HEATING VENTILATION & AIRCONDITIONING

SYSTEM DESIGN

FACULTY CO-ORDINATOR

NEERAJ SHUKLA

CONTENTS

PAGE NO.

1.0 INTRODUCTION AND OVERVIEW - 2

2.0 FUNDAMENTALS OF REFRIGERATION - 5

3.0 BASIC REFRIGERATION SYSTEM - 12

4.0 ELEMENTS OF PSYCHROMETRY - 16

5.0 APPLIED PSYCHROMETRY - 23

6.0 HEAT LOAD ESTIMATION - 33

7.0 HEAT LOAD DATA SHEET & TYPICAL CALCULATION - 47

8.0 HEATING VENTILATION & AIRCONDITIONING SYSTEMS - 52

1.0 INTRODUCTION AND OVERVIEW

INTRODUCTION AND OVERVIEW

A simple definition of air conditioning is the simultaneous control of temperature,

humidity, air movement, and the quality of air in a space.

The use of the conditioned space determines the temperature, humidity, air movement,

and quality of air that must be maintained.

The primary function of air conditioning is to maintain conditions that are (1) conducive

to human comfort, or (2) required by a product, or process within a space. To perform

this function, equipment of the proper capacity must be installed and controlled

throughout the year. The equipment capacity is determined by the actual instantaneous

peak load requirements; type of control is determined by the conditions to be

maintained during peak and partial load. Generally, it is impossible to measure either

the actual peak or the partial load in any given space; these loads must be estimated.

The term ‘refrigeration’ may be defined as the process of removing heat from a

substance under controlled conditions. It also includes the process of reducing and

maintaining the temperature of a body below the general temperature of its

surroundings. In other words, the refrigeration means a continued extraction of heat

from a body whose temperature is already below the temperature of its surroundings.

For example, if some space (say in cold storage) is to be kept at - 2 °C (271 K), we

must continuously extract heat which flows into it due to leakage through the walls and

also the heat which is brought into it with the articles stored after the temperature is

once reduced to - 2 °C (271 K). Thus in a refrigerator, heat is virtually being pumped

from a lower temperature to a higher temperature. According to second law of

Thermodynamics, this process can only be performed with the aid of some external

work. It is obvious that supply of power (say electric motor) is regularly required to drive

a refrigerator. Theoretically, a refrigerator is a reversed heat engine or a heat pump

which pumps heat from a cold body. The substance which works in a heat pump to

extract heat from a cold body and to deliver to a hot body is called a refrigerant.

The refrigeration system is known to the man since the middle of nineteenth century.

The scientists, of the time, developed a few stray machines to achieve some pleasure.

But it paved the way by inviting the attention of scientists for proper studies and

research. They were able to build a reasonably reliable machine by the end of

nineteenth century for refrigeration jobs. But with the advent of efficient rotary

compressors and gas turbines, the science of refrigeration reached the used for the

cooling of storage chambers in which perishable foods, drinks and medicines are

stored. The refrigeration has also wide applications in submarine ships, aircraft and

rockets.

Air conditioning has got wide range of applications and it is very much essential in these

days. Air conditioning is provided for some of the following reasons

1) To improve productivity in offices, factories by maintaining comfort conditions for

persons .

2) To maintain comfortable conditions for working in hotels, labs, etc.,

3) To avoid malfunctioning of some of the control panels in Electrical Control

Buildings.

4) To maintain over pressure inside the premises for avoiding outside (dusty) air in

to the room.

5) To create healthy atmosphere inside the room by supplying filtered air in to the

room.

6) To provide clean, filtered, healthy, comfortable conditions in hospitals etc.,

2.0 FUNDAMENTALS OF

REFRIGERATION

FUNDAMENTALS OF REFRIGERATION

Refrigeration is the process by which heat is removed from a low temperature level and

rejected at a relatively higher temperature level. The Americans define refrigeration in a

somewhat different way, thus, refrigeration is the process by which heat is removed

from a place where it is not required and rejected into a place where it is not

objectionable. This is not, strictly speaking, a proper scientific definition, since it does

not make any mention of its temperature levels. No process can be called

“refrigeration” unless removal of heat is at a temperature lower than the surrounding

temperature.

By nature heat always flows from one body to another body which is relatively at a lower

temperature. This law of nature cannot be altered by any means. Transferring heat

from a low temperature level to a high temperature level is analogous to transferring

water from a lower level to a higher level. Imagine two water tanks, one located at the

ground floor full of water and another empty, located at the roof level of a building. If

water from the ground floor tank is to be transferred to the roof tank, then the only thing

to do is to bring a bucket, place it at a level lower than the ground floor tank and allow

the water to initially drain into the bucket according to the law of nature. The second

step would be to lift this bucket full of water to a level above the roof tank and now allow

the water from the bucket to drain into the roof tank according to the natural flow by

gravity. In the foregoing process, we have used the bucket as the carrier and moved it

up and down, first to a level lower than the ground floor tank and then to a level higher

than the roof tank. Needless to add that in the process some mechanical work has

been performed for lifting the bucket from the lower level to the higher level.

Applying this analogy to the process of refrigeration, it is evident, that we require a

substance as the carrier of heat analogous to the bucket. This substance should be first

brought to a temperature which is lower than the low level temperature so that heat from

the low temperature level will automatically flow into this carrier substance which has

been brought to a still lower level of temperature. After this carrier substance has been

fully loaded with heat it has got to be raised to a temperature which is higher than the

high level temperature so that heat from this carrier will automatically flow according to

the law of nature. The carrier substance referred to above is what is known in

refrigeration parlance as “refrigerant”. We shall now see what a refrigerant is really like.

All volatile liquids including water have property whereby the temperature at which they

evaporate changes according to the pressure it is subjected to. Take water for example.

At normal atmospheric pressure it boils at 100°C (212°F). When the water is subjected

to higher pressure its boiling temperature also becomes higher than 100°C. Likewise, if

the water is subjected to pressures lower than the atmosphere its boiling temperature

also falls below 100°C. In fact water can boil even at as low a temperature as 4°C when

it is kept in vacuum free of air. In this case the only pressure it will have above is its

own vapour pressure. Different volatile substances have different pressure-boiling point

characteristics. For refrigeration purpose the most commonly used refrigerants are

refrigerant 12 and refrigerant 22. By reference to a table giving the properties of these

refrigerants is will be seen that for each pressure there is a corresponding temperature

at which only the refrigerant will boil. It goes without saying that at any given pressure

the temperature at which the liquid refrigerant boils is also the temperature at which the

refrigerant vapour would condense back to liquid form. Whether it is boiling or

condensing all depends on whether it is receiving heat or giving heat. For refrigerant

purpose, therefore, we make use of this natural property of the volatile refrigerant. For

example, if the liquid refrigerant R-22 is by some means or other brought down to an

absolute pressure of say 83.72 psi, then this liquid is now in a position to boil at a

temperature of 40°F. In order to make this liquid at low pressure boil, you will have to

supply heat equivalent to the latent heat of vaporization.

So any body which is above 40°F in temperature can supply this latent heat to make

this liquid boil and turn completely into gaseous form, it goes without saying that the

substance which supplies the heat for boiling the refrigerant will itself naturally cool

down, but in any case not to a temperature below 40°F which is the temperature at

which the refrigerant is boiling. Once the refrigerant liquid has completely vaporized, it

has no further capacity to absorb heat from the heat supplying body, just in the same

way that the bucket had no further capacity to take more water the moment it was full.

In the case of bucket in order to make it useful again, we had to raise this to a higher

level and empty out its contents in the roof tank and then bring it back once again to

take another bucket full from the ground floor tank. In a similar way the completely

vaporized refrigerant will have to be raised to a high temperature level. By raising it to a

higher temperature level, we do not mean that you simply heat up the refrigerant. What

is really to be done is this low pressure refrigerant will have to be compressed in a

compressor to a higher pressure. The pressure to which it is compressed must be such

that the boiling point or conversely the temperature at which the refrigerant vapour

would condense should be higher than the body to which we want to empty out the heat

content of the carrier refrigerant. If, for example, the body to which we want to reject

this heat is water which is at a temperature of 90°F, then the boiling temperature or

condensing temperature of the refrigerant should be higher than 90°F, say 106°F. Now

106°F happens to be the boiling temperature of the refrigerant when it is at a pressure

of 230 psi absolute. This means that the low pressure refrigerant will have to be

compressed to a pressure of 230 psi absolute before it is in a position to empty out its

heat content to the cooling medium, namely, natural water. Note that in the process we

have not violated any natural law concerning heat flow. We have all along allowed heat

to flow from a certain temperature level to a lower temperature level. What we have

done is that we have changed the temperature level of the carrier refrigerant to higher

or lower temperature levels to conform to the natural law of heat flow.

This brings us to a stage where we have to know the definition of certain terms which

are generally used in refrigeration parlance.

Saturation Temperature

For any given refrigerant the temperature at which the liquid refrigerant would boil (or

conversely the refrigerant vapour would condense) when it is subjected to a certain

pressure is defined as the saturation temperature corresponding to that pressure. It is

obvious that at this temperature and pressure, refrigerant in liquid and vapour form kept

in a closed container would be in equilibrium with each other. So long as the pressure

inside this container is maintained steady the liquid portion will vaporize if heat is added

or the vapour portion will condense if heat is removed. On this basis, saturation

temperature for any given pressure is defined as that temperature at which liquid

refrigerant and its vapour remain in contact with each other in equilibrium.

Superheat

We saw in the previous paragraph that liquid refrigerant and its vapour will be in

equilibrium with each other in a closed container at the saturation temperature

corresponding to the pressure. Any addition or removal of heat would only result in

either liquid vaporizing or the vapour condensing, pressure remaining same. However,

if the vaporized refrigerant is separated from the liquid portion, then any heat added to

this refrigerant in vapour form would only go to raise the temperature of the vapour

above its saturation temperature corresponding to its pressure. This is superheated

vapour. Superheat is usually expressed in terms of degrees. When we say 10°

superheat, what we mean is the gaseous refrigerant is at a temperature 10° above the

saturation temperature corresponding to its pressure.

Sub-cooling

In a like manner if the liquid portion of the refrigerant is separated and completely

isolated from the vapour which is in equilibrium with it then any removal of heat from this

refrigerant would lower its temperature to a value below its saturation temperature.

