This paper presents a study on different kinds of air conditioning systems in comparison to existing one to use through of the year. Mainly the system imparts all three regular weather conditions. Like hot and dry, hot and wet and cool and dry. For this the out let condition will be fix ed 25 °C dry bulb temperature (DBT) and 50% relative humidity. In the present paper, for maintaining room condition thermodynamic simulation is being done. For simulation we used excels solver software. If atmospheric condition like relative humidity of air is changed, the year round air conditioning equipments change own parameters such as volume of cellulose cooling pad of evaporating cooler, temperature of cooling coil in hot and dry as well as hot and wet weather conditions, temperature of preheat and reheat coil in cold and dry weather to maintain the room condition.

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THERMODYNAMIC SIMULATION OF YEAR ROUND AIR
CONDITIONING SYSTEM FOR VARIABLE RELATIVE
HUMIDITY OF ATMOSPHERIC AIR
Shankar Kumar1
, S.P.S. Rajput2
, Arvind Kumar3
1
Research scholar, Department of Mechanical Engineering,
Maulana Azad National Institute of Technology, Bhopal, 462051 – India,
2
Professor, Department of Mechanical Engineering,
Maulana Azad National Institute of Technology, Bhopal, 462051 – India,
3
Assistant Professor, Department of Mechanical Engineering,
Maulana Azad National Institute of Technology, Bhopal, 462051 – India,
ABSTRACT
This paper presents a study on different kinds of air conditioning systems in comparison to
existing one to use through of the year. Mainly the system imparts all three regular weather
conditions. Like hot and dry, hot and wet and cool and dry. For this the out let condition will be fixed
25℃ dry bulb temperature (DBT) and 50% relative humidity. In the present paper, for maintaining
room condition thermodynamic simulation is being done. For simulation we used excels solver
software. If atmospheric condition like relative humidity of air is changed, the year round air
conditioning equipments change own parameters such as volume of cellulose cooling pad of
evaporating cooler, temperature of cooling coil in hot and dry as well as hot and wet weather
conditions, temperature of preheat and reheat coil in cold and dry weather to maintain the room
condition.
Keywords: Heating Coil, Cooling Coil, Desiccant Wheel, Cellulose Cooling Pad, Supply Condition.
1. INTRODUCTION
Air conditioning system plays an important role in industry, infrastructure and domestic use.
For different weather condition different air conditioning systems are used. There is no such single
air conditioning system which gives constant output (comfort) throughout the whole year .The
existing air conditioning system works better only in the specified weather condition. While human
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body requires comfort throughout the year. The human body can be considered as thermal machine
with 20% thermal efficiency. The remaining 80% must be disposed off from body to the
surrounding which depends upon temperature and humidity of surrounding otherwise the
accumulation of heat result and causes discomforts in the form of body pain, heat cramp or shocks,
heat stroke, heat exhaustion. There are following psychometric process for air
conditioning.(a)Cooling and humidification(b)Cooling and dehumidification.(c)Heating and
humidification(d)Heating and dehumidification. The year round air conditioning system have
following components.(1) Fan(2)Desiccant wheel (3)Cooling pad(4)Cooling coil(5)Heating coil.
