Unit Syllabus

Introduction to air-conditioning: Psychometric properties & processes – characterization of

sensible and latent heat loads – need for ventilation, consideration of infiltration – load concepts
of RSHF , GSHF, problems concepts of ESHF and ADP temperature.

Requirements of human comfort and concept of effective temp. comfort chart , comfort air-
conditioning, - requirements of industrial air conditioning load calculations.

Unit Objectives:

● Air conditioning, Psychometric properties

● Analyze characterization of sensible and latent heat loads.
● Know need for ventilation, consideration of infiltration,
● Know the concepts of RSHF, GSHF,ESHF and ADP temp.
● Know human comfort, effective temp. comfort chart and comfort air-conditioning.

Unit Outcomes:

After reading this unit, you must be able to:

● Understand Air conditioning, Psychometric properties

● Explain the need for ventilation, consideration of infiltration.
● Distinguish between concepts of RSHF, GSHF,ESHF and ADP temp.
● Debate on ventilation, consideration of infiltration.
● Calculation of cooling loads for human comfort and industrial necessity.

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Unit Lecture Plan

54 Introduction to air conditioning, psychometric Chalk and Board

5.1 properties learning
55 Chalk and Board
Psychometric process, sensible and latent heat loads.
5.2 learning
56 Load concepts of RSHF, GSHF and Problems- Chalk and Board
5.3 Concepts of ESHF and ADP temperature. learning
57 Requirements of human comfort, concept of ICT Tool
5.4 effective temp
58 Comfort chart, comfort air conditioning_ ICT Tool
5.5 Requirement of Industrial Air-Conditioning.
55 Participative
Air conditioning load calculations_ Problems learning
5.6
56 Air conditioning systems - Classification of Chalk and Board
5.7 equipments learning
57 Cooling, heating, humidification - Dehumidification, Experiential
5.8 filters learning
58 Grills and registers_ Fans and blowers - Heat Chalk and Board
5.9 sources learning
59 Think based
Different heat pump circuits.
5.10 Learning

Lecture5.1: Introduction to air conditioning, psychometric

Air conditioning is a process that simultaneously conditions air; distributes it

combined with the outdoor air to the conditioned space; and at the same time controls and
maintains the required space‘s temperature, humidity, air movement, air cleanliness, sound
level, and pressure differential within predetermined limits for the health and comfort of the
occupants, for product processing, or both. The acronym HVAC&R stands for heating,
ventilating, air-conditioning, and refrigerating. The com-bination of these processes is
equivalent to the functions performed by air-conditioning.

Air-Conditioning Systems
An air-conditioning or HVAC&R system consists of components and equipment
arranged in sequential order to heat or cool, humidify or dehumidify, clean and purify,
attenuate objectionable equipment noise, transport the conditioned outdoor air and recirculate
air to the conditioned space, and control and maintain an indoor or enclosed environment at
optimum energy use. The types of buildings which the air-conditioning system serves can be
classified as:
• Institutional buildings, such as hospitals and nursing homes
• Commercial buildings, such as offices, stores, and shopping centers
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 • Residential buildings, including single-family and multifamily low-rise buildings
of three or fewer stories above grade
• Manufacturing buildings, which manufacture and store products
Air conditioners use the vapor compression cycle, a 4-step process.
1. The compressor (in the outside unit) pressurizes a gaseous refrigerant. The
refrigerant heats up during this process.
2. Fans in the outdoor unit blow air across the heated, pressurized gas in the
condensing coil; the refrigerant gas cools and condenses into a liquid.
3. The pressurized liquid is piped inside to the air-handling unit. It enters a throttling
or expansion valve, where it expands and cools.
4. The cold liquid circulates through evaporator coils. Inside air is blown across the
coils and cooled while the refrigerant warms and evaporates. The cooled air is
blown through the ductwork. The refrigerant, now a gas, returns to the outdoor
unit where the process repeats.

Psychometric properties:
Psychometric is the study of the properties of mixtures of air and water vapour.
Atmospheric air is a mixture of many gases plus water vapour and a number of pollutants.
The amount of water vapor and pollutants vary from place to place. The concentration of
water vapor and pollutants decrease with altitude, and above an altitude of about 10 km,
atmospheric air consists of only dry air. The pollutants have to be filtered out before
processing the air. Hence, what we process is essentially a mixture of various gases that
constitute air and water vapour. This mixture is known as moist air.
The moist air can be thought of as a mixture of dry air and moisture. For all practical
purposes, the composition of dry air can be considered as constant. In 1949, a standard
composition of dry air was fixed by the International Joint Committee on Psychrometric data.
It is given in Table.

Methods for estimating properties of moist air:

In order to perform air conditioning calculations, it is essential first to estimate various
properties of air. It is difficult to estimate the exact property values of moist air as it is a
mixture of several permanent gases and water vapour. However, moist air upto 3 atm.
pressure is found to obey perfect gas law with accuracy sufficient for engineering
calculations. For higher accuracy Goff and Gratch tables can be used for estimating moist air
properties. These tables are obtained using mixture models based on fundamental principles
of statistical mechanics that take into account the real gas behaviour of dry air and water
vapour. However, these tables are valid for a barometric pressure of 1 atm. only. Even though
the calculation procedure is quite complex, using the mixture models it is possible to estimate

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moist air properties at 5 other pressures also. However, since in most cases the pressures
involved are low, one can apply the perfect gas model to estimate psychrometric properties.
Basic gas laws for moist air:
According to the Gibbs-Dalton law for a mixture of perfect gases, the total pressure exerted
by the mixture is equal to the sum of partial pressures of the constituent gases. According to
this law, for a homogeneous perfect gas mixture occupying a volume V and at temperature T,
each constituent gas behaves as though the other gases are not present (i.e., there is no
interaction between the gases). Each gas obeys perfect gas equation.Hence, the partial
pressures exerted
by each gas, p1,p2,p3... and the total pressure pt are given by:

where n1,n2,n3,... are the number of moles of gases 1,2,3,...

Applying this equation to moist air

where p = pt= total barometric pressure

pa= partial pressure of dry air
pv= partial pressure of water vapour

Important psychrometric properties:

Dry bulb temperature (DBT) is the temperature of the moist air as measured by a
standard thermometer or other temperature measuring instruments.

Saturated vapour pressure (psat) is the saturated partial pressure of water vapour at
the dry bulb temperature. This is readily available in thermodynamic tables and charts.
ASHRAE suggests the following regression equation for saturated vapour pressure of water,
which is valid for 0 to 1000C.

where psat= saturated vapor pressure of water in kiloPascals

T = temperature in K
The regression coefficients c1 to c6 are given by:
c1 = -5.80022006E+03, c2= -5.516256E+00, c3= -4.8640239E-02
c4 = 4.1764768E-05, c5= -1.4452093E-08, c6= 6.5459673E+00

Relative humidity (Φ) is defined as the ratio of the mole fraction of water vapour in
moist air to mole fraction of water vapour in saturated air at the same temperature and
pressure. Using perfect gas equation we can show that:

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Relative humidity is normally expressed as a percentage. When Φ is 100 percent, the air is
saturated.
Humidity ratio (W): The humidity ratio (or specific humidity) W is the mass of
water associated with each kilogram of dry air1. Assuming both water vapour and dry air to
be perfect gases2, the humidity ratio is given by:

Substituting the values of gas constants of water vapour and air Rv and Ra in the above
equation; the humidity ratio is given by:

For a given barometric pressure pt, given the DBT, we can find the saturated vapour
pressure psat from the thermodynamic property tables on steam. Then using the above
equation, we can find the humidity ratio at saturated conditions, Wsat.
It is to be noted that, W isa function of both total barometric pressure and vapor
pressure of water.
Dew-point temperature:
If unsaturated moist air is cooled at constant pressure, then the temperature at which
the moisture in the air begins to condense is known as dew-point temperature (DPT) of air.
An approximate equation for dew-point temperature is given by:

where Φ is the relative humidity (in fraction). DBT & DPT are in oC. Of course, since from
its definition, the dew point temperature is the saturation temperature corresponding to the
vapour pressure of water vapour, it can be obtained from steam tables or using Eqn.
Degree of saturation μ:
The degree of saturation is the ratio of the humidity ratio W to the humidity ratio of a
saturated mixture Ws at the same temperature and pressure, i.e.,

Enthalpy:
The enthalpy of moist air is the sum of the enthalpy of the dry air and the enthalpy of
the water vapour. Enthalpy values are always based on some reference value. For moist air,
the enthalpy of dry air is given a zero value at 0oC, and for water vapour the enthalpy of
saturated water is taken as zero at 0oC.
The enthalpy of moist air is given by:

where cp= specific heat of dry air at constant pressure, kJ/kg.K

cpw= specific heat of water vapor, kJ/kg.K
t = Dry-bulb temperature of air-vapor mixture, oC
W = Humidity ratio, kg of water vapor/kg of dry air
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 ha= enthalpy of dry air at temperature t, kJ/kg
hg= enthalpy of water vapor3at temperature t, kJ/kg
hfg= latent heat of vaporization at 0oC, kJ/kg
The unit of h is kJ/kg of dry air. Substituting the approximate values of cp and hg, we obtain

Humid specific heat:

From the equation for enthalpy of moist air, the humid specific heat of moist air can
be written as:

where cpm= humid specific heat, kJ/kg.K

cp= specific heat of dry air, kJ/kg.K
cpw = specific heat of water vapor, kJ/kg
W = humidity ratio, kg of water vapor/kg of dry air
Since the second term in the above equation (w.cpw) is very small compared to the
first term, for all practical purposes, the humid specific heat of moist air, cpm can be taken as
1.0216 kJ/kg dry air.K
Specific volume:
The specific volume is defined as the number of cubic meters of moist air per
kilogram of dry air. From perfect gas equation since the volumes occupied by the individual
substances are the same, the specific volume is also equal to the number of cubic meters of
dry air per kilogram of dry air, i.e.,

Lecture 5.2: Psychometric process, sensible and latent heat loads.

a) Sensible cooling:
During this process, the moisture content of air remains constant but its temperature
decreases as it flows over a cooling coil. For moisture content to remain constant, the surface
of the cooling coil should be dry and its surface temperature should be greater than the dew
point temperature of air. If the cooling coil is 100% effective, then the exit temperature of air
will be equal to the coil temperature. However, in practice, the exit air temperature will be
higher than the cooling coil temperature. Figure 28.1 shows the sensible cooling process O-A
on a psychometric chart. The heat transfer rate during this process is given by:

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b) Sensible heating (Process O-B):
During this process, the moisture content of air remains constant and its temperature
increases as it flows over a heating coil. The heat transfer rate during this process is given by:

where c is the humid specific heat (≈1.0216 kJ/kg dry air) and mpma is the mass flow rate of
dry air (kg/s). Figure 28.2 shows the sensible heating process on a psychrometric chart.

c) Cooling and dehumidification (Process O-C):

When moist air is cooled below its dew-point by bringing it in contact with a cold
surface as shown in Fig.28.3, some of the water vapor in the air condenses and leaves the air
stream as liquid, as a result both the temperature and humidity ratio of air decreases as
shown. This is the process air undergoes in a typical air conditioning system. Although the
actual process path will vary depending upon the type of cold surface, the surface
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temperature, and flow conditions, for simplicity the process line is assumed to be a straight
line. The heat and mass transfer rates can be expressed in terms of the initial and final
conditions by applying the conservation of mass and conservation of energy equations as
given below:
By applying mass balance for the water:

It can be observed that the cooling and de-humidification process involves both latent
and sensible heat transfer processes, hence, the total, latent and sensible heat transfer rates (Q,
Qtl and Qs) can be written as:

By separating the total heat transfer rate from the cooling coil into sensible and latent
heat transfer rates, a useful parameter called Sensible Heat Factor (SHF) is defined. SHF is
defined as the ratio of sensible to total heat transfer rate, i.e.,

From the above equation, one can deduce that a SHF of 1.0 corresponds to no latent heat
transfer and a SHF of 0 corresponds to no sensible heat transfer. A SHF of 0.75 to 0.80 is
quite common in air conditioning systems in a normal dry-climate. A lower value of SHF,
say 0.6, implies a high latent heat load such as that occurs in a humid climate. From Fig.28.3,
it can be seen that the slope of the process line O-C is given by:

From the definition of SHF,

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From the above equations, we can write the slope as:

Thus we can see that the slope of the cooling and de-humidification line is purely a
function of the sensible heat factor, SHF. Hence, we can draw the cooling and de-
humidification line on psychrometric chart if the initial state and the SHF are known. In some
standard psychrometric charts, a protractor with different values of SHF is provided. The
process line is drawn through the initial state point and in parallel to the given SHF line from
the protractor as shown in Fig.28.4.

In Fig.28.3, the temperature Ts is the effective surface temperature of the cooling coil, and is
known as apparatus dew-point (ADP) temperature. In an ideal situation, when all the air
comes in perfect contact with the cooling coil surface, then the exit temperature of air will be
same as ADP of the coil. However, in actual case the exit temperature of air will always be
greater than the apparatus dew-point temperature due to boundary layer development as air
flows over the cooling coil surface and also due to temperature variation along the fins etc.
Hence, we can define a by-pass factor (BPF) as:

It can be easily seen that, higher the by-pass factor larger will be the difference
between air outlet temperature and the cooling coil temperature. When BPF is 1.0, all the air
by-passes the coil and there will not be any cooling or de-humidification. In practice, the by-
pass factor can be increased by increasing the number of rows in a cooling coil or by
decreasing the air velocity or by reducing the fin pitch.
Alternatively, a contact factor(CF) can be defined which is given by:

d) Heating and Humidification (Process O-D):

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 During winter it is essential to heat and humidify the room air for comfort. As shown
in Fig.28.5., this is normally done by first sensibly heating the air and then adding water
vapour to the air stream through steam nozzles as shown in the figure.
Mass balance of water vapor for the control volume yields the rate at which steam has
to be added, i.e., mw:

where ma is the mass flow rate of dry air.

From energy balance:

where Qh is the heat supplied through the heating coil and hw is the enthalpy of steam.
Since this process also involves simultaneous heat and mass transfer, we can define a sensible
heat factor for the process in a way similar to that of a cooling and dehumidification process.

e) Cooling & humidification (Process O-E):

As the name implies, during this process, the air temperature drops and its humidity
increases. This process is shown in Fig.28.6. As shown in the figure, this can be achieved by
spraying cool water in the air stream. The temperature of water should be lower than the dry-
bulb temperature of air but higher than its dew-point temperature to avoid condensation
(TDPT< T w < TO).

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It can be seen that during this process there is sensible heat transfer from air to water and
latent heat transfer from water to air. Hence, the total heat transfer depends upon the water
temperature. If the temperature of the water sprayed is equal to the wet-bulb temperature of
air, then the net transfer rate will be zero as the sensible heat transfer from air to water will be
equal to latent heat transfer from water to air. If the water temperature is greater than WBT,
then there will be a net heat transfer from water to air. If the water temperature is less than
WBT, then the net heat transfer will be from air to water. Under a special case when the spray
water is entirely recirculated and is neither heated nor cooled, the system is perfectly
insulated and the make-up water is supplied at WBT, then at steady-state, the air undergoes
an adiabatic saturation process, during which its WBT remains constant. This is the process
of adiabatic saturation. The process of cooling and humidification is encountered in a wide
variety of devices such as evaporative coolers, cooling towers etc.
f) Heating and de-humidification (Process O-F):

This process can be achieved by using a hygroscopic material, which absorbs or

adsorbs the water vapor from the moisture. If this process is thermally isolated, then the
enthalpy of air remains constant, as a result the temperature of air increases as its moisture
content decreases as shown in Fig.28.7. This hygroscopic material can be a solid or a liquid.
In general, the absorption of water by the hygroscopic material is an exothermic reaction, as a
result heat is released during this process, which is transferred to air and the enthalpy of air
increases.

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g) Mixing of air streams:
Mixing of air streams at different states is commonly encountered in many processes,
including in air conditioning. Depending upon the state of the individual streams, the mixing
process can take place with or without condensation of moisture.
i) Without condensation: Figure 28.8 shows an adiabatic mixing of two moist air streams
during which no condensation of moisture takes place. As shown in the figure, when two air
streams at state points 1 and 2 mix, the resulting mixture condition 3 can be obtained from
mass and energy balance.
From the mass balance of dry air and water vapor:

From energy balance:

From the above equations, it can be observed that the final enthalpy and humidity
ratio of mixture are weighted averages of inlet enthalpies and humidity ratios. A generally
valid approximation is that the final temperature of the mixture is the weighted average of the
inlet temperatures. With this approximation, the point on the psychrometric chart
representing the mixture lies on a straight line connecting the two inlet states. Hence, the ratio
of distances on the line, i.e., (1-3)/(2-3) is equal to the ratio of flow rates m a,2/ma,1. The
resulting error (due to the assumption that the humid specific heats being constant) is usually
less than 1 percent.

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ii) Mixing with condensation:
As shown in Fig.28.9, when very cold and dry air mixes with warm air at high relative
humidity, the resulting mixture condition may lie in the two-phase region, as a result there
will be condensation of water vapor and some amount of water will leave the system as liquid
water. Due to this, the humidity ratio of the resulting mixture (point 3) will be less than that at
point 4. Corresponding to this will be an increase in temperature of air due to the release of
latent heat of condensation. This process rarely occurs in an air conditioning system, but this
is the phenomenon which results in the formation of fog or frost (if the mixture temperature is
below 0oC). This happens in winter when the cold air near the earth mixes with the humid
and warm air, which develops towards the evening or after rains.