Such a liquid is called sub-cooled liquid. When we say the liquid is 15° below the

saturation temperature corresponding to its pressure. It is obvious that in the case of

sub-cooled liquid when heat is added it will first rise up in temperature till it reaches its

saturation temperature and thereafter only it will begin to boil as long as it is receiving

heat. Likewise, in the case of superheated gas, when heat is removed from the same it

will first fall down in temperature till it reaches the saturation temperature corresponding

to its pressure (this is generally referred to as de-superheating). Any further removal of

heat after this would result in condensation of the vapour into liquid form at constant

temperature, namely, the saturation temperature. One thing that should be borne in

mind is neither sub-cooling of liquid nor superheating of the vapour is possible when

liquid and its vapour are in contact with each other in equilibrium, because as already

explained earlier, any removal or addition of heat in this case would only respectively

result in condensing of the vapour part or evaporating of the liquid part, at constant

temperature.

Enthalpy

Enthalpy is the terms which denotes the heat content of the refrigerant from a base

saturation temperature of 40°F. At this temperature and the corresponding saturation

pressure the heat content of the liquid has been arbitrarily fixed as 0. It, therefore,

follows that the enthalpy of the liquid above 40°F will be positive and that below 40°F.

will be negative. The enthalpy of the refrigerant when it is in vapour form will be equal

to the enthalpy of the liquid at the same pressure and saturation temperature.

Adiabatic Compression

Any process which is performed without the addition of heat to or removal of heat from

the process is said to be an adiabatic process. Compression of gaseous refrigerant

without addition or removal of heat is called adiabatic compression. The pressure

enthalpy diagram of a refrigerant has also lines showing adiabatic compression. It is,

therefore, possible to find out the enthalpy and temperature of the gas at various

pressures during the course of compression.

3.0 BASIC REFRIGERATION SYSTEM

BASIC REFRIGERATION SYSTEM

The various components which form part of a refrigeration system can be described as

follows:

Evaporator

Let us start from the evaporator. Liquid at high pressure has to be admitted into the

evaporator. In order that this liquid may evaporate at low temperature, it is essential

that the liquid so admitted is simultaneously reduced in pressure. The level to which the

pressure has to be reduced of course is determined by the temp. at which you want this

liquid to evaporate. For example, if you want evaporation of refrigerant 22 at a

temperature of 40°F, the absolute pressure should be brought down to 83.72 psi or if the

evaporation has to be at 10°F the absolute pressure should be brought down to 31.29

psi. The pressures indicated above are the saturation pressures corresponding to the

respective temperatures. This pressure reduction is brought about by the use of what is

known as an expansion valve. The expansion valve is just a needle valve which

throttles the flow of liquid refrigerant thereby bringing about a pressure drop. This

expansion valve can also be hand operated, automatic or thermostatic. Liquid admitted

into the evaporator now needs heat for evaporation. This head is supplied by the air

which is flowing over the evaporator coil. In the process, the air gets cooled and the

liquid refrigerant evaporates.

Compressor

Now if you have got to ensure continuous evaporation at the same temperature, it is

very vital that the vapour evaporating in the coil is removed from it is as rapidly as it is

evaporating. Unless this is done the evaporated vapour will build up a pressure in the

coil which would keep on rising. Any such rise in pressure will naturally raise the

evaporating temperature also since, the evaporation temperature is higher and higher

as the pressure increases. Removal of the evaporator vapour is achieved by

connecting the outlet of the evaporator to the suction side of a refrigerating compressor.

Of course, the compressor has got to be sized so that it has got a volumetric rate of

displacement which matches with the evaporation rate. Thus the evaporation pressure

is maintained as steady and the liquid fed through the expansion valve continues to

evaporate at a steady temperature so long as heat for evaporation is available at an

equally steady rate from the air flowing over.

Condenser

The compressor compresses the vapour and discharges the same into the condenser.

It is in this condenser that the high pressure hot gas delivered by the compressor has to

be condensed. For the purpose of condensing the gas it is necessary that heat is

removed from the hot gas. This removal of heat is achieved by again creating an air

flow over the condenser coil or water flow if water cooled condensers are used. The

heat given up by the refrigerant is picked up by the air or water. The hot gas which has

given up the heat naturally condenses into liquid form at the same pressure. Now let us

see how the pressure built-up in the condenser coil it has got certain definite capacity to

transfer heat from within to the outside air or water for each degree of temperature

difference. We also know that for each 1b of refrigerant which has got to be condensed

into liquid form a definite capacity to transfer heat from within to the outside air or water

of refrigerant which has got to be condensed into liquid form, a definite amount of heat,

namely, the latent heat of condensation has to be removed. If in a refrigeration system

the F-22 circulation is say, 5 1bs/minute, then the amount of heat which has got to be

removed for condensing this refrigerant gas is 5 x latent heat. This means that the

temperature difference between the hot refrigerant gas within the condenser and the air

or water flowing over it should be such that the total amount of heat transferred through

the walls of the condenser tubes just balances with the total amount of heat which has

got to be removed. The condensation rate would, therefore, automatically balance with

the compressor discharge rate as soon as the temperature difference has been built up.

The pressure inside the condenser also which initially starts building up will attain a

steady level when the corresponding saturation temperature results in the desired

temperature difference for creating the desired heat transfer rate. This is called the

condesing temperature of the system. It is obvious that if you use a small size

condenser the temperature difference has necessarily to be higher and hence the

condensing temperature and the corresponding pressure will also have to be relatively

higher.

Receiver

A receiver is a pressure vessel which is used as a storage tank for the condensed liquid

refrigerant leaving the condenser. It is from this receiver that liquid is tapped and sent

to the evaporator through the throttling device or expansion valve. It is not on all

systems that we have a separate liquid receiver. In the case of systems having water

cooled condensers, the shell of the condenser itself serves as a storage vessel for the

liquid refrigerant. In smaller systems even with air cooled condensers, it is possible to

dispense with the use of a receiver if care is taken to charge the system with the correct

amount of refrigerant.

In order that the various components forming part of a refrigeration system can be

designed, it is necessary to make a more scientific study of the entire operations. For

this purpose we have to know the complete properties of the refrigerant concerned

when it is at gaseous form and also in liquid form. The properties of each refrigerant

are shown in what is called a Pressure Enthalpy Diagram.

4.0 ELEMENTS OF PSYCHROMETRY

ELEMENTS OF PSYCHROMETRY

Psychrometry

Since air conditioning, by its very name means treating air with a view to altering its

temperature and moisture content with the use of refrigeration, it is necessary that we

should know how exactly air would behave when it is subjected to cooling, heating,

humidifying or dehumidifying processes. For this purpose, it is necessary to study the

property of air at normal atmospheric pressure in so far as it concerns air conditioning.

Such a study is what is called psychrometry.

For the study of psychrometry, a chart has been devised, which is called Psychrometric

Chart. We will just now see what the various lines of the psychrometric chart are.

Dry Bulb Lines

Any vertical line is a line of constant temperature. Condition of air represented by any

point on this line will have the temperature corresponding to this vertical line. These

lines are called Dry Bulb Lines. By dry bulb what we really mean is dry bulb

temperature i.e., the temperature as recorded by a thermometer which is dry.

Moisture Content

Each horizontal line in the chart is a line of constant moisture content. The condition of

air represented by any point on this line will all have the same moisture content as

applicable to this line. Through any point on the psychrometric chart you can always

draw a horizontal line and a vertical line. Air represented by this point has, therefore, a

dry bulb temperature corresponding to the vertical line and moisture content

corresponding to the horizontal line. It is easy to see that air at any given temperature

can have varying moisture content. Likewise, air containing any given moisture content

can have varying temperature as well.

Saturation Line

The curved line on the extreme left-hand side of the chart is what is called the saturation

line. Condition of air represented by any point on this line is said to be saturated air,

which means that the air is having the maximum possible moisture content in it. It

cannot hold any further moisture.

Wet Bulb Lines

There are number of parallel slant lines which are called wet bulb lines. By wet bulb

temperature what we really mean is the temperature of the air as recorded by a

thermometer with a wet wick on its bulb. You will also understand for the moment that

the air having a certain wet bulb temperature will have a definite heat content although

its dry bulb temperature may be anything.

Relative Humidity Lines

When the air contains its maximum moisture content, we call it saturated air; when it

contains anything less than this maximum limit then it is not saturated air because it has

still capacity to have more moisture. We therefore, say that such air is, say 50%

saturated or 60% saturated. Another term used to denote the percentage saturation is

“relative humidity”. Thus it is one and the same thing whether you say air is 50%

saturated or air has got a relative humidity of 50%. Note that we have used the word

“approximately” because the strict scientific definition of relative humidity is not nearly

the comparison of moisture content. In fact relative humidity is defined as the ratio of

the partial vapour pressure in the air to the maximum vapour pressure that saturated air

will have at this temperature. However, for all practical purpose, this is equal to the ratio

of the actual moisture content present to the maximum moisture it can hold at that

temperature.

Dew Point

We have seen that at any given temperature air has a maximum limit of moisture

holding capacity when it is said to be saturated. For example from the psychrometric

chart we can see that 70°F saturated air can hold a maximum of 110 grains per 1b of

dry air. All temperatures above 70°F, air with the same moisture content will be, say

80%, 90% etc., saturated depending on what its dry bulb temperature would be. If air

with this moisture content and at temperature higher than 70° is cooled down, then its

condition will move along the horizontal 110 grains line, till the temperature falls to 70°F.

70°F and 110 grains / 1b as we have seen corresponds to saturated condition. This is

the temperature at which air with 110 grains of moisture / 1b will begin to shed its

moisture by condensing if you continue to cool the air. This temperature is called the

DEW POINT of the air. Needless to add, it is the moisture content which determines the

dew point. All you have to do is to move horizontally on the psychrometric chart and

read the temperature where you intersect the saturation line.

Enthalpy

We were just now referring to the wet bulb as line of constant heat content of air.

Enthalpy is just another term used in place of “heat content”. Of course, the enthalpies

represented here are all values for samples of air containing 1 lb of dry air.

The amount of moisture content in the air is generally expressed in terms of grains of

moisture per 1b of dry air. For your information, grain is a weight measure. 7000 grains

make 1 lb. When we say that the moisture content is 120 grains, what we mean is there

is 1 lb, of dry air containing 120 grains of moisture. The total weight of this moist air

would, therefore, by 1 + 120/7000 lbs = 1.0171 lb. At any temperature there is a limit to

the maximum moisture holding capacity of air. This limit is something definite and does

not alter except under different atmospheric pressures. At higher and higher

atmospheric pressure the moisture holding capacity at any given temperature becomes

less and less.

At any temperature when air contains the maximum amount of moisture it is said to be

saturated air. When air has attained saturation at any given temperature, it is

impossible to add any further moisture in vapour form.

Short Definitions

Dry-bulb Temperature

The temperature of air as registered by on ordinary temperature.

Wet-bulb Temperature

The temperature registered by a thermometer whose bulb is covered by a wetted wick

and exposed to a current of rapidly moving air.

Dewpoint Temperature

The temperature at which condensation of moisture begins when the air is cooled.