Shankar Kumar et al [1] worked on the actual position of equipment used in year round air
conditioning system. They also studied about the parameters on which the system depends. Zainab
Hasson et al [2] presented an efficient methodology to design modified evaporative air –cooler for
winter air-conditioning in Baghdad city. The performance is reported in terms of effectiveness of
DEC, saturation efficiency of DEC, outlet temperature of air and cooling capacity. They reported a
full numerical model for the concurrent forecast of velocity, temperature and humidity of air flowing
in an air conditioning unit. Kulkarni and Rajput [3] studied about the theoretical performance
analysis of cooling pads of different materials for evaporative cooler. The material have been
considered, rigid cellulose, corrugated paper, corrugated high density polythene and aspen fiber .It
has been observed that the saturation efficiency decreases with increasing mass flow rate of air .It
also seen that material with higher wetted surface area gives higher saturation efficiency. Fatemeh et
al[4]shown the modeling of a desiccant wheel used for dehumidifying the ventilation air of an air-
conditioning system. By the numerical method, the performance of an adiabatic rotary
dehumidification is parametrically studied and the optimums rotational speed is determined by
examine the outlet adsorption –side humidity profiles.Kang and Maclain-cross [5] showed that
dehumidification is the key factor of desiccant cooling system and the cooling COP can be
significant enhanced by enhancing the performance of component.J.J. Jurinak et al and F.E.Nia et
[6] premeditated that, the moisture can be removed from the desiccant by heating it to temperature
around 60°C-90°C and revealing it to a regenerative air stream. Dowdy, J.A.,
Karbash,N.S.[7]obtained the heat and mass transfer coefficient by experimentally for the evaporative
cooling process through various thicknesses of rigid impregnated cellulose evaporative medium.
E.V.Gomez et al [8] studied about the comparison of high grade energy required in air-conditioning
system and evaporating cooling system. The weather data for analysis has taken from IMD Bhopal
[13] for the year 2013.
Nomenclature
T Dry bulb temperature of air
H Enthalpy of air (kj/kg)
Tw Wet bulb temperature
W specific humidity (g/kg)
R Ratio of ambient air mass to Re-circulated air mass
RTH Total heat inter from atmosphere to room (kW)
RTHO Total heat out from room to atmosphere (kW)
RSHF Room sensible heat factor
RSH Room sensible heat
RSHO Room sensible heat out
CLC Cooling load Capacity (TR)
Cpa Specific heat of air at constant pressure (j/kg-k)
Cpv Specific heat of vapour at constant pressure (j/kg-k)
V Volume of cooling pad
B Bi-pass factor

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Tc Cooling coil temperature
Trh Temperature of re-heat coil
Tph Temperature of pre-heat coil
Trd Regeneration temperature of desiccant wheel
P % mixing of re-circulated air to atmospheric air
Mm Mass flow rate of air in duct after mixing (kg/s) of ambient air and re-circulating air
RDW Radius of desiccant wheel
dt thickness of desiccant coating(mm)
Dh Hydraulic diameter of a channel (mm)
N Desiccant wheel speed
Greek letters
φ Relative humidity of air(%)
ρ Density of air(kg/m³)
ɳ Saturation efficiency of evaporative cooler (%)
ɛ Efficiency of desiccant wheel.
Subscripts
Ph pre-heating coil
Rh re-heating coil
c cooling coil
phi pre- heating coil inlet
pho pre –heating coil outlet
di desiccant wheel inlet
do desiccant wheel outlet
ei evaporative cooler inlet
eo evaporative cooler outlet
ci cooling coil inlet
co cooling coil outlet
rhi re-heat coil inlet
rho re-heat coil outlet
a ambient
s supply
r room
m mixed air
2. METHODOLOGY
Following assumptions have been taken for analysis.
2.1. Assumption
(1)Axial heat conduction and water vapour diffusion in the air is negligible.
(2)Axial molecular diffusion within the desiccant is negligible.
(3)The channels are assumed adiabatic and impermeable.
(4)The heat and mass transfer coefficient are constant.
(5)The adsorption heat per kilogram of adsorbed water is constant.
(6)The wet bulb temperature before and after the evaporative cooler will be constant.
(7)The cooling coil and heating coil be only used for sensible cooling and sensible heating.
(8)The mass flow rate (or air velocity) will be same throughout the duct.

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2.2 Equations
2.2.1 Equation of specific humidity for mixed air.
=
×
(1)
2.2.2 Equation of DBT for mixed air [9].
=
( . × )
( . . × )
(2)
2.2.3 Equation of pre-heat coil temperature [9].
=
( × )
( )
(3)
2.2.4 Equation of enthalpy for supply air [9].
(2.2.4.1)For cool and dry weather condition.
