hva cprogramme material
DESCRIPTION
Provides introduction to HVACTRANSCRIPT
HEATING VENTILATION & AIRCONDITIONING
SYSTEM DESIGN
FACULTY CO-ORDINATOR
NEERAJ SHUKLA
CONTENTS
PAGE NO.
1.0 INTRODUCTION AND OVERVIEW - 2
2.0 FUNDAMENTALS OF REFRIGERATION - 5
3.0 BASIC REFRIGERATION SYSTEM - 12
4.0 ELEMENTS OF PSYCHROMETRY - 16
5.0 APPLIED PSYCHROMETRY - 23
6.0 HEAT LOAD ESTIMATION - 33
7.0 HEAT LOAD DATA SHEET & TYPICAL CALCULATION - 47
8.0 HEATING VENTILATION & AIRCONDITIONING SYSTEMS - 52
1.0 INTRODUCTION AND OVERVIEW
INTRODUCTION AND OVERVIEW
A simple definition of air conditioning is the simultaneous control of temperature,
humidity, air movement, and the quality of air in a space.
The use of the conditioned space determines the temperature, humidity, air movement,
and quality of air that must be maintained.
The primary function of air conditioning is to maintain conditions that are (1) conducive
to human comfort, or (2) required by a product, or process within a space. To perform
this function, equipment of the proper capacity must be installed and controlled
throughout the year. The equipment capacity is determined by the actual instantaneous
peak load requirements; type of control is determined by the conditions to be
maintained during peak and partial load. Generally, it is impossible to measure either
the actual peak or the partial load in any given space; these loads must be estimated.
The term ‘refrigeration’ may be defined as the process of removing heat from a
substance under controlled conditions. It also includes the process of reducing and
maintaining the temperature of a body below the general temperature of its
surroundings. In other words, the refrigeration means a continued extraction of heat
from a body whose temperature is already below the temperature of its surroundings.
For example, if some space (say in cold storage) is to be kept at - 2 °C (271 K), we
must continuously extract heat which flows into it due to leakage through the walls and
also the heat which is brought into it with the articles stored after the temperature is
once reduced to - 2 °C (271 K). Thus in a refrigerator, heat is virtually being pumped
from a lower temperature to a higher temperature. According to second law of
Thermodynamics, this process can only be performed with the aid of some external
work. It is obvious that supply of power (say electric motor) is regularly required to drive
a refrigerator. Theoretically, a refrigerator is a reversed heat engine or a heat pump
which pumps heat from a cold body. The substance which works in a heat pump to
extract heat from a cold body and to deliver to a hot body is called a refrigerant.
The refrigeration system is known to the man since the middle of nineteenth century.
The scientists, of the time, developed a few stray machines to achieve some pleasure.
But it paved the way by inviting the attention of scientists for proper studies and
research. They were able to build a reasonably reliable machine by the end of
nineteenth century for refrigeration jobs. But with the advent of efficient rotary
compressors and gas turbines, the science of refrigeration reached the used for the
cooling of storage chambers in which perishable foods, drinks and medicines are
stored. The refrigeration has also wide applications in submarine ships, aircraft and
rockets.
Air conditioning has got wide range of applications and it is very much essential in these
days. Air conditioning is provided for some of the following reasons
1) To improve productivity in offices, factories by maintaining comfort conditions for
persons .
2) To maintain comfortable conditions for working in hotels, labs, etc.,
3) To avoid malfunctioning of some of the control panels in Electrical Control
Buildings.
4) To maintain over pressure inside the premises for avoiding outside (dusty) air in
to the room.
5) To create healthy atmosphere inside the room by supplying filtered air in to the
room.
6) To provide clean, filtered, healthy, comfortable conditions in hospitals etc.,
2.0 FUNDAMENTALS OF
REFRIGERATION
FUNDAMENTALS OF REFRIGERATION
Refrigeration is the process by which heat is removed from a low temperature level and
rejected at a relatively higher temperature level. The Americans define refrigeration in a
somewhat different way, thus, refrigeration is the process by which heat is removed
from a place where it is not required and rejected into a place where it is not
objectionable. This is not, strictly speaking, a proper scientific definition, since it does
not make any mention of its temperature levels. No process can be called
“refrigeration” unless removal of heat is at a temperature lower than the surrounding
temperature.
By nature heat always flows from one body to another body which is relatively at a lower
temperature. This law of nature cannot be altered by any means. Transferring heat
from a low temperature level to a high temperature level is analogous to transferring
water from a lower level to a higher level. Imagine two water tanks, one located at the
ground floor full of water and another empty, located at the roof level of a building. If
water from the ground floor tank is to be transferred to the roof tank, then the only thing
to do is to bring a bucket, place it at a level lower than the ground floor tank and allow
the water to initially drain into the bucket according to the law of nature. The second
step would be to lift this bucket full of water to a level above the roof tank and now allow
the water from the bucket to drain into the roof tank according to the natural flow by
gravity. In the foregoing process, we have used the bucket as the carrier and moved it
up and down, first to a level lower than the ground floor tank and then to a level higher
than the roof tank. Needless to add that in the process some mechanical work has
been performed for lifting the bucket from the lower level to the higher level.
Applying this analogy to the process of refrigeration, it is evident, that we require a
substance as the carrier of heat analogous to the bucket. This substance should be first
brought to a temperature which is lower than the low level temperature so that heat from
the low temperature level will automatically flow into this carrier substance which has
been brought to a still lower level of temperature. After this carrier substance has been
fully loaded with heat it has got to be raised to a temperature which is higher than the
high level temperature so that heat from this carrier will automatically flow according to
the law of nature. The carrier substance referred to above is what is known in
refrigeration parlance as “refrigerant”. We shall now see what a refrigerant is really like.
All volatile liquids including water have property whereby the temperature at which they
evaporate changes according to the pressure it is subjected to. Take water for example.
At normal atmospheric pressure it boils at 100°C (212°F). When the water is subjected
to higher pressure its boiling temperature also becomes higher than 100°C. Likewise, if
the water is subjected to pressures lower than the atmosphere its boiling temperature
also falls below 100°C. In fact water can boil even at as low a temperature as 4°C when
it is kept in vacuum free of air. In this case the only pressure it will have above is its
own vapour pressure. Different volatile substances have different pressure-boiling point
characteristics. For refrigeration purpose the most commonly used refrigerants are
refrigerant 12 and refrigerant 22. By reference to a table giving the properties of these
refrigerants is will be seen that for each pressure there is a corresponding temperature
at which only the refrigerant will boil. It goes without saying that at any given pressure
the temperature at which the liquid refrigerant boils is also the temperature at which the
refrigerant vapour would condense back to liquid form. Whether it is boiling or
condensing all depends on whether it is receiving heat or giving heat. For refrigerant
purpose, therefore, we make use of this natural property of the volatile refrigerant. For
example, if the liquid refrigerant R-22 is by some means or other brought down to an
absolute pressure of say 83.72 psi, then this liquid is now in a position to boil at a
temperature of 40°F. In order to make this liquid at low pressure boil, you will have to
supply heat equivalent to the latent heat of vaporization.
So any body which is above 40°F in temperature can supply this latent heat to make
this liquid boil and turn completely into gaseous form, it goes without saying that the
substance which supplies the heat for boiling the refrigerant will itself naturally cool
down, but in any case not to a temperature below 40°F which is the temperature at
which the refrigerant is boiling. Once the refrigerant liquid has completely vaporized, it
has no further capacity to absorb heat from the heat supplying body, just in the same
way that the bucket had no further capacity to take more water the moment it was full.
In the case of bucket in order to make it useful again, we had to raise this to a higher
level and empty out its contents in the roof tank and then bring it back once again to
take another bucket full from the ground floor tank. In a similar way the completely
vaporized refrigerant will have to be raised to a high temperature level. By raising it to a
higher temperature level, we do not mean that you simply heat up the refrigerant. What
is really to be done is this low pressure refrigerant will have to be compressed in a
compressor to a higher pressure. The pressure to which it is compressed must be such
that the boiling point or conversely the temperature at which the refrigerant vapour
would condense should be higher than the body to which we want to empty out the heat
content of the carrier refrigerant. If, for example, the body to which we want to reject
this heat is water which is at a temperature of 90°F, then the boiling temperature or
condensing temperature of the refrigerant should be higher than 90°F, say 106°F. Now
106°F happens to be the boiling temperature of the refrigerant when it is at a pressure
of 230 psi absolute. This means that the low pressure refrigerant will have to be
compressed to a pressure of 230 psi absolute before it is in a position to empty out its
heat content to the cooling medium, namely, natural water. Note that in the process we
have not violated any natural law concerning heat flow. We have all along allowed heat
to flow from a certain temperature level to a lower temperature level. What we have
done is that we have changed the temperature level of the carrier refrigerant to higher
or lower temperature levels to conform to the natural law of heat flow.
This brings us to a stage where we have to know the definition of certain terms which
are generally used in refrigeration parlance.
Saturation Temperature
For any given refrigerant the temperature at which the liquid refrigerant would boil (or
conversely the refrigerant vapour would condense) when it is subjected to a certain
pressure is defined as the saturation temperature corresponding to that pressure. It is
obvious that at this temperature and pressure, refrigerant in liquid and vapour form kept
in a closed container would be in equilibrium with each other. So long as the pressure
inside this container is maintained steady the liquid portion will vaporize if heat is added
or the vapour portion will condense if heat is removed. On this basis, saturation
temperature for any given pressure is defined as that temperature at which liquid
refrigerant and its vapour remain in contact with each other in equilibrium.
Superheat
We saw in the previous paragraph that liquid refrigerant and its vapour will be in
equilibrium with each other in a closed container at the saturation temperature
corresponding to the pressure. Any addition or removal of heat would only result in
either liquid vaporizing or the vapour condensing, pressure remaining same. However,
if the vaporized refrigerant is separated from the liquid portion, then any heat added to
this refrigerant in vapour form would only go to raise the temperature of the vapour
above its saturation temperature corresponding to its pressure. This is superheated
vapour. Superheat is usually expressed in terms of degrees. When we say 10°
superheat, what we mean is the gaseous refrigerant is at a temperature 10° above the
saturation temperature corresponding to its pressure.
Sub-cooling
In a like manner if the liquid portion of the refrigerant is separated and completely
isolated from the vapour which is in equilibrium with it then any removal of heat from this
refrigerant would lower its temperature to a value below its saturation temperature.
