DESIGN OF AN EFFETIVE LOW PRESSURE VAV AIR DISTRIUTION … Guide to... · 2017. 2. 22. · ume (VAV) system can be one of the most energy – efficient and comfortable systems for

DESIGN OF AN EFFECTIVE LOW PRESSURE VAV AIR DISTRIBUTION SYSTEM


ABSTRACT

The aim of this paper is to provide the air conditioning engineer with

a clear set of guidelines for use in the design of the air distribution

system for a low pressure Variable Air Volume system.

The following aspects are considered:

1. Different types of terminal outlets and why some work better than

others.

2. A brief comparison between pressure-dependent and pressure-

independent systems.

3. Duct design and methods of sizing.

4. The static regain principle and duct static pressure control.

5. Commissioning procedures.

In considering these aspects, practical guidelines for good design

and installation are provided. The paper also highlights many of the

common pitfalls, which may result in a system which does not live up

to expectation.

Keywords: Variable Air Volume, Air Distribution, Static Pressure

& Diffusers.

1. INTRODUCTION

A properly designed, installed and commissioned Variable Air Vol-

ume (VAV) system can be one of the most energy – efficient and

comfortable systems for occupied zones. Indoor air quality, noise

and overall comfort are generally excellent with a system that per-

forms as it is meant to.

VAV technology has the distinct advantage of flexibility and adapta-

bility that no other system can offer.

However, systems do not always work as they should and this could

be due to any number of reasons. It is important to note however

that the shortcomings of the applied VAV technology are not intrin-

sic or generic, and they should have limited impact if the design,

installation and operation of the system are properly addressed. In

almost all cases, problems and complaints result from errors or omis-

sions in the design, construction and operation of the system. These

can and should be corrected.

This paper addresses specifically the subject of AIR DISTRIBUTION,

although it is difficult to treat the subject of air distribution without

getting quickly involved with the system itself. But, because time

and space do not allow, this discussion will be restricted to matters

pertaining to air distribution only.

The principles involved in the subject of air movement are not always

clearly understood by air conditioning engineers. This paper aims to

provide guidelines, which will help the engineer avoid the common

pitfalls that often result in a system that does not live up to expecta-

tion.

Thirty years ago dual duct systems were common and although they

were not at all energy or cost efficient, they worked well because

they were generally over-engineered and could cope with almost any

condition, climatic or occupancy related.

Today’s systems need to be much more effective in terms of cost,

energy consumption and indoor air quality. Occupants of modern

buildings are also more demanding and have higher expectations of

the environmental control system.

To meet this demand, the air conditioning engineer needs to devel-

op a better understanding of the theory and practice of VAV sys-

tems. This paper, therefore, seeks to provide the information neces-

sary to design a top class, effective air distribution system, a vital link

in the chain of requirements for the complete installation to be suc-

cessful.

2. COMPARISON OFTWO MAIN TYPES OF

VAV SYSTEMS

As a means of briefly comparing the two common types of VAV

systems, only the components, advantages and disadvantages will be

considered. Although by no means comprehensive, this comparison

should highlight the salient differences between the two systems.

2.1. Pressure Independent System

2.1.1. Essential Components

The more conventional VAV system, also known as a single duct,

pressure independent system, consists of the following.

2.1.1.1. Cooling/Heating coils, filters etc, items common to all types

of A/C systems.

2.1.1.2. Supply air fan, often with variable speed control, capable of

providing duct pressures, at the fan discharge, in the re-

gion of 500Pa to 1500Pa.

2.1.1.3. Supply air ducting, designed for medium pressure and

relatively high velocities.

2.1.1.4. Volume control terminal units that serve occupied zones by

distributing the supply air through a group (typically 2 to 8)

of fixed aperture outlets.

2.1.1.5. A temperature controller, which will control the supply air

quantity for a particular zone, to satisfy the cooling/heating

demand for that zone.

2.1.2. Advantages of Pressure Independent Systems

2.1.2.1 The relatively high duct pressures and velocities allow the

duct size (and cost) to be reduced.