Two types of heat are important in air conditioning. These include:

1. Sensible Heat
2. Latent Heat
SENSIBLE HEAT

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Sensible heat is heat exchanged by a thermodynamic system that changes the temperature of
the system without changing some variables such as volume or pressure. As the name
implies, sensible heat is the heat that you can feel. The sensible heat possessed by an object is
evidenced by its temperature. As temperature increases, the sensible heat content also
increases. However, for a given change in sensible heat content, all objects do not change
temperature by the same amount. Each substance has its own characteristics relationship
between heat content and temperature. The proportionality constant between temperature rise
and change in heat content is called the specific heat, measured in calories per gram per
degree Celsius or joules per kilogram per kelvin. Water for example has a specific heat of 1
Cal/g/oC. In general, the gain in heat is accompanied by either a change in volume or a
change in pressure (e.g. the water in the pot swells somewhat as you heat it; if you heat gas in
a fixed volume, its pressure goes up).
LATENT HEAT
This is the energy absorbed or released by a thermodynamic system during a constant
temperature process. Examples include ice melting or water boiling. When a solid turns into a
liquid (melts) or a liquid turns into a gas (evaporates), the loosening of attraction among the
molecules requires energy. If you raise ice from -20oC to 0oC, you put in sensible heat. If you
keep adding heat to the ice, it melts but its temperature is constant, the sensible heat of
ice/water system is not increasing but you continue to add heat energy to it. Energy is
conserved, such that the extra heat tears apart the frozen ice molecules and sets them loose as
a liquid. The liquid is therefore storing this energy in a form that you cannot sense. This
energy is call latent heat.
To melt all the ice, you have to pump in quite a bit of heat, but you cannot sense any change
in the heat content because ice/water system remains at 0oC. Only after all the ice has melted
does the temperature of the water rise. At this point, the heat you put in is once again creating
a change in sensible heat. However, although you cannot feel it, the liquid has stored all that
latent heat. The only way you will observe the latent heat is if you try to transform the water
back to ice. If you take the temperature down to 0oC, that alone will not freeze water; you
must keep pulling out heat until you have removed every joule of latent heat. Only then will
all the water freeze and you can begin to remove more sensible heat and lower the
temperature of the system below 0oC.
Heat energy is conserved no matter how the phase change occurs. If you put heat into water,
it can evaporate. It evaporates on its own even if you do not add heat. The water will cool off;
i.e. some of the sensible heat is lost and converted to latent heat. Conversely, if you cool off
some water vapor, it can condense into liquid. If it condenses on its own, it will give off
(sensible) heat and get warmer. Therefore sensible heat can be felt while latent heat is the
type of heat that cannot be felt.

Lecture 5.3: : Load concepts of RSHF, GSHF and Problems- Concepts of ESHF and
ADP temperature.

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 Room Sensible Heat Factor - RSHF - is defined as sensible heat load divided
by total heat load in a room.
The Room Sensible Heat Factor - RSHF - expresses the ratio between sensible heat load and
total heat load in a room. It can be expressed as:
RSHF = Qsr / Qtr (1)
Where RSHF = room sensible heat factor
Qsr = sensible heat load in room (kW, Btu/hr)
Qtr = total heat load in room (kW, Btu/hr)
(1) can be modified to:
RSHF = cp (tr - ti) / (hr - hi) (2)
where
cp = specific heat capacity of air (kJ/kg.oC)
tr = room temperature (oC)
ti = inlet temperature (oC)
hr = room enthalpy (kJ/kg)
hi = inlet enthalpy (kJ/kg)

Room Sensible Heat Factor - RSHF

The sensible heat generated within a room is 4 kW. The latent heat generated is 1.4
kW.
The Room Sensible Heat Factor can be calculated as:
RSHF = (4 kW) / ((4 kW) + (1.4 kW))
= 0.75

Load concepts of GSHF:

Great Sensible Heat Factor is the ratio between sensible and total heat in a cooling coil
A cooling coil should have the capacity to handle the heat load generated in the room in
addition the capacity to cool down the fresh make up air. The Great Sensible Factor is
independent of the total air flow and can be expressed as
GSHF = [Qs + mf cp (tf - tr)] / [Qt + mf (hf - hr)] (1)
Where GSHF = Great Sensible Heat Factor
Qs = sensible heat generated in the room (kW)
mf = fresh air mass flow (make up air) (kg/s)
cp = specific heat of air (kg/kg.oC)
tf = temperature in fresh air (oC)
tr = temperature in room air (oC)
Qt = total heat generated in the room (kW)
hf = specific enthalpy of fresh air (kJ/kgoC)
hr = specific enthalpy of air in the room (kJ/kgoC)

Great Sensible Heat Factor - GSHF

The sensible heat generation in a room is 4 kW and the latent heat generated is 1.4
kW. The room is cooled with an air flow of 1000 kg/h.
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The fresh (out door) air temperature is 28oC with a specific enthalpy of 60 kJ/kg and a
humidity ratio of 12.7 g/kg.
The room temperature is 22oC with a specific enthalpy of 45 kJ/kg and a specific humidity of
9.1 g/kg.
The Great Sensible Heat Factor can be calculated like
GSHF = [(4000 W) + (1000 kg/h)(1/3600 s/h)(1000 J/kgoC)(28oC - 22oC)] / [(4000 W + 1400
W) + (1000 kg/h)(1/3600 s/h)(60000 J/kg - 45000 J/kg)]
= 0.59

Concept of ESHF (Effective Sensible Heat Factor):

It is the ratio of effective room sensible heat to the effective room total heat.
ESHF= ERSH/ERTH.

Concept of ADP (Apparatus dew point):

ADP is the coil surface dew point temperature required to accomplish a
cooling/dehumidifying process.

An ADP is essential to any cooling or dehumidifying process. The chart below shows an
impossible process because there is no ADP

Relationship of ADP to Supply Air and Chilled Water temperatures

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Bypass Factor:
Some of the air flowing through the coil impinges on the water tubes or the fins and is
cooled to the ADP. Other air passes through unchanged.

The percentage of air that passes through the coil unchanged is called the bypass factor
Apparatus Dew Point (ADP) is the Cooling coil (Evaporator) temperature. It can be
determined in the following ways:
(i) Apparatus Dew Point (ADP) is the point where GSHF line meets the saturation curve.
GSHF line is drawn from the mixture condition parallel to the GSHF line obtained with the
help of GSHF and fixed point of 27oC and 50 % RH n psychrometric chart. It is applied when
outside air and room recirculated air first mixed and then passed over the cooling coil.
(ii) Apparatus Dew Point (ADP) is the point where OASHF line meets the saturation curve.
OASHF line is drawn from the outside design condition (ODC condition) parallel to the
OASHF line. It is applicable when outside air is first cooled and then mixed with the
recirculated room air.
(iii) Apparatus Dew Point (ADP) is the point where ERSHF line meets the saturation curve.
ERSHF line is drawn from the room condition (IDC condition) parallel to the ERSHF line

Lecture 5.4: Requirements of human comfort, concept of effective temp

Selection of inside design conditions:

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The required inside design conditions depend on the intended use of the building. Air
conditioning is required either for providing suitable comfort conditions for the occupants
(e.g. comfort air conditioning), or for providing suitable conditions for storage of perishable
products (e.g. in cold storages) or conditions for a process to take place or for products to be
manufactured (e.g. industrial air conditioning). The required inside conditions for cold
storage and industrial air conditioning applications vary widely depending on the specific
requirement. However, the required inside conditions for comfort air conditioning systems
remain practically same irrespective of the size, type, location, use of the air conditioning
building etc., as this is related to the thermal comfort of the human beings.
Thermal comfort:
Thermal comfort is defined as ―that condition of mind which expresses satisfaction with the
thermal environment‖. This condition is also some times called as ―neutral condition‖, though
in a strict sense, they are not necessarily same. A living human body may be likened to a heat
engine in which the chemical energy contained in the food it consumes is continuously
converted into work and heat. The process of conversion of chemical energy contained in
food
into heat and work is called as ―metabolism‖. The rate at which the chemical energy is
converted into heat and work is called as ―metabolic rate‖. Knowledge of metabolic rate of
the occupants is required as this forms a part of the cooling load of the air conditioned
building. Similar to a heat engine, one can define thermal efficiency of a human being as the
ratio of useful work output to the energy input. The thermal efficiency of a human being can
vary from 0% to as high as 15-20% for a short duration. By the manner in which the work is
defined, for most of the light activities the useful work output of human beings is zero,
indicating a thermal efficiency of 0%. Irrespective of the work output, a human body
continuously generates heat at a rate varying from about 100 W (e.g. for a sedentary person)
to as high as 2000W (e.g. a person doing strenuous exercise). Continuous heat generation is
essential, as the temperature of the human body has to be maintained within a narrow range
of temperature, irrespective of the external surroundings. A human body is very sensitive to
temperature. The body temperature must be maintained within a narrow range to avoid
discomfort, and within a somewhat wider range, to avoid danger from heat or cold stress.
Studies show that at neutral condition, the temperatures should be:
Skin temperature, tskin≈33.7oC
Core temperature, tcore≈36.8oC
At other temperatures, the body will feel discomfort or it may even become lethal. It
is observed that when the core temperature is between 35 to 39oC, the body experiences only
a mild discomfort. When the temperature is lower than 35oC or higher than 39oC, then
people suffer major loss in efficiency. It becomes lethal when the temperature falls below
31oC or rises above 43oC. This is shown in Fig. 29.2.

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Concepts of Effective temperature:
Indices for thermal comfort:
It is seen that important factors which affect thermal comfort are the activity, clothing, air
DBT, RH, air velocity and surrounding temperature. It should be noted that since so many
factors are involved, many combinations of the above conditions provide comfort. Hence to
evaluate the effectiveness of the conditioned space, several comfort indices have been
suggested. These indices can be divided into direct and derived indices. The direct indices are
the dry bulb temperature, humidity ratio, air velocity and the mean radiant temperature
(Tmrt).

