Effect of regeneration temperatures in the exergetic performances of the developed desiccant-evaporative air-conditioning system

Accepted Manuscript Effect of Regeneration Temperatures in the Exergetic Performances of the Developed Desiccant-Evaporative Air-Conditioning System Napoleon Enteria, Hiroshi Yoshino, Rie Takaki, Akashi Mochida, Akira Satake, Ryuichiro Yoshie PII:

S0140-7007(13)00209-0

DOI:

10.1016/j.ijrefrig.2013.08.005

Reference:

JIJR 2584

To appear in:

International Journal of Refrigeration

Received Date: 14 June 2012 Revised Date:

29 July 2013

Accepted Date: 2 August 2013

Please cite this article as: Enteria, N., Yoshino, H., Takaki, R., Mochida, A., Satake, A., Yoshie, R., Effect of Regeneration Temperatures in the Exergetic Performances of the Developed DesiccantEvaporative Air-Conditioning System, International Journal of Refrigeration (2013), doi: 10.1016/ j.ijrefrig.2013.08.005. This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting proof before it is published in its final form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.

ACCEPTED MANUSCRIPT

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Effect of Regeneration Temperatures in the Exergetic Performances of the Developed Desiccant-Evaporative Air-Conditioning System

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Napoleon Enteriaa,b*, Hiroshi Yoshinob, Rie Takakic, Akashi Mochidab, Akira Sataked, Ryuichiro Yoshiee

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a

Enteria Grün Energietechnik, Davao 8000, Philippines Faculty of Engineering, Tohoku University, Sendai 980-8579, Japan c Faculty of Engineering, Akita Prefectural University, Akita 010-0195, Japan d Technical Research Institute, Maeda Corporation, Tokyo 179-8914, Japan e Faculty of Engineering, Tokyo Polytechnic University, Atsugi 243-0297, Japan b

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11 Abstract

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The developed desiccant-evaporative air-conditioning system was evaluated using the exergetic method under controlled environmental conditions to determine the performances of the whole system and its components.

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Percentage contributions of exergy destruction of system components at different regeneration temperatures and reference temperatures were determined. Exergy destruction coefficient of different components at different regeneration and reference temperatures were presented. It was shown that exergetic performances varied with respect to the regeneration and reference temperatures. The exergetic performances based on thermal, electric, total exergy input, first definition and second definition efficiencies were shown.

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Based on the results, reference and regeneration temperatures affected the determination of the system performances and its components. It was shown that air-heating coil, air fans and desiccant wheel contributed to large percentage of exergy destruction. Hence, the mentioned components should be given attention for further improvement of the system performances.

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Keywords: Desiccant dehumidification; Evaporative cooling; Air handling system; Exergy Analysis

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________________________________________________

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*Corresponding Author (N. Enteria):

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Napoleon Enteria (IIR Associate Member) Enteria Grün Energietechnik, Davao 8000, Philippines EMail: [email protected]; Tel/Fax No: +63-82-305-2226

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Nomenclatures

38

AHC/HC

air-heating coil

39

Cp

specific heat, kJ·kg-1·K-1

40

DEC/EC

direct evaporative cooler

41

DW

desiccant wheel

42

DP

depletion number

43

E

end

44

EA

exit air

45

EAF

exit air fan

46

HEvap

enthalpy of evaporation, kJ·kg-1

47

HX [1]

primary heat exchanger (big)

48

HX [2]

secondary heat exchanger (small)

49

h

enthalpy, kJ·kg-1

50

IE

electric current, A

51

Irr

irreversibility, kW

52

L,M,N

variables

53

ṁ

mass flow rate, kg·s-1

54

OA

outdoor air

55

OAF

outdoor air fan

56

P

pressure, Pa

57

Qɺ

heat transfer rate, kW

58

Qɺ X

59

R

60

RA

61

RegT

regeneration temperature

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RT

reference temperature

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Ṡ

entropy rate, kW·K-1

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input exergy in air-heating coil, kW gas constant, kJ·kg-1·K-1