ℎ =
(! × "#)
!
(4)
(2.2.4.2) For hot and dry as well as hot and wet weather condition.
ℎ =
(! × ")
!
=
(! × $. % ×&'&)
!
(5)
2.2.5 Equation of DBT for supply air [9].
(2.2.5.1)For cool and dry weather condition.
= ( +
( *"#)
( . + $ +% ,×! )
(6)
(2.2.5.2) For hot and dry as well as hot and wet weather condition.
= ( −
*"
( . + $ +% ,×! )
(7)
2.2.6 Equation for saturation efficiency of evaporative cooler [3, 10].
. =
( /0 / )
( /0 1/0)
(8)
2.2.7 Equation for volume of cooling pad.[3]
2 =
(& /0×& 3)×! ×45( 6)
( 7×$, )
(9)
2.2.8 Equation of re-heats coil temperature.
( =
( × / 8)
( )
(10)
2.2.9 Equation for percentage mixing of re-circulated air with atmospheric air.
9 =
( :/ <)
=
( )
(11)
2.2.10 Equation for temperature of cooling coil.
= =
( 8 × / )
( )
(12)

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2.2.11 Area of desiccant wheel in which mixed air is passed
> =
?×( @ )<
(13)
2.2.12 Equation for velocity of air
A =
!
(B ×C)
(14)
2.2.13 Equation for outlet DBT of desiccant wheel [11].
DE = F₁(H) F₂( ) F₃(KL)F₄( (D)F₅ ( )F₆(P )F₇(A) (15)
Where
P₁ (N)= −0.0002H + 0.0112H + 0.420
F₂( ) = −0.0001( )² + 0.0275 + 0.7993
F₃(KL) = −18.79(KL)² + 7.92KL + 1.75
p4( (D)=-0.0004( (D)²+0.1255 (D+0.6757
p5(Wm)=594.48(Wm/1000)²+26.76(Wm/1000)+3.79
p6(P )=-0.039(P )³+0.026(P )²+0.603P +0.0912
p7(U)=-0.060U+0.7973
2.2.14 Equation for efficiency of desiccant wheel [10].
ɛ = k1(N)k2( ) k3(dt)k4( (D)k5 ( )k6(P )k7(U) (16)
Where
k1(N) =-0.0001N2
+0.0042N+0.4474
k2( ) =-0.0001( )2
-0.0031 +0.8353
k3(dt) =-21.67(dt)2
+6.93dt+1.34
k4( (D) =-0.0001( (D)2
+0.0355 (Trd) -0.4924
k5( ) =592.77( /1000)²-41.23( /1000)+1.283
k6(P ) =-0.0572(P )3
+0.0933(P )2
+0.6139P -0.0922
k7(U) =-0.0611U+0.8376
2.2.15 Equation for specific humidity of outlet air for desiccant wheel [11]
DE = ( − ɛ ) (17)
3. SYSTEM DESCRIPTION
The positions of equipments used in year round air conditioning system are shown in the
diagram of year round air conditioning system(Fig.1).The analysis is done on the concept that,
supply air condition is depends on the room conditions, mass flow rate, CLC, and RSHF. We take
the room conditions be constant. So the supply condition will be depends on mass flow rate, CLC
and RSHF.

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Fig.1. Diagram of year round air conditioning system
4. RESULTS AND DISCUSSION
The result is found for the variation in relative humidity of ambient air for all the weather
conditions. The results and their discussions are given below.