The mean radiant temperature Tmrt affects the radiative heat transfer and is defined (in K) as:

where:
Tg= Globe temperature measured at steadystate by a thermocouple placed at the center of a
black painted, hollow cylinder (6‖ dia)
kept in the conditioned space, K. The reading of thermocouple results from a balance of
convective and radiative heat exchanges between
the surroundings and the globe
Ta= Ambient DBT, K
V = Air velocity in m/s, and
C = A constant, 0.247 X 109
The derived indices combine two or more direct indices into a single factor. Important
derived indices are the effective temperature, operative temperature, heat stress index,
Predicted Mean Vote (PMV), Percent of People Dissatisfied (PPD) etc.

Effective temperature (ET):

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 This factor combines the effects of dry bulb temperature and air humidity into a single
factor. It is defined as the temperature of the environment at 50% RH which results in the
same total loss from the skin as in the actual environment. Since this value depends on other
factors such as activity, clothing, air velocity and Tmrt, a Standard Effective Temperature
(SET) is defined for the following conditions:
clothing = 0.6 clo
Activity = 1.0 met
Air velocity = 0.1 m/s
Tmrt= DBT (in K)
Operative temperature (Top):
This factor is a weighted average of air DBT and Tmrt into a single factor. It is given by:

where hr and hc are the radiative and convective heat transfer coefficients and Tamb is the
DBT of air.

Comfort chart:
ASHRAE has defined a comfort chart based on the effective and operative temperatures.
Figure 29.3 shows the ASHRAE comfort chart with comfort zones for summer and winter
conditions. It can be seen from the chart that the comfort zones are bounded by effective
temperature lines, a constant RH line of 60% and dew point temperature of 2oC. The upper
and lower limits of humidity (i.e. 60 % RH and 2oC DPT, respectively) are based on the
moisture content related considerations of dry skin, eye irritation, respiratory health and
microbial growth. The comfort chart is based on statistical sampling of a large number of
occupants with activity levels less than 1.2 met. On the chart, the region where summer and
winter comfort zones overlap, people in winter clothing feel slightly warm and people in
summer clothing feel slightly cool. Based on the chart ASHARE makes the following
recommendations:
Inside design conditions for Winter:
Top between 20.0 to 23.5oC at a RH of 60%
Top between 20.5 to 24.5oC at a DPT of 2oC
Inside design conditions for Summer:
Top between 22.5 to 26.0oC at a RH of 60%
Top between 23.5 to 27.0oC at a DPT of 2oC

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Table 29.2 shows the recommended comfort conditions for different seasons and clothing
suitable at 50 % RH, air velocity of 0.15 m/s and an activity level of ≤1.2 met.

The above values may be considered as recommended inside design conditions for
comfort air conditioning. It will be shown later that the cost of air conditioning (initial plus
running) increases as the required inside temperature increases in case of winter and as the
required inside condition decreases in case of summer. Hence, air conditioning systems
should be operated at aslow a temperature as accept able in winter and as high a temperature
as acceptable in summer. Use of suitable clothing and maintaining suitable air velocities in
the conditioned space can lead to reduced cost of air conditioning. For example, in summer
the clothing should be minimal at a socially acceptable level, so that the occupied space can
be maintained at higher temperatures. Similarly, by increasing air velocity without causing
draft, one can maintain the occupied space at a higher temperature in summer. Similarly, the
inside temperatures can be higher for places closer to the equator (1oC rise in ET is allowed
for each 5oreduction in latitude). Of course, the above recommendations are for normal
activities. The required conditions change if the activity levels are different. For example,
when the activity level is high (e.g. in gymnasium), then the required indoor temperatures
will be lower. These special considerations must be kept in mind while fixing the inside
design conditions. Prof. P.O. Fanger of Denmark has carried out pioneering and detailed
studies on thermal comfort and suggested comfort conditions for a wide variety of situations.

Lecture 5.5: Comfort chart, comfort air conditioning_ Requirement of Industrial Air-
Conditioning.
146
Low-energy buildings, offices, museums, sports facilities, schools, clinics, hotels, banks,
historical buildings and many more. With comfort air conditioning, the focus is on people.
Our technology is based on the respective requirements of a project, but simultaneously
always looks for the most efficient method with the lowest consumption of energy. For
example, we cool with water in order to save electrical energy, or make use of sorption-based
air conditioning, with which you can carry out dehumidification by means of heat, e.g. from
solar thermal energy or process waste heat. It is even possible to store excess solar heat for an
indefinite period without any losses for the purposes of dehumidification.
Comfort conditioning as the name implies is solely to provide a comfortable
environment for the majority of occupants. Humans are reasonably tolerant to humidity and
may be comfortable from a range of between 55% and 20% relative humidity at normal
comfort temperatures. It is therefore common when specifying to limit the humidity in
summer and not specify a limit in winter. Typically therefore a specification would state an
internal condition of 22°C / 50% relative humidity being maintained at 30°C / 20°C wet bulb
external conditions in summer. In winter the specification may typically be 21°C internal
temperature at -3°C saturated outside air temperature.

Requirements of industrial air conditioning:

Industrial conditioning is provided generally for a process which requires a closely
controlled atmosphere. A typical specification may be that an internal environment is
required of say 21°C ±0.5°C and 50% relative humidity ±2.5% at all external conditions. It
will be seen therefore that the industrial conditions for clearly defined limits rather than
comfort conditioning which is based on statistical surveys of occupants feelings.
On the air side, industrial air conditioners are one or more characteristics: temperature
level, the need for precise humidity control, increasing the level of filtration and removal of
contaminants. Industrial air conditioning appears in applications such as printing presses,
where humidity control is essential for proper registration of multi-color printing; textile
enterprises in order to prevent fibers from hacking and reduce static electricity; manufacture
of photographic materials; clean rooms for manufacturing electronic equipment; and in
different laboratories. One of the differences between the refrigeration and comfort for the
industrial air-conditioning may be in more high precision temperature control, generally
required for industrial air conditioning.

Winter Air Conditioning System

In winter air conditioning, the air is heated which is generally followed by
humidification. The schematic arrangement of the system is shown in Fig.
The outside air flows through a damper and mixes up with the re-circulated air (which
is obtained from the conditioned space). The mixed air passes through a filter to remove dirt,
dust and other impurities. The air now passes through a preheat coil in order to prevent the
possible freezing of water and to control the evaporation of water in the humidifier. After
that, the air is made to pass through a reheat coil to bring the air to the designed dry bulb
temperature. Now, the conditioned air is supplied to the conditioned space by a fan. From the
conditioned space, a part of the used air is exhausted to the atmosphere by the exhaust fans or
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ventilators. The remaining part of the used air (Known as re-circulated air) is again
conditioned as shown in Fig.
The outside air is sucked and made to mix with re-circulated air, in order to make up
for the loss of conditioned (or used) air through exhaust fans or ventilation from the
conditioned space.

Summer Air Conditioning System

It is the most important type of air conditioning, in which the air is cooled and generally
dehumidified. The schematic arrangement of a typical summer air conditioning system is
shown in Fig.
The outside air flows through the damper and mixes up with re-circulated air (which
is obtained from the conditioned space) The mixed air passes through a filter to remove dirt
dust and other impurities. The air now passes through a cooling coil. The coil has a
temperature much below the required dry bulb temperature of the air in the conditioned
space. The cooled air passes through a perforated membrane and loses its moisture in the
condensed from which is collected in a sump. After that, the air is made to pass through a
heating coil which heats up the air slightly. This is done to bring the air to the designed dry
bulb temperature and relative humidity.
Now the conditioned air is supplied to the conditioned space by a fan. From the conditioned
space, a part of the used air is exhausted to the atmosphere by the exhaust fans or ventilators.
The remaining part of the used air (Known as re-circulated air) is again conditioned as shown
in Fig. 18.5. The outside air is sucked and made to mix with the re-circulated air in order to
make up for the loss of conditioned (or used) air through exhaust fans or ventilation from the
conditioned space.

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Year-Round Air Conditioning System
The year-round air conditioned systems should have equipment for both the summer
and winter air conditioning. The schematic arrangement of a modern summer year-round air
conditioning system is shown in Fig.
The outside air flows through the damper and mixed up with the re-circulated air
(which is obtained from the conditioned space). The mixed air passes through a filter to
remove dirt, dust and other impurities. In summer air conditioning, the cooling coil operates
to cool the air to the desired value. The dehumidification is obtained by operating the cooling
coil at a temperature lower than the dew point temperature (apparatus dew point). In winter,
the cooling coil is made inoperative and the heating coil operates to heat the air. The spray
type humidifier is also made use of in the dry season to humidify the air.

149
Unitary Air Conditioning System
In this system, factory assembled air conditioners air installed in or adjacent to the
space to be conditioned. The unitary air conditioning systems are of the following two types:
1. Window units
These are self-contained units of small capacity of 1 TR to 3 TR, and are mounted in a
window or through the wall. They are employed to condition the air of one room only. If the
room is bigger in size, then two or more units are installed.

2. Vertical packed units

150
These are also self-contained units of bigger capacity of 5 to 20 TR and are installed adjacent
to the space to be conditioned. This is very useful for conditioning the air of a restaurant,
bank or small office. The unitary air conditioning system may be adopted for winter, summer
or year-round air conditioning.