2.1.2.2. Cost saving, when a volume control terminal (VAV box) is

used with large number of outlets.

2.1.2.3. System is tolerant of inadequacies in duct design and in-

stallation because, as the name implies, it is not dependent

on accurate control of duct pressure; the volume control

terminal units will compensate for wide variations in duct

pressures.


2.1.3. Disadvantages of Pressure Independent Systems

2.1.3.1 Excessive fan power required to develop high duct pres-

sures (Power absorbed is proportional to pressure).

2.1.3.2. Associated with the above, are the higher noise levels re-

quiring additional attenuation.

2.1.3.3. Lack of flexibility. It is difficult to split into two a zone

served by a single VAV box and still maintain individual

control in each sub-zone.

2.1.3.4. Because the terminal outlet has fixed opening size, there is

a risk of cold air “dumping” and hot air stratification at

reduced airflows. The aspect will be examined in more

detail later.

2.1.3.5. Costs can be high in a situation that requires a large num-

ber of small offices to be served by individually controlled

terminals.

2.2. Low Pressure Systems with Variable Geometry

Outlets

2.2.1. Essential Components

The alternative to the above, also known as a pressure dependent

system, consists of the following.

2.2.1.1 Cooling/Heating coils, filters etc, as above.

2.2.1.2. Supply air fan, also generally variable speed, capable of

supplying duct pressures, at the fan discharge, in the re-

gion of 250Pa. An explanation of this will be given later.

2.2.1.3. Supply air ducting, designed for low pressures and veloci-

ties (below 10m/s).

2.2.1.4. A means of duct static pressure control.

2.2.1.5. Variable volume outlets, which vary the supply air volume

at the point of discharge into the occupied space.

2.2.1.6. A temperature controller that will control the supply air

volume of a particular outlet (or group of outlets) in ac-

cordance with the cooling/heating demands for that space.

2.2.2. Advantages of Pressure Dependent Systems

2.2.2.1 Lower fan power requirements for the lower duct pressures.

2.2.2.2 Lower fan sound power levels, associates with the above.

2.2.2.3 Flexibility. Each outlet can be individually controlled; A

group of outlets operated by a single temperature control-

ler may easily be split, merely by the addition of another

controller (in the case of electronically controlled units).

2.2.2.4. Because the duct static pressure is controlled and kept

constant, it is possible to mix constant volume and variable

volume diffusers on the same supply duct system.

2.2.2.5. No risk of cold air “dumping’ or hot air stratification be-

cause of the variable geometry nature of the outlet. This

feature will be more closely examined later.

2.2.2.6. Can be more cost effective for applications that require

each outlet to be individually controlled.

2.2.3. Disadvantages of Pressure Independent Systems

2.2.3.1 Duct sizes need to be larger to compensate for lower air

velocities.

2.2.3.2. For optimum performance, it is important to pay close

attention to the design and installation of the supply air

duct system. A poorly designed duct system will compro-

mise the “equal pressure” requirement for the outlets.

2.2.3.3. System can be more expensive in situations where a large

number of outlets serve a common space or zone, such as

an open plan office.

It may be stated at this point that because the system is based upon

the principal of constant duct static pressure, it is not necessary to

monitor the air velocity at any point in the system. Supply air vol-

ume will always be proportional to damper opening.

3. COMPARISON OF TERMINAL OUTLETS

The function of an air diffuser is to supply cold or warm air to an

occupied space evenly, without causing excessive air movement at

any particular point in the room while at the same time providing

near-uniform temperatures throughout the occupied zone. To do

this, it must introduce air above the occupied zone at a velocity high

enough to mix well with room air, such that it slows down to a

harmless speed before reaching the occupied zone.

It takes energy to produce this mixing and this energy can only

come from the velocity of the primary air stream itself. This mixing,

otherwise known as entrainment or induction, is a function of dis-

charge velocity and length of perimeter from which the air is dis-

charged. So for example, a round nozzle will generate little induc-

tion when compared with a long slot type diffuser, for a given open-

ing size and airflow rate. Therefore, to optimize the induction, the

discharge velocity and exit perimeter length need to be maximized.