return air

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start SA

supply air

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t

time, s

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T

temperature, K

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work rate, kW

Greek symbol

71

ε

efficiency

72

ψ

specific flow exergy, kJ·kg-1

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ω

humidity ratio, kgVapor·kgAir-1

74

relative humidity, %

75 Subscripts

77

a

air

78

Des

destruction

79

e

exit

80

EX

exergy

81

Gen

generation

82

HX

heat exchanger

83

i

84

l

85

ma

86

N

87

o

outlet

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r

reference condition

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Sys

system

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inlet

liquid

moist air node

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v

vapour

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w

water

92 1. Introduction

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Desiccant-based air-conditioning system is one of the most promising alternatives to the refrigerant-based air-conditioning system (Enteria and Mizutani, 2011; Choudhury et al., 2013). The main advantage of the desiccant-based air-conditioning system is the natural air dehumidification process called sorption process. The typical desiccant air-conditioning system consists of the desiccant wheel, heat wheel and direct evaporative coolers installed in the supply and return air streams (Henning et al., 2001; Panaras et al., 2011; Hurdogan et al., 2010 and Bourdoukan et al., 2008). Fixed bed (Enteria et al., 2011; Cho and Kato, 2010) and rotating desiccant dehumidifier (Finocchiaro et al., 2012; Hurdogan et al., 2010) are the most common desiccant dehumidifier. However, rotating desiccant dehumidifier (desiccant wheel) is preferable for application due to its simplicity (Enteria and Mizutani, 2011).

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It is shown that determination of the exergetic performances of any thermal energy system gives more detailed and level playing field in the analysis of the system thermal performances (Bejan, 2006). In fact there are several researches in the exergetic analyses on renewable energy systems (Hepbasli, 2008; Li and Yang, 2010), thermal energy systems (Rosen, 2001; Kaushik et al., 2011), thermal system components (Fakheri, 2010), air-handling systems (Caliskan et al., 2011; Koroneos et al., 2010; Marletta, 2010; Zhai et al., 2009) and buildings (Torio et al., 2009). Exergetic analyses for the desiccant-evaporative air-conditioning system (Kanoglu et al., 2004; La et al., 2012), hybrid system (Pons and Kodama, 2000; Hurdogan et al., 2011) solar-desiccant air-conditioning system (Enteria et al., 2012) are studied.

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In the exergetic evaluation of the desiccant air-conditioning system, it is shown that performances of the system components affected the system performance (Lavan et al., 1982). It is shown that higher regeneration temperature affected the system performance (Kodama et al., 2000). Shen and Worek, 1996 shows that there is an optimal regeneration temperature in which the system performance is high. La et al., 2012 show that there is trade-off between the regeneration temperature and the exergy of the supply air that affect the system performance. This study presented the exergetic evaluation of the developed desiccant air-conditioning system at different regeneration temperatures using steady state conditions. The objective of the study is to determine the exergetic performances of the developed system and its components for further improvement; also, to evaluate the developed system performances in comparison with different desiccant-based air-conditioning systems.

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2. Experimentation

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2.1. System overview

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2.1.1. System description

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The system consists of the desiccant wheel, sensible heat exchangers, and a direct evaporative cooler (Enteria et al., 2011). Figs. 1 and 2 show the developed desiccant-evaporative air-conditioning system diagram and experimental set-up made. In Fig. 1, the air from state 1 to state 3 is dehumidified by desiccant wheel (DW). The air from state 3 is pre-cooled by a sensible heat exchanger (HX [1]) and split into two streams of air (states 4 and 4’). The air from state 4’ is cooled by a direct evaporative cooler 4

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(DEC). The air of state 4 is sensibly cooled by smaller heat exchanger (HX [2]) from state 4 to state 5. The air at state 5 becomes the supply air. The air at state 5 has the same humidity ratio as the processed air at state 3. The air at state 6 is mixed with the return air (state 7) for the pre-heating of hot air shown at state 9. The hot air at state 9 is heated by the air heating coil (AHC) to become the regeneration air (state 10). The regeneration air at state 10 is used to remove the moisture in the desiccant wheel (DW). The air at state 11 has higher humidity ratio due to regeneration process in the desiccant wheel. The air at state 11 is exhausted by fan (EAF) to state 12.

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The system uses a honeycomb silica-gel coated desiccant wheel with 0.4 m diameter and 0.2 m thickness. The surface areas for air dehumidification side and wheel regeneration side are equal. The installed heat exchangers are cross-flow heat exchanges (sinusoidal flutes). The heat exchangers are made of paper coated with wax to avoid the transfer of moisture. The big heat exchanger (HX [1]) has dimensions of height 0.4 m, width 0.3 m, and length 0.3 m. The smaller heat exchanger (HX [2]) has dimensions of height 0.4 m, width 0.15 m, and length 0.15 m. The direct evaporative cooler (DEC) is a gravity flowing film type made of corrugated paper.