Central Air Conditioning System

This is the most important type of air conditioning system, which is adopted, when the
cooling capacity required is 25 TR or more. The central air conditioning system is also
adopted when the air flow is more than 300 m min or different zones in a building are to air
conditioned.

Lecture 5.6: Air conditioning load calculations_ Problems

Air conditioning load calculations:

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 Cooling load calculations: As mentioned before, load calculations involve a
systematic and stepwise procedure that takes into account all the relevant building energy
flows. The cooling load experienced by a building varies in magnitude from zero (no cooling
required) to a maximum value. The design cooling load is a load near the maximum
magnitude, but is not normally the maximum. Design cooling load takes into account all the
loads experienced by a building under a specific set of assumed conditions.
The assumptions behind design cooling load are as follows:
1. Design outside conditions are selected from a long-term statistical database. The
conditions will not necessarily represent any actual year, but are representative of the location
of the building. Design data for outside conditions for various locations of the world have
been collected and are available in tabular form in various handbooks.
2. The load on the building due to solar radiation is estimated for clear sky conditions.
3. The building occupancy is assumed to be at full design capacity.
4. All building equipment and appliances are considered to be operating at a
reasonably representative capacity.
The total building cooling load consists of heat transferred through the building
envelope (walls, roof, floor, windows, doors etc.) and heat generated by occupants,
equipment, and lights. The load due to heat transfer through the envelope is called as external
load, while all other loads are called as internal loads. The percentage of external versus
internal load varies with building type, site climate, and building design. The total cooling
load on any building consists of both sensible as well as latent load components. The sensible
load affects dry bulb temperature, while the latent load affects the moisture content of the
conditioned space.
Buildings may be classified as externally loaded and internally loaded. In externally
loaded buildings the cooling load on the building is mainly due to heat transfer between the
surroundings and the internal conditioned space. Since the surrounding conditions are highly
variable in any given day, the cooling load of an externally loaded building varies widely. In
internally loaded buildings the cooling load is mainly due to internal heat generating sources
such as occupants or appliances or processes. In general the heat generation due to internal
heat sources may remain fairly constant, and since the heat transfer from the variable
surroundings is much less compared to the internal heat sources, the cooling load of an
internally loaded building remains fairly constant. Obviously from energy efficiency and
economics points of view, the system design strategy for an externally loaded building should
be different from an internally loaded building. Hence, prior knowledge of whether the
building is externally loaded or internally loaded is essential for effective system design.
Estimation of cooling load involves estimation of each of the above components from
the given data. In the present chapter, the cooling load calculations are carried out based on
the CLTD/CLF method suggested by ASHRAE. For more advanced methods such as TFM,
the reader should refer to ASHRAE and other handbooks.
Estimation of external loads:
a) Heat transfer through opaque surfaces:
This is a sensible heat transfer process. The heat transfer rate through opaque surfaces
such as walls, roof, floor, doors etc. is given by:
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 Qopaque = CLTD.A.UQ
where U is the overall heat transfer coefficient and A is the heat transfer area of the surface
on the side of the conditioned space. CLTD is the cooling load temperature difference.
For interior air conditioned rooms surrounded by non-air conditioned spaces, the
CLTD of the interior walls is equal to the temperature difference between the surrounding
non-air conditioned space and the conditioned space. Obviously, if an air conditioned room is
surrounded by other air conditioned rooms, with all of them at the same temperature, the
CLTD values of the walls of the interior room will be zero.
Estimation of CLTD values of floor and roof with false ceiling could be tricky. For
floors standing on ground, one has to use the temperature of the ground for estimating CLTD.
However, the ground temperature depends on the location and varies with time. ASHRAE
suggests suitable temperature difference values for estimating heat transfer through ground. If
the floor stands on a basement or on the roof of another room, then the CLTD values for the
floor are the temperature difference across the floor (i.e., difference between the temperature
of the basement or room below and the conditioned space). This discussion also holds good
for roofs which have non-air conditioned rooms above them. For sunlit roofs with false
ceiling, the U value may be obtained by assuming the false ceiling to be an air space.
However, the CLTD values obtained from the tables may not exactly fit the specific roof.
Then one has to use his judgement and select suitable CLTD values.

b) Heat transfer through fenestration:

Heat transfer through transparent surface such as a window, includes heat transfer by
conduction due to temperature difference across the window and heat transfer due to solar
radiation through the window. The heat transfer through the window by convection is
calculated using Eq.(35.3), with CLTD being equal to the temperature difference across the
window and A equal to the total area of the window. The heat transfer due to solar radiation
through the window is given by:

where Aunshaded is the area exposed to solar radiation, SHGFmax and SC are the
maximum Solar Heat Gain Factor and Shading Coefficient, respectively, and CLF is the
Cooling Load Factor. As discussed in a previous chapter, the unshaded area has to be
obtained from the dimensions of the external shade and solar geometry. SHGFmax and SC
are obtained from ASHRAE tables based on the orientation of the window, location, month
of the year and the type of glass and internal shading device
The Cooling Load Factor (CLF) accounts for the fact that all the radiant energy that
enters the conditioned space at a particular time does not become a part of the cooling load1
instantly. As solar radiation enters the conditioned space, only a negligible portion of it is
absorbed by the air particles in the conditioned space instantaneously leading to a minute
change in its temperature. Most of the radiation is first absorbed by the internal surfaces,
which include ceiling, floor, internal walls, furniture etc. Due to the large but finite thermal
capacity of the roof, floor, walls etc., their temperature increases slowly due to absorption of
solar radiation. As the surface temperature increases, heat transfer takes place between these

153
surfaces and the air in the conditioned space. Depending upon the thermal capacity of the
wall and the outside temperature, some of the absorbed energy due to solar radiation may be
conducted to the outer surface and may be lost to the outdoors. Only that fraction of the solar
radiation that is transferred to the air in the conditioned space becomes a load on the building,
the heat transferred to the outside is not a part of the cooling load. Thus it can be seen that the
radiation heat transfer introduces a time lag and also a decrement factor depending upon the
dynamic characteristics of the surfaces. Due to the time lag, the effect of radiation will be felt
even when the source of radiation, in this case the sun is removed. The CLF values for
various surfaces have been calculated as functions of solar time and orientation and are
available in the form of tables in ASHRAE Handbooks. Table 35.2 gives typical CLF values
for glass with interior shading.

c) Heat transfer due to infiltration:

Heat transfer due to infiltration consists of both sensible as well as latent components.
The sensible heat transfer rate due to infiltration is given by:

where hfg is the latent heat of vaporization of water, Wo and Wi are the outdoor and indoor
humidity ratio, respectively.
The infiltration rate depends upon several factors such as the tightness of the building
that includes the walls, windows, doors etc and the prevailing wind speed and direction. As
mentioned before, the infiltration rate is obtained by using either the air change method or the
crack method.
The infiltration rate by air change method is given by:

154
where ACH is the number of air changes per hour and V is the gross volume of the
conditioned space in m3 . Normally the ACH value varies from 0.5 ACH for tight and well-
sealed buildings to about 2.0 for loose and poorly sealed buildings. For modern buildings the
ACH value may be as low as 0.2 ACH. Thus depending upon the age and condition of the
building an appropriate ACH value has to be chose, using which the infiltration rate can be
calculated.
The infiltration rate by the crack method is given by:

where A is the effective leakage area of the cracks, C is a flow coefficient which depends on
the type of the crack and the nature of the flow in the crack, ΔP is the difference between
outside and inside pressure (Po-Pi) and n is an exponent whose value depends on the nature
of the flow in the crack. The value of n varies between 0.4 to 1.0, i.e., 0.4 ≤ n ≤ 1.0. The
pressure difference ΔP arises due to pressure difference due to the wind (ΔPwind), pressure
difference due to the stack effect (ΔPstack) and pressure difference due to building
pressurization (ΔPbld), i.e.,

Semi-empirical expressions have been obtained for evaluating pressure difference due
to wind and stack effects as functions of prevailing wind velocity and direction, inside and
outside temperatures, building dimensions and geometry etc.
Representative values of infiltration rate for different types of windows, doors walls
etc. have been measured and are available in tabular form in air conditioning design
handbooks.
d) Miscellaneous external loads:
In addition to the above loads, if the cooling coil has a positive by-pass factor (BPF >
0), then some amount of ventilation air directly enters the conditioned space, in which case it
becomes a part of the building cooling load. The sensible and latent heat transfer rates due to
the by-passed ventilation air can be calculated using equations (35.5) and (35.6) by replacing
with , where is the ventilation rate and BPF is the by-pass factor of the
cooling coil.
In addition to this, sensible and latent heat transfer to the building also occurs due to
heat transfer and air leakage in the supply ducts. A safety factor is usually provided to
account for this depending upon the specific details of the supply air ducts.
If the supply duct consists of supply air fan with motor, then power input to the fan
becomes a part of the external sensible load on the building. If the duct consists of the electric
motor, which drives the fan, then the efficiency of the fan motor also must be taken into
account while calculating the cooling load. Most of the times, the power input to the fan is
not known a priori as the amount of supply air required is not known at this stage. To take
this factor into account, initially it is assumed that the supply fan adds about 5% of the room
sensible cooling load and cooling loads are then estimated. Then this value is corrected in the
end when the actual fan selection is done.