4.1: Hot and dry weather condition
Φa (%) 10 11 12 13 14 15 16 17
Tm(°C) 32.4713 32.4895 32.5097 32.5289 32.5511 32.5673 32.5924 32.6146
Twm(°C) 18.9 19.09 19.28 19.47 19.66 19.84 20.03 20.23
Teo(°C) 29.23 29.86 30.5 31.13 31.77 32.34 32.97 33.67
Tc(°C) 21.1861 21.0286 20.8686 20.7111 20.5511 20.4086 20.251 20.0761
Ts(°C) 22.7949 22.7949 22.7949 22.7949 22.7949 22.7949 22.794 22.7949
Wm(g/kg) 9.1176 9.372 9.6272 9.8828 10.1388 10.3956 10.652 10.908
Weo(g/kg) 10.33 10.33 10.33 10.33 10.33 10.33 10.33 10.33
Ws(g/kg) 10.33 10.33 10.33 10.33 10.33 10.33 10.33 10.33
η (%) 23.8836 19.6242 15.1911 10.7125 6.05931 1.78623 -3.0053 -8.5217
V(cm³) 524.63 431.036 333.486 235.104 132.8832 39.23918 -65.977 -186.39
Table1: variation of output parameter w.r.t. relative humidity of ambient air at 44.8°C DBT
of ambient air, altitude be 523m , 25°C DBT and 50% relative humidity of room,60 % mixing of re-
circulated air , 0.9 RSHF ,0.2 BPF factor and 1.25kg/s mass flow rate.

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Variation in Tm, Twm, Teo, Tc, and Ts w.r.t. φa
Figure (2) indicates variation in temperature w.r.t relative humidity of ambient air
respectively. The inlet temperature of evaporative cooler or the mixed air temperature increases with
the increase in value of DBT of ambient air. The value of Twm also increases with increase in relative
humidity of ambient air. The result shows that the supply condition be constant because the supply
condition is depends on the mass flow rate ,room conditions, RTH or cooling load and RSHF value
and these parameters are taken to be constant . The cooling coil temperature decreases, since the
supply air DBT is constant and the inlet of cooling coil that is outlet of evaporative cooler DBT of air
increases and the bypass factor be constant. The value of cooling pad decreases with increase in DBT
of ambient air (shown in table 2), since mixed air specific humidity or inlet air specific humidity of
evaporative cooler increases and outlet specific humidity is constant and which is only possible when
volume of the cooling pad decreases. Because volume or thickness decreases then the air has lesser
contact time with water for same mass flow rate. The increase in relative humidity only be allowed
up to that limit at which the specific humidity of inlet air of evaporative cooler will be less than
outlet air specific humidity. If both will be same then at that condition the volume of evaporative
cooler becomes zero. That is to say no requirement of evaporative cooler in this case .If inlet specific
humidity of air is more than outlet specific humidity then other system be applicable.
4.2: Hot and wet weather condition
φₐ(%) 91 92 93 94 95 96 97
Tm(°C) 27.59778 27.6037 27.60962 27.61556 27.62152 27.62744 27.63338
Tdo(°C) 59.60345 59.68513 59.76714 59.84947 59.93215 60.01511 60.09842
Twdo(°C) 24.02 24.09 24.16 24.23 24.3 24.37 24.44
Teo(°C) 47.63 47.92 48.21 48.45 48.74 49.03 49.32
Tc(°C) 16.58614 16.51364 16.44114 16.38114 16.30864 16.23614 16.16364
Ts(°C) 22.79491 22.79491 22.79491 22.79491 22.79491 22.79491 22.79491
Wm(g/kg) 16.86454 16.98519 17.10595 17.22682 17.3478 17.46891 17.59013
Wdo(g/kg) 5.478724 5.557505 5.636652 5.716162 5.796026 5.876248 5.956815
Weo(g/kg) 10.33 10.33 10.33 10.33 10.33 10.33 10.33
Ws(g/kg) 10.33 10.33 10.33 10.33 10.33 10.33 10.33
ɛ 0.675133 0.672803 0.670486 0.668182 0.665893 0.663617 0.661355
η(%) 33.6489 33.0526 32.4574 32.0035 31.4103 30.818 30.2269
V(cm³) 491.1157 486.7043 482.1564 480.0388 475.2284 470.2816 465.1998
15
17
19
21
23
25
27
29
31
33
35
9 11 13 15 17 19
Temperature(°C)
φₐ
Figure:2
Tm
Twm
Teo
Tc
Ts

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Table2: variation of output parameter w.r.t. relative humidity of ambient air at 30°C DBT of
ambient air, 523m of altitude , 25°C DBT and 50% relative humidity of air conditioned room ,
60% mixing of re-circulated air, 1TR CLC Value, 0.9 RSHF and 0.2 BPF, 1.25kg/s mass flow
rate,15 rph and 90°C of regeneration temperature of desiccant wheel.