3.1. Fixed Aperture Outlets

In case of the conventional VAV box system, the airflow rate is con-

trolled some distance upstream of the point of air discharge and the

high escape velocities from the damper device cannot be used di-

rectly to generate room air entrainment or induction. This is because

the associated outlets have a fixed aperture and their discharge ve-

locity is proportional to volume.

Using the energy formula for a moving body :

Kinetic energy, E = ½ MV2

Where: M = Mass & V = Velocity

SUPPLYAIR DUCT

VOLUMECONTROL UNIT

FIXED APERTUREAIR OUTLET

FLEXIBLE DUCT

Figure 1: Fixed Aperture Outlet.


In the case of the fixed aperture outlet, M and V change at the same

rate as flow and therefore it can be seen that energy changes with

the cube of the volume flow rate. The relationship between energy

and volume flow rate through the normal range of volume control is

shown in Table 1.

Obviously at a flow rate of 33% there is very little energy left to gen-

erate the induction of secondary room air into the supply air stream.

This drastically reduced energy leads to the “dumping” of cold air

onto occupants below.

3.2. Variable Geometry Outlets

In the case of variable geometry diffusers, the flow rate is controlled

by changing the outlet area at the point of discharge. This has dis-

tinct advantages. The first is the regain of static pressure at the out-

let under reduced air volume conditions, which may be explained

using the square law principle.

P2/P1 = (V2/V1)2

Figure 2 shows a typical supply air duct, flexible duct connection and

a variable geometry diffuser with motorized damper actuator.

The static pressure in duct “A” remains constant at all times by virtue

of the static pressure control system, which is examined in greater

detail later. The volume control damper in the diffuser varies the

discharge aperture in accordance with the demand of the room ther-

mostat.

As the flow rate diminishes; the static pressure loss due to friction in

the flexible duct reduces in proportion to the square of the flow rate.

So, at the minimum air condition, there is in fact more static pressure

available at the discharge to increase the jet velocity, which in turn

enhances the room air induction rate. This increase in available static

pressure is depicted in Figure 3.

AT MAXIMUM AIRFLOW

Static pressure at A = 60Pa

Velocity at B = V1 = 5.3m/s

Pressure loss due to friction = P1 = 16Pa (friction in flexible

duct)

Static pressure at B = 0.24 – 0.064 = 44Pa (available to force air

through aperture)

AT 33% AIRFLOW

Static pressure at A = 60Pa

Velocity at B = V2 = 1.75m/s

Static pressure loss P2 at 33% airflow

Using square law:

P2 = P1 x (V2/V1)2 = 60 x (1.75/5.3)2 = 1.7Pa (friction in

flexible duct)

Static pressure B = 60-1.7 = 58.3Pa (available to force air

through aperture)

If the energy equation is applied to variable geometry diffusers, it will

be seen that as flow is reduced, only the mass of moving air is re-

duced. Discharge velocity is maintained and in fact is slightly in-

creased because of the regain of static pressure as shown above.

This results in the following comparative energy relationship.

Table 1: Comparative Energy for Variable Geometry Outlets

Comparing these results leaves little doubt of the greater effective-

ness of the variable geometry type of diffuser for VAV air distribu-

tion.

4. DUCT DESIGN

When designing a low pressure VAV system, which uses variable

geometry diffusers, correct duct design is of the utmost importance

for a successful installation. While good duct design is a relatively

SUPPLY AIR DUCT

VARIABLEGEOMETRYDIFFUSER

INDUCEDROOM AIR

Figure 2: Variable Geometry Outlet.