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The system operation is carried on using a controlled chamber (Chamber OA; see Fig. 1). The control of the return air conditions (T and x) is done by means of controlled Chamber RA. The measurements of the air flow rates (outdoor air, supply air, return air, and exit air) are determined by means of the orifices in the air duct with pressure-difference sensors and temperature sensors. The control of air flow rates (outdoor air, supply air, return air, and exit air) is done by the air dampers attached to the air ducts (See Enteria et al., 2010). The monitoring and control of the desiccant wheel rotational speed is through the use of rotational laser sensor attached to the wheel belt. The control of the direct evaporative cooler is by means of the relative humidity sensor attached at the downstream side (state 5’) through closed loop control system. The control, monitoring and data gathering of the experimentation are made by means of the sensors attached to the experimental test-rig of the desiccant-evaporative cooling system and connected to the personal computer through data logger.

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2.2. Experimental set-up

162

2.2.1. Assembly

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The total experimental set-up has two controlled chambers (OA and RA) to control the condition of the air (temperature and humidity) as mentioned above (Section 2.1.4). Between the two controlled chambers, a working chamber in which the desiccant-evaporative cooling system is located with air ducts connected to the two controlled chambers (See Fig. 1). The air temperatures in the controlled chambers are controlled by the electric heater and air cooler using the heat pump. The heat pump serves also as the dehumidifier for the air, in case air dehumidification is needed. The humidification process of the air is done through spray-type air humidifier. The operation of the two controlled chambers is meanly accomplished by passing the raw outdoor air to the dehumidifier and cooler/heater device. In case low humidity is needed, dehumidification process operates. However, if the air becomes cool due to dehumidification process, electric heater operates to increase the temperature. In the case when higher humidity of air is needed, the raw outside air passes the humidifier as shown in the diagram. The mixing chamber is installed after the air processing devices to control and homogenize the air prior to supplying it to the working chamber (See Enteria et al., 2007 for more detailed explanation).

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2.1.2. Components description

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2.1.3. System operation

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2.2.2. Instrumentations

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All the attached sensors in the developed desiccant-evaporative cooling system are evaluated and calibrated in the controlled box (Eyela, Co.) prior to their installation. The evaluated and calibrated sensors in the box are the dew point temperature sensors and the thermocouple temperature sensors. Based on the calibration, all the reading is within the specified value given by the manufacturers. Figure 1 shows the location and type of the sensors attached to the system. The measurement of the air humidity ratio is done by means of the dew point temperature sensors (Yokogawa Denshikiki Co., Ltd.) (±0.5 oC) and dry bulb temperature sensors (fabricated thermocouples). The fabricated thermocouples are from JIS VT3, O 0.32 mm (±0.5 oC). Proper installations of dew point temperature and thermocouple probes are observed based on our previous experiences (Enteria et al., 2010). The measurement of the desiccant wheel rotation is by the installed rotational laser sensor attached to the wheel belt (Omron, Co., Ltd.). The measurement of air flow rates is done using the installed orifices with digital pressure difference sensors (Nagano Keiki, Co., Ltd.) (±3 Pa). The sensors attached to the desiccant-evaporative air-conditioning system are connected to the data logger (Eto Denki Co.). The data logger is then connected to the personal computer (Dell Computer). In this measurement the data gathering is made in every 30 seconds.

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In the experimentation, the air conditions for the outdoor air (Point 1) are set at value of 30oC dry bulb temperature (2% accuracy) and 60% relative humidity (10% accuracy) (16.1g·kg-1). This is the standard summer testing condition in Japan. The return air (RA) is set at value of 26oC dry bulb temperature (1% accuracy) and 55% relative humidity (1% accuracy) (11.6 g·kg-1). This is the standard indoor air condition in Japan. The volumetric flow rates for the outdoor air (OA) and exit air (EA) are set at 200 m3·h-1. The volumetric flow rates for the return air (RA) and supply air (SA) is set 100 m3·h-1. The evaporative water inlet temperature is measured at 21.6oC average and controlled by water heater to mimic the temperature of supply water pipe. The system is evaluated using the four different regeneration temperatures that can be supported by solar energy or low temperature thermal energy source from waste-heat. The regeneration temperatures used in the experimentation are 60oC, 65oC, 70oC and 75oC. Fig. 3 shows the actual states of air inside the desiccant-evaporative cooling system. Table 1 shows the measured air flow rates for the different regeneration temperatures presented in Fig. 3. Table 2 shows the values for the specific enthalpy, specific exergy and specific entropy for the different states inside the system for four cases of regeneration temperatures with reference state of 25oC and 15g/kg.