Relative Humidity

Ratio of the actual water vapor pressure of the air to the saturated water vapor pressure

of the air at the same temperature.

Specific Humidity or Moisture Content

The weight of water vapor in grains or pounds of moisture per pound of dry air.

Enthalpy

A thermal property indicating the quantity of heat in the air above an arbitrary datum. In

BTU per pound of dry air. The datum for dry air is 0 °F and, for moisture content, 32 °F

water.

Enthalpy Deviation

Enthalpy indicated above, for any given condition, is the enthalpy of saturation. It

should be corrected by the enthalpy deviation due to the air not being in the saturated

state. Enthalpy deviations in BTU per pound of dry air. Enthalpy deviation is applied

where extreme accuracy is required : however, on normal air conditioning estimates it is

omitted.

Specific Volume

The cubic feet of the mixture per pound of dry air.

Sensible Heat Factor

The ratio of sensible to total heat.

Alignment Circle

Located at 80 °F db and 50% rh and used in conjunction with the sensible heat factor to

plot the various air conditioning process lines.

Pounds of Dry Air

The basis for all pyschrometric calculations, remains constant during all psychrometric

processes. The dry-bulb, wet-bulb, and dewpoint temperatures and the relative

humidity are so related that if two properties are known, all other properties shown may

then be determined. When air is saturated, dry-bulb, wet-bulb, and dewpoint

temperatures are all equal.

5.0 APPLIED PSYCHROMETRY

APPLIED PSYCHROMETRY

Let us now see how the various air conditioning procedures will be represented on a

Psychrometric chart.

1. Sensible Heating

By sensible heating, we mean adding heat to air whereby the entire heat added

goes to raise the temperature of the air. It is obvious that in such a process there

is no change in the moisture content of the air. In other words, during sensible

heating process the air retains a constant moisture content and accordingly, its

condition will move on a horizontal line corresponding to its constant moisture

content. Since heat is being added during such process, its enthalpy also rises.

Therefore, during the heating process the wet bulb temperature of the air will also

rise, because as we have already seen, it is the wet bulb temperature lines which

are identified as constant enthalpy lines.

2. Addition of Moisture

Likewise, if moisture is somehow or the other added to the air without adding any

sensible heat, the process would be represented by a vertical line corresponding

to its dry bulb temperature. In this case also, since the moisture added carries

with it the latent heat of vaporization of water, the heat content of the air also

rises and hence its wet bulb temperature also rises.

3. Heating and Humidifying

If heat is added so that part of it goes to raise the temperature and the remaining

part goes to vaporize water and add it to the air, such a process is called heating

and humidifying.

4. Cooling and Dehumidifying

Cooling and dehumidifying is just the reverse of heating and humidifying. On a

psychrometric chart such a process will also be represented in the same manner

as for heating and humidifying, the only difference being the arrows representing

the direction of movement of conditions would be just reverse.

5. Evaporative Cooling

Evaporative cooling is the process by which air is simply subjected to a spray of

re-circulated water just as in the experiment described earlier, the only difference

being, we do not provide an infinite number of spray banks as in the experiment.

The chamber with the banks of spray is called an Air Washer. Air so subjected

would of course tend to get saturated and change out at a temperature equal to

its wet bulb temperature. However, since we do not provide adequate number of

spray banks to completely humidify, the air comes out not at 100% humidity but

somewhat lower than that. Needless to say, since this process is adiabatic, the

air has constant enthalpy throughout the process and hence its condition moves

along the line representing its wet bulb temperature.

Psychrometry as Applied to Airconditioning

It now remains for us to study psychrometry as applied to air conditioning

process. We will only see for the present what the heat load form is like and

also the various sections into which it is divided. It is only after you understand

this that you will be in a better position to understand psychrometry as applied to

air conditioning.

When a space is maintained at a temperature below the atmospheric

temperature surrounding the space, then there is a transfer of heat from outside

into the conditioned area, which tends to raise the inside temperature unless this

heat is removed as fast as it enters this space. Then you have heat or any other

appliances which may be in the space. All such heat which are either transmitted

into the room or generated from within due to occupants and appliances which

tend to raise the inside temperature are termed as room sensible heat. In the

like manner, the occupants within the room also release moisture from their body

into the room. There may be other sources inside the conditioned area which

add up more moisture into the atmosphere. If the space has not only to be

maintained at a particular temperature, but also to be held within certain limits of

relative humidity, then it is necessary that such moisture gain inside the room

should also be removed just as rapidly. By removal of moisture what we have

really mean is condensing this moisture from the air and discarding it outside.

For condensing the moisture, you have to remove the latent heat of vaporisation

of water. So instead of stating that we have got to remove moisture gained, we

state this in terms of the corresponding total amount of latent heat to be removed

for condensing that quantity of moisture gained. This is also expressed in terms

of heat units viz. BTU. So the heat to be removed per hour for condensing the

moisture is termed as room latent heat. So what we really mean by heat load is

the room sensible heat and the room latent heat that are to be removed from

within the space at a calculated rate to effect the gain of sensible and latent

heats into the conditioned space.

In heat load, there is one more source which contributes to the room sensible

and room latent heat loads. This is on account of infiltration of fresh air directly

into the conditioned space and bypass of certain amount of fresh air that is

normally taken into the system through the air handling apparatus. The form is

designed so that the room sensible heat, latent heat and the additional load due

to outside air, not forming part of room load are all calculated separately.

Here, we have used the term “Bypass”. You must understand what exactly the

meaning of the term “Bypass” is. For removing sensible heat and latent heat at

the same rate at which they are being gained within the conditioned space,

conditioned air is admitted within this space at a predetermined temperature and

humidity condition such that this air would absorb the room sensible and room

latent heat loads and in the process attain a final condition which is exactly equal

to the condition to be maintained in the room. This is achieved by continuously

drawing from within the room certain amount of air and adding to it a certain

percentage of fresh air for ventilation and cooling and dehumidifying this mixture

in a cooling coil. It is this treated air, which is supplied back into the conditioned

area. On account of some free passages in between the fins and tubes a small

percentage of the air comes out on the other side of the coil without undergoing

any change. It is this, which we terms as bypass of air. As far as the portion of

the air, which is actually re-circulated from the room is concerned, bypass will

have no influence on the ultimate result. It only means that some air has been

withdrawn from the room and just put back into the same room without any

change in its condition either upward or downward. But, what really influences is

the bypass of the fresh air, which is also passed through the cooling coil along

with the re-circulated air. Since this outside air is at a much higher temperature

and humidity conditions than the conditioned space, entry of such bypass air

would tend to upset the room conditions unless this bypass air is also brought

down to the room condition. The general formula for arriving at the exact air

quantity is:

cfm = =

However, you must realise it is not merely the selection of the condition of the

supply air that is important. We have also to consider how air can be cooled

down to the selected condition in a cooling apparatus. In a cooling coil in which

air is cooled, there is no practical means of ensuring that the air leaving the coil

would be at the exact temperature and humidity condition corresponding any

condition selected by us on the sensible heat factor line. However, there is one

temperature and humidity condition which is very easy to keep under control.

This is the condition which lies not only on the sensible heat factor line but also

on the saturation line on the psychrometric chart. In other words, if the sensible

heat factor line is extended till it meets the saturation line, then the condition

represented by the point of intersection of these two lines is the one condition

which can be under our control. This temperature is called apparatus dew point.

Bypass Factor

The problem becomes a bit more complicated because in every cooling coil there

is always a small percentage of the total cfm which escapes totally untreated.

When outside air taken into the system bypasses the coil, it will tend to raise the

room temperature and humidity conditions above the desired level. It is,

therefore, necessary to take into consideration the effect of bypass right at the

time of making the heat load calculations.

PSYCHROMETRIC FORMULAS

A. AIR MIXING EQUATIONS (Outdoor and Return Air)

tm = (1)

hm = (2)

Wm = (3)

B. COOLING LOAD EQUATIONS

ERSH = RSH + (BF) (OASH) + RSHS* (4)

ERLH = RLH + (BF) (OALH) + RLHS* (5)

ERTH = ERLH + ERSH (6)

TSH = RSH + OASH + RSHS* (7)

TLH = RLH + OALH + RLHS* (8)

GTH = TSH + TLH + GLHS* (9)

RSH = 1.08 x cfmsa x (trm - tsa) (10)

RLH = 0.68 x cfmsa x (Wrm - Wsa) (11)

RTH = 4.45 x cfmsa x (hrm - hsa) (12)

RTH = RSH + RLH (13)

OASH = 1.08 x cfmoa x (toa - trm) (14)

OALH = 0.68 x cfmoa x (Woa - Wrm) (15)

OATH = 4.45 x cfmoa x (hoa - hrm) (16)

OATH = OASH + OALH (17)

(BF) (OATH) = (BF) (OASH) + (BF) (OALH) (18)

ERSH = 1.08 x cfmda x (trm - tadp)(1-BF) (19)

ERLH = 0.68 x cfmda x (Wrm - Wadp)(1-BF) (20)

ERTH = 4.45 x cfmda x (hrm - hadp)(1-BF) (21)

TSH = 1.08 x cfmda x (tedb - tldp) ** (22)

TLH = 0.68 x cfmda x (Wea - Wta) ** (23)

GTH = 4.45 x cfmda x (hea - hta) ** (24)

C. SENSIBLE HEAT FACTOR EQUATIONS

RSHF = = (25)

ESHF = = (26)

GSHF = = (27)

D. BYPASS FACTOR EQUATIONS

BF = ;(1-BF) = (28)

BF = ;(1-BF) = (29)

BF = ;(1-BF) = (30)

E. TEMPERATURE EQUATIONS AT APPARATUS

tedb ** = (31)

tldb = tadp + BF (tedb – tadp) (32)

tewb and tlwb correspond to the calculated values of hea and hla on the

psychrometric chart.

hea ** = (33)

hta = hadp + BF (hea – hadp) (34)

F. TEMPERATURE EQUATIONS FOR SUPPLY AIR

tsa = t4m – (35)

G. AIR QUANTITY EQUATIONS

cfmda = (36)

cfmda = (37)

cfmda = (38)

cfmda = (39)

cfmda = (40)

cfmda = (41)

cfmsa = (42)

cfmsa = (43)

cfmsa = (44)

cfmba = cfmsa - cfmda (45)

Note : cfmda will be less than cfmsa only when air is physically bypassed around

the conditioning apparatus.

cfmsa = cfmoa + cfmra (46)

H. DERIVATION OF AIR CONSTANTS

1.08 = .224 X

Where .224 = Specific heat of moist air at 70 F db and 50% rh,

Btu/(deg F) (lb dry air)

60 = min/hr

13.5 = Specific volume of moist air at 70 F db and 50% rh

.68 = X

where 60 = min/hr

13.5 = Specific volume of moist air at 70 F db and 50% rh

1076 = Average heat removal required to condensate one

pound of water vapor from the room air.

