fans in power generation. They are suitable for direct-on-line or star-delta starting. The rotational speed is dependant on the number of winding poles (2, 4, 6, 8 etc.) and the supply frequency. For a 50 Hz supply basic synchronous speeds for induction motors are 3000 rpm (2 pole), 1500 rpm (4 pole), 1000 rpm (6 pole), 750 rpm (8 pole) etc. Induction motors run below synchronous speed by an amount which varies with size, type and load. The difference between synchronous and actual speeds is known as slip. The range in efficiency varies considerably with the motor size with small, 100 W motors being 50% efficient and large high efficiency motors (HEM) of 100+ kW being 97% efficient. 3-phase squirrel cage induction motors in sizes 1.1 kW and above are classified EFF3, EFF2 or EFF1 according their efficiency with EFF1 being the most efficient. The efficiency rating increases with motor size. A4.1.2

Single phase motor

Shaded pole

Torque

Ts = starting torque Tsa = ‘saddle’ torque Tb = breakdown torque

Tb

Single phase

These are motors of a relatively simple construction made up to ratings of about 75 watts output. They are cost efficient, operate at relatively high slip and have a low efficiency in the region of 10% to 30%.

Shaded pole motor

Split phase capacitor

Tsa Ts System characteristic 0

0

Speed

Figure A4.1 Example of speed-torque diagram and comparison of motor characteristics

These have separate main and auxiliary windings sometimes referred to as running and starting windings. Large motors have a built-in centrifugal switch which automatically switches out the starting winding at a preset speed in the starting cycle. A more common type used to drive small and medium size fans is the capacitor start and run type. These are very similar to the split phase type

Appendix 4: Electric motors with the addition of a capacitor in circuit with the auxiliary winding to provide a substantial increase in starting torque. The capacitor and auxiliary starting winding stays in circuit after the starting cycle and assists the main running winding. Efficiencies are between 40 and 70% Commutator or universal These are normally used for special purposes where speeds in excess of 2850 rpm are required (e.g. vacuum cleaner fans, some hand dryer fans etc.) Because of the continuous contact between brushes and armature this type is not very suitable where long running periods are required.

A4.2

Direct current (DC) and direct current electric commutation DC(EC) motors

Small direct current motors are commonly used to drive fans in electronic and telecommunication applications where only a DC supply is available. They are available for supply voltages of 5, 12, 24 or 48 V DC and have a permanent magnet rotor. Other voltages, such as 110 V DC for traction applications, are available. Brushed commutators gave way to brushless electronics long ago, providing the necessary alternating current through the stator winding. Motor sizes are available up to 500 W output. They are not restricted to synchronous speeds. The speed is defined by the magnetic field strength and is limited by the current carrying capacity of the electronics and motor winding. Efficiencies are between 50 to 90% depending on motor size. Large direct current motors, commonly referred to as EC motors, are DC motors with AC supply input. The operating principle and technology are the same as for brushless commutated DC motors with integrated AC–DC converters. As with a DC motor, higher than synchronous speeds are possible. Motors sizes are up to 5.5 kW and efficiencies are 80–90%.

A4.3

General features of motors

A4.3.1

Construction

41 A4.3.3

Airstream rated motors

A fan motor used for direct drive of a fan impeller, whether axial or centrifugal, is sometimes situated directly in the airstream entering or leaving the impeller. In such a case use can be made of the cooling effect of the air to obtain a higher output from the motor than would be obtainable in still air. Such a motor is referred to as airstream rated. In an installation using this type care must be taken to ensure the airflow over the motor is not restricted. The reduced cooling effect from reduced speed of variable speed control of fans also needs to be considered. A4.3.4

Motor speed control

There are a number of techniques available to vary the speed of the motor. Not all are suitable to all types of motor and care should be taken to prevent permanent damage. The fan supplier will advise what methods of speed control are suitable for the motor supplied with the fan. The following gives a short overview of the most common techniques. Star/delta See Figure A4.2. Three-phase motors windings are normally wired in delta configuration resulting in 400 V across each winding. When wired in star the effective voltage across each winding is now 230 V reducing the power to the motor and the output speed. They are limited to two speeds being a function of the number of winding poles. Efficiency reduces as a variable voltage technique is being applied.

400 V Stator

Rotor

400 V Rotor

Stator 400 V

Delta (high speed)

400 V

Star (low speed)

Figure A4.2 Delta and star motor winding configurations

There is a wide variation in construction between motor sizes. Medium and large size squirrel cage AC induction motors are of internal rotor design and are typically classified according to the distance from mounting feet to centre line of the output shaft and have an enclosed housing. Small AC induction motors, shaded pole, DC and EC motors can be of either internal rotor or external rotor design, with either enclosed housings or open frame design. The degree of protection against ingress of dust and moisture are all classified in the same manner with an ingress protection (IP) rating(A4.1). A4.3.2

Insulation class

Motor windings fall into various insulation classes according to the maximum operating temperature of the windings.