Figure 3: Static Pressure at Outlet versus Volume Flow Rate

FLOW RATE FIXED

APERTURE

VARIABLE

GEOMETRY

100% 100% 100%

75% 42% 76%

50% 12,5% 54%

33% 3,6% 40%

COMPARATIVE ENERGY


simple task, it is probably the least understood aspect of a VAV in-

stallation. As stated previously, the conventional high-pressure sys-

tem is very tolerant of duct design deficiencies because there is usu-

ally substantially more pressure available throughout the system than

is actually required. It is precisely this type of over-design philoso-

phy that creates problems if applied to low-pressure systems, results

in high noise levels or excessive air at terminals under minimum air

conditions.

Duct sizing is usually accomplished by one of the following methods.

1. EQUAL FRICTION

2. STATIC REGAIN

The equal friction method is the more common one and, as the

name implies, results in a system in which the duct static pressure

reduces at a constant rate down the length of the duct. So for ex-

ample if a duct is 30m long and is designed for a friction rate of

0.1Pa/m, the static pressure at the end of the duct will be 30Pa lower

than at the beginning. This is for a simple straight duct and with a

few bends and fittings the static pressure loss could easily double to

60Pa. This method is fine for constant volume systems where manu-

al duct dampers may be used to throttle the airflow to achieve a

balanced system.

All a throttling damper in fact does is destroy static pressure, which

results in lower airflow rates. It is important to understand that a

volume control damper is primarily a static pressure reducing device

– the airflow cannot be reduced unless the static pressure loss is

increased. Volume flow rate through a diffuser is therefore directly

related to static pressure in the duct. To get more air out of a dif-

fuser, reduce the static pressure loss by opening the throttling damp-

er.

The equal friction method of duct sizing will work satisfactorily for

low pressure systems with variable geometry diffusers only if the

duct runs are short or if the duct velocities are kept low (below 5m/

s). If the duct run is short, the static pressure loss from beginning to

end of the duct will not amount to much and if the velocities are

kept low, the friction rate per metre of duct is very low (± .33Pa/m),

resulting in small static pressure losses.

For more complex systems, it is essential to use the static regain

method of duct sizing. It is important to do the duct design correct-

ly from the outset because there is no cure for a duct that has been

undersized.

A low pressure VAV system utilizing variable geometry diffusers relies

on having the same constant static pressure at the take-off to each

outlet. This being achieved by using the static regain method of

duct sizing. This paper does not seek to explain how static regain

duct sizing is done; for this purpose there are various software pro-

grams available from various sources.

One objection to the use of this method to size ducting is that it

results in larger and more expensive ducting. While this is true, the

extent of the increase is often overestimated. The weight of sheet

metal required for a system designed by static regain is approximate-

ly 13% more than the system designed by equal friction. However,

the marginal increase in first cost, is essentially offset by the cost of

reduced balancing time and operating costs.

To put the size issue into perspective, the following illustrates the

relationship between air volume and duct size. Volume is propor-

tional to the square of the duct dimension i.e. to increase the volume

of air carried in a duct by 50%, a typical duct size would have to

increase from 20x16 to 24x20. It is the area that increases by 50%,

not the duct dimensions. Similarly, significant reductions in air veloc-

ity require only modest increases in duct size.

The basic principle of the static regain method is to size a duct run

so that the increase in static pressure at each take-off just offsets the

loss due to friction in the succeeding section of duct. Static regain

occurs when air slows down. A brief explanation of this is as follows:

In a perfect system where friction is ignored, the Total pressure of

the air remains constant as it travels through a diverging section of

duct from A to B.

Now P total = P static + P velocity. As the velocity, and therefore

velocity pressure, reduces from point A to point B, the static pressure

must increase simultaneously to maintain total pressure constant. In

reality, we have to contend with friction and this reduces the static

regain by a factor, preventing a full recovery of pressure.

In practice this means that the air velocity is systematically reduced

from the first take-off or branch duct all the way to the last take-off.

Generally a size reduction of less than 2 inches is regarded as being

uneconomical and not essential. Towards the end of the duct run

the duct size could become quite small and in this case a 1 inch

reduction may be sufficient to justify its inclusion. The use of a duct

smaller than 150mm x 200mm is not recommended.