Estimation of internal loads:

155
 The internal loads consist of load due to occupants, due to lighting, due to equipment
and appliances and due to products stored or processes being performed in the conditioned
space.
a) Load due to occupants:
The internal cooling load due to occupants consists of both sensible and latent heat
components. The rate at which the sensible and latent heat transfer take place depends mainly
on the population and activity level of the occupants. Since a portion of the heat transferred
by the occupants is in the form of radiation, a Cooling Load Factor (CLF) should be used
similar to that used for radiation heat transfer through fenestration. Thus the sensible heat
transfer to the conditioned space due to the occupants is given by the equation:

Table shows typical values of total heat gain from the occupants and also the sensible
heat gain fraction as a function of activity in an air conditioned space. However, it should be
noted that the fraction of the total heat gain that is sensible depends on the conditions of the
indoor environment. If the conditioned space temperature is higher, then the fraction of total
heat gain that is sensible decreases and the latent heat gain increases, and vice versa.

The value of Cooling Load Factor (CLF) for occupants depends on the hours after the
entry of the occupants into the conditioned space, the total hours spent in the conditioned
space and type of the building. Values of CLF have been obtained for different types of
buildings and have been tabulated in ASHRAE handbooks.
Since the latent heat gain from the occupants is instantaneous the CLF for latent heat
gain is 1.0, thus the latent heat gain due to occupants is given by:

b) Load due to lighting:

Lighting adds sensible heat to the conditioned space. Since the heat transferred from
the lighting system consists of both radiation and convection, a Cooling Load Factor is used
to account for the time lag. Thus the cooling load due to lighting system is given by:

The usage factor accounts for any lamps that are installed but are not switched on at
the time at which load calculations are performed. The ballast factor takes into account the
load imposed by ballasts used in fluorescent lights. A typical ballast factor value of 1.25 is
taken for fluorescent lights, while it is equal to 1.0 for incandescent lamps. The values of
CLF as a function of the number of hours after the lights are turned on, type of lighting

156
fixtures and the hours of operation of the lights are available in the form of tables in
ASHRAE handbooks.
c) Internal loads due to equipment and appliances:
The equipment and appliances used in the conditioned space may add both sensible as
well as latent loads to the conditioned space. Again, the sensible load may be in the form of
radiation and/or convection. Thus the internal sensible load due to equipment and appliances
is given by:

The installed wattage and usage factor depend on the type of the appliance or
equipment. The CLF values are available in the form of tables in ASHARE handbooks.
The latent load due to appliances is given by:

Table 35.4 shows typical load of various types of appliances.

For other equipment such as computers, printers etc, the load is in the form of sensible
heat transfer and is estimated based on the rated power consumption. The CLF value for these
equipment may be taken as 1.0 as the radiative heat transfer from these equipment is
generally negligible due to smaller operating temperatures. When the equipment are run by
electric motors which are also kept inside the conditioned space, then the efficiency of the
electric motor must be taken into account. Though the estimation of cooling load due to
appliance and equipment appears to be simple as given by the equations, a large amount of
uncertainty is introduced on account of the usage factor and the difference between rated
(nameplate) power consumption at full loads and actual power consumption at part loads.
Estimation using nameplate power input may lead to overestimation of the loads, if the
equipment operates at part load conditions most of the time.
If the conditioned space is used for storing products (e.g. cold storage) or for carrying
out certain processes, then the sensible and latent heat released by these specific products and
or the processes must be added to the internal cooling loads. The sensible and latent heat
release rate of a wide variety of live and dead products commonly stored in cold storages are
available in air conditioning and refrigeration handbooks. Using these tables, one can
estimate the required cooling capacity of cold storages.
Thus using the above equations one can estimate the sensible (Qs,r), latent (Ql,r) and
total cooling load (Qt,r) on the buildings. Since the load due to sunlit surfaces varies as a
function of solar time, it is preferable to calculate the cooling loads at different solar times
and choose the maximum load for estimating the system capacity. From the sensible and total
cooling loads one can calculate the Room Sensible Heat Factor (RSHF) for the building. As
discussed in an earlier chapter, from the RSHF value and the required indoor conditions one
can draw the RSHF line on the psychrometric chart and fix the condition of the supply air.

157
Lecture 5.9: Grills and registers_ Fans and blowers - Heat sources
A decorative covering for an outlet or intake is known as grill and a grill provided with
damper is known as Register. The proper air supply to an air-conditioned room is made
through grills or registers. The grill or register can be located in the floor, high side of the
wall or in the ceiling. The essential requirement of the supply point is that the air stream
coming out should not strike the occupants before it has lost its high velocity. The upper wall
side location is more suitable as it has no chance to strike an occupant anywhere in the room
and an additional advantage is that it does not interfere with the furniture arrangement. The
disadvantage of wall register is the requirement of high velocity for the spaces that have great
width.
Floor to ceiling air distribution provides excellent heating conditions (winter air-conditioner)
to the large spaces like assemblies, theatres and auditoriums having crowded occupancies.
The major disadvantage of this is the picking up of dust by the air as the air flows over and
around the floor. Ceiling to floor air distribution is excellent for cooling large spaces as
theatre· and auditoriums. The advantage of this system is the natural downwards flow of cool
air. This system eliminates the draft conditions. Ceiling to ceiling openings provide excellent
flow of air for heating as well as for cooling applications. Inlet and outlet openings are
constructed as one unit. Inlet air opening to the sides and outlet opening near the centre of
ceiling causes the air to spread to all sections of the room.
Design of the Grill. It is necessary to consider the required throw and the velocity of air
through
the grill to determine the size of the grill for a particular room. A single supply opening in the
upper portion of the wall side is not desirable as the noise from the excessive velocity is
objectionable· The formcalculating the *throw is given by the expression.

The values of the constant Kare different for the different vane angles.
The throw of the air should not be more than the room side which is parallel to the flow of air
we
air supply is made from the-grill located in the side-wall, otherwise cold air may turn
downvard and the floor causing uncomfortable drafts.

Factors Affecting Grill Performance

1. Vanes. Straight vanes produce an included angle of 14 to 24°. Covering vanes spread the
air similar to the straight vanes. The throw in this case is greater than the straight vanes.
Diverging vanes produce an angular spread which has a marked effect on the direction of air
flow and distance travelled. The throw in this case is smaller than the other two. Two grills

158
of same area will have smaller induction rate and longer throw than the single grill of same
area.
2. Outlet Types. The type of the outlets may be perforated one, fixed bar grill. adjustable-bar
grill or slotted outlets as per the requirements. The type of the outlet used has prominent
effect on the induction of the air and throw achieved.

Lecture 5.10: Different heat pump circuits.

The heat pump is a heat multiplier. It takes warm air and makes it hot air. This is done by
compressing the air and increasing its temperature. Heat pumps have received more attention
since the fuel embargo of 1974. Energy conservation has become a more important concern
for everyone. If a device can be made to take heat from the air and heat a home or
commercial building, it is very useful to many people.

The heat pump can take the heat generated by a refrigeration unit and use it to heat a house or
room. Most take the heat from outside the home and move it indoors (see Figure). This unit
can be used to air-condition the house in the summer and heat it in the winter by taking the
heat from the outside air and moving it inside.
On mild-temperature heating days, the heat pump handles all heating needs. When the
outdoor temperature reaches the balance point of the home—that is, when the heat loss is
equal to the heat-pump heating capacity—the two-stage indoor thermostat activates the
furnace (a secondary heat source, in most cases electric heating elements). As soon as the
furnace is turned on, a heat relay de-energizes the heat pump.
When the second-stage (furnace) need is satisfied and the plenum temperature has cooled to
below 90 and 100 degrees F, the heat-pump relay turns the heat pump back on and controls
the conditioned space until the second-stage operation is required again. Figure shows the
heat-pump unit. The optional electric- heat unit shown in Figure 13.13 is added in geographic
locations where needed. This particular unit can provide 23,000 to 56,000 Btus per hour
(Btuh) and up to 112,700 Btu with the addition of electric heat.

159
If the outdoor temperature drops below the setting of the low-temperature compressor
monitor, the control shuts off the heat pump completely, and the furnace handles all the
heating needs.
During the defrost cycle, the heat pump switches from heating to cooling. To prevent cool air
from being circulated in the house when heating is needed, the control automatically turns on
the furnace to compensate for the heat-pump defrost cycle (see Figure 13.14). When supply
air temperature climbs above 110 to 120 degrees F, the defrost-limit control turns off the
furnace and keeps indoor air from getting too warm.
If, after a defrost cycle, the air downstream of the coil rises above 115 degrees F, the closing
point of the heat-pump relay, the compressor will stop until the heat exchanger has cooled
down to 90 to 100 degrees F, as it does during normal cycling operation between furnace and
heat pump.
During summer cooling, the heat pump works as a normal split system, using the furnace
blower as the primary air mover (see Figure).
In a straight heat-pump/supplementary-electric-heater application, at least one outdoor
thermostat is re- quired to cycle the heaters as the outdoor temperature drops. In the system
shown here, the indoor thermo- stat controls the supplemental heat source (furnace). The
outdoor thermostat is not required.
Since the furnace is serving as the secondary heat source, the system does not require the
home rewiring usually associated with supplemental electric strip heating.
Special Requirements of Heat-Pump Systems
The installation, maintenance, and operating efficiency of the heat-pump system are like
those of no other comfort system. A heat-pump system requires the same air quantity for
heating and cooling. Because of this, the air-moving capability of an existing furnace is
extremely important. It should be carefully checked be- fore a heat pump is added. Heating
and load calculations must be accurate. System design and installation must be precise and
according to the manufacturer‘s suggestions.