Variation in Tm, Tdo, Twdo, Teo, Tc, and Ts w.r.t. φa
Figure (3) indicates variation in temperature w.r.t relative humidity of ambient air
respectively. The inlet temperature of desiccant wheel or mixed air temperature increases with the
increase in relative humidity of ambient air. The value of Twdo and Tdo also increases with increase in
relative humidity of ambient air. The result shows that the supply condition be constant because the
supply condition is depends on the mass flow rate ,Room conditions, RTH or cooling load and
RSHF value and these parameters are taken to be constant. The cooling coil temperature decreases,
since the supply air DBT is constant and the inlet of cooling coil that is outlet of evaporative cooler
DBT of air increases and the bypass factor be constant. From table (2), the value of specific humidity
of outlet air of evaporative cooler is constant, because the specific humidity of supply air is constant
and it is equal to specific humidity of outlet air of evaporative cooler. The value of effectiveness of
desiccant wheel decreases with increase in relative humidity of ambient air. Because at constant
mass flow rate and regeneration temperature ,if more humidified air pass through the desiccant wheel
then less absorption phenomena is done. The value of cooling pad decreases with increase in relative
humidity of ambient air. The outlet air specific humidity of desiccant wheel or inlet air specific
humidity of evaporative cooler increases and outlet specific humidity of evaporating cooler is
constant and which is only possible when volume of the cooling pad decreases. Because as volume
or thickness is decreases then the air has lesser contact time with water for same mass flow rate. The
increase in specific humidity only be allowed up to that limit at which the specific humidity of inlet
air of evaporative cooler will be less than outlet air specific humidity. If both will be same, at that
condition the volume of evaporative cooler becomes zero. That is to say no requirement of
evaporative cooler in this case .The saturation efficiency of evaporative cooler decreases as volume
of cooling pad decreases.
0
10
20
30
40
50
60
70
90 91 92 93 94 95 96 97 98 99 100
Temperature(°C)
φₐ
Figure:3
Tm
Tdo
Twdo
Teo
Tc
Ts

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4.3: Cold and dry weather condition
φₐ 50 55 60 65 70 75 80
Tm(°C) 17.95571 17.95687 17.95799 17.95922 17.96034 17.9615 17.96266
Tpho(°C) 26.83 26.55 26.22 25.88 25.55 25.22 24.88
Tph(°C) 29.04857 28.69828 28.2855 27.86019 27.44741 27.03463 26.60934
Teo(°C) 19 19 19 19 19 19 19
Trh(°C) 29.24315 29.24315 29.24315 29.24315 29.24315 29.24315 29.24315
Ts(°C) 27.19452 27.19452 27.19452 27.19452 27.19452 27.19452 27.19452
Wm(g/kg) 7.652604 7.785032 7.9176 8.05028 8.183156 8.316148 8.44928
Wpho(g/kg) 7.652604 7.785032 7.9176 8.05028 8.183156 8.316148 8.44928
Weo(g/kg) 10.88304 10.88304 10.88304 10.88304 10.88304 10.88304 10.88304
Ws(g/kg) 10.88304 10.88304 10.88304 10.88304 10.88304 10.88304 10.88304
η(%) 72.29917 71.56398 70.64579 69.63563 68.58639 67.46204 66.21622
V(cm³) 2098.883 2089.334 2069.007 2043.499 2016.003 1984.618 1947.074
Table 3: variation of output parameter w.r.t. relative humidity of ambient air at 7°C DBT of
ambient air, altitude be 523m , 25°C DBT and 50% relative humidity of room,60 % mixing of re-
circulated air , 0.9 RSHF ,0 .2 BPF ,3.5KW RTHO value , 1.25kg/s mass flow rate and 16°C wet
bulb temperature of outlet air for pre-heat coil.