Such a liquid is called sub-cooled liquid. When we say the liquid is 15° below the
saturation temperature corresponding to its pressure. It is obvious that in the case of
sub-cooled liquid when heat is added it will first rise up in temperature till it reaches its
saturation temperature and thereafter only it will begin to boil as long as it is receiving
heat. Likewise, in the case of superheated gas, when heat is removed from the same it
will first fall down in temperature till it reaches the saturation temperature corresponding
to its pressure (this is generally referred to as de-superheating). Any further removal of
heat after this would result in condensation of the vapour into liquid form at constant
temperature, namely, the saturation temperature. One thing that should be borne in
mind is neither sub-cooling of liquid nor superheating of the vapour is possible when
liquid and its vapour are in contact with each other in equilibrium, because as already
explained earlier, any removal or addition of heat in this case would only respectively
result in condensing of the vapour part or evaporating of the liquid part, at constant
temperature.
Enthalpy
Enthalpy is the terms which denotes the heat content of the refrigerant from a base
saturation temperature of 40°F. At this temperature and the corresponding saturation
pressure the heat content of the liquid has been arbitrarily fixed as 0. It, therefore,
follows that the enthalpy of the liquid above 40°F will be positive and that below 40°F.
will be negative. The enthalpy of the refrigerant when it is in vapour form will be equal
to the enthalpy of the liquid at the same pressure and saturation temperature.
Adiabatic Compression
Any process which is performed without the addition of heat to or removal of heat from
the process is said to be an adiabatic process. Compression of gaseous refrigerant
without addition or removal of heat is called adiabatic compression. The pressure
enthalpy diagram of a refrigerant has also lines showing adiabatic compression. It is,
therefore, possible to find out the enthalpy and temperature of the gas at various
pressures during the course of compression.
3.0 BASIC REFRIGERATION SYSTEM
BASIC REFRIGERATION SYSTEM
The various components which form part of a refrigeration system can be described as
follows:
Evaporator
Let us start from the evaporator. Liquid at high pressure has to be admitted into the
evaporator. In order that this liquid may evaporate at low temperature, it is essential
that the liquid so admitted is simultaneously reduced in pressure. The level to which the
pressure has to be reduced of course is determined by the temp. at which you want this
liquid to evaporate. For example, if you want evaporation of refrigerant 22 at a
temperature of 40°F, the absolute pressure should be brought down to 83.72 psi or if the
evaporation has to be at 10°F the absolute pressure should be brought down to 31.29
psi. The pressures indicated above are the saturation pressures corresponding to the
respective temperatures. This pressure reduction is brought about by the use of what is
known as an expansion valve. The expansion valve is just a needle valve which
throttles the flow of liquid refrigerant thereby bringing about a pressure drop. This
expansion valve can also be hand operated, automatic or thermostatic. Liquid admitted
into the evaporator now needs heat for evaporation. This head is supplied by the air
which is flowing over the evaporator coil. In the process, the air gets cooled and the
liquid refrigerant evaporates.
Compressor
Now if you have got to ensure continuous evaporation at the same temperature, it is
very vital that the vapour evaporating in the coil is removed from it is as rapidly as it is
evaporating. Unless this is done the evaporated vapour will build up a pressure in the
coil which would keep on rising. Any such rise in pressure will naturally raise the
evaporating temperature also since, the evaporation temperature is higher and higher
as the pressure increases. Removal of the evaporator vapour is achieved by
connecting the outlet of the evaporator to the suction side of a refrigerating compressor.
Of course, the compressor has got to be sized so that it has got a volumetric rate of
displacement which matches with the evaporation rate. Thus the evaporation pressure
is maintained as steady and the liquid fed through the expansion valve continues to
evaporate at a steady temperature so long as heat for evaporation is available at an
equally steady rate from the air flowing over.
Condenser
The compressor compresses the vapour and discharges the same into the condenser.
It is in this condenser that the high pressure hot gas delivered by the compressor has to
be condensed. For the purpose of condensing the gas it is necessary that heat is
removed from the hot gas. This removal of heat is achieved by again creating an air
flow over the condenser coil or water flow if water cooled condensers are used. The
heat given up by the refrigerant is picked up by the air or water. The hot gas which has
given up the heat naturally condenses into liquid form at the same pressure. Now let us
see how the pressure built-up in the condenser coil it has got certain definite capacity to
transfer heat from within to the outside air or water for each degree of temperature
difference. We also know that for each 1b of refrigerant which has got to be condensed
into liquid form a definite capacity to transfer heat from within to the outside air or water
of refrigerant which has got to be condensed into liquid form, a definite amount of heat,
namely, the latent heat of condensation has to be removed. If in a refrigeration system
the F-22 circulation is say, 5 1bs/minute, then the amount of heat which has got to be
removed for condensing this refrigerant gas is 5 x latent heat. This means that the
temperature difference between the hot refrigerant gas within the condenser and the air
or water flowing over it should be such that the total amount of heat transferred through
the walls of the condenser tubes just balances with the total amount of heat which has
got to be removed. The condensation rate would, therefore, automatically balance with
the compressor discharge rate as soon as the temperature difference has been built up.
The pressure inside the condenser also which initially starts building up will attain a
steady level when the corresponding saturation temperature results in the desired
temperature difference for creating the desired heat transfer rate. This is called the
condesing temperature of the system. It is obvious that if you use a small size
condenser the temperature difference has necessarily to be higher and hence the
condensing temperature and the corresponding pressure will also have to be relatively
higher.
Receiver
A receiver is a pressure vessel which is used as a storage tank for the condensed liquid
refrigerant leaving the condenser. It is from this receiver that liquid is tapped and sent
to the evaporator through the throttling device or expansion valve. It is not on all
systems that we have a separate liquid receiver. In the case of systems having water
cooled condensers, the shell of the condenser itself serves as a storage vessel for the
liquid refrigerant. In smaller systems even with air cooled condensers, it is possible to
dispense with the use of a receiver if care is taken to charge the system with the correct
amount of refrigerant.
In order that the various components forming part of a refrigeration system can be
designed, it is necessary to make a more scientific study of the entire operations. For
this purpose we have to know the complete properties of the refrigerant concerned
when it is at gaseous form and also in liquid form. The properties of each refrigerant
are shown in what is called a Pressure Enthalpy Diagram.
4.0 ELEMENTS OF PSYCHROMETRY
ELEMENTS OF PSYCHROMETRY
Psychrometry
Since air conditioning, by its very name means treating air with a view to altering its
temperature and moisture content with the use of refrigeration, it is necessary that we
should know how exactly air would behave when it is subjected to cooling, heating,
humidifying or dehumidifying processes. For this purpose, it is necessary to study the
property of air at normal atmospheric pressure in so far as it concerns air conditioning.
Such a study is what is called psychrometry.
For the study of psychrometry, a chart has been devised, which is called Psychrometric
Chart. We will just now see what the various lines of the psychrometric chart are.
Dry Bulb Lines
Any vertical line is a line of constant temperature. Condition of air represented by any
point on this line will have the temperature corresponding to this vertical line. These
lines are called Dry Bulb Lines. By dry bulb what we really mean is dry bulb
temperature i.e., the temperature as recorded by a thermometer which is dry.
Moisture Content
Each horizontal line in the chart is a line of constant moisture content. The condition of
air represented by any point on this line will all have the same moisture content as
applicable to this line. Through any point on the psychrometric chart you can always
draw a horizontal line and a vertical line. Air represented by this point has, therefore, a
dry bulb temperature corresponding to the vertical line and moisture content
corresponding to the horizontal line. It is easy to see that air at any given temperature
can have varying moisture content. Likewise, air containing any given moisture content
can have varying temperature as well.
Saturation Line
The curved line on the extreme left-hand side of the chart is what is called the saturation
line. Condition of air represented by any point on this line is said to be saturated air,
which means that the air is having the maximum possible moisture content in it. It
cannot hold any further moisture.
Wet Bulb Lines
There are number of parallel slant lines which are called wet bulb lines. By wet bulb
temperature what we really mean is the temperature of the air as recorded by a
thermometer with a wet wick on its bulb. You will also understand for the moment that
the air having a certain wet bulb temperature will have a definite heat content although
its dry bulb temperature may be anything.
Relative Humidity Lines
When the air contains its maximum moisture content, we call it saturated air; when it
contains anything less than this maximum limit then it is not saturated air because it has
still capacity to have more moisture. We therefore, say that such air is, say 50%
saturated or 60% saturated. Another term used to denote the percentage saturation is
“relative humidity”. Thus it is one and the same thing whether you say air is 50%
saturated or air has got a relative humidity of 50%. Note that we have used the word
“approximately” because the strict scientific definition of relative humidity is not nearly
the comparison of moisture content. In fact relative humidity is defined as the ratio of
the partial vapour pressure in the air to the maximum vapour pressure that saturated air
will have at this temperature. However, for all practical purpose, this is equal to the ratio
of the actual moisture content present to the maximum moisture it can hold at that
temperature.
Dew Point
We have seen that at any given temperature air has a maximum limit of moisture
holding capacity when it is said to be saturated. For example from the psychrometric
chart we can see that 70°F saturated air can hold a maximum of 110 grains per 1b of
dry air. All temperatures above 70°F, air with the same moisture content will be, say
80%, 90% etc., saturated depending on what its dry bulb temperature would be. If air
with this moisture content and at temperature higher than 70° is cooled down, then its
condition will move along the horizontal 110 grains line, till the temperature falls to 70°F.
70°F and 110 grains / 1b as we have seen corresponds to saturated condition. This is
the temperature at which air with 110 grains of moisture / 1b will begin to shed its
moisture by condensing if you continue to cool the air. This temperature is called the
DEW POINT of the air. Needless to add, it is the moisture content which determines the
dew point. All you have to do is to move horizontally on the psychrometric chart and
read the temperature where you intersect the saturation line.
Enthalpy
We were just now referring to the wet bulb as line of constant heat content of air.
Enthalpy is just another term used in place of “heat content”. Of course, the enthalpies
represented here are all values for samples of air containing 1 lb of dry air.
The amount of moisture content in the air is generally expressed in terms of grains of
moisture per 1b of dry air. For your information, grain is a weight measure. 7000 grains
make 1 lb. When we say that the moisture content is 120 grains, what we mean is there
is 1 lb, of dry air containing 120 grains of moisture. The total weight of this moist air
would, therefore, by 1 + 120/7000 lbs = 1.0171 lb. At any temperature there is a limit to
the maximum moisture holding capacity of air. This limit is something definite and does
not alter except under different atmospheric pressures. At higher and higher
atmospheric pressure the moisture holding capacity at any given temperature becomes
less and less.
At any temperature when air contains the maximum amount of moisture it is said to be
saturated air. When air has attained saturation at any given temperature, it is
impossible to add any further moisture in vapour form.
Short Definitions
Dry-bulb Temperature
The temperature of air as registered by on ordinary temperature.
Wet-bulb Temperature
The temperature registered by a thermometer whose bulb is covered by a wetted wick
and exposed to a current of rapidly moving air.
Dewpoint Temperature
The temperature at which condensation of moisture begins when the air is cooled.