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3. Thermodynamic analysis

212

3.1. General formulation

213 214 215

The evaluation of the desiccant-evaporative air-conditioning system is based on the governing concept of the laws of thermodynamics. The general formulations for exergy analyses of the thermal system are presented in Eqs. (1) to (20) and applied in the desiccant-evaporative cooling system thermodynamic

216

model shown in Fig. 1. In these formulations, the reference states are the Tr,

217

The general mass rate balance for the open system is

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2.2.3. Cases

r,

ωr, Pr.

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N

M

∑ mɺ

j,I

∑ mɺ

−

j =1

The general energy rate balance for the open system is

∑ (mɺ h ) = ∑Wɺ N

j,I

+

224

225 226

227

i =1

  T 1 −  r T j =1    K, j

∑

 Qɺ K , j +  

∑ (mɺ jψ j )I = ∑Wɺi,O − ∑ (mɺ iψ i )O M

M

j =1

i =1

i =1

The general entropy generation rate for the open system is SɺGen =

 QS , k   T  k =1  S ,k 

∑ (mɺ i si )O − ∑ (mɺ j s j )I − ∑  N

M

i =1

j =1

L

(3)

(4)

Where the general exergy rate of destruction or the rate of irreversibility is Eɺ X , Dest = I rr = Tr SɺGen =

∑ (Eɺ x, j )I − ∑ (Eɺ x,i )O N

M

j =1

i =1

(5)

The general exergetic efficiency for the open system is M M  (mɺ iψ i )O   Wɺ i ,O +  i =1  i =1 N  N  T  1 −  r   Qɺ K , j + mɺ jψ   TK , j   j =1  j =1

εX =

∑

∑(

∑ (Eɺ )

)

j I

=

M

i =1 N

x ,i O

∑ (Eɺ ) j =1

(6)

x, j I

RD =

EP

The relative destruction of exergy for the open system is presented as

(Eɺ X , Dest )t

⋅ 100

C

∑

(7)

Eɺ X , Dest

AC C

230

(2)

i i O

i =1

N

∑ 229

∑ (mɺ h )

The general exergy rate balance for the open system is

∑

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+

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223

M

i ,O

j I

i =1

N

222

M

j

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∑ Qɺ j =1

221

(1)

dt

i =1

N

220

dmSys

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=

i ,O

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t =1

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The Van Gool’s improvement potential (IP) for the open system is presented as (Hepbasli, A., 2008)  IP = (1 − ε X ) ⋅   

∑ (Eɺ ) − ∑ (Eɺ ) N

j =1

x, j I

M

i =1

x ,i O

   

(8)

The Van Gool’s improvement potential (IP) is the maximum possible improvement potential that could be done for any processes through minimization of the exergy destruction or the difference between the inlet exergy and outlet exergy. 7

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The sustainability index (SI) for the open system can be expressed as (Rosen et al., 2008)

∑(Eɺ x, j )I N

237

SI =

j =1

∑(Eɺ x, j )I − ∑(Eɺ x,i )O N

M

j =1

i =1

∑(Eɺ ) N

1 1 = = = 1 − ε X DP

j =1 C

x, j I

(9)

∑ Eɺ X ,Dest t =1

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The sustainability index is an inverse of the depletion number (Dp); which is the relationship between the exergy destruction of the exergy input. When the exergy destruction decreases even with the same exergy input due to the increase of the efficiency, the sustainability index increases. It means that less fuel is consumed particularly when using conventional sources.