Variation in Tm, Tph, Tpho, Teo, Trh, and Ts w.r.t. φa
Figure (4) indicates variation in temperature w.r.t relative humidity of ambient air. The inlet
temperature of pre-heating coil or the mixed air temperature increases with the increase in relative
humidity of ambient air. For fixed WBT of outlet air of pre-heating coil, Tpho decreases with increase
of relative humidity. The value of Tph decreases which depends on the value of Tpho and BPF of
heating coil. But the BPF is constant. The result shows that the supply condition be constant. The re-
heating coil temperature constant, since the supply air DBT is constant and the inlet of re-heating
coil that is outlet of evaporative cooler DBT of air is constant and the bypass factor be also constant..
The value of cooling pad decreases with increase in relative humidity of ambient air(shown in table
:3), since inlet air specific humidity of evaporative cooler increases and outlet specific humidity be
constant, which is only possible when volume of the cooling pad decreases. Because volume or
thickness decreases then the air has lesser contact time with water for same mass flow rate. The
value of specific humidity of outlet air of evaporative cooler is constant, because the specific
humidity of supply air is constant and it is equal to specific humidity of outlet air of evaporative
cooler. The increase in relative humidity only be allowed up to that limit at which the specific
humidity of inlet air of evaporative cooler will be less than outlet air specific humidity. If both will
be same than at that condition the volume of evaporative cooler becomes zero. That is to say no
requirement of evaporative cooler in this case
15
20
25
30
45 55 65 75 85
Temperature(°C)
φₐ
Figure:4
Tm
Tph
Tpho
Teo
Trh
Ts

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5. CONCLUSION
The analysis of year round air conditioning system depends mainly the performance of the
desiccant wheel, evaporative cooler, heating coil and cooling coil. The supply air condition remain
same corresponding to increase in relative humidity of ambient air because its affect is reduced by
changing in its own parameters of equipment used for the system like efficiency of desiccant wheel
,saturation efficiency of evaporative cooler, volume of cooling pad ,temperature of heating and
cooling coil. The value of DBT of supply air is 22.79°C for hot and dry as well as hot and wet
weather condition and 27.19°C for dry and cool weather condition. The specific humidity of supply
air remain constant with respect to increases in relative humidity of ambient air and its value is 10.33
g/kg for dry and hot in addition to hot and wet weather condition where as 10.88g/kg for cold and
dry weather condition at constant mass flow rate of 1.25kg/s. The volume of cellulose cooling pad
decreases 524.63 to -186.39 cm3
for hot and dry, 491.11 to 465.19 cm3
for hot and wet and 2098.88
to 1947.07 cm3
for cold and dry weather condition with increase in relative humidity of ambient air
(10 to 17%, 91 to 97%, and 50 to 80% respectively).The saturation efficiency also decreases 23.88 to
-8.52% for hot and dry, 33.64 to 30.22% for hot and wet and 72.29 to 66.21% for cold and dry
weather condition with increase in relative humidity of ambient air (10 to 17%, 91 to 97%, and 50 to
80% respectively).The negative or zero value of saturation efficiency or volume of cooling pad
means no requirement of evaporative cooler. The efficiency of desiccant wheel decreases (67.1 to
66.1%) with increase in relative humidity of ambient air (91 to 97%) in hot and wet weather
condition. The cooling coil temperature decreases with increase in relative humidity for hot and dry
as well as hot and wet weather condition. The value of pre-heating coil decreases where as
temperature of reheating coil remains constant with the increase in relative humidity for cold and dry
weather condition.
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