Under certain conditions, the static regain method produces some

unexpected results although there is a perfectly logical explanation

for these. For example, if the take-offs are far apart, the frictional

pressure loss is relatively large and a duct size reduction may not be

required – the reduced flow rate after a take-off in the same size

duct results in sufficient slowing down of the air to produce the re-

quired static regain.

5. FAN SELECTION

One of the main advantages of the low pressure VAV air distribution

system is the reduced fan power requirement. To make the most of

this feature, it is important to have a clear understanding of how to

calculate the total system pressure against which the fan must oper-

ate. Space does not allow a detailed analysis of every component,

but the following guidelines will provide the engineer with the basic

information needed to successfully predict the fan requirements.

5.1. Simple Systems

A simple system is shown in Figure 5. Return air ducting may or may

not be required, often depending on the size of the system and

whether it is possible for the return air to find its way back to the air

handling unit directly through corridors etc. However, there will be

some pressure loss associated with the return air, even if only

through a louver.

A

B

Figure 4: Air Travelling through a diverging duct.


The fan will be required to overcome the resistance through the

following elements:

a) Return air ducting/louver etc.

B) Pressure losses inside the AHU, e.g. air filter, cooling/heating

coil, entry & exit losses.

c) Static pressure loss due to friction in the first section of duct

from A to B.

d) The static pressure in the remainder of the duct system. This is

the pressure at which the diffusers are selected and is generally

in the range 0.12 – 0.32 ins. wg.

The pressure losses inside the AHU may be obtained from the ven-

dor.

The pressure loss from A to B is obtained by calculation and depends

on air velocity and equivalent duct length, which takes into account

number of bends etc.

From point B onwards, the fan “sees” only the static pressure re-

quired to overcome the friction through the flexible duct and the air

outlet terminal, which is typically around 0.24 ins. wg.

This can be summarized as follows, for a typical simple system:

Inches wg (Static):

a) Return air components 100

b) Air handling equipment 400

c) Duct friction (A-B) 40

d) System pressure 60

----

600Pa

The fan selection is generally carried out on the basis of total pres-

sure (Static plus velocity pressure). Based on the air volume and the

size of the fan discharge, the velocity pressure at the fan discharge

may be calculated as follows:

P velocity = 0.50r (V)2

Where; r = air density in kg/m³

V = velocity in m/s

Fan total pressure is simply the sum of static and velocity pressures.

Note that the size of the header duct (A to B) is the same for both

equal friction and static regain method of duct sizing. Based on the

information given in Table 3, select a starting velocity appropriate to

the particular system and calculate the pressure loss in this first sec-

tion of duct. Ideal starting velocities are in the range 7 – 9m/s.

Table 2: Recommended maximum duct velocities for low pres

sure systems

5.2. Large Systems

For a larger system where, for example, a single AHU serves a num-

ber of floors of a building, a slightly different approach is usually

taken. This is done by dividing the air distribution system into the

most conveniently selected low pressure supply duct zones, fed from

medium pressure main ducts or risers, via branch duct dampers

which control the static pressure in the branch ducts.

A

BCD

Figure 5: A Simple Air Distribution System. TABLE 2: RECOMMENDED MAXIMUM DUCT VELOCITIES FOR LOW VELOCITY SYSTEM (m/s)

APPLICATION

CONTROLLING

FACTOR: NOISE

GENERATION

CONTROLLING FACTOR : DUCT FRICTION

MAIN DUCTS BRANCH DUCTS

MAIN DUCTS SUPPLY RETURN SUPPLY RETURN

RESIDENCES 0.015 0.025 0.020 0.015 0.015

APARTMENTS

HOTEL BEDROOMS

HOSPITAL BEDOOMS

0.025 0.038 0.033 0.030 0.025

PRIVATE OFFICES

DIRECTORS ROOMS

LIBRARIES

0.030 0.051 0.038 0.041 0.030

THEATRES

AUDITORIUMS 0.020 0.033 0.028 0.025 0.020

GENERAL OFFICES

HIGH CLASS STORES

RESTAURANTS BANKS 0.038 0.051 0.038 0.041 0.030

AVERAGE STORES

CAFETERIAS 0.046 0.051 0.038 0.041 0.030

INDUSTRIAL 0.064 0.076 0.046 0.056 0.038

S

S

S

R

STATIC PRESURESENSOR

DAMPER

BRANCH DUCT

Figure 6. Larger, more complex air distribution system.