Relative Humidity
Ratio of the actual water vapor pressure of the air to the saturated water vapor pressure
of the air at the same temperature.
Specific Humidity or Moisture Content
The weight of water vapor in grains or pounds of moisture per pound of dry air.
Enthalpy
A thermal property indicating the quantity of heat in the air above an arbitrary datum. In
BTU per pound of dry air. The datum for dry air is 0 °F and, for moisture content, 32 °F
water.
Enthalpy Deviation
Enthalpy indicated above, for any given condition, is the enthalpy of saturation. It
should be corrected by the enthalpy deviation due to the air not being in the saturated
state. Enthalpy deviations in BTU per pound of dry air. Enthalpy deviation is applied
where extreme accuracy is required : however, on normal air conditioning estimates it is
omitted.
Specific Volume
The cubic feet of the mixture per pound of dry air.
Sensible Heat Factor
The ratio of sensible to total heat.
Alignment Circle
Located at 80 °F db and 50% rh and used in conjunction with the sensible heat factor to
plot the various air conditioning process lines.
Pounds of Dry Air
The basis for all pyschrometric calculations, remains constant during all psychrometric
processes. The dry-bulb, wet-bulb, and dewpoint temperatures and the relative
humidity are so related that if two properties are known, all other properties shown may
then be determined. When air is saturated, dry-bulb, wet-bulb, and dewpoint
temperatures are all equal.
5.0 APPLIED PSYCHROMETRY
APPLIED PSYCHROMETRY
Let us now see how the various air conditioning procedures will be represented on a
Psychrometric chart.
1. Sensible Heating
By sensible heating, we mean adding heat to air whereby the entire heat added
goes to raise the temperature of the air. It is obvious that in such a process there
is no change in the moisture content of the air. In other words, during sensible
heating process the air retains a constant moisture content and accordingly, its
condition will move on a horizontal line corresponding to its constant moisture
content. Since heat is being added during such process, its enthalpy also rises.
Therefore, during the heating process the wet bulb temperature of the air will also
rise, because as we have already seen, it is the wet bulb temperature lines which
are identified as constant enthalpy lines.
2. Addition of Moisture
Likewise, if moisture is somehow or the other added to the air without adding any
sensible heat, the process would be represented by a vertical line corresponding
to its dry bulb temperature. In this case also, since the moisture added carries
with it the latent heat of vaporization of water, the heat content of the air also
rises and hence its wet bulb temperature also rises.
3. Heating and Humidifying
If heat is added so that part of it goes to raise the temperature and the remaining
part goes to vaporize water and add it to the air, such a process is called heating
and humidifying.
4. Cooling and Dehumidifying
Cooling and dehumidifying is just the reverse of heating and humidifying. On a
psychrometric chart such a process will also be represented in the same manner
as for heating and humidifying, the only difference being the arrows representing
the direction of movement of conditions would be just reverse.
5. Evaporative Cooling
Evaporative cooling is the process by which air is simply subjected to a spray of
re-circulated water just as in the experiment described earlier, the only difference
being, we do not provide an infinite number of spray banks as in the experiment.
The chamber with the banks of spray is called an Air Washer. Air so subjected
would of course tend to get saturated and change out at a temperature equal to
its wet bulb temperature. However, since we do not provide adequate number of
spray banks to completely humidify, the air comes out not at 100% humidity but
somewhat lower than that. Needless to say, since this process is adiabatic, the
air has constant enthalpy throughout the process and hence its condition moves
along the line representing its wet bulb temperature.
Psychrometry as Applied to Airconditioning
It now remains for us to study psychrometry as applied to air conditioning
process. We will only see for the present what the heat load form is like and
also the various sections into which it is divided. It is only after you understand
this that you will be in a better position to understand psychrometry as applied to
air conditioning.
When a space is maintained at a temperature below the atmospheric
temperature surrounding the space, then there is a transfer of heat from outside
into the conditioned area, which tends to raise the inside temperature unless this
heat is removed as fast as it enters this space. Then you have heat or any other
appliances which may be in the space. All such heat which are either transmitted
into the room or generated from within due to occupants and appliances which
tend to raise the inside temperature are termed as room sensible heat. In the
like manner, the occupants within the room also release moisture from their body
into the room. There may be other sources inside the conditioned area which
add up more moisture into the atmosphere. If the space has not only to be
maintained at a particular temperature, but also to be held within certain limits of
relative humidity, then it is necessary that such moisture gain inside the room
should also be removed just as rapidly. By removal of moisture what we have
really mean is condensing this moisture from the air and discarding it outside.
For condensing the moisture, you have to remove the latent heat of vaporisation
of water. So instead of stating that we have got to remove moisture gained, we
state this in terms of the corresponding total amount of latent heat to be removed
for condensing that quantity of moisture gained. This is also expressed in terms
of heat units viz. BTU. So the heat to be removed per hour for condensing the
moisture is termed as room latent heat. So what we really mean by heat load is
the room sensible heat and the room latent heat that are to be removed from
within the space at a calculated rate to effect the gain of sensible and latent
heats into the conditioned space.
In heat load, there is one more source which contributes to the room sensible
and room latent heat loads. This is on account of infiltration of fresh air directly
into the conditioned space and bypass of certain amount of fresh air that is
normally taken into the system through the air handling apparatus. The form is
designed so that the room sensible heat, latent heat and the additional load due
to outside air, not forming part of room load are all calculated separately.
Here, we have used the term “Bypass”. You must understand what exactly the
meaning of the term “Bypass” is. For removing sensible heat and latent heat at
the same rate at which they are being gained within the conditioned space,
conditioned air is admitted within this space at a predetermined temperature and
humidity condition such that this air would absorb the room sensible and room
latent heat loads and in the process attain a final condition which is exactly equal
to the condition to be maintained in the room. This is achieved by continuously
drawing from within the room certain amount of air and adding to it a certain
percentage of fresh air for ventilation and cooling and dehumidifying this mixture
in a cooling coil. It is this treated air, which is supplied back into the conditioned
area. On account of some free passages in between the fins and tubes a small
percentage of the air comes out on the other side of the coil without undergoing
any change. It is this, which we terms as bypass of air. As far as the portion of
the air, which is actually re-circulated from the room is concerned, bypass will
have no influence on the ultimate result. It only means that some air has been
withdrawn from the room and just put back into the same room without any
change in its condition either upward or downward. But, what really influences is
the bypass of the fresh air, which is also passed through the cooling coil along
with the re-circulated air. Since this outside air is at a much higher temperature
and humidity conditions than the conditioned space, entry of such bypass air
would tend to upset the room conditions unless this bypass air is also brought
down to the room condition. The general formula for arriving at the exact air
quantity is:
cfm = =
However, you must realise it is not merely the selection of the condition of the
supply air that is important. We have also to consider how air can be cooled
down to the selected condition in a cooling apparatus. In a cooling coil in which
air is cooled, there is no practical means of ensuring that the air leaving the coil
would be at the exact temperature and humidity condition corresponding any
condition selected by us on the sensible heat factor line. However, there is one
temperature and humidity condition which is very easy to keep under control.
This is the condition which lies not only on the sensible heat factor line but also
on the saturation line on the psychrometric chart. In other words, if the sensible
heat factor line is extended till it meets the saturation line, then the condition
represented by the point of intersection of these two lines is the one condition
which can be under our control. This temperature is called apparatus dew point.
Bypass Factor
The problem becomes a bit more complicated because in every cooling coil there
is always a small percentage of the total cfm which escapes totally untreated.
When outside air taken into the system bypasses the coil, it will tend to raise the
room temperature and humidity conditions above the desired level. It is,
therefore, necessary to take into consideration the effect of bypass right at the
time of making the heat load calculations.
PSYCHROMETRIC FORMULAS
A. AIR MIXING EQUATIONS (Outdoor and Return Air)
tm = (1)
hm = (2)
Wm = (3)
B. COOLING LOAD EQUATIONS
ERSH = RSH + (BF) (OASH) + RSHS* (4)
ERLH = RLH + (BF) (OALH) + RLHS* (5)
ERTH = ERLH + ERSH (6)
TSH = RSH + OASH + RSHS* (7)
TLH = RLH + OALH + RLHS* (8)
GTH = TSH + TLH + GLHS* (9)
RSH = 1.08 x cfmsa x (trm - tsa) (10)
RLH = 0.68 x cfmsa x (Wrm - Wsa) (11)
RTH = 4.45 x cfmsa x (hrm - hsa) (12)
RTH = RSH + RLH (13)
OASH = 1.08 x cfmoa x (toa - trm) (14)
OALH = 0.68 x cfmoa x (Woa - Wrm) (15)
OATH = 4.45 x cfmoa x (hoa - hrm) (16)
OATH = OASH + OALH (17)
(BF) (OATH) = (BF) (OASH) + (BF) (OALH) (18)
ERSH = 1.08 x cfmda x (trm - tadp)(1-BF) (19)
ERLH = 0.68 x cfmda x (Wrm - Wadp)(1-BF) (20)
ERTH = 4.45 x cfmda x (hrm - hadp)(1-BF) (21)
TSH = 1.08 x cfmda x (tedb - tldp) ** (22)
TLH = 0.68 x cfmda x (Wea - Wta) ** (23)
GTH = 4.45 x cfmda x (hea - hta) ** (24)
C. SENSIBLE HEAT FACTOR EQUATIONS
RSHF = = (25)
ESHF = = (26)
GSHF = = (27)
D. BYPASS FACTOR EQUATIONS
BF = ;(1-BF) = (28)
BF = ;(1-BF) = (29)
BF = ;(1-BF) = (30)
E. TEMPERATURE EQUATIONS AT APPARATUS
tedb ** = (31)
tldb = tadp + BF (tedb – tadp) (32)
tewb and tlwb correspond to the calculated values of hea and hla on the
psychrometric chart.
hea ** = (33)
hta = hadp + BF (hea – hadp) (34)
F. TEMPERATURE EQUATIONS FOR SUPPLY AIR
tsa = t4m – (35)
G. AIR QUANTITY EQUATIONS
cfmda = (36)
cfmda = (37)
cfmda = (38)
cfmda = (39)
cfmda = (40)
cfmda = (41)
cfmsa = (42)
cfmsa = (43)
cfmsa = (44)
cfmba = cfmsa - cfmda (45)
Note : cfmda will be less than cfmsa only when air is physically bypassed around
the conditioning apparatus.
cfmsa = cfmoa + cfmra (46)
H. DERIVATION OF AIR CONSTANTS
1.08 = .224 X
Where .224 = Specific heat of moist air at 70 F db and 50% rh,
Btu/(deg F) (lb dry air)
60 = min/hr
13.5 = Specific volume of moist air at 70 F db and 50% rh
.68 = X
where 60 = min/hr
13.5 = Specific volume of moist air at 70 F db and 50% rh
1076 = Average heat removal required to condensate one
pound of water vapor from the room air.