7000 = Grains per pound

4.45 =

where 60 = min/hr

13.5 = Specific volume of moist air at 70 F db and 50% rh.

6.0 HEAT LOAD ESTIMATION

HEAT LOAD ESTIMATION

Introduction

The primary objective is to provide a convenient consistent, and accurate method of

calculating heating and cooling loads and to enable the designer to select systems that

meet the requirements for efficient energy utilization and are also responsive to

environmental needs.

The ability to estimate loads more accurately due to changes in the calculation

procedure provides a lessened margin of error. Therefore, it becomes increasingly

important to survey and check more carefully the load sources, each item in the load

and the effects of the system type on the load. This tightening up on the hidden safety

factors occurs for a number of reasons. There is greater emphasis, by standards and

codes, on sizing equipment closer to the expected loads, as determined by outside

design weather conditions. Also the suggested indoor design temperatures are now

usually 75 °F for cooling and 72 °F for heating. Installed lighting levels are being

reduced and the calculations are using lighting loads closer to the actual loads. All of

these factors require that the designer introduce any margin of safety by a positive

action, rather than rely on an assumed hidden margin.

Purpose of Load Calculations

Load calculations can be used to accomplish one or more of the following objectives :

7) Provide information for equipment selection and HVAC system design

8) Provide data for evaluation of the optimum possibilities for load reduction.

9) Permit analysis of partial loads as required for system design, operation and

control

These objectives can be obtained not only by making accurate load calculations but

also by understanding the basis for the loads. There a brief description of cooling and

heating loads are included.

Principles of Cooling Loads

In airconditioning design there are three distinct but related heat flow rates, each of

which varies which varies with time:

10) Heat Gain or Loss

11) Cooling load or Heating Load

12) Heat Extraction or Heat Addition Rate

Heat Gain, or perhaps more correctly, instantaneous rate of heat gain, is the rate at

which heat enters or is generated within a space at a given instant of time. There are

two ways that heat gain is classified. They are the manner in which heat enters the

space and the type of heat gain.

The manner in which a load source enters a space is indicated as follows:

13) Solar radiation through transparent surfaces such as windows

14) Heat conduction through exterior walls and roofs

15) Heat conduction through interior partitions ceilings and floors

16) Heat generated within the space by occupants, lights, appliances, equipment and

processes

17) Loads as a result of ventilation and infiltration of outdoor air

18) Other miscellaneous heat gains

The types of heat gain are sensible and latent. Proper selection of cooling and

humidifying equipment is made by determining whether the heat gain is sensible or

latent. Sensible heat gain is the direct addition of heat to an enclosure, apart from any

change in the moisture content, by any or all of the mechanisms of conduction,

convection and radiation. When moisture is added to the space, for example, by vapor

emitted by the occupants, there is an energy quantity associated with that moisture

which must be accounted for.

If a constant humidity ratio is to be maintained in the enclosure, then water vapor must

be condensed out in the cooling apparatus at a rate equal to its rate of addition in the

space. The amount of energy required to do this is essentially equal to the product of

the rate of condensation per hour and the latent heat of condensation. This product is

called the latent heat gain.

As a further example, the infiltration of outdoor air with a high dry-bulb temperature and

a high humidity ratio, and the corresponding escape of room air at a lower dry-bulb

temperature and a lower humidity ratio, would increase both the sensible heat gain and

the latent heat gain of the space.

The proper design of an airconditioning system requires the determination of the

sensible heat gain in the space, the latent heat gain in the space, and a value for the

total load, sensible plus latent, of the outdoor air used for ventilation.

The sensible cooling load is defined as the rate at which heat must be removed from the

space to maintain the room air temperature at a constant value. The summation of all

instantaneous sensible heat gains at a specific time does not necessarily equal the

sensible cooling load for the space at that time. The latent load however is essentially

an instantaneous cooling load. That part of the sensible heat gain which occur by

radiation is partially absorbed by the surfaces and contents of the space and is not felt

by the room air until sometimes later. The radiant energy must first be absorbed by the

surface that enclose the space such as walls and floor and by furniture and other

objects. As soon as these surfaces and objects become warmer than the air some heat

will be transferred to the air in the room by convection. The heat storage capacity of the

building components and item such as walls, floors and furniture governs the rate at

which their surface temperatures increase for a given radiant input. Thus, the interior

heat storage capacity governs the relationship between the radiant portion of the

sensible heat gain and how it contributes to the cooling load. The thermal storage effect

can be important in determining the cooling equipment capacity.

The actual total cooling load is generally less than the peak total instantaneous heat

gain thus requiring smaller equipment than would be indicated by the heat gain. If the

design is based on the instantaneous heat gain, the rest of the system may be

oversized as well.

Heat extraction rate is the rate at which heat is removed from the conditioned space.

Normal control systems operating in conjunction with the intermittent operation of the

cooling equipment will cause a “swing” in room temperature. There, the room air

temperature is constant only at those rare times when the heat extraction rate equals

the cooling load. Consequently, the computation of the heat extraction rate gives a

more realistic value of energy removal at the cooling equipment than does just the

instantaneous value of the cooling load provided the control system is simulated

properly. The determination of the heat extraction rate must include the characteristics

of the cooling equipment and the operating schedule of thee equipment, in addition to

the various sources of cooling load.

If the equipment is operated some what longer before and after the peak load periods,

and / or the temperature in the space is allowed to rise a few degrees at the peak

periods during the cooling operation (floating temperature), a reduction in the design

equipment capacity my be made. A smaller system operating for longer periods at

times of peak loads will produce a lower first cost to the customer with commensurate

lower demand charges and lower operating costs. Generally, equipment sized to more

nearly meet the cooling requirements result in a more efficient, better operating system

particularly when is at a partially loaded condition.

Usually a fraction of the sensible heat gain does not appear a cooling load, but instead

is shifted to the surroundings. The fraction Fc depends upon the thermal conductance

between the room air and the surroundings. It may be also considered as a adjustment

factor which results when the load components as superimposed.

The adjustment factor, Fc is calculated by the following equation.

Fc = 1 - 0.02 KT

Where KT the unit length conductance between the room air as surroundings in Btu / (hr.

ft2 F), is given by

KT = 1/LF (UWAW + UowAow + UcAc)

Where

LF = Length of the exterior walls of the room, ft.

U = U-value of room enclosure element (subscript w for window, ow

for outside wall and c for corridor), Btu (hr. ft2 F)

A = Area of the specific element

If the cooling load component has already been obtained by the technique used in this

manual, multiply that result by the calculated Fc factor.

The adjustment factor should be used only for individual small spaces or zones. It is not

to be used for block loads nor for industrial applications.

Diversity of Cooling Loads

Diverting of cooling load results from not using part of the load on a design day.

Therefore diversity factors are factors of usage and are applied to the refrigeration

capacity of large airconditioning systems. These factors vary with location, type, and

size of applicant and are based entirely on the judgment and experience of the

engineer.

Generally, diversity factors can be applied on loads from people and lights; there is

neither 100% occupancy nor total lighting at the time of such other peak loads as peak

solar and transmission loads. The reductions in cooling loads from nonuse are real and

should be accounted for.

In addition to the factors for people are lights a factor should also be applied to the

machinery load in industrial buildings. For instance, electric motors may operate at a

continuous overload, or may operate continuously at less than the rate capacity or may

operate intermittently. It is advisable to measure the power input whenever possible;

this will provide a diversity factor. It is also possible to determine a diversity factor for a

large existing building by reviewing the maximum electrical demand and monthly energy

consumption obtained from the utility bills.

Principles and Procedures for Calculating Heating Load

The peak heating requirements may occur either at night during unoccupied hours or in

the morning pickup period following a shutdown. Therefore a number of calculations

are helpful in making a proper equipment selection and system design.

Information Required (Input)

Before a cooling or heating load can be properly estimated a complete survey must be

made of the physical data. The more exact the information that can be obtained about

space characteristics, heat load sources, location of equipment and services, weather

data, etc. the more accurate will be the load estimate.

Required Input - External Loads - Cooling

For calculation of the outdoor loads the input information should include:

19) Orientation and dimensions of building components.

20) Construction materials for roof, walls, ceiling, interior partitions, floors and

fenestration

21) Size and use of space to be conditioned

22) Surrounding conditions outdoors and in adjoining spaces

Required Input - Heating Load

The input for calculation of heating load is essentially the same as that for the cooling

load. However, it may not be necessary to calculate the internal sources and solar heat

gain.

In heat load estimation we compute

a. Room sensible and latent heat gains due to transmission, sunlight, occupancy

and other internal sources of heat.

b. Grand total heat comprising total room load under (1) plus additional loads due to

outside air intake, heat gains in return air ducts, in chilled water distribution

systems, pumping horse power load, etc.

Room load estimation under (a) is required for computing the condition and quantity of

supply air while the grand total heat under (b) is required for terminating the total

capacity of the cooling system.

In this discussion, we will confine ourselves to transmission gains and related subjects

only. There are certain similarities between heat transmission through barriers and

electric current transmission through conductors. We will use this similarly wherever

required for better understanding of the subject. The well known formula relating to

transmission of electric current is:

I =

Where “I” = current in Amperes, “V” = Voltage & “R”, the Resistance of the conductor.

In this formula, if (I) / (R) is considered as the conductance of the conductor, say “C”,

then the formula can be rewritten as:

I = V x C

For transmission of heat through a barrier, the motive force corresponding to “V” is

temperature difference between the two sides of the barrier. The formula for rate of

heat transmission per hour H is:

H = A x U x (T)

Where T is the temperature difference in °F and A is the area of the barrier in sq.ft. and

U is the overall heat transmission coefficient expressed in BTU/Hr/Sq.ft/°F temperature

difference. The product (AxU) corresponds to the conductance “C” of the electric

conductor.

Thermal conductivity of any material is the heat transmitted through the material

expressed as BTU/Hr/Sq.ft/Inch thickness/°F temperature difference and is referred to

by the symbol “K”. It K is the conductivity of the material, then 1/K is the resistance of

the material of 1 sq.ft. cross section and 1” thickness. If the thickness “t” inches, the

resistance becomes (t) / (K) per sq.ft.

In electrical system, resistance connected in series are added to find the total

resistance.

Similarly, if a barrier is made up of several materials, the individual resistances of the

components have to be added to arrive at the total barrier resistance. If a barrier is

made up of, say, three materials having thermal conductivities K1, K2 & K3, the total

thermal resistance of the barrier is:

t1/K1 + t2/K2 + t3/K3

Where t1, t2 & t3 are the thicknesses of the barriers.