This layout reduces the size of the riser duct where space may be

limited. The riser duct should also be sized using the static regain

method, especially for high-rise buildings where the length of the

duct is significant. For such riser ducts noise is often less of a deter-

mining factor and initial velocities of up to 10m/s may be used quite

safely. If the static pressure in the low pressure branch duct is in the

region of 40 to 70Pa, then the static pressure just upstream of the

pressure controlling branch duct damper need be no more than 100

to 200Pa above pressure. This would normally eliminate the need for

sound attenuators after the Pressure Control Damper. Both riser

duct and branch duct static pressure are controlled from static pres-

sure sensors, positioned at a point about one half to two thirds of

the distance between the duct damper or supply air fan and the end

of the duct section.

6. BYPASS DAMPERS

Bypass dampers may also be used effectively for the control of duct

static pressure, especially in smaller systems where fan power is not

significant. In this case, fan power saving is not possible as the vol-

ume of air through the fan is not reduced as the diffusers close

down.

6.1 Sizing of Bypass Damper

It must be borne in mind - if the outlet diffusers are able to turn

down to 33% of maximum, then the bypass damper must be sized to

handle 67% of the total supply air quantity. It is recommended that

the bypass damper be sized for an average face velocity of between

800 & 1000 fpm. At higher face velocities the pressure drop across

the fully open damper will increase and insufficient air will pass

through the damper, causing the static pressure in the duct to rise

above the required level. If the size of the bypass damper is restrict-

ed or limited as a result of restricted available space, it may be nec-

essary to include a face damper to operate in conjunction with the

bypass damper. This is shown in Figure 7.

6.2 Type of Pressure Control Damper

The modulating damper used for pressure control should be of the

type which has airfoil shaped vanes. This type has a near-linear air

transfer characteristics, unlike opposed blade or parallel blade type

dampers, which have non-linear characteristics. The airfoil blade type

provides much more stable control, especially at minimum air vol-

umes.

7. BALANCING THE AIR DISTRIBUTION

SYSTEM

A well designed duct system, which has been sized using the static

regain method, is essentially self balancing. No attempt must be

made to balance the airflow by means of manually operated damp-

ers placed in the duct spin-in collars. There is a very good reason

for this. Manual dampers add series resistance to the flow and this

resistance changes as the square of the flow. Therefore at full flow

(100%) the manual damper will offer a high resistance (say 45Pa, for

example), whereas at 33% flow, the resistance will reduce to (33/100)

² or 1/9th of 45Pa i.e. 5Pa. This means that at full flow the damper

will reduce the static pressure to the diffuser by 45Pa. Whereas, at

minimum flow the static pressure at the diffuser will rise by 40Pa

thereby resulting in unacceptably high noise levels and possible

space over-cooling.

If is far better to put the extra effort into correct duct design and

effect cost savings in terms of both hardware and labor, not to men-

tion the benefit of a system which operates at low noise level and

efficiently.

Note that flexible ducts that are excessive in length or excessively

looped have the same effect as a manual damper.

8. COMMISSIONING

The discussion here is restricted to the commissioning of the static

pressure control system and variable geometry diffusers. It also as-

sumes that the system has been checked for obvious faults such as

ruptured or disconnected flexible ducting, air leaks, etc. Avoid com-

missioning a system when warm air is being supplied by the air han-

dling unit, especially when the balancing hood method is used to

measure air volumes. The “stack effect” created by the warm air in

the hood will artificially reduce volume flow through a diffuser.

8.1. Simple System

A simple system would consist of a single AHU with variable speed

fan serving a single run of duct with a number of variable geometry

diffusers.