7000 = Grains per pound
4.45 =
where 60 = min/hr
13.5 = Specific volume of moist air at 70 F db and 50% rh.
6.0 HEAT LOAD ESTIMATION
HEAT LOAD ESTIMATION
Introduction
The primary objective is to provide a convenient consistent, and accurate method of
calculating heating and cooling loads and to enable the designer to select systems that
meet the requirements for efficient energy utilization and are also responsive to
environmental needs.
The ability to estimate loads more accurately due to changes in the calculation
procedure provides a lessened margin of error. Therefore, it becomes increasingly
important to survey and check more carefully the load sources, each item in the load
and the effects of the system type on the load. This tightening up on the hidden safety
factors occurs for a number of reasons. There is greater emphasis, by standards and
codes, on sizing equipment closer to the expected loads, as determined by outside
design weather conditions. Also the suggested indoor design temperatures are now
usually 75 °F for cooling and 72 °F for heating. Installed lighting levels are being
reduced and the calculations are using lighting loads closer to the actual loads. All of
these factors require that the designer introduce any margin of safety by a positive
action, rather than rely on an assumed hidden margin.
Purpose of Load Calculations
Load calculations can be used to accomplish one or more of the following objectives :
7) Provide information for equipment selection and HVAC system design
8) Provide data for evaluation of the optimum possibilities for load reduction.
9) Permit analysis of partial loads as required for system design, operation and
control
These objectives can be obtained not only by making accurate load calculations but
also by understanding the basis for the loads. There a brief description of cooling and
heating loads are included.
Principles of Cooling Loads
In airconditioning design there are three distinct but related heat flow rates, each of
which varies which varies with time:
10) Heat Gain or Loss
11) Cooling load or Heating Load
12) Heat Extraction or Heat Addition Rate
Heat Gain, or perhaps more correctly, instantaneous rate of heat gain, is the rate at
which heat enters or is generated within a space at a given instant of time. There are
two ways that heat gain is classified. They are the manner in which heat enters the
space and the type of heat gain.
The manner in which a load source enters a space is indicated as follows:
13) Solar radiation through transparent surfaces such as windows
14) Heat conduction through exterior walls and roofs
15) Heat conduction through interior partitions ceilings and floors
16) Heat generated within the space by occupants, lights, appliances, equipment and
processes
17) Loads as a result of ventilation and infiltration of outdoor air
18) Other miscellaneous heat gains
The types of heat gain are sensible and latent. Proper selection of cooling and
humidifying equipment is made by determining whether the heat gain is sensible or
latent. Sensible heat gain is the direct addition of heat to an enclosure, apart from any
change in the moisture content, by any or all of the mechanisms of conduction,
convection and radiation. When moisture is added to the space, for example, by vapor
emitted by the occupants, there is an energy quantity associated with that moisture
which must be accounted for.
If a constant humidity ratio is to be maintained in the enclosure, then water vapor must
be condensed out in the cooling apparatus at a rate equal to its rate of addition in the
space. The amount of energy required to do this is essentially equal to the product of
the rate of condensation per hour and the latent heat of condensation. This product is
called the latent heat gain.
As a further example, the infiltration of outdoor air with a high dry-bulb temperature and
a high humidity ratio, and the corresponding escape of room air at a lower dry-bulb
temperature and a lower humidity ratio, would increase both the sensible heat gain and
the latent heat gain of the space.
The proper design of an airconditioning system requires the determination of the
sensible heat gain in the space, the latent heat gain in the space, and a value for the
total load, sensible plus latent, of the outdoor air used for ventilation.
The sensible cooling load is defined as the rate at which heat must be removed from the
space to maintain the room air temperature at a constant value. The summation of all
instantaneous sensible heat gains at a specific time does not necessarily equal the
sensible cooling load for the space at that time. The latent load however is essentially
an instantaneous cooling load. That part of the sensible heat gain which occur by
radiation is partially absorbed by the surfaces and contents of the space and is not felt
by the room air until sometimes later. The radiant energy must first be absorbed by the
surface that enclose the space such as walls and floor and by furniture and other
objects. As soon as these surfaces and objects become warmer than the air some heat
will be transferred to the air in the room by convection. The heat storage capacity of the
building components and item such as walls, floors and furniture governs the rate at
which their surface temperatures increase for a given radiant input. Thus, the interior
heat storage capacity governs the relationship between the radiant portion of the
sensible heat gain and how it contributes to the cooling load. The thermal storage effect
can be important in determining the cooling equipment capacity.
The actual total cooling load is generally less than the peak total instantaneous heat
gain thus requiring smaller equipment than would be indicated by the heat gain. If the
design is based on the instantaneous heat gain, the rest of the system may be
oversized as well.
Heat extraction rate is the rate at which heat is removed from the conditioned space.
Normal control systems operating in conjunction with the intermittent operation of the
cooling equipment will cause a “swing” in room temperature. There, the room air
temperature is constant only at those rare times when the heat extraction rate equals
the cooling load. Consequently, the computation of the heat extraction rate gives a
more realistic value of energy removal at the cooling equipment than does just the
instantaneous value of the cooling load provided the control system is simulated
properly. The determination of the heat extraction rate must include the characteristics
of the cooling equipment and the operating schedule of thee equipment, in addition to
the various sources of cooling load.
If the equipment is operated some what longer before and after the peak load periods,
and / or the temperature in the space is allowed to rise a few degrees at the peak
periods during the cooling operation (floating temperature), a reduction in the design
equipment capacity my be made. A smaller system operating for longer periods at
times of peak loads will produce a lower first cost to the customer with commensurate
lower demand charges and lower operating costs. Generally, equipment sized to more
nearly meet the cooling requirements result in a more efficient, better operating system
particularly when is at a partially loaded condition.
Usually a fraction of the sensible heat gain does not appear a cooling load, but instead
is shifted to the surroundings. The fraction Fc depends upon the thermal conductance
between the room air and the surroundings. It may be also considered as a adjustment
factor which results when the load components as superimposed.
The adjustment factor, Fc is calculated by the following equation.
Fc = 1 - 0.02 KT
Where KT the unit length conductance between the room air as surroundings in Btu / (hr.
ft2 F), is given by
KT = 1/LF (UWAW + UowAow + UcAc)
Where
LF = Length of the exterior walls of the room, ft.
U = U-value of room enclosure element (subscript w for window, ow
for outside wall and c for corridor), Btu (hr. ft2 F)
A = Area of the specific element
If the cooling load component has already been obtained by the technique used in this
manual, multiply that result by the calculated Fc factor.
The adjustment factor should be used only for individual small spaces or zones. It is not
to be used for block loads nor for industrial applications.
Diversity of Cooling Loads
Diverting of cooling load results from not using part of the load on a design day.
Therefore diversity factors are factors of usage and are applied to the refrigeration
capacity of large airconditioning systems. These factors vary with location, type, and
size of applicant and are based entirely on the judgment and experience of the
engineer.
Generally, diversity factors can be applied on loads from people and lights; there is
neither 100% occupancy nor total lighting at the time of such other peak loads as peak
solar and transmission loads. The reductions in cooling loads from nonuse are real and
should be accounted for.
In addition to the factors for people are lights a factor should also be applied to the
machinery load in industrial buildings. For instance, electric motors may operate at a
continuous overload, or may operate continuously at less than the rate capacity or may
operate intermittently. It is advisable to measure the power input whenever possible;
this will provide a diversity factor. It is also possible to determine a diversity factor for a
large existing building by reviewing the maximum electrical demand and monthly energy
consumption obtained from the utility bills.
Principles and Procedures for Calculating Heating Load
The peak heating requirements may occur either at night during unoccupied hours or in
the morning pickup period following a shutdown. Therefore a number of calculations
are helpful in making a proper equipment selection and system design.
Information Required (Input)
Before a cooling or heating load can be properly estimated a complete survey must be
made of the physical data. The more exact the information that can be obtained about
space characteristics, heat load sources, location of equipment and services, weather
data, etc. the more accurate will be the load estimate.
Required Input - External Loads - Cooling
For calculation of the outdoor loads the input information should include:
19) Orientation and dimensions of building components.
20) Construction materials for roof, walls, ceiling, interior partitions, floors and
fenestration
21) Size and use of space to be conditioned
22) Surrounding conditions outdoors and in adjoining spaces
Required Input - Heating Load
The input for calculation of heating load is essentially the same as that for the cooling
load. However, it may not be necessary to calculate the internal sources and solar heat
gain.
In heat load estimation we compute
a. Room sensible and latent heat gains due to transmission, sunlight, occupancy
and other internal sources of heat.
b. Grand total heat comprising total room load under (1) plus additional loads due to
outside air intake, heat gains in return air ducts, in chilled water distribution
systems, pumping horse power load, etc.
Room load estimation under (a) is required for computing the condition and quantity of
supply air while the grand total heat under (b) is required for terminating the total
capacity of the cooling system.
In this discussion, we will confine ourselves to transmission gains and related subjects
only. There are certain similarities between heat transmission through barriers and
electric current transmission through conductors. We will use this similarly wherever
required for better understanding of the subject. The well known formula relating to
transmission of electric current is:
I =
Where “I” = current in Amperes, “V” = Voltage & “R”, the Resistance of the conductor.
In this formula, if (I) / (R) is considered as the conductance of the conductor, say “C”,
then the formula can be rewritten as:
I = V x C
For transmission of heat through a barrier, the motive force corresponding to “V” is
temperature difference between the two sides of the barrier. The formula for rate of
heat transmission per hour H is:
H = A x U x (T)
Where T is the temperature difference in °F and A is the area of the barrier in sq.ft. and
U is the overall heat transmission coefficient expressed in BTU/Hr/Sq.ft/°F temperature
difference. The product (AxU) corresponds to the conductance “C” of the electric
conductor.
Thermal conductivity of any material is the heat transmitted through the material
expressed as BTU/Hr/Sq.ft/Inch thickness/°F temperature difference and is referred to
by the symbol “K”. It K is the conductivity of the material, then 1/K is the resistance of
the material of 1 sq.ft. cross section and 1” thickness. If the thickness “t” inches, the
resistance becomes (t) / (K) per sq.ft.
In electrical system, resistance connected in series are added to find the total
resistance.
Similarly, if a barrier is made up of several materials, the individual resistances of the
components have to be added to arrive at the total barrier resistance. If a barrier is
made up of, say, three materials having thermal conductivities K1, K2 & K3, the total
thermal resistance of the barrier is:
t1/K1 + t2/K2 + t3/K3
Where t1, t2 & t3 are the thicknesses of the barriers.