160
161
162
Special Requirements of Heat-Pump Systems
The installation, maintenance, and operating efficiency of the heat-pump system are like
those of no other comfort system. A heat-pump system requires the same air quantity for
heating and cooling. Because of this, the air-moving capability of an existing furnace is
extremely important. It should be carefully checked before a heat pump is added. Heating and
load calculations must be accurate. System design and installation must be precise and
according to the manufacturer‘s suggestions.

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The air-distribution system and diffuser location are equally important. Supply ducts must be
properly sized and insulated. Adequate return air is also required. Heating-supply air is cooler
than with other systems. This is quite noticeable to homeowners accustomed to gas or oil
heat. This makes diffuser location and system balancing critical.
Heat-Pump Combinations
There are four ways to describe the heat-pump methods of transporting heat into the house:
• Air-to-air: the most common method and the type of system previously described
• Air-to-water: this method uses two different types of heat exchangers: warmed refrigerant
flows through pipes to a heat exchanger in the boiler, and heated water flows into radiators
located within the heated space
• Water-to-water: this type uses two water-to-refrigerant heat exchangers; heat is taken from
the water source (well water, lakes, or the sea) and passed on by the refrigerant to the water
used for heating; the reverse takes place in the cooling system
• Water-to-air: well water furnishes the heat by warming the refrigerant in the heat-exchanger
coil; the refrigerant, compressed, flows to the top of the unit, where a fan blows air past the
heat exchanger
Each type of heat pump has its advantages and disadvantages. The electrical connections and
controls are used to do the job properly. Before attempting to work on this type of equipment,
make sure you have a complete schematic of the electrical wiring and know all the
component parts of the system.

Test Questions

a. Fill in the blanks type of questions

1. All specific psychrometric properties of moist air are based on _________. (unit
mass of moist air)
2. The minimum number of independent properties to be specified for fixing the state of
moist air is ______________. (three)
3. When the enthalpy of air is equal to the enthalpy of saturated air at the wetted surface
temperature, then the wet bulb temperature of air ______________. (remains
constant)
4. During cooling and humidification process, the enthalpy of air may
_________________ depending upon the temperature of the wet surface. (increase,
decrease or remain constant)
5. The BPF of the coil ______________ as the fin pitch increases. (increases)
6. Sensible heat factor is _______ for air conditioning plants operating in humid
climates. (low)
7. During sensible cooling of air, dry and wet bulb temperatures ______________ but
dew point temperature remains constant. (decrease)
8. During sensible cooling of air, both dry bulb and wet bulb temperatures
____________. (decrease)
9. _________ dry bulb temperatures and ____________ moisture content are
recommended for winter. (Lower, Lower)

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10. To maintain thermal comfort, the DBT of air should be _______________ as the
temperature of the surrounding surfaces decrease. (increased)

b. Multiple choice questions

1. Which of the following statements are TRUE?

a) Straight-line law is applicable to any fluid-air mixtures
b) Straight-line law is applicable to any water-air mixtures only
c) Straight-line holds good as long as the Prandtl number is close to unity
d) Straight-line holds good as long as the Lewis number is close to unity
Ans.: b) and d)

2. Which of the following statements are TRUE?

a) When the dry bulb temperature is equal to dew point temperature, the relative
humidity of air-water mixture is 1.0
b) All specific psychrometric properties of moist air are based on unit mass of water
vapor
c) All specific psychrometric properties of moist air are based on unit mass of dry air
d) All specific psychrometric properties of moist air are based on unit mass of moist
air
Ans.: a) and d)

3. Which of the following statements are TRUE?

a) Thermodynamic WBT is a property of moist air, while WBT as measured by wet
bulb thermometer is not a property
b) Both the thermodynamic WBT and WBT as measured by wet bulb thermometer
are properties of moist air
c) Under no circumstances, dry bulb and wet bulb temperatures are equal
d) Wet bulb temperature is always lower than dry bulb temperature, but higher than
dew point temperature
Ans.: a)

4. Which of the following statements are TRUE?

a) The metabolic rate depends mainly on age of the human being
b) The metabolic rate depends mainly on the activity level of the human being
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 c) The metabolic rate depends mainly on the sex of the human being
d) All of the above
Ans: b)
5. Which of the following statements are TRUE?
a) To maintain thermal comfort, the DBT of air should be increased as its moisture
content increases
b) To maintain thermal comfort, the DBT of air should be decreased as air velocity
increases
c) To maintain thermal comfort, the DBT of air should be increased as the
temperature of the surrounding surfaces decrease
d) All of the above
Ans: c)

6. Which of the following statements are TRUE?

a) Surrounding air velocity affects convective heat transfer from the body only
b) Surrounding air velocity affects evaporative heat transfer from the body only
c) Surrounding air velocity affects both convective and evaporative heat transfers
from the body
d) Moisture content of the air affects both convective and evaporative heat transfers
from the body
Ans: c)

7. Which of the following statements are TRUE?

a) Steady state methods are justified in case of heating load calculations as the outside
temperatures during winter are normally very low
b) Steady state methods are justified in case of heating load calculations as the peak
load normally occurs before sunrise
c) Steady state methods are justified in case of heating load calculations as the outside
temperature variation is normally low during winter months
d) Neglecting internal heat sources while calculating heating loads underestimates the
required capacity
Ans.: b) and c)

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 8. Which of the following statements are TRUE?
a) An all year air conditioning system has to be switched from winter mode to
summer mode as the outdoor temperature exceeds the outdoor temperature at balance
point
b) An all year air conditioning system has to be switched from summer mode to
winter mode as the outdoor temperature exceeds the outdoor temperature at balance
point
c) The outdoor temperature at balance point increases as the amount of insulation
increases
d) The outdoor temperature at balance point decreases as the amount of insulation
increases
Ans.: a) and d)

9. Which of the following statements are TRUE?

a) During sensible cooling of air, both dry bulb and wet bulb temperatures decrease
b) During sensible cooling of air, dry bulb temperature decreases but wet bulb
temperature remains constant
c) During sensible cooling of air, dry and wet bulb temperatures decrease but dew
point temperature remains constant
d) During sensible cooling of air, dry bulb, wet bulb and dew point temperatures
decrease
Ans.: a) and c)

10. Which of the following statements are TRUE?

a) The sensible heat factor for a sensible heating process is 1.0
b) The sensible heat factor for a sensible cooling process is 0.0
c) Sensible heat factor always lies between 0.0 and 1.0
d) Sensible heat factor is low for air conditioning plants operating in humid climates

c. True or False questions

1. The metabolic rate depends mainly on age of the human being (F)
2. Straight-line law is applicable to any fluid-air mixtures (F)

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 3. Straight-line law is applicable to any water-air mixtures only (T)
4. Straight-line holds good as long as the Prandtl number is close to unity (T)
5. Straight-line holds good as long as the Lewis number is close to unity (F)
6. Steady state methods are justified in case of heating load calculations as the peak load
normally occurs before sunrise. (T)
7. The sensible heat factor for a sensible heating process is 1.0. (T)
8. During sensible cooling of air, dry bulb temperature decreases but wet bulb
temperature remains constant. (F)
9. Wet bulb temperature is always lower than dry bulb temperature, but higher than dew
point temperature. (F)
10. To maintain thermal comfort, the DBT of air should be increased as the temperature
of the surrounding surfaces decrease. (T)

Review Questions

a. Objective type of questions (Very short notes)

1. What do you understand by the term ‗psychometry‘.
2 .Define the following:
(a) specific humidity
(b) Absolute humidity
(c) Relative humidity
(d) Dew point temperature
3. What is a sling psychrometer ? Make a neat sketch and explain its use.
4. What is the difference between ‗wet bulb temp‘ and ‗thermodynamic wet bulb
temperature?
5.Prove that enthalpy of the humid air remains constant along a wet bulb temperature line
on the psychrometric chart.
6. When is dehumidification of air necessary and how it is achieved.
7. Write a short note on by-pass factor for cooling processes.
8. Define sensible heat factor.
9. Show dehumidification of moist air by cooling on the psychrometric chart.
10. Show adiabatic mixing of two air streams on the psychrometric chart.