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The specific flow exergy is

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ψ = (h − hr ) − Tr (s − sr )

 Tw   − RvTr ln φ r  Tr 

ψ w = C P ,w (Tw − Tr ) − Tr C P ,w ln 

specific exergy of moist air is (Hepbasli, A., 2008)

 Tma    T   P   − 1 − ln ma   + (1 + 1.607ω )Tr ln ma  +  Tr    Pr   Tr  

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ω  (1 + 1.607 wr )  RgTr (1 + 1.607ω )ln  + 1.607 ln    (1 + 1.7607ω )  ωr  (12)

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(11)

φr is the equivalent relative humidity at reference temperature ( Tr ) and humidity ratio ( ω r ). The ψ ma = (C p , g + ωC p , v )Tr 

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(10)

The specific exergy of incompressible substance such as water is (Marletta, 2010)

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The specific enthalpy and entropy of water, water vapor, dry air and moist air are (Pons, M., Kodama, A., 2000). hw = C P, w (Tw − 273.15)

(13)

253

hv = 2502 + 1 . 90 (Tv − 273 . 15 )

(14)

254

ha = 1.005(Ta − 273.15)

(15)

255

hma = ha + xhv

(16)

256

 T  sw = 4.19ln w   273.15

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(17)

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 T  sv = 9.158+ 1.90ln w   273.15 

(18)

258

 T  sa = 1.005ln a   273.15

(19)

259

x  0.622     s ma = s a − 0.287 ln  + x s w − 0.462 ln   x + 0.622    x + 0.622 

3.2. System analysis

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3.2.1. Desiccant Wheel (DW) Mass balance:

263

ɺ 2 = mɺ 3 = mɺ OA m

264

ɺ 10 = mɺ 11 = mɺ EA m

265

(B)

.  .   E X ,Dest  =  E E  + [mɺ EA (ψ 10 −ψ 11 )] + [mɺ OA (ψ 2 −ψ 3 )]   DW   DW

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Exergy destruction rate:

(C)

Exergy efficiency

(ε X )DW

268

=

[mɺ OA (ψ 3 − ψ 2 )]

.   E E  + [mɺ EA (ψ 10 − ψ 11 )]   DW

271

(21) (22)

(23)

(24)

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(A)

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(20)

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3.2.2. Sensible Heat Exchanger (HX [1]/HX [2]) (A)

Mass balance:

ɺ 3 = (mɺ 4' + m ɺ 4 ) = mɺ OA m

(25)

273

ɺ 8 = mɺ 9 = mɺ EA m

(26)

274

ɺ 4 = mɺ 5 = mɺ SA m

(27)

275

ɺ 5' = mɺ 6 = (mɺ OA − mɺ SA ) m

(28)

276

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(B)

Exergy destruction rate:

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  (ψ + ψ 4 )    .  = mɺ OA ψ 3 −  4 '  E X , Dest    + [mɺ EA (ψ 8 − ψ 9 )] 2   HX (1)   

(29)

278

.  = [mɺ SA (ψ 4 − ψ 5 )] + [(mɺ OA − mɺ SA )(ψ 5' − ψ 6 )]  E X , Dest    HX (2 )

(30)

(C)

Exergy efficiency:

[mɺ EA (ψ 9 −ψ 8 )]   (ψ 4' + ψ 4 )   ψ −

(ε X )HX (1) =

280

mɺ OA  

3

 

2

(31)

 

(ε X )HX (2) = [(mOA − mSA )(ψ 6 −ψ 5' )] [mɺ SA (ψ 4 −ψ 5 )] ɺ

281

ɺ

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3.2.3. Direct Evaporative Cooler (DEC) (A) Mass balance:

285

mɺ 4' = mɺ 5' = (mɺ OA − mɺ SA )

286

mɺ W ,I = mɺ W ,E = mɺ W

(B)

.   E X , Dest  = [(mɺ OA − mɺ SA )(ψ 5' −ψ 4' )] − (mɺ Wψ W )   EC

288 289

Exergy destruction:

(C)

Exergy efficiency:

(ε X )EC

290

=

[(mɺ OA − mɺ SA )ψ 5' ] [mɺ Wψ W ] + [(mɺ OA − mɺ SA )ψ 4' ]

AC C

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(32)

(33) (34)

(35)

(36)

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3.2.4. Air Heating Coil (AHC)

293

(A) Mass balance:

294

ɺT = m ɺL mɺ S = m

(37)

295

ɺ9 = m ɺ10 = m ɺ EA m

(38)

296

(B)