The first step would be to drive all the diffusers to the fully open

position. If the system has been designed for a specific volume di-

versification factor, open only enough of the VAV diffusers to allow

the maximum simultaneous air volume for which the system has

been designed, to flow through the duct. Now select the diffuser

requiring the highest pressure to satisfy its design volume (in this

system probably I or J furthest from the AHU) and measure the sup-

ply air volume from this diffuser using a correctly sized balancing

hood. Adjust the static pressure in the duct until the desired air

volume for this diffuser has been achieved.

FACE DAMPER(OPTIONAL)BYPASS

DAMPER

AHU

Figure 7: Face & Bypass Dampers.

AHU

ACE

BDFH

GI

J

Figure 8: Commissioning a Simple System.


Once this diffuser is satisfied, it is safe to assume that all diffusers

further upstream will deliver no less than the design air quantity. A

few random spot checks should confirm this. In operation the sys-

tem becomes self-balancing as each diffuser adjusts to the required

room load.

8.2. Larger, More Complex Systems

Larger systems would generally consist of an AHU which supplies air

to a high pressure header duct, which in turn serves a number of

branch ducts. Each branch duct would have an independent pres-

sure control damper to control the static pressure in that duct,

throughout the range of airflow volumes.

The aim is to set the system up in such a way that the supply air fan

static pressure is sufficient to supply the design air quantity to the

diffuser furthest from the AHU, under the most demanding condition,

i.e. when all diffusers served by the AHU are fully open. If a diversifi-

cation factor has been used, it must be applied as stated earlier.

The first task is to select the “index” branch duct. This is usually the

branch duct furthest from the AHU, or the duct most likely to be

starved if the fan does not supply sufficient air. The damper serving

this duct must be driven to the fully open position. Next, fully open

all the diffusers on this branch duct (presumably all these diffusers

will be serving a common zone and there will be no diversity factor),

and select the diffuser requiring the highest pressure to satisfy its

design volume. Now adjust the static pressure in the header/main

riser duct to the point where the selected diffuser meets the design

airflow requirement. Finally, without changing anything in the branch

duct, adjust the branch duct static pressure controller so that the

pressure in that branch will be controlled at the level that exists

when these measurements are made.

With the header/main duct static pressure now set, each of the

branch duct static pressure controllers may be adjusted using the

procedure suggested for a simple system.

At this stage the system is fully commissioned. However, the system

may be tested by monitoring duct pressures, while changing airflow

rates at various points in the system. In a system operating satisfac-

torily one would expect duct pressures to vary by less than 10% as

airflows vary from 100% down to 30%.

9. COST COMPARISONS

There is no simple cost comparison that can be made between a

conventional VAV box system and a low-pressure system using varia-

ble geometry diffusers. This is because the cost depends on whether

the system is designed to serve relatively large zones where a large

number of outlets may be operated by single temperature controller,

or whether the building has many small offices, each requiring an

individually controlled outlet.

The first type of application favors the VAV box system, as a large

number of low cost, fixed aperture outlets may be connected to a

single volume control VAV box. However, where all or most of the

outlets are required to have their own temperature controller for

individual comfort, the variable geometry diffuser system becomes

more cost effective. As can be expected, there is a point between

these two extremes at which the costs break even and the choice

would depend on other factors such as flexibility and running costs.

Appendix 1 shows the results of a study undertaken recently by a

contracting company in Philadelphia USA. It was found that the

break-even point occurred when the system required an average of

about 6 outlets per VAV box. Above this the VAV box system is

likely to be more cost effective while at 5 outlets or less (on average)

the low pressure system with variable geometry outlets is more eco-

nomical. It must be stressed that this may be used only as a guide

although similar investigations in Malaysia, Israel and Australia have

confirmed this finding.

The study also revealed, predictably, that the annual operating costs

were some 37% lower for the low-pressure system than for the VAV

box system.