Film Coefficient

In addition to the resistance of the various components of a barrier, we have to consider

one more resistance offered by a film of air (or fluid if the barrier is in a fluid) which

clings on to the barrier surfaces. This resistance is more when the air is still and is

relatively less when there is wind velocity. Like thermal conductivity, the heat

transmission capacity of a film is expressed as the rate of heat of transfer in

BTU/Hr/Sq.ft/°F temperature difference (Note that this differs from thermal conductivity

in the sense it is not related to any film thickness as in the case of materials). This is

called the film coefficient and is expressed by the symbol “f”. The reciprocal of “f” is the

thermal resistance of the film. “f1” denotes the film coefficient on the interior surface of

the barrier and “f0” denotes the film coefficient on the exterior surface of the barrier.

The resistance of the complete barrier is:

1/f1 + t1/K1 + t2/K2 + t3/K3 + 1/f0

If “U” is the overall heat transmission of the barrier in BTU/Hr/Sq.ft./°F, then 1/0 is the

overall thermal resistance of the barrier.

1/U = 1/f1 + t1/K1 + t2/K2 + t3/K3 + 1/f0

U = 1/(1/f1 + t1/K1 + t2/K2 + t3/K3 + 1/f0)

Storage Effect

Suppose T0 is the temperatures on both sides of the barrier. There will be no heat

transmission through the barrier and the temperatures at all points within the barrier will

also be the same. There is, therefore, no temperature gradient. Now suppose the

temperature on one side of the barrier is raised from T0 to T7, do you think that heat

transmission through the barrier will commence immediately? No, since all points within

the barrier is at the same temperature, no heat can flow through any interior section.

The first thing that happens is the outermost layer of the barrier absorbs the heat from

the outside and rises in temperature. Heat then flows over to the next layer of the

barrier because of the temperature difference between the first and second layers. The

second layer also will first rise in temperature before heat begins to flow over to the third

layer. Thus progressively all the layers within the barrier rise in temperature thereby

establishing the total temperature gradient from one side to the other side of the barrier.

It is only after the complete gradient has been established that heat will begin to flow to

the other side of the barrier. The temperature at various points within the barrier will

now be as determined by the gradient.

Attic Spaces

Whenever a false ceiling is provided in a room, the space enclosed between the false

ceiling and the concrete ceiling is called ATTIC SPACE. If the attic space is not

ventilated the entire space within the attic will assume an intermediate temperature

which will be more than the room temperature and less than the outside temperature.

This temperature can be worked out as follows:

Ag = area of the concrete ceiling

A1 = area of the false ceiling

U0 = “U” factor of the concrete ceiling

Uf = “U” factor of the false ceiling

T2 = Outside temperature

T1 = Inside temperature

T = Temperature of the attic space

When steady heat transmission from outside to inside takes place through the attic

space, then the rate of flow of heat from outside into the attic space is equal to the rate

of flow of heat from the attic space into the room, i.e.,

Ac.Uc.(T2 – T) = Af.Uf.(T – T1)

“T” can therefore be calculated from this equation. After “T” has been worked out, the

transmission load into the room from the ceiling can be worked out by substituting the

value of “T” in the above equation.

Solar Gain

Solar Gain, as the name implies, comes from direct sunlight. There are two kinds of

solar gains:

a. Radiation from sun which directly enters the conditioned space through glass and

absorbed by objects in the room and then by the air within the room. The effect

of such gain into the space is felt almost immediate.

The amount of radiation for various exposures and time of the day and year are

given in tables for the various latitudes on the earth. Depending on the type of

glass, about 5 to 6% of the radiation is reflected while the rest pass into the

room. Solar gain is not confined merely to the side which directly faces the sun.

You get solar heat even from other sides through glasses, but to a much smaller

degree. This is diffused radiation.

b. Solar & Transmission Gain

This is due to transmission through sunlit walls whose temperature rises above

the ambient temperature due to absorption of direct radiation and hence causes

a larger temperature differential than the ambient temperature. The equivalent

temperature difference that is to be taken are given in tables, taking into

consideration the exposure, sun time and storage effect.

c. Transmission Gain through Glass & Partition

In addition to solar gain through glass, you have also to work out transmission

gain through glass due to temperature difference. Transmission through

partitions between conditioned and non-conditioned areas are worked out on the

basis of actual temperature difference. No storage effect apply for these cases.

d. Internal Load

This comprises load from:

23) Occupancy: The sensible/latent heat gains from people are given in

tables, based on the nature of their activities in the room.

24) Lights: Lighting is generally specified in terms of watts per sq.ft. The total

watt has to be converted into BTU/Hr by multiplying by conversion factors.

25) Appliances: Electrical, gas burners, steam generation, etc.

26) Electric Motors: Applies generally in some of industrial applications. This

load will have to be properly analysed by discussion with user and

appropriate diversity factors should be applied for estimating the actual

load. Convert the HP into BTU/Hr.

We shall now briefly lay down the procedure for heat load estimating with explanations

wherever required.

1. Collect architect’s drawings for the building giving all details and dimensions of

walls, floors, windows, etc. If such drawings are not available, survey the place

and get the particulars.

2. For every application, there are certain things which the ultimate user has to

specify. These are:

27) Temperature & humidity conditions to be maintained inside the space and

tolerance.

28) Occupancy – i.e. maximum no. of people likely to occupy the space and the

nature of their activity.

29) Lighting load and other internal source of heat generation.

30) Period of operation – e.g. 8 a.m. to 7 p.m. or 10 a.m. to 8 p.m. etc.

31) For industrial application you require also the HP load in the conditioned

space and diversity factor thereon.

32) Minimum ventilation required.

3. Outside Design Conditions

33) For comfort air conditioning application, use the mean maximum DB

temperature & the WB temperature which occurs simultaneously with the

assumed DB.

34) For industrial applications where temperatures and humidities are to be

maintained within very close tolerance through the year, tank the maximum

DB and the simultaneously occurring WB temperature.

4. For all applications make a second load estimate for monsoon conditions.

5. For applications where the conditioned spaces are spread over very vast floor

areas, divide the entire area into convenient zones and make load estimates.

6. Occupancy - In certain applications a diversity factor may have to be used even

in respect of occupancy. Examples are: Office areas where a separate

conference room is also provided. The conference room may be designed for a

large number of people. But you must realize that it is mostly the people in the

office who go into conferences and hence any occupancy in the conference room

brings about an equal reduction in the occupancy in other areas of the office.

7.0 HEAT LOAD DATA SHEET &

TYPICAL CALCULATION

TYPICAL DIVERSITY FACTORS FOR LARGE BUILDINGS

(APPLY TO REFRIGERATION CAPACITY)

DIVERSITY FACTOR

PEOPLE LIGHTS

Office 0.75 to 0.90 0.70 to 0.85

Apartment, Hotel 0.40 to 0.60 0.30 to 0.50

Department storage 0.80 to 0.90 0.90 to 1.0

Industrial 0.85 to 0.95 0.80 to 0.90

Fresh air requirement - 2.0 air changes / hr.

or

10 CFM per person

Design conditions

a - indoor - 70 °F ± 2 °F DBT ; 55% ± 5% RH

b - Outdoor - 103 °F DBT ; 82 ° F WBT

U - FACTOR CALCULATIONS

a. Exposed Walls

Total resistance RT = Ro + X1 R1 + X2 R2 + X3 R3 + Ri

= 0.25 + 12.5 x 0.2 + 230 x 0.2 + 12.5 x 0.2 + 0.68 25 25 25

= 2.97 hr. ft2. °F /BTU

Overall heat transfer Co-efficient

= 1_ = 1_RT 2.97

= 0.337 BTU / hr. ft2. °F

b. Partitions

RT = Ri + X1 R1 + X2 R2 + X3 R3 + Ri

= 0.68 + 12.5 x 0.2 + 230 x 0.2 + 12.5 x 0.2 + 0.68 25 25 25

= 3.40 hr. ft2. °F /BTU

Overall heat transfer Co-efficient

= 1_ = 1_RT 3.40

= 0.294 BTU / hr. ft2. °F

c. Roof exposed to sun

RT = Ri + X1 R1 + X2 R2 + Ri

= 0.25 + 150 x 0.2 + 50 x 4.0 + 0.92 25 25

= 10.37 hr. ft2. °F /BTU

Overall heat transfer Co-efficient

= 1_ = 1_RT 10.37

= 0.096 BTU / hr. ft2. °F

8.0 HEATING, VENTILATION &

AIRCONDITIONING SYSTEMS

HEATING, VENTILATION & AIRCONDITIONING SYSTEMS

A. AIR CONDITIONING SYSTEMS

Airconditioning is defined as the simultaneous control of temperature, humidity,

quality and movement of air in a conditioned space or building.

An air conditioning system is therefore, defined as an arrangement of equipment

which will air condition a space or a building. Thus, a complete air conditioning

system includes a means of refrigeration, one or more heat transfer units, air

filters, a means of air transport and distribution, an arrangement for piping the

refrigerant and heating medium, and controls to regulate the proper capacity and

operation of these components.

The items outlined above are considered to be the components of a complete air

conditioning system.

There has been a tendency by many designers to classify an air conditioning

system by referring to one of its components. For example, the airconditioning

system in a building may include a dual duct air transport arrangement to

distribute the conditioned air and is then referred to as a dual duct system. This

classification makes no reference to the type of refrigeration, the piping

arrangement or the type of controls.

For the purpose of classification, the following definitions will be used:

An Airconditining unit is understood to consist of heat transfer surface for

heating and cooling, a fan for air circulation, means of cleaning the air, a motor, a

drive, and a casing.

A self-contained airconditioning unit is understood to be an airconditioning

unit that is complete with compressor, condenser, controls, and a casing.

An air handling unit consist of a fan heat transfer surface, a motor, a drive and

a casing

A remote air handling unit or a remote air conditioning unit is a unit located

outside of the conditioned space which it serves.

The most common types of refrigeration machines, classified according to their type of operation are (1) mechanical compression, (2) absorption and (3) vacuum.

Apart from the above types the airconditioning system are generally clarified is

to following categories:

1. Window (room) airconditioners

2. Split airconditioning units

3. Packaged airconditioning units

4. Centralised airconditioning plant - DX system

5. Centralised airconditioning plant – chilled water system

The details of the above are further detailed in the subsequent pages.

The types of refrigeration machines which are further explained as under:

Mechanical Compression machines may be divided into reciprocating,

centrifugal, and rotary types.

The term “heat pump” is occasionally used to describe a refrigeration machine.

However, a heat pump is a refrigeration cycle – either reciprocating, rotary or

centrifugal - in which the cooling effect as well as the heat rejected is used to

furnish cooling or heating to the air conditioning units, either simultaneously or

separately.