Exergy destruction rate:

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.   E X , Dest  = [mɺ L (ψ S − ψ T )] + [mɺ EA (ψ 9 − ψ 10 )]   AH

297 (C)

Exergy efficiency: ɺ (ε X ) AH = [mEA (ψ 10 −ψ 9 )] [mɺ L (ψ T −ψ S )]

299

(40)

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(39)

301

3.2.5. Air Fan (OAF and EAF)

302

(A) Mass balance:

mɺ 1 = mɺ 2 = mɺ OA = mɺ F (OAF )

304

mɺ 11 = mɺ 12 = mɺ EA = mɺ F ( EAF )

305

(B)

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303

SC

300

Exergy destruction:

(41) (42)

306

.  .  = EE  − [mɺ F (OAF ) (ψ 2 − ψ 1 )]  E X , Dest    F (OAF )   F (OAF )

(43)

307

.  .  = EE  − [mɺ F ( EAF ) (ψ 12 − ψ 11 )]  E X , Dest    F ( EAF )   F ( EAF )

(44)

Exergy efficiency:

(ε X )F (OAF ) =

310

(ε X )F (EAF ) =

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[mɺ F (OAF ) (ψ 2 − ψ 1 )] .   E E   F (OAF ) 

[mɺ F (EAF ) (ψ 12 −ψ 11 )] .   E E    F (EAF )

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(45)

(46)

The calculation of the electric energy consumed by the electric fan is based on the .  m F (OAF ) (h2 − h1 )  =

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.   E E    F (OAF )

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  EE    F ( EAF ) .

η M (OAF )

.   m F ( EAF ) (h12 − h11 ) =

η M ( EAF )

(47)

(48)

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(A) Electric exergetic coefficient of performance (EEXCOP)

[(ε

]

)

X Sys Electric

=

[mɺ SA (ψ 1 − ψ 5 )]

(49)

       E E  +  E E  +  E E    DW   F (OAF )   F (EAF ) .

.

.

317

(B) Thermal exergetic coefficient of performance (TEXCOP)

318

[(ε

319

(C) Total exergetic coefficient of performance (SEXCOP)

[(ε

)

]

X Sys Thermal

)

]

X Sys Total

=

321 322

[mɺ SA (ψ 1 − ψ 5 )]

(50)

Qɺ X

[mɺ SA (ψ 1 − ψ 5 )]

.  .  . Qɺ X +  E E  +  E E  +  E  + (mɺ Wψ W )   DW   F (OAF )   F (EAF )

(51)

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3.2.6. System exergetic performances

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The calculation of thermal exergy delivered to the air-heating coil (HC) is based on . .   T  Q X = m F C p (TI − TO ) − Tr ln I  (52)  TO  

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3.2.7. System exergetic efficiencies

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The actual thermal coefficient of performance is simply the ratio of the cooling produced to the thermal energy supplied as .

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.

m EA (h10 − h9 )

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COP A =

(53)

The Carnot coefficient of performance for any heat driven air-conditioning system is shown based on the diagram shown in Fig. 1 is

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m SA (h1 − h5 )

 T  T  COPC = 1 − 1  7   T10  T1 − T7 

(54)

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The COPC is the highest potential of the coefficient of performance for any heat driven air-conditioning system which is mostly applied in closed cycles.

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The proposed reversible Carnot coefficient of performance for the open cycle desiccant air-conditioning system is expressed as (Lavan et al., 1982)

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 T COPRe v = 1 − c  Ts

 Te   T − T  c e

   

(55)

Where Ts, Te and Tc are equivalent temperatures for the heat source, evaporator and condenser and are expressed based on Fig. 1

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Ts =

339

Te =

h9 − h10 s 9 − s10

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(56)

.

.

.

.

.

.

m SA h5 − m RA h7 + m W (5→ 7 ) h w

(57)

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(ε X )1 =

COPA COPRe v

(ε X )2

=

m SA (ψ 1 − ψ 5 ) .

m EA (ψ 10 − ψ 9 )

(60)

EP

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(59)

For comparison purposes with different studies (Dincer and Rosen, 2013) in thermal driven air-conditioning systems, the exergy efficiency is expressed based on the ratio of the cooling produced to the thermal exergy supplied as (exergy efficiency in the second definition). .