Pole change The speed of AC squirrel cage motors is defined by the number of winding poles. There are some motor designs with multiple motor windings wound with different speeds. Switching between these windings provides a variation in speed. Up to three speeds are possible. Variable voltage Reducing the applied voltage will reduce the output torque and increase the slip resulting in a reduced speed. This practice is not suitable for standard low resistance squirrel cage motors of medium to large machines. A high resistance rotor is required. Small motors are often designed with high resistance rotors for variable speed control. The variable voltage can be achieved in a number of ways:

42

Fan application guide Motor

L

Main winding

Auxiliary winding

5-step transformer

Rotor N Figure A4.3 5-step transformer connected to a single phase motor

—

dropping the voltage across a resistor (limited to very small motors)

—

dropping the voltage using an auto transformer typically giving 3 to 5 speeds (see Figures A4.3 and A4.4)

—

triac control, where the supply voltage is phasechopped producing a variable voltage of 25 to 100% (see Figures A4.5 and A4.6).

The motor efficiency reduces considerably across the speed control range. Motor L1

Tapped winding

5-step transformer Rotor Stator windings

L2

Tapped windings are only found with small single phase AC shaded pole or split phase capacitor motors. The tapped winding is akin to the tappings of an auto transformer. The effect is similar to varying the supply voltage. Motor efficiency reduces with reduced speed. Variable frequency

5-step transformer

See Figure A4.7. This is an efficient form of speed control for AC squirrel cage induction motors. As the motor speed is a function of the number of motor poles and supply frequency, varying the supply frequency will change the fan speed. The supply voltage is also reduced proportionally. The frequency and voltage is varied to the motor by use of pulse width modulation (PWM) of the voltage supply.

L3 Figure A4.4 5-step transformer connected to a 3-phase motor

Triac control

L

L1

Motor

Auxiliary winding

Mains L2 L3

Main winding

Rectifier

Rotor

N Figure A4.5 Phase angle control of a single phase motor (triac) DC capacitor

Thyristor control Motor L1

Inverter

Rotor

L2

Stator windings

L3 Motor Figure A4.6 Phase angle control of a 3-phase motor (thyristor)

Figure A4.7 Frequency control of a 3-phase motor

Appendix 4: Electric motors The drive varies the PWM characteristics to effect a change in frequency and applied voltage. The speed can be infinitely varied with the minimum speed being defined by the strength of the motor output torque and the maximum speed limited by the maximum output of the motor.

43 Capacitor The speed of small split phase AC motors can be varied by changing the capacitor wiring configuration or by use of an additional capacitor wired in-line. This is not suitable for all motors and care should be taken before applying this technique.

Control line DC ( EC )

motors are very similar to variable frequency drives. To effect a change in motor speed a low voltage control signal is applied to the control line input of the motor to vary the magnetic field strength. Speed is infinitely variable with maximum speed set by the maximum power output of the motor. The efficiency is nearly constant across the speed range.

Reference A4.1

BS EN 60529: 1992: Specification for degrees of protection provided by enclosures (IP code) (London: British Standards Institution) (1992)

44

Index adjustable shading 5 AC motors 20, 40–41 access points 28 acoustic attenuators 12–14, 28 acoustic characteristics see noise acoustic enclosures 28 acoustic lining 13, 20 acoustic terms 24–25 aerofoil blade centrifugal fans 6, 13 air flow rate see volume flow air intakes, entry coefficients 33–35 air pressure see pressure air turbulence 14–15, 26–27 air volume see volume flow airstream rated motors 41 anemometer measurements 32, 37 applications fan types 6–9 selection for 11, 38–40 attenuators, acoustic 12–14, 28 axial fans 1–2 inlet design 14 performance 13 types 7–8, 9 back-to-back impellers 4 backup fans 29 backward curved blades 3 backward inclined blades 3, 4, 6, 31 balance 27, 29 bearings, maintenance 29 belt driven systems 20, 29 blade angle regulation 21 blowers see high pressure fans breakout noise 26 BS 848: Part 1:1997 32, 35, 38 Building Regulations Part L 2006 12, 19 bypass control 21–22 capture hoods 37 cascade turning vanes 15 casings 4, 8, 29 centrifugal fans 3–5 inlet design 15 performance 13 types 6–7, 8, 9 characteristic curves 6–9, 11–12, 31 cleaning 29 Coanda effect 1 commissioning 28 conical inlet pressure tappings 32 control methods 19–22 effect on noise levels 27 parallel operation 23 series operation 23–24 control systems efficiency 38 maintenance 29 controllers 20, 22, 27 cross flow fans 8 dampers 21, 22, 23 DC motors 20, 41, 43 decibels 24–25 delta motor winding 41 differential pressure measurement 35, 37 dirt removal 29 double inlet centrifugal fans 4 drive belts 20, 29 ducted propeller fans 2, 13 ductwork acoustic lining 13, 20 airflow measurement 36–37