Reciprocating or rotary compressor can be used in systems that circulate the

refrigerant through remote direct expansion heat transfer surfaces. Alternately

they can be used in conjunction with a water chilling heat exchanger, to produce

chilled water for circulation through remote heat transfer surfaces that cool and

dehumidify the air.

Centrifugal refrigeration machines are generally not suitable for circulating and

expanding the liquid refrigerant in remote heat exchanges surfaces. Centrifugal

machines are therefore used only to chill water or brine for circulation through

remote heat exchange surfaces.

Absorption machine cycles are similarly to mechanical compression machine

cycles only to the extent that both cycles evaporate and condense a refrigerant

liquid. They differ in the mechanical compression cycle use purely mechanical

processes, while the absorption cycle uses physiochemical processes to produce

the refrigeration effect.

Vacuum refrigeration machines, such as steam jet and water vapor units, are

seldom used in modern airconditioning systems.

1. Room Airconditioner

This is the simplest form of an air conditioning system. It has a hermetically

sealed motor compressor assembly, an air cooled condenser coil, an evaporator

coil, condenser fan and evaporator fan. It has a capillary tube in place of an

expansion valve for metering refrigerant flow to the evaporator. Room air

conditioners are generally made of capacities ranging from 3/4 ton to 1-1/2 tons

suitable for operation on 230 V, single phase, 50 cycles supply. It is completely

factory assembled and can be straightaway plugged into power supply when

installed.

Application

Generally used for small office rooms, shops and residential rooms where the

where the load will generally be within 1-1/2 tons. Sometimes, these units are

used in multiple for larger areas.

Advantages

The main advantage is that the unit can be switched ON and OFF as required.

Where multiple units are used, there is no fear of a total breakdown of air

conditioning since it is most unlikely that all the units will breakdown

simultaneously.

Disadvantages

The hermetically sealed compressor is susceptible to burn out when the supply

voltage fluctuates widely and whenever such burnouts occur, the whole system

has to be thoroughly cleaned before a new compressor can be fitted. The life of

the unit is generally between 10 to 15 years only.

2. Packages type air conditioner

These are larger versions of Room Air Conditioners except that they are

generally made with water-cooled condensers. They can also be made with air

cooled condensers either built in with the package or for remote installation.

They are generally made in capacities ranging from 5 to 10 tons. Units with

water cooled condensers require condenser water circulating system and cooling

tower. The units may also require external duct work for air distribution. This unit

operates on 400 V, phase, 50 cycles supply.

Application

These units are most ideal where the load is between 5 to 20 tons. Sometimes,

they are also used for much larger loads by using more number of units

interconnected on the supply air side.

Advantages

Installation and commissioning can be done in the shortest possible time since

the field work involved only relates to condenser water piping, air distribution

system and electrical wiring. When multiple units are used for larger areas, the

number of units in operation can be varied according to the load requirement,

thereby saving on power consumption.

Disadvantages

These units also have hermetically sealed motor compressor assemblies and

hence have the same disadvantage as Room Air Conditioners.

3. General plant - DX systems

The system consists of an open type compressor ranging in capacity from 5 tons

to 120 tons operating on refrigerant 22. They are motor driven either through belt

drive or direct coupling. They can also be driven by diesel engine, but then only

by direct drive. Belt drive should not be used when diesel engine is used. They

are generally with water-cooled condensers even though they can also be built

with air-cooled condensers.

Application

The DX system of Central Plant is perhaps the most widely used system for

medium loads between 20 and 100 tons. It can be used for almost all types of

application.

Advantages

The DX system is perhaps the most efficient of all system from a thermo dynamic

point of view since the heat transfer is directly between the conditioned air and

the refrigerant. The open type compressors used for these systems have built in

capacity controls to take care of load fluctuations. Plants of any capacity can be

built with DX systems using multiple compressors, condensers and evaporators.

Although it is preferable to keep each compressor with its condenser and

evaporator as a single unit, these plants can also be built with interconnection

between them on the refrigerant side. Such interconnections naturally provide

more flexibility in operation.

Disadvantages

DX systems should not be used where air distribution through duct work has to

be carried out from a central air handling unit to various zones because of fire

hazard. Therefore, a single air handling unit should necessarily be confined to a

single zone. Where there are multiple zones use of DX system is permissible

only when separate DX plants are used for each zone without any

interconnection on the air distribution size. Where the building condition has got

number of floors one above the other, DX systems could be considered only if it

is possible to install separate Central Plant for each floor. Of course, such

decision would involve installation of the plant on upper floors where vibration

and other problems have to be effectively tackled in order to eliminate

transmission of vibrations to the occupied zones. Cost wise also, such individual

systems in each floor may prove to be much higher.

4. Central plant - Chilled Water System

A central chilled water system is made up on one or more water chilling plants.

Each water chilling plant may be built with either one or two compressors to work

with one or two chillers (DX chiller or shell and tube flooded chiller) and one or

two water cooled condensers. Each such water chilling unit is field assembled

on structural framework with the necessary refrigerant pipes so as to make a

compact assembly. Where such multiple water chilling units are used, they are

generally interconnected on the water side both in the condenser circulating

system and chilled water circulating system.

Application

Multiple water chilling units with reciprocating compressors are generally suitable

for multistoreyed office buildings where the load is between 100 and 300 tons.

However, there is no bar against using more number of water chilling units with

reciprocating compressors even for loads higher than 300 tons. For loads

exceeding 300 tons. Water chilling units with centrifugal compressors would be

preferable.

5. Chilled Water for Process Cooling

Advantages

The best advantage of a chilling water system in that the Central Plant can be

installed in as remote a location as desired from the conditioned areas. In fact,

they can even be built in a remote plant room with chilled water piping either

underground or overhead running to all the zones where air handling units are

installed. This system provides maximum flexibility in operation since the air

handling units serving individual zones can be cut off from the chilled water

circulating system whenever air conditioning is not required in any particular

zone. Since each zone will have its own air handling unit, no interconnecting

duct work will be required thereby eliminating all possibilities of fire spreading

from one zone to another. In the case of large hotels, fan coil units in

individual

rooms can be switched off whenever the room is not occupied. Individual

temperature controls can be provided for each zone or individual rooms by

regulating the chilled water flow through the coil either thermostatically or

manually.

Disadvantages

On application where the load is small, this system would prove very much

costlier than the DX system. Another disadvantage is that since one more heat

transfer medium viz. chilled water, has been introduced, the heat transfer is now

from air to water and then from water to refrigerant. This naturally lowers the

evaporating temperature as compared to a DX system for the same load. Hence

the power consumption will be relatively higher than that for DX system.

6. Air conditioning System for Operation Theatres

It is desirable that a DX system is used for each Operation Theatre. However, in

large hospitals, if there are several operation theatres located in various floors,

there is no bar against using a central chilled water system, but exclusively for

the operation theatre only. Each operation theatre must however have individual

air treatment units with pre-filters on the air suction side of fan and

supplementary microves filters on the air discharge side of the fan. For operation

theatres no re-circulation of room air is permitted. You should, therefore,

estimate the heat load on the basis of 100% fresh air.

7. Screw Chillers

Refer to figure showing single line diagram of refrigeration cycle for the above

and for piping schematics. Each vertical screw compressor discharges hot, high

pressure gas through a discharge service valve (A) (or check valve in multiple

compressor units) into the condenser, where it condenses outside tubes,

rejecting heat to cooling tower water flowing inside the tubes. The liquid

refrigerant drains to the bottom of the condenser and exits into the economizer

feed line.

The refrigerant flows through the economizer feed ball valve (B), dropping its

pressure, causing it to flash. It then flows into the flash economizer tank (C)

which is at an intermediate pressure between condenser and evaporator, liquid is

centrifugally separated from the flash gas and the liquid drains to the bottom of

the tank, exits via the economizer drain line, and passes through the economizer

drain ball valve (D). Both economizer ball valves are actuated by a modutrol

motor (U) that adjusts flow to maintain an appropriate refrigerant level in the

evaporator, determined by a liquid level float switch (V).

From the drain line, liquid refrigerant flows into the flooded evaporator, where it

boils, cooling the water flowing inside evaporator tubes. Vapor from the boiling

refrigerant flows up the suction pipes through a shut-off valve (E) (optional),

suction check valve (F) and suction filter (G) (inside compressor) into the

compressor where it is compressed and starts cycle again.

Vapor flows from the top of flash economizer into the compressor at the vapor

injection port, which feeds it into the compressor part way through the

compression process. Check valve (H) prevents backflow at shutdown in multi

compressor units. Al compressors operate in parallel on a common evaporator

and condenser.

8. Ice Storage Systems for Airconditioning Applications

Use of Ice for Airconditioning

Building air conditioning in summer daylight hours is one of the largest

contributors to electrical utility demand peaks. Typically between 2-4 PM in the

afternoon when solar loading peaks, more air conditioners are needed to

maintain comfortable environments in buildings. Add to this the electricity utilized

by lighting, computers, building subsystems plus other equipment and the utility

is faced with a peak load condition dictating that it bring on-line additional, more

costly peak power generating sources to handle the load.

Traditional air conditioning systems operate during the day to meet cooling

demand and remain idle at night. Chillers are selected to satisfy the maximum

demand, which occurs only a few hours per year, and thus spend the majority of

their operational life at reduced capacity and low efficiency.

The ice storage system, which is suitable for any A/C application, allows installed

chiller capacity (and size of other components) to be significantly reduced –

typically between 40% and 60%. This enables efficient and real energy

management whilst taking advantage of low tariff electricity.

Large commercial users whose air conditioning loads contribute to the utility

peaking problem are assessed an added charge typically based on their highest

15 minutes window of peak demand for electricity. This is called a “demand

charge” which in many areas of the country can account for as much as 40

percent of the building owner’s total electrical bill.

The use of ice storage to minimize peak energy usage is not a new or

experimental idea. It has been used for years on applications with short peak

energy usage such as churches, meeting facilities and theaters. On these

applications, however, the longer peak uses were handled by conventional

rooftop cooling or water chilling / air handling systems.

Now, however, there is renewed interest in a broad use of ice making and

storage systems by both users and utility companies as the best way of offsetting

rising demand loads and resulting utility cost increases.

Ice storage systems can not only cut operating costs substantially, but they can

also reduce capital outlays when systems are properly applied for both new and

existing buildings and commercial and industrial types. Simply stated, engineers

can specify smaller chillers operating 24 hours a day rather than larger chillers

operating 10-12 hours a day and cut the capital outlay for air conditioning

equipment substantially.

An ice storage system can utilize either a load shifting or a load leveling strategy

to significantly lower demand charges during the cooling season. Because this

lowers energy demand, it substantially lowers the total energy costs. It typically

utilizes a standard packaged chiller to produce ice at night or during off-peak

periods when the building’s electrical needs are at a minimum. The ice is stored

in modular tanks to provide cooling ton-hours to help meet the buildings cooling

load requirements the following day. By doing so, it minimizes the peak energy

usage during the utility daylight peaking period.