Film Coefficient
In addition to the resistance of the various components of a barrier, we have to consider
one more resistance offered by a film of air (or fluid if the barrier is in a fluid) which
clings on to the barrier surfaces. This resistance is more when the air is still and is
relatively less when there is wind velocity. Like thermal conductivity, the heat
transmission capacity of a film is expressed as the rate of heat of transfer in
BTU/Hr/Sq.ft/°F temperature difference (Note that this differs from thermal conductivity
in the sense it is not related to any film thickness as in the case of materials). This is
called the film coefficient and is expressed by the symbol “f”. The reciprocal of “f” is the
thermal resistance of the film. “f1” denotes the film coefficient on the interior surface of
the barrier and “f0” denotes the film coefficient on the exterior surface of the barrier.
The resistance of the complete barrier is:
1/f1 + t1/K1 + t2/K2 + t3/K3 + 1/f0
If “U” is the overall heat transmission of the barrier in BTU/Hr/Sq.ft./°F, then 1/0 is the
overall thermal resistance of the barrier.
1/U = 1/f1 + t1/K1 + t2/K2 + t3/K3 + 1/f0
U = 1/(1/f1 + t1/K1 + t2/K2 + t3/K3 + 1/f0)
Storage Effect
Suppose T0 is the temperatures on both sides of the barrier. There will be no heat
transmission through the barrier and the temperatures at all points within the barrier will
also be the same. There is, therefore, no temperature gradient. Now suppose the
temperature on one side of the barrier is raised from T0 to T7, do you think that heat
transmission through the barrier will commence immediately? No, since all points within
the barrier is at the same temperature, no heat can flow through any interior section.
The first thing that happens is the outermost layer of the barrier absorbs the heat from
the outside and rises in temperature. Heat then flows over to the next layer of the
barrier because of the temperature difference between the first and second layers. The
second layer also will first rise in temperature before heat begins to flow over to the third
layer. Thus progressively all the layers within the barrier rise in temperature thereby
establishing the total temperature gradient from one side to the other side of the barrier.
It is only after the complete gradient has been established that heat will begin to flow to
the other side of the barrier. The temperature at various points within the barrier will
now be as determined by the gradient.
Attic Spaces
Whenever a false ceiling is provided in a room, the space enclosed between the false
ceiling and the concrete ceiling is called ATTIC SPACE. If the attic space is not
ventilated the entire space within the attic will assume an intermediate temperature
which will be more than the room temperature and less than the outside temperature.
This temperature can be worked out as follows:
Ag = area of the concrete ceiling
A1 = area of the false ceiling
U0 = “U” factor of the concrete ceiling
Uf = “U” factor of the false ceiling
T2 = Outside temperature
T1 = Inside temperature
T = Temperature of the attic space
When steady heat transmission from outside to inside takes place through the attic
space, then the rate of flow of heat from outside into the attic space is equal to the rate
of flow of heat from the attic space into the room, i.e.,
Ac.Uc.(T2 – T) = Af.Uf.(T – T1)
“T” can therefore be calculated from this equation. After “T” has been worked out, the
transmission load into the room from the ceiling can be worked out by substituting the
value of “T” in the above equation.
Solar Gain
Solar Gain, as the name implies, comes from direct sunlight. There are two kinds of
solar gains:
a. Radiation from sun which directly enters the conditioned space through glass and
absorbed by objects in the room and then by the air within the room. The effect
of such gain into the space is felt almost immediate.
The amount of radiation for various exposures and time of the day and year are
given in tables for the various latitudes on the earth. Depending on the type of
glass, about 5 to 6% of the radiation is reflected while the rest pass into the
room. Solar gain is not confined merely to the side which directly faces the sun.
You get solar heat even from other sides through glasses, but to a much smaller
degree. This is diffused radiation.
b. Solar & Transmission Gain
This is due to transmission through sunlit walls whose temperature rises above
the ambient temperature due to absorption of direct radiation and hence causes
a larger temperature differential than the ambient temperature. The equivalent
temperature difference that is to be taken are given in tables, taking into
consideration the exposure, sun time and storage effect.
c. Transmission Gain through Glass & Partition
In addition to solar gain through glass, you have also to work out transmission
gain through glass due to temperature difference. Transmission through
partitions between conditioned and non-conditioned areas are worked out on the
basis of actual temperature difference. No storage effect apply for these cases.
d. Internal Load
This comprises load from:
23) Occupancy: The sensible/latent heat gains from people are given in
tables, based on the nature of their activities in the room.
24) Lights: Lighting is generally specified in terms of watts per sq.ft. The total
watt has to be converted into BTU/Hr by multiplying by conversion factors.
25) Appliances: Electrical, gas burners, steam generation, etc.
26) Electric Motors: Applies generally in some of industrial applications. This
load will have to be properly analysed by discussion with user and
appropriate diversity factors should be applied for estimating the actual
load. Convert the HP into BTU/Hr.
We shall now briefly lay down the procedure for heat load estimating with explanations
wherever required.
1. Collect architect’s drawings for the building giving all details and dimensions of
walls, floors, windows, etc. If such drawings are not available, survey the place
and get the particulars.
2. For every application, there are certain things which the ultimate user has to
specify. These are:
27) Temperature & humidity conditions to be maintained inside the space and
tolerance.
28) Occupancy – i.e. maximum no. of people likely to occupy the space and the
nature of their activity.
29) Lighting load and other internal source of heat generation.
30) Period of operation – e.g. 8 a.m. to 7 p.m. or 10 a.m. to 8 p.m. etc.
31) For industrial application you require also the HP load in the conditioned
space and diversity factor thereon.
32) Minimum ventilation required.
3. Outside Design Conditions
33) For comfort air conditioning application, use the mean maximum DB
temperature & the WB temperature which occurs simultaneously with the
assumed DB.
34) For industrial applications where temperatures and humidities are to be
maintained within very close tolerance through the year, tank the maximum
DB and the simultaneously occurring WB temperature.
4. For all applications make a second load estimate for monsoon conditions.
5. For applications where the conditioned spaces are spread over very vast floor
areas, divide the entire area into convenient zones and make load estimates.
6. Occupancy - In certain applications a diversity factor may have to be used even
in respect of occupancy. Examples are: Office areas where a separate
conference room is also provided. The conference room may be designed for a
large number of people. But you must realize that it is mostly the people in the
office who go into conferences and hence any occupancy in the conference room
brings about an equal reduction in the occupancy in other areas of the office.
7.0 HEAT LOAD DATA SHEET &
TYPICAL CALCULATION
TYPICAL DIVERSITY FACTORS FOR LARGE BUILDINGS
(APPLY TO REFRIGERATION CAPACITY)
DIVERSITY FACTOR
PEOPLE LIGHTS
Office 0.75 to 0.90 0.70 to 0.85
Apartment, Hotel 0.40 to 0.60 0.30 to 0.50
Department storage 0.80 to 0.90 0.90 to 1.0
Industrial 0.85 to 0.95 0.80 to 0.90
Fresh air requirement - 2.0 air changes / hr.
or
10 CFM per person
Design conditions
a - indoor - 70 °F ± 2 °F DBT ; 55% ± 5% RH
b - Outdoor - 103 °F DBT ; 82 ° F WBT
U - FACTOR CALCULATIONS
a. Exposed Walls
Total resistance RT = Ro + X1 R1 + X2 R2 + X3 R3 + Ri
= 0.25 + 12.5 x 0.2 + 230 x 0.2 + 12.5 x 0.2 + 0.68 25 25 25
= 2.97 hr. ft2. °F /BTU
Overall heat transfer Co-efficient
= 1_ = 1_RT 2.97
= 0.337 BTU / hr. ft2. °F
b. Partitions
RT = Ri + X1 R1 + X2 R2 + X3 R3 + Ri
= 0.68 + 12.5 x 0.2 + 230 x 0.2 + 12.5 x 0.2 + 0.68 25 25 25
= 3.40 hr. ft2. °F /BTU
Overall heat transfer Co-efficient
= 1_ = 1_RT 3.40
= 0.294 BTU / hr. ft2. °F
c. Roof exposed to sun
RT = Ri + X1 R1 + X2 R2 + Ri
= 0.25 + 150 x 0.2 + 50 x 4.0 + 0.92 25 25
= 10.37 hr. ft2. °F /BTU
Overall heat transfer Co-efficient
= 1_ = 1_RT 10.37
= 0.096 BTU / hr. ft2. °F
8.0 HEATING, VENTILATION &
AIRCONDITIONING SYSTEMS
HEATING, VENTILATION & AIRCONDITIONING SYSTEMS
A. AIR CONDITIONING SYSTEMS
Airconditioning is defined as the simultaneous control of temperature, humidity,
quality and movement of air in a conditioned space or building.
An air conditioning system is therefore, defined as an arrangement of equipment
which will air condition a space or a building. Thus, a complete air conditioning
system includes a means of refrigeration, one or more heat transfer units, air
filters, a means of air transport and distribution, an arrangement for piping the
refrigerant and heating medium, and controls to regulate the proper capacity and
operation of these components.
The items outlined above are considered to be the components of a complete air
conditioning system.
There has been a tendency by many designers to classify an air conditioning
system by referring to one of its components. For example, the airconditioning
system in a building may include a dual duct air transport arrangement to
distribute the conditioned air and is then referred to as a dual duct system. This
classification makes no reference to the type of refrigeration, the piping
arrangement or the type of controls.
For the purpose of classification, the following definitions will be used:
An Airconditining unit is understood to consist of heat transfer surface for
heating and cooling, a fan for air circulation, means of cleaning the air, a motor, a
drive, and a casing.
A self-contained airconditioning unit is understood to be an airconditioning
unit that is complete with compressor, condenser, controls, and a casing.
An air handling unit consist of a fan heat transfer surface, a motor, a drive and
a casing
A remote air handling unit or a remote air conditioning unit is a unit located
outside of the conditioned space which it serves.
The most common types of refrigeration machines, classified according to their type of operation are (1) mechanical compression, (2) absorption and (3) vacuum.
Apart from the above types the airconditioning system are generally clarified is
to following categories:
1. Window (room) airconditioners
2. Split airconditioning units
3. Packaged airconditioning units
4. Centralised airconditioning plant - DX system
5. Centralised airconditioning plant – chilled water system
The details of the above are further detailed in the subsequent pages.
The types of refrigeration machines which are further explained as under:
Mechanical Compression machines may be divided into reciprocating,
centrifugal, and rotary types.
The term “heat pump” is occasionally used to describe a refrigeration machine.
However, a heat pump is a refrigeration cycle – either reciprocating, rotary or
centrifugal - in which the cooling effect as well as the heat rejected is used to
furnish cooling or heating to the air conditioning units, either simultaneously or
separately.