Fan application guide ductwork (continued) effect on noise levels 26 effect on performance 14 dust laden air 38, 40 dust removal 29 duty cycle 39 duty point see operating point dynamic pressure 30 measurement 32, 35, 36–37 efficiency 12, 29 axial fans 7–8, 9 centrifugal fans 3, 4, 6–7, 8, 9 definitions 31 measurement 38 mixed flow fans 7, 9 motors 40, 41 related to noise and vibration 13, 27 electric motors 40–43 electrical supplies 12, 20–21, 29 entry coefficients 33–35 explosive atmospheres 40 fan curves 6–9, 11–12, 31 fan laws 9–10 fan types 6–9 flexible ducting 14 flow measuring hoods 37 flow rate see volume flow flow resistance 10–11, 14, 15, 38 forward curved centrifugal fans 7, 13, 21 guards, inlet 15, 17, 29 guide vanes 23 high frequency noise 12, 13 high pressure fans 5–6 housings 4, 8, 29 ‘hunting’ 23 impeller efficiency 31 ingress protection (IP) rating 41 inlet cones 14, 15, 16 inlet design 14–18, 26 inlet guards 15, 17, 29 inlet rings 18 inlet swirl vanes 21 inlet tapers 15, 16 inlet vanes 4–5, 21 input power 30 instability see ‘hunting’; stall installation 28 effect on noise levels 26 effect on performance 14–18 types 32 insulation class (motors) 41 inverters 20–21 ISO 5801: 1997 32, 35, 38 isolating dampers 23 jet fans 2 low frequency noise 12, 14 lubrication 29 maintenance 29 manufacturers’ data 12, 32 noise ratings 25, 26 mechanical vibration 26, 27 mixed flow fans 2–3, 7, 9 motors 20, 40–43 positions 39 multi-stage fans 22 noise 12–14

noise (continued) application and installation effects 26 control methods 26–28 human perception 25 ratings 25–26 units of measurement 24–25 obstructions, effect on performance 17, 26 operating point 10, 23, 29 operating principles 1, 3 operational performance see performance characteristics outlet design effect on noise levels 26 effect on performance 15, 17 positions 39 output power 10–11, 30–31 overall efficiency 30 overload protection 29 painting, housings 29 parallel operation 22–23 peak efficiency 2, 4, 6–9 related to noise and vibration 13, 27 performance characteristics 9–11, 13, 14 axial fans 2, 7–8, 9 centrifugal fans 3, 4, 6–7, 8, 9 characteristic curves 6–9, 11–12, 31 installation and system effects 14–18 mixed flow fans 7, 9 in operation 29 see also efficiency pitot tubes 35, 36 plate bladed centrifugal fans, performance 13 plenum fans 9 PWM (pulse width modulation) 27, 42–43 positioning 15, 16, 28 power characteristics 6–9, 9–11 power, definitions 30–31 pressure change 9–10, 32–33 pressure, definitions 30 pressure loss 10–11, 18, 35 pressure measurement 32, 35 propeller fans 7, 13 pulse width modulation 27, 42–43 radial bladed centrifugal fans 3, 6 resistance to flow 10–11, 14, 15, 38 roof ventilators 8–9 rotating vane anemometers 37 running speed 12 selection for application 11, 38–40 sensors, control 22, 29 series operation 23–24 servicing 29 SFP (specific fan power) 12, 31 silencers 12–14, 28 slip 40 soft starting 12 sound attenuators 12–14, 28 specific fan power (SFP) 12, 31 speed regulation 20–21, 27, 41–43 splitters 15 stability see ‘hunting’; stall stall 2, 4, 12, 31 standards 32, 38 star motor winding 41 static pressure 30 measurement 32–33, 35–36 swirl 14 swirl vanes 21 synchronous speed 12, 40 system resistance 10–11, 14, 15, 38

Index tangential flow fans 5 tapped windings 42 temperature, operating 11, 12, 29, 39 testing 12, 32–38 total pressure 30 measurement 35 tube-axial fans 2, 8, 13 tubular centrifugal fans 8 turbo-blowers 6

45 turbulence 14–15, 26–27 turning vanes 15 vane-axial fans 2, 8 variable frequency drives 20, 27 variable output sensors 22 variable pitch fans 2, 3, 21 variable pitch inlet vanes 21 variable speed systems 20–21, 27, 41–43

variable voltage motor controllers 20, 27 velocity pressure see dynamic pressure vibration 26, 27 volume control dampers 21, 22 volume flow 9–11, 19, 30 measurement 32–37 wall rings, effect on performance 15, 16 wiring, electrical 12, 29