Full Storage Or Partial Storage?

Two load management strategies are possible with ice storage systems. When

utility rates call for complete load shifting, a conventionally sized chiller can be

used to shift the entire load into off-peak hours. This is called a full storage

system and is used most often in existing building renovation or retrofit

applications using existing installed chiller capacity.

In new construction, a partial, storage system is usually the most practical and

cost effective load management strategy. In this load leveling method, the chiller

is sized to run continuously except for scheduled preventive maintenance down

time. It usually charges the ice storage tanks at night and cools the load directly

during the daytime peak hours with help from stored cooling capacity.

This will greatly reduce the installed chiller capacity and its required capital

expenditure, as well as the demand charge for electricity to run the chiller during

utility peaking periods. Typically reductions can be 50 percent or more.

How the Ice Storage System Works

A common ice storage system is a modular, insulated tank. Tanks are typically

available in several ton-hour rated sizes. Typically at night a mild concentration

of glycol-water solution (typically 25 percent ethylene glycol based industrial

coolant such as Dow Chemical Company Dowtherm SR-1 or Union Carbine

Corporation’s UCAR Thermofluid 17) from a standard packaged air conditioning

water chiller system circulates through the heat exchanger and extracts heat until

eventually all the water in the tank is frozen solid. The ice is built uniformly

throughout the tank.

Typical schematic flow diagrams for a partial storage system are shown in

figure1&2. At night, the water-glycol solution circulates through the chiller and the

ice bank heat exchanger, bypassing the air handling coil that supplies

conditioned air to occupied building spaces. During the day, the solution is

cooled by the ice bank from 52 F to 34 F. A temperature modulation valve set at

44 F in a bypass loop around the ice bank permits a sufficient quantity of 52 F

fluid to bypass the ice bank permits a sufficient quantity of 52 F fluid to bypass

the ice bank, mix with the 34F fluid, and achieve the desired 44 F temperature.

The 44 F fluid then enters the coil, where it cools air from approximately 75 F to

55 F. The fluid then leaves the coil at an elevated temperature (approximately

60F) and enters the water chiller where it is cooled 60F to 52 F.

It is important to note that while making ice at night, the chiller must cool the

water- glycol solution down to 26 F, rather than producing 44 F water required for

conventional air conditioning systems.

Chillers with air-cooled condensing also benefit from cooler outdoor ambient dry

bulb temperatures to lower the system condensing temperature at night.

The temperature modulating valve in the bypass loop has the added advantage

of providing excellent capacity control. During mild temperature days, typically in

the spring and fall, the chiller will often be capable of providing all the necessary

cooling capacity for the building without the use of cooling capacity from the ice

storage system. When the building’s actual cooling load is equal to or less than

the chiller capacity at the time, all of the system coolant will flow through the

bypass loop as shown in fig.3

It is important that the coolant chosen by an ethylene glycol-based industrial

coolant, such as Dowtherm SR-1 or UCAR Thermofluid 17, which is specially

formulated for low viscosity and good heat transfer properties. Either of these

fluids contain a multi-component corrosion inhibitor which is effective with most

materials of construction including aluminium, copper, silver solder and plastics.

Further, they contain no anti-leak agents and produce no films to interfere with

heat transfer efficiency. They also permit use of standard pumps, seals and air

handling coils. It should be noted, however, that because of the slight difference

in heat transfer properties between water and the mild glycol solution, the cooling

coil capacities will need to be increased by approximately 5 percent. It is also

important that the water and glycol solution be thoroughly mixed before the

solution is placed into the system.

The use of ice storage system technology opens new doors to other economic

opportunities in system design. These offer significant potential for not only first-

cost savings but also operating cost savings that should be evaluated on a life

cycle cost basis using a computerized economic analysis program.

Back-Up

Most A/C and refrigeration systems require some form of stand-by, or back-up,

facility to protect against system failure and costly lost production time. The ice

storage system, is an ideal, efficient solution for these applications. The ice

storage system offers rapid response back-up in the form of an independent,

static technology solution which ensures the highest degree of reliability.

Advantages of Ice Storage Systems

· Reduced installed plant capacity.

· Reduced electrical installation for lower investment and saving in demand

charges.

· Reduced installed cooling tower capacity incase of water cooled system.

· Reduced installed D.G. set capacity.

· Better plant utilization with longer equipment life and lower operating costs.

· Use of off-peak energy for lower energy bill, where differential tariff is applicable.

Applications of Ice Storage System

· Air-conditioning of industrial and commercial buildings - Offices, Hotels,

Shopping Complexes, Supermarkets, etc.

· Air-conditioning of data-processing centers, hospitals, telephone exchanges, etc.

requiring added system reliability and security.

· Dairy plants, Breweries, Food Processing, Bottling Plants, Chemical and

Fertilizer Plants, Pharmaceuticals, etc.

DUCTING DESIGN

The satisfactory distribution of conditioned air requires a well designed and energy

efficient air transport system with appropriate ducts and fans plus air treatment and

control devices.

The various method of duct designs, proper fan selection and control and methods of air

distribution system control for acceptable comfort and air quality in the conditioned

spaces are some of the points to be discussed.

The various methods of duct designing are

a. Constant Velocity method

b. Equal friction method

c. Static regain method

Classification of Ducts

Supply and return duct systems are classified with respect to the velocity and pressure

of the air within the duct.

Velocity

There are tow types of air transmission systems used for airconditioning application.

They are called Coventional or Low Velocity and High Velocity system. The dividing line

between these systems is rather nebulous but, for the purpose of this section, the

following initial supply air velocities are offered as a guide.

1. Commercial comfort air conditioning

a. Low velocity – upto 2500 fpm normally between 1200 & 220 fpm

b. High velocity – above 2500 fpm

2. Factory comfort airconditioning

a. Low velocity - upto 2500 normally between 2200 and 2500 fpm.

b. High velocity - above 2500 to 5000 fpm

Normally return air systems for both low and high velocity supply air systems are

designed as low velocity systems. The velocity range for commercial and factory

comfort application is as follows:

1. Commercial comfort airconditioning - low velocity upto 2000 fpm. Normally

between 1500 and 1800 fpm.

2. Factory comfort airconditioning - low velocity upto 2500 fpm. Normally between

1800 and 2200 fpm.

Pressure

Air distribution systems are divided into three pressure categories; low, medium and

high. These divisions have the same pressure ranges as Class I, II & III fans and

indicated:

1. Low pressure - upto 3¾ inch wg - class I fan

2. Medium pressure - from 3¾ to 6 ¾ inch wg - class II fan

3. High pressure - from 6 ¾ to 12 ¾ inch wg - class III

These pressure ranges are total pressure, including the losses through the air handling

apparatus, ductwork and the air terminal in the space.

The choice of design method depends almost entirely upon the size of the ductwork

installation. Small duct systems (homes, shops or a few office rooms) are commonly

designed by the velocity method. Large high pressure systems are most frequently

designed by computer software programs using the static regain method. Duct

arrangements between these two extremes are nearly always laid out by the equal

friction method. Sometimes a duct arrangement will be designed by a combination of

two methods. For instance, the trunk duct will be laid out by the static regain method

and the branch duct runs designed by the equal friction method.

In designing ductwork, a new term called “unit friction” will be utilized which means the

friction loss per 100 ft of duct work equivalent length.

Regardless of the duct design method chosen by the air transport system designer, the

final design and duct layout will likely result from the use of computerized duct design

and drafting programs available that are based on algorithms from the ASHRAE hand

book of fundamentals and other test data from SMACNA.

Recommended and maximum Duct Velocities for Conventional System

DESIGNATION RECOMMEND VELOCITIES, FPMRESIDENCES SCHOOLS, THEATERS,

PUBLIC BUILDINGSINDUSTRIAL

BUILDINGOutdoor air intakes1 500 500 500Filters1 250 300 350Heating coils1,2 450 500 600Cooling coils1 450 500 600Air washers1 500 500 500Fan outlets 1000 - 1600 1300 - 2000 1600 - 2400Main ducts2 700 - 900 1000 - 1300 1200 - 1800Branch ducts2 600 600 - 900 800 - 1000Branch risers2 500 600 - 700 800

Maximum velocities, FPMOutdoor air intakes1 800 900 1200Filters1 300 350 350Heating coils1,2 500 600 700Cooling coils1 450 500 600Air washers 500 500 500Fan outlets 1700 1500 - 2200 1700 - 2800Main ducts2 800 - 1200 1100 - 1600 1300 - 2200Branch ducts2 700 - 1000 800 - 1300 1000 - 1800Branch risers2 650 - 800 800 - 1200 1000 - 1600

1 These velocities are for total face area, not the net free area : other velocities in table

are for net free are2 For low velocity systems only.

@ 1965 American society of heating, refrigerating and airconditioning engineers, inc.

reprinted by permission for ASHRAE guide and data book.

PIPING DESIGN

The water piping system are divided into once thru and re-circulating types. In a once

thru system water passes thru the equipment only once and is discharged. In a re-

circulating system water is not discharged, but flows in a repeating circuit from the heat

exchanger to the refrigeration equipment and back to the heat exchanger.

Open and Closed

Both types are further classified as open or closed systems. An open system is one in

which the water flows into a reservoir open to the atmosphere; cooling towers and air

washers are examples of reservoirs open to the atmosphere. A closed system is one in

which the flow of water is not exposed to the atmosphere at any point. This system

usually contains an expansion tank that is open to the atmosphere but the water area

exposed is insignificant.

Water Piping Design

There is a friction loss in any pipe thru which water is flowing. This loss depends on the

following factors:

1. Water velocity

2. Pipe diameter

3. Interior surface roughness

4. Pipe length

System pressure has not effect on the head loss of the equipment in the system.

However, higher than normal system pressures may dictate the use of heavier pipe,

fittings and valves along with specially designed equipment.

To properly design a water piping system, the engineer must evaluate not only the pipe

friction loss by the loss thru valves, fittings and other equipment. In addition to these

friction losses, the use of diversity in reducing the water quantity and pipe size is to be

considered in designing the water piping system.

Pipe Friction Loss

The pipe friction loss in a system depends on water velocity, pipe diameter, interior

surface roughness and pipe length. Varying any one of these factors influences the

total friction loss in the pipe.

Most air conditioning applications use either steel pipe or copper tubing in the piping

system.

Charts enclosed are for schedule 40 pipe upto 24 inch diameter. Chart shows the

friction losses for closed re-circulation piping systems and for once thru / open re-

circulation piping systems.

These charts show water velocity, pipe or tube diameter, and water quantity, in addition

to the friction rate per 100 ft of equivalent pipe length. Knowing any two of these

factors, the other two can be easily determined from the chart. The effect of inside

roughness of the pipe or tube is considered in all these values.

