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
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
b) To improve productivity in offices, factories by maintaining comfort
conditions for persons .
c) To maintain comfortable conditions for working in hotels, labs, etc.,
d) To avoid malfunctioning of some of the control panels in Electrical Control
e) To maintain over pressure inside the premises for avoiding outside (dusty)
air in to the room.
f) To create healthy atmosphere inside the room by supplying filtered air in
to the room.
g) To provide clean, filtered, healthy, comfortable conditions in hospitals etc.,
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
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.
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.
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.
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 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
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.
BASIC REFRIGERATION SYSTEM
The various components which form part of a refrigeration system can be
described as follows:
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.
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.
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.
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
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
ELEMENTS OF 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
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.
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
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.
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
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.
The temperature of air as registered by on ordinary temperature.
The temperature registered by a thermometer whose bulb is covered by a wetted
wick and exposed to a current of rapidly moving air.
The temperature at which condensation of moisture begins when the air is
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.
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 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.
The cubic feet of the mixture per pound of dry air.
Sensible Heat Factor
The ratio of sensible to total heat.
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.
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
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
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.
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
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
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
where 60 = min/hr
13.5 = Specific volume of moist air at 70 F db and 50% rh.
HEAT LOAD ESTIMATION
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
h) Provide information for equipment selection and HVAC system design
i) Provide data for evaluation of the optimum possibilities for load reduction.
j) Permit analysis of partial loads as required for system design, operation
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:
k) Heat Gain or Loss
l) Cooling load or Heating Load
m) 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:
n) Solar radiation through transparent surfaces such as windows
o) Heat conduction through exterior walls and roofs
p) Heat conduction through interior partitions ceilings and floors
q) Heat generated within the space by occupants, lights, appliances,
equipment and processes
r) Loads as a result of ventilation and infiltration of outdoor air
s) 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
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
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
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
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)
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
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
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:
t) Orientation and dimensions of building components.
u) Construction materials for roof, walls, ceiling, interior partitions, floors and
v) Size and use of space to be conditioned
w) 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:
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 = VxC
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
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.
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
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)
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.
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, 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
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:
x) Occupancy: The sensible/latent heat gains from people are given
in tables, based on the nature of their activities in the room.
y) 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
z) Appliances: Electrical, gas burners, steam generation, etc.
aa)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:
bb)Temperature & humidity conditions to be maintained inside the space
cc) Occupancy – i.e. maximum no. of people likely to occupy the space
and the nature of their activity.
dd)Lighting load and other internal source of heat generation.
ee)Period of operation – e.g. 8 a.m. to 7 p.m. or 10 a.m. to 8 p.m. etc.
ff) For industrial application you require also the HP load in the
conditioned space and diversity factor thereon.
gg)Minimum ventilation required.
3. Outside Design Conditions
hh)For comfort air conditioning application, use the mean maximum DB
temperature & the WB temperature which occurs simultaneously with
the assumed DB.
ii) 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
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 DIVERSITY FACTORS FOR LARGE BUILDINGS
(APPLY TO REFRIGERATION CAPACITY)
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.
10 CFM per person
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 = R o + X1 R 1 + X2 R 2 + X3 R 3 + R i
= 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_
= 0.337 BTU / hr. ft2. °F
RT = R i + X1 R 1 + X2 R 2 + X3 R 3 + R i
= 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_
= 0.294 BTU / hr. ft2. °F
c. Roof exposed to sun
RT = R i + X1 R 1 + X2 R 2 + R i
= 0.25 + 150 x 0.2 + 50 x 4.0 + 0.92
= 10.37 hr. ft2. °F /BTU
Overall heat transfer Co-efficient
= 1_ = 1_
8.0 HEATING, VENTILATION &
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
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.
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.
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
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
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.
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.
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
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.
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.
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.
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
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
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
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
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
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
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
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
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
• 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
• Use of off-peak energy for lower energy bill, where differential tariff is
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.
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.
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
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.
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
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
Recommended and maximum Duct Velocities for Conventional System
DESIGNATION RECOMMEND VELOCITIES, FPM
RESIDENCES SCHOOLS, THEATERS, INDUSTRIAL
PUBLIC BUILDINGS BUILDING
Outdoor air intakes1 500 500 500
Filters1 250 300 350
Heating coils1,2 450 500 600
Cooling coils1 450 500 600
Air washers1 500 500 500
Fan outlets 1000 - 1600 1300 - 2000 1600 - 2400
Main ducts2 700 - 900 1000 - 1300 1200 - 1800
Branch ducts2 600 600 - 900 800 - 1000
Branch risers2 500 600 - 700 800
Maximum velocities, FPM
Outdoor air intakes1 800 900 1200
Filters1 300 350 350
Heating coils1,2 500 600 700
Cooling coils1 450 500 600
Air washers 500 500 500
Fan outlets 1700 1500 - 2200 1700 - 2800
Main ducts2 800 - 1200 1100 - 1600 1300 - 2200
Branch ducts2 700 - 1000 800 - 1300 1000 - 1800
Branch 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 are
2 For low velocity systems only.
@ 1965 American society of heating, refrigerating and airconditioning engineers,
inc. reprinted by permission for ASHRAE guide and data book.
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
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
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.
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
Recommend Water Velocity
SERVICE VELOCITY RANGE
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
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
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 m 3/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
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 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.
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
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
Q = (cf)CvAv (2)
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
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,
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
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
2. Inlet openings should not be obstructed by buildings, signboards or
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
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
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
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
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.