348

(58)

The exergy efficiency can be expressed as the ratio of the actual coefficient of performance for the thermal driven air-conditioning system to the reversible coefficient of performance shown as (Dincer and Rosen, 2013). It is called as the exergy efficiency in the first definition here.

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. . .  m OA h1 − m EA h12 −  m w(5→7 ) + mw(4'→5') hw   Tc = . . .  m OA s1 − m EA s12 −  m w(5→7 ) + mw(4'→5') s w  

SC

m SA s 5 − m RA s 7 + m W (5→ 7 ) s w

4. Results and discussion

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4.1. Components performances

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4.1.1. Outdoor air fan (OAF)

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Fig. 4a shows the inlet exergy rate (electricity) supplied for the operation of the outdoor air fan (OAF). The electricity (exergy rate) increases as the regeneration temperature increases. This is due to the increase of the flow resistances inside the system as inside air temperature increases. Fig. 4b shows the exergy destruction rate which is increasing as the regeneration temperature increases due to the increase of irreversibility as fan power requirement increases as frictional resistance increases to maintain the air flow rate. Fig. 4c shows the exergetic efficiency of the fan in which fan efficiency is almost the same for different regeneration temperatures. Fig. 4d shows the improvement potential for the outdoor air fan

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(OAF) in which the improvement potential increases as the regeneration temperature increases; however, the increase is less than 20% even for the highest regeneration temperature (75oC).

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4.1.2. Exit air fan (EAF)

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The inlet exergy rate for the EAF is much lower than for the case of the OAF due to the high frictional losses for the supply air stream compared to the return air stream shown in the system diagram (Fig. 5a) (Please see Fig. 1). In addition, high pressure frictional losses may happen in the outdoor air duct and supply air duct that contributed to the high inlet exergy rate for the OAF compared to the EAF. As the regeneration temperature increases, the inlet exergy rate increases except for 70oC and 75oC which are almost the same. Fig. 5b shows the exergy destruction rate for the EAF in which exergy destruction increases as the regeneration temperature increases except for the case of 70oC and 75oC (almost the same). Fig. 5c shows the exergetic efficiency of the EAF showing that the efficiency is almost the same, except that there is a small change as the regeneration temperature increases. Fig. 5d shows the improvement potential for the EAF for the different regeneration temperatures. The trend of the improvement potential follows the trend of the results of the exergy efficiency in which there is a variation for the cases of 70oC and 75oC.

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4.1.3. Air-heating coil (AHC/HC)

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Fig. 6a presented the inlet exergy rate for the thermal supply for the system. It shows the increase of thermal exergy as the regeneration temperature increases. Fig. 6b shows the exergy destruction rate for the air-heating coil showing the increase of destruction as the regeneration temperature increases. Fig. 6c shows the exergetic efficiency of the air-heating coil at different regeneration temperatures. As shown, the exergy efficiency increases as the regeneration temperature increases by 10%. Fig. 6d shows the improvement potential for the air-heating for different regeneration temperatures. It shows that the improvement potential increases as the regeneration temperature increases.

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4.1.4. Desiccant wheel (DW)

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It shows the inlet exergy rate shown in Fig. 7a. The inlet exergy rate increases as the regeneration temperature increases at approximately 15%. Fig. 7b shows the exergy destruction rate with the increase of 15% as the regeneration temperature increases. Fig. 7c shows the exergetic efficiency for different regeneration temperatures. The exergy efficiency is the same for all the regeneration temperatures. This result is due to the rate of exergy destruction which is the same to the inlet exergy rate shown in Figs. 7a and 7b. The exergy efficiency of the desiccant wheel is between 45% and 75% depending on different reference temperature. Van den Bulck et al. (1988) shows that the maximum exergy efficiency for desiccant dehumidifiers is around 85%. On the other hand, Kanoglu et al. (2004) get 76.1%. It means the installed desiccant wheel exergy efficiency has reasonable result. In addition, it shows that at 65 to 75oC, the efficiency is the highest or the optimum. Hence, Shen and Worek (1996) shows that there is an optimum regeneration for desiccant wheel. Fig. 7d shows the desiccant wheel improvement potential for different regeneration temperatures. It shows that for higher regeneration temperature, the improvement potential for desiccant wheel is also higher.