The water quantity is determined from the airconditioning load and the water velocity by

pre-determined recommendations. These two factors are used to establish pipe size

and friction rate.

Water Velocity

The velocities recommended for water piping depend on two conditions;

1. The service for which the pipe is to be used.

2. The effect of erosion.

The design of the water piping system is limited by the maximum permissible flow

velocity.

Recommend Water Velocity

SERVICE VELOCITY RANGE (FPS)

Pump discharge 8 - 12

Pump suction 4 - 7

Drain line 4 - 7

Header 4 - 15

Riser 3 - 10

General service 3 - 10

City water 3 - 7

B. VENTILATION SYSTEM

Outdoor air that flows through a building either intentionally as ventilation air or

unintentionally as infiltration (and exfiltration) is important for two reasons.

Dilution with outdoor air is a primary means of controlling indoor air contaminants

and the energy associated with heating or cooling this outdoor air is a significant,

if not a major, load on the heating and air-conditioning system. For maximum

load conditions, it is essential to know the magnitude of this air flow to properly

size equipment; for average conditions, to properly estimate average or seasonal

energy consumption; and for minimum conditions, to assure proper control of

indoor contaminants. In larger buildings, it is important to know ventilation

effectiveness. Knowledge of smoke circulation patterns can be crucial in the

event of fire.

Ventilation occurs by two means, natural and forced, Natural ventilation can be

classified as (1) infiltration or (2) controlled. Manually controlled natural

ventilation is the ventilation from operable windows, doors or other openings in

the buildings envelope. The latter is an important means of ventilation in

residences in mild weather when infiltration is minimal or in warm climates to

avoid air conditioning costs.

Forced ventilation is mandatory in larger buildings where a minimum amount of

outdoor air is required for occupant comfort. Air contaminant measurement

technology has advanced to include alternate methods designed to assure that

indoor air quality meets specified conditions. These methods permit the amount

of outdoor air to vary according to the actual requirements of occupants in the

space.

This chapter focuses on envelope or shell-dominated buildings; i.e., residences

or small commercial buildings in which the energy load is determined by the

construction and performance of the building envelope. The physical principles

discussed also apply to large buildings. However, in large buildings, ventilation

energy load and indoor air quality conditions depend more on ventilation system

design that on building envelope performance.

Ventilation Requirements

The amount of ventilation needed has been debated for over a century, and the

different rationales developed have produced radically different ventilation

standards. Considerations such as the amount of air expel exhaled air, moisture

removal from indoor air and control of carbon dioxide (CO2) were each primary

criteria used at different times during the nineteenth century.

This research investigated the ventilation rates required to keep body-generated

odors below an acceptable level in rooms with comfortable levels of temperature

and humidity. It was found that the required ventilation rates varied considerably,

depending on the cleanliness of the subjects and the number present in the

room. Researches also found that CO2 concentration was not a good indicator of

the ventilation rate above 17 m3h per person; the CO2 concentration was almost

always lower than expected for a given ventilation rate. However, below 17

m3/hr per person the discrepancies were not so great and in fact the current

rationale for the 8.5 m3/h per person minimum outside air requirement in

Standard 62 is based on the CO2 produced by an individual depends on diet and

activity level. A representative value of CO2 production by a sedentary individual

who eats a normal diet is 0.019 m3/hr.

Types of Ventilation

Several techniques are possible to achieve the ventilation specified in the

standards; in decreasing order of desirability they are : (1) forced ventilation that

affords automatic control, (2) natural ventilation with manual control.

Forced ventilation is rarely used in envelope-dominated structures. However,

tighter, more energy conserving buildings with less infiltration require mechanical

ventilation systems. When coupled with an air-to-air heat exchanger, adequate

ventilation is provided at lower operating cost.

Natural ventilation is driven by pressures from wind and indoor-outdoor

temperature differences, causing air movement. This type of ventilation is

characterized by occupant control. Airflow through openable windows, doors and

other design openings can be used provide adequate ventilation for contaminant

dilution and temperature control.

Natural Ventilation

Natural or passive ventilation occurs because of wind and thermal pressure that

produce a flow of outdoor air through openable windows, doors and other

controllable openings. This is in contrast to infiltration, airflow through the

unintentional openings of a buildings. Natural ventilation can be used effectively

for both temperature and contaminant control. Temperature control by natural

ventilation conserves energy during the cooling season and is particularly

effective in mild climates. The arrangement, location and control of ventilation

openings should be designed to combine the driving forces of wind and

temperature.

Natural Ventilation Openings

Types of natural ventilation openings include: (1) windows, doors monitor

openings and skylights, (2) roof ventilators, (3) stacks connecting to registers and

(4) specially designed inlet or outlet openings.

Windows transmit light and provide ventilation when open. They may open by

sliding vertically or horizontally; by tilting on horizontal pivots at no near the

center ; or by swinging on pivots at the top, bottom or side. The type of pivoting

is an important consideration for weather protection.

Roof ventilators are determined to provide a weatherproof air outlet. Capacity is

determined by the ventilators offer to air flow; its ability to use kinetic wind energy

to induce flow by centrifugal or ejector action; and the height of the draft.

Natural draft or gravity roof ventilators can be stationary, pivoting or oscillating,

and rotating selection criteria are: ruggedness; corrosion-resistance: storm

proofing features; dampers and operating mechanism; possibility of noise;

original cost; and maintenance. Natural ventilators can supplement power-driven

supply fans the motors need only be energized when the natural exhaust

dampers or dampers controlled by thermostat or wind velocity.

A roof ventilator should be positioned so that it receives the full, unrestricted

wind. Turbulence created by surrounding obstructions, including higher adjacent

buildings, impairs a ventilator’s ejector action. The ventilator inlet should be

conical or bell mounted to give a high flow coefficient. The opening area at the

inlet should be increased if screens. Building air inlets at lower levels should be

larger than the combined throat areas of all roof ventilators.

Stacks or vertical flues should be located where wind can act on them from any

direction. Without wind, stack effect alone removes air from the room with the

inlets.

Required Flow

The ventilation flow needed to remove a given amount of heat from a building

can be calculated from Eq. (1) if the quantity of heat to be removed and the

average indoor-out-door temperature difference are known.

Q =

Where

Q = Air flow removed, m3/hr

H = Heat removed, W

Cp = Specific heat of air at constant pressure, 1 KJ.K

T1-T0 = Average indoor-outdoor temperature difference, K

cf1 = Conversion factor, 0.28

cf1 = Conversion factor, 0.34

Flow Caused by Wind

Factors that affect ventilation wind forces include average speed, prevailing

direction, seasonal and daily variation in speed and direction, and local

obstructions such as nearby buildings, hills trees and shrubbery.

Wind speeds are usually lower in summer than in winter; frequency from various

directions differs in summer and winter. There are relatively few places where

speed falls below half the average for more than a few hours a month. Therefore

natural ventilating systems are often designed for wind speeds of half the

average seasonal velocity.

Equation (2) shows the quantity of air forced through ventilation inlet openings by

wind or determines the proper size of openings to produce given airflow rates:

Q = (cf)CvAv (2)

Where

Q = Air flow m3/hr

A = Free area of inlet openings, m2

Wind speed, m/s

Cv = Effectiveness of openings (Cv) is assumed to be 0.50 to

0.60 for perpendicular winds

cf = Conversion factor, 3600

Inlets should face directly into the prevailing wind direction. If they are not

advantageously placed flow will be less than in the equation: if unusually well

placed flow will be slightly more. Desirable outlet locations are (1) on the

leeward side of the building directly opposite the inlet (2) on the roof.

Ventilation and Infiltration

In the pressure area caused by a flow discontinuity of the wind, (3) on the

adjacent to the windward face where low pressure areas occur, (4) in a monitor

on the leeward side, (5) in roof ventilators or (6) by stacks. Refer to Chapter 14

for a general description of wind on a building.

Flow Caused by Thermal Forces

If there is not significant building internal resistance, the flow caused by stack

effect is:

Q = (cf)A[h(T1 – T0)/T1]1/2

Where

Q = Air flow, m3/hr

A = Free area of inlets or outlets (assumed equal), m3

H = Height from lower opening to NPL, m

T1 = Average temperature of indoor air in height h,

K[=t(deg.c)+273.15]

T0 = Temperature of outdoor air, K

cf = Conversion factor, including a value of 65% for

effectiveness of openings; this should be 50% if conditions

are not favourable (cf = 10360)

The height h is the distance from the lower opening to the neutral pressure level.

Natural Ventilation Guidelines

Several general rules should be observed in designing for ventilation:

1. Systems using natural ventilation should be designed for effective

ventilation regardless of wind direction. Ventilation must be adequate

when the wind does not come from the prevailing direction.

2. Inlet openings should not be obstructed by buildings, signboards or

indoors partitions.

3. Greatest flow per unit area of total opening is obtained by inlet and outlet

openings of nearly equal areas.

4. The neutral pressure level tends to move to the level of any single

openings, resulting in pressure reduction across the opening. Two

openings on opposite sides of a space increase the ventilation flow. If the

openings are at the same level and near the ceiling, much of the flow may

bypass the occupied level and the ineffective in diluting contaminants at

the occupied level.

5. There must be vertical distance between openings for temperature

difference to produce natural ventilation; the greater the vertical distance,

the greater the ventilation.

6. Openings in the vicinity of the NPL are least effective for thermally induced

ventilation.

7. Openings with areas much larger than calculated are sometimes desirable

when anticipating increased occupancy or very hot weather. The

openings should be accessible to and operable by occupants.

8. When both wind the stack pressures act together, even without

interference, estimated resulting airflow is not equal to the two flows

separately. Flow through any openings is proportional to the square root

of the sum of the squares of the two flows calculated separately.

C. FORCED VENTILATION

This involves forced supply systems, forced exhaust systems or both, depending

on the requirements.

This is done by fans of various types, including propeller fans, axial flow fans and

centrifugal fans. Propeller fans are generally wall mounted type and cater to

small capacity / small pressure static requirements. Axial fans can either be duct

mounted or wall mounted type and cater to medium capacity requirement.

Centrifugal fans, which are a separate topic by themselves, cater to a wide range

of capacity and static pressure requirements.

The later two types of fans can be hooked up to a supply or exhaust duct system.

They can also be hooked up to an air washer / fan-filter system.

Forced ventilation systems can also be classified into dry or wet systems. Dry

systems involve the use of fans alone or with filter banks for dust removal.

Wet ventilation involves the use of fans with filters and a water spray / water

logged fill arrangement which will humidity hot, dry air and cool it. These

systems are suitable for hot and dry areas and are not effective in high humidity /

coastal areas.