Reciprocating or rotary compressor can be used in systems that circulate the
refrigerant through remote direct expansion heat transfer surfaces. Alternately
they can be used in conjunction with a water chilling heat exchanger, to produce
chilled water for circulation through remote heat transfer surfaces that cool and
dehumidify the air.
Centrifugal refrigeration machines are generally not suitable for circulating and
expanding the liquid refrigerant in remote heat exchanges surfaces. Centrifugal
machines are therefore used only to chill water or brine for circulation through
remote heat exchange surfaces.
Absorption machine cycles are similarly to mechanical compression machine
cycles only to the extent that both cycles evaporate and condense a refrigerant
liquid. They differ in the mechanical compression cycle use purely mechanical
processes, while the absorption cycle uses physiochemical processes to produce
the refrigeration effect.
Vacuum refrigeration machines, such as steam jet and water vapor units, are
seldom used in modern airconditioning systems.
1. Room Airconditioner
This is the simplest form of an air conditioning system. It has a hermetically
sealed motor compressor assembly, an air cooled condenser coil, an evaporator
coil, condenser fan and evaporator fan. It has a capillary tube in place of an
expansion valve for metering refrigerant flow to the evaporator. Room air
conditioners are generally made of capacities ranging from 3/4 ton to 1-1/2 tons
suitable for operation on 230 V, single phase, 50 cycles supply. It is completely
factory assembled and can be straightaway plugged into power supply when
installed.
Application
Generally used for small office rooms, shops and residential rooms where the
where the load will generally be within 1-1/2 tons. Sometimes, these units are
used in multiple for larger areas.
Advantages
The main advantage is that the unit can be switched ON and OFF as required.
Where multiple units are used, there is no fear of a total breakdown of air
conditioning since it is most unlikely that all the units will breakdown
simultaneously.
Disadvantages
The hermetically sealed compressor is susceptible to burn out when the supply
voltage fluctuates widely and whenever such burnouts occur, the whole system
has to be thoroughly cleaned before a new compressor can be fitted. The life of
the unit is generally between 10 to 15 years only.
2. Packages type air conditioner
These are larger versions of Room Air Conditioners except that they are
generally made with water-cooled condensers. They can also be made with air
cooled condensers either built in with the package or for remote installation.
They are generally made in capacities ranging from 5 to 10 tons. Units with
water cooled condensers require condenser water circulating system and cooling
tower. The units may also require external duct work for air distribution. This unit
operates on 400 V, phase, 50 cycles supply.
Application
These units are most ideal where the load is between 5 to 20 tons. Sometimes,
they are also used for much larger loads by using more number of units
interconnected on the supply air side.
Advantages
Installation and commissioning can be done in the shortest possible time since
the field work involved only relates to condenser water piping, air distribution
system and electrical wiring. When multiple units are used for larger areas, the
number of units in operation can be varied according to the load requirement,
thereby saving on power consumption.
Disadvantages
These units also have hermetically sealed motor compressor assemblies and
hence have the same disadvantage as Room Air Conditioners.
3. General plant - DX systems
The system consists of an open type compressor ranging in capacity from 5 tons
to 120 tons operating on refrigerant 22. They are motor driven either through belt
drive or direct coupling. They can also be driven by diesel engine, but then only
by direct drive. Belt drive should not be used when diesel engine is used. They
are generally with water-cooled condensers even though they can also be built
with air-cooled condensers.
Application
The DX system of Central Plant is perhaps the most widely used system for
medium loads between 20 and 100 tons. It can be used for almost all types of
application.
Advantages
The DX system is perhaps the most efficient of all system from a thermo dynamic
point of view since the heat transfer is directly between the conditioned air and
the refrigerant. The open type compressors used for these systems have built in
capacity controls to take care of load fluctuations. Plants of any capacity can be
built with DX systems using multiple compressors, condensers and evaporators.
Although it is preferable to keep each compressor with its condenser and
evaporator as a single unit, these plants can also be built with interconnection
between them on the refrigerant side. Such interconnections naturally provide
more flexibility in operation.
Disadvantages
DX systems should not be used where air distribution through duct work has to
be carried out from a central air handling unit to various zones because of fire
hazard. Therefore, a single air handling unit should necessarily be confined to a
single zone. Where there are multiple zones use of DX system is permissible
only when separate DX plants are used for each zone without any
interconnection on the air distribution size. Where the building condition has got
number of floors one above the other, DX systems could be considered only if it
is possible to install separate Central Plant for each floor. Of course, such
decision would involve installation of the plant on upper floors where vibration
and other problems have to be effectively tackled in order to eliminate
transmission of vibrations to the occupied zones. Cost wise also, such individual
systems in each floor may prove to be much higher.
4. Central plant - Chilled Water System
A central chilled water system is made up on one or more water chilling plants.
Each water chilling plant may be built with either one or two compressors to work
with one or two chillers (DX chiller or shell and tube flooded chiller) and one or
two water cooled condensers. Each such water chilling unit is field assembled
on structural framework with the necessary refrigerant pipes so as to make a
compact assembly. Where such multiple water chilling units are used, they are
generally interconnected on the water side both in the condenser circulating
system and chilled water circulating system.
Application
Multiple water chilling units with reciprocating compressors are generally suitable
for multistoreyed office buildings where the load is between 100 and 300 tons.
However, there is no bar against using more number of water chilling units with
reciprocating compressors even for loads higher than 300 tons. For loads
exceeding 300 tons. Water chilling units with centrifugal compressors would be
preferable.
5. Chilled Water for Process Cooling
Advantages
The best advantage of a chilling water system in that the Central Plant can be
installed in as remote a location as desired from the conditioned areas. In fact,
they can even be built in a remote plant room with chilled water piping either
underground or overhead running to all the zones where air handling units are
installed. This system provides maximum flexibility in operation since the air
handling units serving individual zones can be cut off from the chilled water
circulating system whenever air conditioning is not required in any particular
zone. Since each zone will have its own air handling unit, no interconnecting
duct work will be required thereby eliminating all possibilities of fire spreading
from one zone to another. In the case of large hotels, fan coil units in
individual
rooms can be switched off whenever the room is not occupied. Individual
temperature controls can be provided for each zone or individual rooms by
regulating the chilled water flow through the coil either thermostatically or
manually.
Disadvantages
On application where the load is small, this system would prove very much
costlier than the DX system. Another disadvantage is that since one more heat
transfer medium viz. chilled water, has been introduced, the heat transfer is now
from air to water and then from water to refrigerant. This naturally lowers the
evaporating temperature as compared to a DX system for the same load. Hence
the power consumption will be relatively higher than that for DX system.
6. Air conditioning System for Operation Theatres
It is desirable that a DX system is used for each Operation Theatre. However, in
large hospitals, if there are several operation theatres located in various floors,
there is no bar against using a central chilled water system, but exclusively for
the operation theatre only. Each operation theatre must however have individual
air treatment units with pre-filters on the air suction side of fan and
supplementary microves filters on the air discharge side of the fan. For operation
theatres no re-circulation of room air is permitted. You should, therefore,
estimate the heat load on the basis of 100% fresh air.
7. Screw Chillers
Refer to figure showing single line diagram of refrigeration cycle for the above
and for piping schematics. Each vertical screw compressor discharges hot, high
pressure gas through a discharge service valve (A) (or check valve in multiple
compressor units) into the condenser, where it condenses outside tubes,
rejecting heat to cooling tower water flowing inside the tubes. The liquid
refrigerant drains to the bottom of the condenser and exits into the economizer
feed line.
The refrigerant flows through the economizer feed ball valve (B), dropping its
pressure, causing it to flash. It then flows into the flash economizer tank (C)
which is at an intermediate pressure between condenser and evaporator, liquid is
centrifugally separated from the flash gas and the liquid drains to the bottom of
the tank, exits via the economizer drain line, and passes through the economizer
drain ball valve (D). Both economizer ball valves are actuated by a modutrol
motor (U) that adjusts flow to maintain an appropriate refrigerant level in the
evaporator, determined by a liquid level float switch (V).
From the drain line, liquid refrigerant flows into the flooded evaporator, where it
boils, cooling the water flowing inside evaporator tubes. Vapor from the boiling
refrigerant flows up the suction pipes through a shut-off valve (E) (optional),
suction check valve (F) and suction filter (G) (inside compressor) into the
compressor where it is compressed and starts cycle again.
Vapor flows from the top of flash economizer into the compressor at the vapor
injection port, which feeds it into the compressor part way through the
compression process. Check valve (H) prevents backflow at shutdown in multi
compressor units. Al compressors operate in parallel on a common evaporator
and condenser.
8. Ice Storage Systems for Airconditioning Applications
Use of Ice for Airconditioning
Building air conditioning in summer daylight hours is one of the largest
contributors to electrical utility demand peaks. Typically between 2-4 PM in the
afternoon when solar loading peaks, more air conditioners are needed to
maintain comfortable environments in buildings. Add to this the electricity utilized
by lighting, computers, building subsystems plus other equipment and the utility
is faced with a peak load condition dictating that it bring on-line additional, more
costly peak power generating sources to handle the load.
Traditional air conditioning systems operate during the day to meet cooling
demand and remain idle at night. Chillers are selected to satisfy the maximum
demand, which occurs only a few hours per year, and thus spend the majority of
their operational life at reduced capacity and low efficiency.
The ice storage system, which is suitable for any A/C application, allows installed
chiller capacity (and size of other components) to be significantly reduced –
typically between 40% and 60%. This enables efficient and real energy
management whilst taking advantage of low tariff electricity.
Large commercial users whose air conditioning loads contribute to the utility
peaking problem are assessed an added charge typically based on their highest
15 minutes window of peak demand for electricity. This is called a “demand
charge” which in many areas of the country can account for as much as 40
percent of the building owner’s total electrical bill.
The use of ice storage to minimize peak energy usage is not a new or
experimental idea. It has been used for years on applications with short peak
energy usage such as churches, meeting facilities and theaters. On these
applications, however, the longer peak uses were handled by conventional
rooftop cooling or water chilling / air handling systems.
Now, however, there is renewed interest in a broad use of ice making and
storage systems by both users and utility companies as the best way of offsetting
rising demand loads and resulting utility cost increases.
Ice storage systems can not only cut operating costs substantially, but they can
also reduce capital outlays when systems are properly applied for both new and
existing buildings and commercial and industrial types. Simply stated, engineers
can specify smaller chillers operating 24 hours a day rather than larger chillers
operating 10-12 hours a day and cut the capital outlay for air conditioning
equipment substantially.
An ice storage system can utilize either a load shifting or a load leveling strategy
to significantly lower demand charges during the cooling season. Because this
lowers energy demand, it substantially lowers the total energy costs. It typically
utilizes a standard packaged chiller to produce ice at night or during off-peak
periods when the building’s electrical needs are at a minimum. The ice is stored
in modular tanks to provide cooling ton-hours to help meet the buildings cooling
load requirements the following day. By doing so, it minimizes the peak energy
usage during the utility daylight peaking period.