10. CONCLUSION

Although there is much more that may be said about air distribution,

from the information presented in this paper the air conditioning

engineer should have a clear understanding of the basic principles

involved in the process of VAV air distribution. It should also be

clear that a low pressure, pressure dependent system using variable

geometry diffusers offers one of the most effective ways to provide

excellent room conditions for human comfort. This system also

meets the challenge of providing state-of-the-art technology at af-

fordable cost, without compromising individual comfort or indoor air

quality.

By paying close attention to the potential pitfalls highlighted in this

paper, the engineer can be confident of being able to design and

install an effective air distribution system.

11. ACKNOWLEDGEMENT

The author wishes to thank the Directors of Rickard Air Diffusion

(Pty) Ltd for permission to publish this paper as well as for making

available the time and information resources without which this pa-

per would not have been possible.

12. BIBLIOGRAPHY

1. Chen, S.Y.S. and Demster, S.J: Variable Air Volume Systems for

Environmental Quality.

McGraw – Hill Book Company (1996)

2. ASHRAE: ASHRAE Handbook – 1997 Fundamentals, American

Society of Heating, Refrigeration and Air Conditioning Engineers,

Inc., Chapter 31 & 32 (1997)

3. Carrier Air Conditioning Company: Handbook of Air Condition-

ing Design, Part 2. McGraw – Hill Book Company (1965)

EAST ZONE

INTERNAL ZONE

WEST ZONE

BRANCH DUCTSTATIC PRESSURESENSOR

PRESSURE CONTROLDAMPER

MAIN DUCTSTATIC PRESSURECONTROL

AIR HANDLINGUNIT

Figure 9: Commissioning a Larger, more Complex System.


APPENDIX 1.

1. Fan Powered VAV Boxes:

1.1 External zones - 17 x VAV Boxes cooling/heating $ 13,617

1.2 Internal zones - 8 x VAV Boxes cooling only $ 4,848

1.3 VAV box installation $ 1,530

1.4 Supply air diffusers and balancing dampers $ 4,500

1.5 System Balancing $ 2,250

1.6 Air Handling Unit $ 30,000

1.7 Analog controls Included

1.8 Hot water coil valves $ 1,530

1.9 Hot water piping – supply and installation $ 8,000

1.10 Electrical wiring to VAV boxes and controls $ 5,125

Total cost: $ 71,400

Annual operating costs assuming 75% AHU motor efficiency

and 50% fan powered VAV box motor efficiency $ 12,320

2. Standard VAV Boxes:

2.1 External zones - 17 x VAV Boxes cooling/heat $ 7,939

2.2 Internal zones - 8 x VAV Boxes cooling only $ 2,176

2.3 VAV box installation $ 1,530

2.4 Supply air diffusers and balancing dampers $ 4,500

2.5 System Balancing $ 2,250


2.7 Analog controls Included

2.8 Hot water coil valves $ 1,530


2.10 Electrical wiring to VAV boxes and controls $ 3,125


Annual operating costs assuming 75% AHU motor efficiency $ 7,603


3. Thermally Powered Variable Geometry VAV Diffusers:

3.1 External zones-68 x VSD 7-4 S24 heat/cool diffusers $ 11,812

3.2 Internal zones-32 x VSD 7-4 S24 cool only diffusers $ 4,202

3.3 Two Pressure control dampers and controls $ 1,306

3.4 Four hot water duct heating coils $ 1,000

3.5 Four hot water heating coil valves $ 800


3.7 VAV diffuser balancing $ 500


3.9 Damper electrical wiring $ 350



4. Electronically Controlled Variable Geometry VAV Diffusers:

4.1 Ext. zones-17 x VSD 7-1 S24 heat/cool master diffusers $ 4,939

4.2 Int. zones-8 x VSD 7-1 S24 cool only master diffusers $ 2,200

4.3 75 x VSD 7-1 S24 slave diffusers $ 12,000

4.4 Two Pressure control dampers and controls $ 1,306

4.5 Four hot water duct heating coils $ 1,000

4.6 Four hot water heating coil valves $ 800


4.8 VAV diffuser balancing $ 500


4.10 Damper electrical wiring $ 3,125



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