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b. Analytical type questions
1. Establish the following expression for air-vapour mixture
Specific humidity W = 0.622 x Pv / Pb – Pv where Pv = partial pressure of water
vapour, and
Pb = Barometric pressure
2. What is fog ? Show on the psychrometric chart when two air streams yield fogged state
of air.
3. With the help of psychrometric chart, explain the following processes:
(a) Sensible heating and sensible cooling process.
(b) Heating and dehumidification process.
(c) Cooling and humidification process.
(d) Cooling and dehumidification process.
4. Prove that the partial pressure of water vapour in the atmospheric air remains constant
as long as the specific humidity remains constant.
5. When is dehumidification of air necessary and how it is achieved.
6. Write a short-note on by-pass factor for cooling coils.
7. An unsaturated air stream is undergoing adiabatic saturation process and the outgoing
air is saturated. Draw the system schematically. Represent the process on T-s and
psychrometric charts .Find for the incoming stream of air, humidity ratio from the
adiabatic saturation equation obtained from heat balance.
8. Prove that the partial pressure of water vapour in the atmospheric air remains constant
as long as the specific humidity remains constant.
9. Explain adiabatic mixing of two air streams. For calculating mass ratio.

c. Essay type Questions

1. Derive psychrometric relations.
2. Derive an expression for calculation of enthalpy of moist air.
3.Explain clearly ―Thermodynamic wet bulb temp.‖
4. Eplain sensible heating and sensible cooling
5. Explain bypass factor of heating and cooling coils and derive an expression for both the
cases and also efficiency of heating and cooling coils.
6. Explain humidification and dehumidification and various methods for obtaining both.

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7. What is meant by ―Adiabatic mixing of two air streams‖ ? Derive an expression for
that.
8. What is human comfort and state the factors affecting human comfort.
9. What is comfort air-conditioning.How does it differ it from normal air-
conditioning.Write the various air-conditioning methods.
10. Explain sensible heat factor and room sensible heat factor and grand sensible sensible
heat factor. And effective room sensible heat factor.

d. Problems

1. The readings from a sling psychrometer are as follows :

Dry bulb temp. =30C Wet bulb temp = 20C, Barometric reading =740 mm of Hg.
Using steam tables , determine (a) Dew point temp. (b) Relative humidity (c) Specific
humidity (d)Degree of saturation (e) vapour density (f) enthalpy of mixture per kg of dry
air.
2. On a particular day, the atmospheric air was found to have a dry bulb temp. of 30C
and wet bulb temp. of 18C .The barometric pressure was observed to be 756 mm of Hg.
Using the tables of psychrometric properties of air, determine the relative humidity, the
specific humidity, the dew point temp. the enthalpy of air per kg of dry air and the
volume of mixture per kg of dry air.

3. The humidity ratio of atmospheric air at 28C dry bulb temp. and 760 mm of mercury is
0.016kg / kg of dry air. Determine (a) partial pressure of water vapour (b) relative
humidity (c) dew point temp. (d) specific enthalpy (e) vapour density.

4. A room 7mx4mx4m is occupied by an air-water vapour mixture at 38C.The

atmospheric pressure is 1 bar and the relative humidity is 70% . Determine the humidity
ratio, dew point, mass of dry air and mass of water vapour .If the mixture of air-water
vapour is further cooled at constant pressure until the temp. is 10C find the amount of
water vapour condensed.

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 5. Atmospheric air at 0.965 bar enters the adiabatic saturator.The wet bulb temp. is 20C
and dry bulb temp. is 31C during adiabatic saturationprocess. Determine (a) humidity
ratio of the entering air (b) vapour pressure and relative humidity (c) dew point temp.

6. In a heating application, moist air enters a steam heating coil at 10C, 50% relative
humidity and leaves at 30C. Determine the sensible heat transfer, if mass flow rate of air
is 100 kg of dry air per second. Also determine the steam mass flow rate if steam enters
saturated at 100C and condensate leaves T 80C.

7. A quantity of air having a volume of 300 m3 at 30C wet bulb temp. is heated to 40C
dry bulb temp. Estimate the amount of heat added , final relative humidity and wet bulb
temp. The air pressure is 1.01325 bar.

8. Atmospheric air wth dry bulb temp of 28C and a wet bulb temp of 17C without
changing its moisture content. Find original relative humidity, final relative humidity,
final wet bulb temp.

9. The moist air is heated by steam condensing inside the tubes of a heating coil . The part
of air passes through the coil and part is by-passed around the coil. The barometric
pressure is 1 bar . determine (a) the air per minute which by-pass the coil (b) the heat
added by the coil.

10. The atmospheric air at 60 mm of Hg, dry bulb temp. 15C and wet bulb temp 11C
enters a heating coil whose temp.is 41C. Assume by-pass factor of heating coil as 0.5 ,
determine dry bulb temp, wet bulb temp., and relative humidity of the air leaving the coil
.Also determine the sensible heat added to the air per kg of dry air.

h. Case study
Estimate cooling loadfor an auditorium of capacity 400 persons

Skill Building Exercises/Assignments

Prepare a report on existing air conditioning systems used in various types of aircrafts
and also mention the basis on which respective conditioning system is choosed.

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Previous Questions (Asked by JNTUK from the concerned Unit)

1. What is the difference between wet bulb temperature and thermodynamic wet bulb
temperature?

2. Explain the difference between comfort air-conditioning and industrial air-

conditioning.
3. Define the term `` effective temperature `` and explain its importance in air
conditioning system. Describe the factors which affect effective temperature.
4. Define room sensible heat factor. How room sensible heat factor line is drawn on the
psychrometric chart?
5. Explain the procedure to draw a grand sensible heat factor line on a psychrometric
chart. What do you understand by effective room sensible heat factor?
6. Why ventilation is required? Explain why different ventilation standards for different
purposes are recommended?
7. With the help of psychrometric chart explain the following process and give the
important characteristic features of adiabatic cooling and humidification process.
8. Explain the concept of effective sensible heat factor for room to be air conditioned.
How is it useful to find the ADP for fixed room design condition?
9. Define the `` human comfort `` and explain the factors which affect human comfort.
10. A refrigerating plant of 28 kW capacity has its evaporation temperature -80C and
condenser temperature of 300C. The refrigerant, R-12 is sub-cooled 50C before entering
the expansion valve and the vapour is superheated 60C before leaving the evaporator coil.
The compression of the refrigerant in the compressor is isentropic. If there is a suction
pressure drop of 0.2 bar through the valves; and discharge pressure drop through the
valve of 0.1 bar, determine the C.O.P. of the plant, theoretical piston displacements/min
and the heat removed in the condenser. Solve the problem with the help of P-h chart. Give
also a diagrammatic sketch of this cycle on the T-s chart.

GATE Questions
1. Dew point temperature is the temperature at which condensation begins when the air
is cooled at constant.
(a) volume (b) entropy (c) pressure (d) enthalpy

2. For air with a relative humidity of 80%

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(a) the dry bulb temperature is less than the wet bulb temperature
(b) the dew point temperature is less than wet bulb temperature
(c) the dew point and wet bulb temperatures are equal
(d) the dry bulb and dew point temperatures are equal

3. The statements concern Psychrometric chart.

1. Constant relative humidity lines are uphill straight lines to the right
2. Constant wet bulb temperature lines are downhill straight lines to the right
3. Constant specific volume lines are downhill straight lines to the right
4. Constant enthalpy lines are coincident with constant wet bulb temperature lines
Which of the statements are correct?
(a) 2 and 3 (b) 1 and 2 (c) 1 and 3 (d) 2 and 4

4. When atmospheric air is heated at constant pressure, then which one is not correct.
(a) humidity ratio does not change
(b) relative humidity increases
(c) dew point temperature does not change
(d) wet bulb temperature increases

5. During chemical dehumidification process of air

(a) dry bulb temperature and specific humidity decrease
(b) dry bulb temperature increases and specific humidity decreases
(c) dry bulb temperature decreases and specific humidity increases
(d) dry bulb temperature and specific humidity increase

6. Water at 42°C is sprayed into a stream of air at atmospheric pressure, dry bulb
temperature of 40oC and a wet bulb temperature of 20oC. The air leaving the spray
humidifier is not saturated. Which of the following statements is true?
(a) Air gets cooled and humidified (b) air gets heated and humidified
(c) Air gets heated and dehumidified (d) Air gets cooled and dehumidified

7. Air (at atmospheric pressure) at a dry bulb temperature of 40°C and wet bulb
temperature of 20°C is humidified in an air washer operating with continuous water
recirculation. The wet bulb depression (i.e. the difference between the dry and wet bulb

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temperatures) at the exit is 25% of that at the inlet. The dry bulb temperature at the exit
of the air washer is closest to
(A) 100C (B) 200C (C) 250C (D) 300C

8. The statements concern Psychrometric chart

1. Constant relative humidity lines are uphill straight lines to the right
2. Constant wet bulb temperature lines are downhill straight lines to the right
3. Constant specific volume lines are downhill straight lines to the right
4. Constant enthalpy lines are coincident with constant wet bulb temperature lines
Which of the statements are correct?
(a) 2 and 3 (b) 1 and 2 (c) 1 and 3 (d) 2 and 4

9. In a sample of moist air at standard atmospheric pressure of 101.325 kPa and 26°C the
partial pressure of water vapour is 1.344 kPa. If the saturation pressure of water vapour
is 3.36 kPa at 26� C, then what are the humidity ratio and relative humidity of moist air
sample?
(a) 0.00836 and 1.32% (b) 0.00836 and 40%
(c) 0.01344 and 1.32% (d) 0.01344 and 40%

10. If the volume of moist air with 50% relative humidity is isothermally reduced to half
its original volume, then relative humidity of moist air becomes.
(a) 25 % (b) 60 % (c) 75 % (d) 100 %

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