Full Storage Or Partial Storage?
Two load management strategies are possible with ice storage systems. When
utility rates call for complete load shifting, a conventionally sized chiller can be
used to shift the entire load into off-peak hours. This is called a full storage
system and is used most often in existing building renovation or retrofit
applications using existing installed chiller capacity.
In new construction, a partial, storage system is usually the most practical and
cost effective load management strategy. In this load leveling method, the chiller
is sized to run continuously except for scheduled preventive maintenance down
time. It usually charges the ice storage tanks at night and cools the load directly
during the daytime peak hours with help from stored cooling capacity.
This will greatly reduce the installed chiller capacity and its required capital
expenditure, as well as the demand charge for electricity to run the chiller during
utility peaking periods. Typically reductions can be 50 percent or more.
How the Ice Storage System Works
A common ice storage system is a modular, insulated tank. Tanks are typically
available in several ton-hour rated sizes. Typically at night a mild concentration
of glycol-water solution (typically 25 percent ethylene glycol based industrial
coolant such as Dow Chemical Company Dowtherm SR-1 or Union Carbine
Corporation’s UCAR Thermofluid 17) from a standard packaged air conditioning
water chiller system circulates through the heat exchanger and extracts heat until
eventually all the water in the tank is frozen solid. The ice is built uniformly
throughout the tank.
Typical schematic flow diagrams for a partial storage system are shown in
figure1&2. At night, the water-glycol solution circulates through the chiller and the
ice bank heat exchanger, bypassing the air handling coil that supplies
conditioned air to occupied building spaces. During the day, the solution is
cooled by the ice bank from 52 F to 34 F. A temperature modulation valve set at
44 F in a bypass loop around the ice bank permits a sufficient quantity of 52 F
fluid to bypass the ice bank permits a sufficient quantity of 52 F fluid to bypass
the ice bank, mix with the 34F fluid, and achieve the desired 44 F temperature.
The 44 F fluid then enters the coil, where it cools air from approximately 75 F to
55 F. The fluid then leaves the coil at an elevated temperature (approximately
60F) and enters the water chiller where it is cooled 60F to 52 F.
It is important to note that while making ice at night, the chiller must cool the
water- glycol solution down to 26 F, rather than producing 44 F water required for
conventional air conditioning systems.
Chillers with air-cooled condensing also benefit from cooler outdoor ambient dry
bulb temperatures to lower the system condensing temperature at night.
The temperature modulating valve in the bypass loop has the added advantage
of providing excellent capacity control. During mild temperature days, typically in
the spring and fall, the chiller will often be capable of providing all the necessary
cooling capacity for the building without the use of cooling capacity from the ice
storage system. When the building’s actual cooling load is equal to or less than
the chiller capacity at the time, all of the system coolant will flow through the
bypass loop as shown in fig.3
It is important that the coolant chosen by an ethylene glycol-based industrial
coolant, such as Dowtherm SR-1 or UCAR Thermofluid 17, which is specially
formulated for low viscosity and good heat transfer properties. Either of these
fluids contain a multi-component corrosion inhibitor which is effective with most
materials of construction including aluminium, copper, silver solder and plastics.
Further, they contain no anti-leak agents and produce no films to interfere with
heat transfer efficiency. They also permit use of standard pumps, seals and air
handling coils. It should be noted, however, that because of the slight difference
in heat transfer properties between water and the mild glycol solution, the cooling
coil capacities will need to be increased by approximately 5 percent. It is also
important that the water and glycol solution be thoroughly mixed before the
solution is placed into the system.
The use of ice storage system technology opens new doors to other economic
opportunities in system design. These offer significant potential for not only first-
cost savings but also operating cost savings that should be evaluated on a life
cycle cost basis using a computerized economic analysis program.
Back-Up
Most A/C and refrigeration systems require some form of stand-by, or back-up,
facility to protect against system failure and costly lost production time. The ice
storage system, is an ideal, efficient solution for these applications. The ice
storage system offers rapid response back-up in the form of an independent,
static technology solution which ensures the highest degree of reliability.
Advantages of Ice Storage Systems
· Reduced installed plant capacity.
· Reduced electrical installation for lower investment and saving in demand
charges.
· Reduced installed cooling tower capacity incase of water cooled system.
· Reduced installed D.G. set capacity.
· Better plant utilization with longer equipment life and lower operating costs.
· Use of off-peak energy for lower energy bill, where differential tariff is applicable.
Applications of Ice Storage System
· Air-conditioning of industrial and commercial buildings - Offices, Hotels,
Shopping Complexes, Supermarkets, etc.
· Air-conditioning of data-processing centers, hospitals, telephone exchanges, etc.
requiring added system reliability and security.
· Dairy plants, Breweries, Food Processing, Bottling Plants, Chemical and
Fertilizer Plants, Pharmaceuticals, etc.
DUCTING DESIGN
The satisfactory distribution of conditioned air requires a well designed and energy
efficient air transport system with appropriate ducts and fans plus air treatment and
control devices.
The various method of duct designs, proper fan selection and control and methods of air
distribution system control for acceptable comfort and air quality in the conditioned
spaces are some of the points to be discussed.
The various methods of duct designing are
a. Constant Velocity method
b. Equal friction method
c. Static regain method
Classification of Ducts
Supply and return duct systems are classified with respect to the velocity and pressure
of the air within the duct.
Velocity
There are tow types of air transmission systems used for airconditioning application.
They are called Coventional or Low Velocity and High Velocity system. The dividing line
between these systems is rather nebulous but, for the purpose of this section, the
following initial supply air velocities are offered as a guide.
1. Commercial comfort air conditioning
a. Low velocity – upto 2500 fpm normally between 1200 & 220 fpm
b. High velocity – above 2500 fpm
2. Factory comfort airconditioning
a. Low velocity - upto 2500 normally between 2200 and 2500 fpm.
b. High velocity - above 2500 to 5000 fpm
Normally return air systems for both low and high velocity supply air systems are
designed as low velocity systems. The velocity range for commercial and factory
comfort application is as follows:
1. Commercial comfort airconditioning - low velocity upto 2000 fpm. Normally
between 1500 and 1800 fpm.
2. Factory comfort airconditioning - low velocity upto 2500 fpm. Normally between
1800 and 2200 fpm.
Pressure
Air distribution systems are divided into three pressure categories; low, medium and
high. These divisions have the same pressure ranges as Class I, II & III fans and
indicated:
1. Low pressure - upto 3¾ inch wg - class I fan
2. Medium pressure - from 3¾ to 6 ¾ inch wg - class II fan
3. High pressure - from 6 ¾ to 12 ¾ inch wg - class III
These pressure ranges are total pressure, including the losses through the air handling
apparatus, ductwork and the air terminal in the space.
The choice of design method depends almost entirely upon the size of the ductwork
installation. Small duct systems (homes, shops or a few office rooms) are commonly
designed by the velocity method. Large high pressure systems are most frequently
designed by computer software programs using the static regain method. Duct
arrangements between these two extremes are nearly always laid out by the equal
friction method. Sometimes a duct arrangement will be designed by a combination of
two methods. For instance, the trunk duct will be laid out by the static regain method
and the branch duct runs designed by the equal friction method.
In designing ductwork, a new term called “unit friction” will be utilized which means the
friction loss per 100 ft of duct work equivalent length.
Regardless of the duct design method chosen by the air transport system designer, the
final design and duct layout will likely result from the use of computerized duct design
and drafting programs available that are based on algorithms from the ASHRAE hand
book of fundamentals and other test data from SMACNA.
Recommended and maximum Duct Velocities for Conventional System
DESIGNATION RECOMMEND VELOCITIES, FPMRESIDENCES SCHOOLS, THEATERS,
PUBLIC BUILDINGSINDUSTRIAL
BUILDINGOutdoor air intakes1 500 500 500Filters1 250 300 350Heating coils1,2 450 500 600Cooling coils1 450 500 600Air washers1 500 500 500Fan outlets 1000 - 1600 1300 - 2000 1600 - 2400Main ducts2 700 - 900 1000 - 1300 1200 - 1800Branch ducts2 600 600 - 900 800 - 1000Branch risers2 500 600 - 700 800
Maximum velocities, FPMOutdoor air intakes1 800 900 1200Filters1 300 350 350Heating coils1,2 500 600 700Cooling coils1 450 500 600Air washers 500 500 500Fan outlets 1700 1500 - 2200 1700 - 2800Main ducts2 800 - 1200 1100 - 1600 1300 - 2200Branch ducts2 700 - 1000 800 - 1300 1000 - 1800Branch risers2 650 - 800 800 - 1200 1000 - 1600
1 These velocities are for total face area, not the net free area : other velocities in table
are for net free are2 For low velocity systems only.
@ 1965 American society of heating, refrigerating and airconditioning engineers, inc.
reprinted by permission for ASHRAE guide and data book.
PIPING DESIGN
The water piping system are divided into once thru and re-circulating types. In a once
thru system water passes thru the equipment only once and is discharged. In a re-
circulating system water is not discharged, but flows in a repeating circuit from the heat
exchanger to the refrigeration equipment and back to the heat exchanger.
Open and Closed
Both types are further classified as open or closed systems. An open system is one in
which the water flows into a reservoir open to the atmosphere; cooling towers and air
washers are examples of reservoirs open to the atmosphere. A closed system is one in
which the flow of water is not exposed to the atmosphere at any point. This system
usually contains an expansion tank that is open to the atmosphere but the water area
exposed is insignificant.
Water Piping Design
There is a friction loss in any pipe thru which water is flowing. This loss depends on the
following factors:
1. Water velocity
2. Pipe diameter
3. Interior surface roughness
4. Pipe length
System pressure has not effect on the head loss of the equipment in the system.
However, higher than normal system pressures may dictate the use of heavier pipe,
fittings and valves along with specially designed equipment.
To properly design a water piping system, the engineer must evaluate not only the pipe
friction loss by the loss thru valves, fittings and other equipment. In addition to these
friction losses, the use of diversity in reducing the water quantity and pipe size is to be
considered in designing the water piping system.
Pipe Friction Loss
The pipe friction loss in a system depends on water velocity, pipe diameter, interior
surface roughness and pipe length. Varying any one of these factors influences the
total friction loss in the pipe.
Most air conditioning applications use either steel pipe or copper tubing in the piping
system.
Charts enclosed are for schedule 40 pipe upto 24 inch diameter. Chart shows the
friction losses for closed re-circulation piping systems and for once thru / open re-
circulation piping systems.
These charts show water velocity, pipe or tube diameter, and water quantity, in addition
to the friction rate per 100 ft of equivalent pipe length. Knowing any two of these
factors, the other two can be easily determined from the chart. The effect of inside
roughness of the pipe or tube is considered in all these values.
The water quantity is determined from the airconditioning load and the water velocity by
pre-determined recommendations. These two factors are used to establish pipe size
and friction rate.
Water Velocity
The velocities recommended for water piping depend on two conditions;
1. The service for which the pipe is to be used.
2. The effect of erosion.
The design of the water piping system is limited by the maximum permissible flow
velocity.
Recommend Water Velocity
SERVICE VELOCITY RANGE (FPS)
Pump discharge 8 - 12
Pump suction 4 - 7
Drain line 4 - 7
Header 4 - 15
Riser 3 - 10
General service 3 - 10
City water 3 - 7
B. VENTILATION SYSTEM
Outdoor air that flows through a building either intentionally as ventilation air or
unintentionally as infiltration (and exfiltration) is important for two reasons.
Dilution with outdoor air is a primary means of controlling indoor air contaminants
and the energy associated with heating or cooling this outdoor air is a significant,
if not a major, load on the heating and air-conditioning system. For maximum
load conditions, it is essential to know the magnitude of this air flow to properly
size equipment; for average conditions, to properly estimate average or seasonal
energy consumption; and for minimum conditions, to assure proper control of
indoor contaminants. In larger buildings, it is important to know ventilation
effectiveness. Knowledge of smoke circulation patterns can be crucial in the
event of fire.
Ventilation occurs by two means, natural and forced, Natural ventilation can be
classified as (1) infiltration or (2) controlled. Manually controlled natural
ventilation is the ventilation from operable windows, doors or other openings in
the buildings envelope. The latter is an important means of ventilation in
residences in mild weather when infiltration is minimal or in warm climates to
avoid air conditioning costs.
Forced ventilation is mandatory in larger buildings where a minimum amount of
outdoor air is required for occupant comfort. Air contaminant measurement
technology has advanced to include alternate methods designed to assure that
indoor air quality meets specified conditions. These methods permit the amount
of outdoor air to vary according to the actual requirements of occupants in the
space.
This chapter focuses on envelope or shell-dominated buildings; i.e., residences
or small commercial buildings in which the energy load is determined by the
construction and performance of the building envelope. The physical principles
discussed also apply to large buildings. However, in large buildings, ventilation
energy load and indoor air quality conditions depend more on ventilation system
design that on building envelope performance.
Ventilation Requirements
The amount of ventilation needed has been debated for over a century, and the
different rationales developed have produced radically different ventilation
standards. Considerations such as the amount of air expel exhaled air, moisture
removal from indoor air and control of carbon dioxide (CO2) were each primary
criteria used at different times during the nineteenth century.
This research investigated the ventilation rates required to keep body-generated
odors below an acceptable level in rooms with comfortable levels of temperature
and humidity. It was found that the required ventilation rates varied considerably,
depending on the cleanliness of the subjects and the number present in the
room. Researches also found that CO2 concentration was not a good indicator of
the ventilation rate above 17 m3h per person; the CO2 concentration was almost
always lower than expected for a given ventilation rate. However, below 17
m3/hr per person the discrepancies were not so great and in fact the current
rationale for the 8.5 m3/h per person minimum outside air requirement in
Standard 62 is based on the CO2 produced by an individual depends on diet and
activity level. A representative value of CO2 production by a sedentary individual
who eats a normal diet is 0.019 m3/hr.
Types of Ventilation
Several techniques are possible to achieve the ventilation specified in the
standards; in decreasing order of desirability they are : (1) forced ventilation that
affords automatic control, (2) natural ventilation with manual control.
Forced ventilation is rarely used in envelope-dominated structures. However,
tighter, more energy conserving buildings with less infiltration require mechanical
ventilation systems. When coupled with an air-to-air heat exchanger, adequate
ventilation is provided at lower operating cost.
Natural ventilation is driven by pressures from wind and indoor-outdoor
temperature differences, causing air movement. This type of ventilation is
characterized by occupant control. Airflow through openable windows, doors and
other design openings can be used provide adequate ventilation for contaminant
dilution and temperature control.
Natural Ventilation
Natural or passive ventilation occurs because of wind and thermal pressure that
produce a flow of outdoor air through openable windows, doors and other
controllable openings. This is in contrast to infiltration, airflow through the
unintentional openings of a buildings. Natural ventilation can be used effectively
for both temperature and contaminant control. Temperature control by natural
ventilation conserves energy during the cooling season and is particularly
effective in mild climates. The arrangement, location and control of ventilation
openings should be designed to combine the driving forces of wind and
temperature.
Natural Ventilation Openings
Types of natural ventilation openings include: (1) windows, doors monitor
openings and skylights, (2) roof ventilators, (3) stacks connecting to registers and
(4) specially designed inlet or outlet openings.
Windows transmit light and provide ventilation when open. They may open by
sliding vertically or horizontally; by tilting on horizontal pivots at no near the
center ; or by swinging on pivots at the top, bottom or side. The type of pivoting
is an important consideration for weather protection.
Roof ventilators are determined to provide a weatherproof air outlet. Capacity is
determined by the ventilators offer to air flow; its ability to use kinetic wind energy
to induce flow by centrifugal or ejector action; and the height of the draft.
Natural draft or gravity roof ventilators can be stationary, pivoting or oscillating,
and rotating selection criteria are: ruggedness; corrosion-resistance: storm
proofing features; dampers and operating mechanism; possibility of noise;
original cost; and maintenance. Natural ventilators can supplement power-driven
supply fans the motors need only be energized when the natural exhaust
dampers or dampers controlled by thermostat or wind velocity.
A roof ventilator should be positioned so that it receives the full, unrestricted
wind. Turbulence created by surrounding obstructions, including higher adjacent
buildings, impairs a ventilator’s ejector action. The ventilator inlet should be
conical or bell mounted to give a high flow coefficient. The opening area at the
inlet should be increased if screens. Building air inlets at lower levels should be
larger than the combined throat areas of all roof ventilators.
Stacks or vertical flues should be located where wind can act on them from any
direction. Without wind, stack effect alone removes air from the room with the
inlets.
Required Flow
The ventilation flow needed to remove a given amount of heat from a building
can be calculated from Eq. (1) if the quantity of heat to be removed and the
average indoor-out-door temperature difference are known.
Q =
Where
Q = Air flow removed, m3/hr
H = Heat removed, W
Cp = Specific heat of air at constant pressure, 1 KJ.K
T1-T0 = Average indoor-outdoor temperature difference, K
cf1 = Conversion factor, 0.28
cf1 = Conversion factor, 0.34
Flow Caused by Wind
Factors that affect ventilation wind forces include average speed, prevailing
direction, seasonal and daily variation in speed and direction, and local
obstructions such as nearby buildings, hills trees and shrubbery.
Wind speeds are usually lower in summer than in winter; frequency from various
directions differs in summer and winter. There are relatively few places where
speed falls below half the average for more than a few hours a month. Therefore
natural ventilating systems are often designed for wind speeds of half the
average seasonal velocity.
Equation (2) shows the quantity of air forced through ventilation inlet openings by
wind or determines the proper size of openings to produce given airflow rates:
Q = (cf)CvAv (2)
Where
Q = Air flow m3/hr
A = Free area of inlet openings, m2
Wind speed, m/s
Cv = Effectiveness of openings (Cv) is assumed to be 0.50 to
0.60 for perpendicular winds
cf = Conversion factor, 3600
Inlets should face directly into the prevailing wind direction. If they are not
advantageously placed flow will be less than in the equation: if unusually well
placed flow will be slightly more. Desirable outlet locations are (1) on the
leeward side of the building directly opposite the inlet (2) on the roof.
Ventilation and Infiltration
In the pressure area caused by a flow discontinuity of the wind, (3) on the
adjacent to the windward face where low pressure areas occur, (4) in a monitor
on the leeward side, (5) in roof ventilators or (6) by stacks. Refer to Chapter 14
for a general description of wind on a building.
Flow Caused by Thermal Forces
If there is not significant building internal resistance, the flow caused by stack
effect is:
Q = (cf)A[h(T1 – T0)/T1]1/2
Where
Q = Air flow, m3/hr
A = Free area of inlets or outlets (assumed equal), m3
H = Height from lower opening to NPL, m
T1 = Average temperature of indoor air in height h,
K[=t(deg.c)+273.15]
T0 = Temperature of outdoor air, K
cf = Conversion factor, including a value of 65% for
effectiveness of openings; this should be 50% if conditions
are not favourable (cf = 10360)
The height h is the distance from the lower opening to the neutral pressure level.
Natural Ventilation Guidelines
Several general rules should be observed in designing for ventilation:
1. Systems using natural ventilation should be designed for effective
ventilation regardless of wind direction. Ventilation must be adequate
when the wind does not come from the prevailing direction.
2. Inlet openings should not be obstructed by buildings, signboards or
indoors partitions.
3. Greatest flow per unit area of total opening is obtained by inlet and outlet
openings of nearly equal areas.
4. The neutral pressure level tends to move to the level of any single
openings, resulting in pressure reduction across the opening. Two
openings on opposite sides of a space increase the ventilation flow. If the
openings are at the same level and near the ceiling, much of the flow may
bypass the occupied level and the ineffective in diluting contaminants at
the occupied level.
5. There must be vertical distance between openings for temperature
difference to produce natural ventilation; the greater the vertical distance,
the greater the ventilation.
6. Openings in the vicinity of the NPL are least effective for thermally induced
ventilation.
7. Openings with areas much larger than calculated are sometimes desirable
when anticipating increased occupancy or very hot weather. The
openings should be accessible to and operable by occupants.
8. When both wind the stack pressures act together, even without
interference, estimated resulting airflow is not equal to the two flows
separately. Flow through any openings is proportional to the square root
of the sum of the squares of the two flows calculated separately.
C. FORCED VENTILATION
This involves forced supply systems, forced exhaust systems or both, depending
on the requirements.
This is done by fans of various types, including propeller fans, axial flow fans and
centrifugal fans. Propeller fans are generally wall mounted type and cater to
small capacity / small pressure static requirements. Axial fans can either be duct
mounted or wall mounted type and cater to medium capacity requirement.
Centrifugal fans, which are a separate topic by themselves, cater to a wide range
of capacity and static pressure requirements.
The later two types of fans can be hooked up to a supply or exhaust duct system.
They can also be hooked up to an air washer / fan-filter system.
Forced ventilation systems can also be classified into dry or wet systems. Dry
systems involve the use of fans alone or with filter banks for dust removal.
Wet ventilation involves the use of fans with filters and a water spray / water
logged fill arrangement which will humidity hot, dry air and cool it. These
systems are suitable for hot and dry areas and are not effective in high humidity /
coastal areas.