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4.1.5. Primary heat exchanger (HX1)

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Fig. 8a shows the inlet exergy rate which shows an increase as the regeneration temperature increases. The increase is almost 10%. Fig. 8b shows the exergy destruction rate for the primary heat exchanger which is increasing as the regeneration temperature increases. Fig. 8c shows the exergetic efficiency of the primary heat exchanger. It shows the decrease of the efficiency as the regeneration temperature increases due to the high rate of exergy destruction presented in Fig. 8b. This is due to the increase of

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irreversibility as heat transfer increases. Fig. 8d shows the improvement potential for the primary heat exchanger showing the increase as the regeneration temperature increases.

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4.1.6. Secondary heat exchanger (HX2)

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Fig. 9a shows the inlet exergy rate for the secondary heat exchanger at different regeneration temperatures. It shows that the inlet exergy rate increases with the regeneration temperature. Fig. 9b shows the change of the exergy destruction rate as the regeneration temperature increases. It shows that the exergy destruction rate increases as the regeneration temperature increases. It is evident for higher regeneration temperature. With such phenomenon, the exergetic efficiency decreases as the regeneration temperature increases as presented in Fig. 9c. Hence, the improvement potential for the heat exchanger increases as the regeneration temperature increases as shown in Fig. 9d. Comparing the two heat exchangers (HX1 and HX2), it appears that the primary heat exchanger has lower exergy efficiency compared to the secondary heat exchanger as primary heat exchanger is operating at higher temperature. With respect to the reference temperatures, the sudden drop of exergy efficiency of the secondary heat exchanger as the reference temperature increases is affected by the condition of Point 5’ (See Fig. 1) condition from the evaporative cooler outlet condition.

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4.1.7. Direct evaporative cooler (EC)

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Fig. 10a shows the inlet exergy rate for the evaporative cooler. As shown, the inlet exergy rate increases as the regeneration temperature increases due to the increase of the water consumption as saturation required more water as Point 4’ temperature increases (See Figs. 1 and 3). The exergy destruction rate is presented in Fig. 10b showing the increase as the regeneration temperature increases. Comparing the inlet exergy rate increase and the exergy destruction rate (Figs. 10a and 10b), the exergy destruction rate is higher due to the increase of the irreversibility of adding moisture to the air during the evaporative cooling process. This resulted to lowering of the exergy efficiency of the evaporative cooler as the regeneration temperature increases shown in Fig. 10c. Furthermore, the exergy efficiency decreases until the reference temperature reaches 20o to 25oC then moves up. Entropy generation decreases with the increasing reference temperature. Also, Reference temperature is not the only reason for this decrease. Exergy destruction affects the entropy generation as well as reference temperature (Caliskan et al., 2011). With such phenomenon, the improvement potential for the evaporative cooler increases as the regeneration temperature increases (Fig. 10d).

432

4.2. Exergy destruction

433

4.2.1. Exergy destruction contribution

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Fig. 11 shows the exergy destruction contribution of the developed desiccant-evaporative air-conditioning system at different regeneration temperatures and reference temperatures. The air-heating coil (AHC) contributed to high percentage of the total exergy destruction: the same trend for exergy destruction for the case of Hurdogan et al., 2011which shows that system thermal source and air fans are the main contributors. The destruction contribution is followed by the outdoor air fan (OAF) and then by the desiccant wheel (DW). As the regeneration temperature increases, the contribution of the air-heating coil (AHC) to exergy destruction increases. Kodama et al., 2009 shows that heat source and desiccant wheel are major sources of exergy destruction (air fans are not considered in their analysis). The exergy destruction contribution of the desiccant wheel (DW) increases as the regeneration temperature increases. The trend for the exergy destruction contributions for the heat exchangers (HX1 and HX2) and direct evaporative cooler (EC) increases as the regeneration temperature increases due to the increase of the components irreversibility discussed in Section 4.1. Furthermore, based on the different reference temperatures, it shows the increase of the air-heating coil exergy destruction contribution.

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4.2.2. Exergy destruction coefficient

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Fig. 12 shows the exergy destruction coefficient for different components of the developed desiccant-evaporative air-conditioning system at different regeneration temperatures and reference temperatures. The trend of the exergy destruction coefficient is the same as the exergy destruction contribution presented in Fig. 11. The air-heating coil (HC) has the largest exergy destruction coefficient followed by the outdoor air fan (OAF) and desiccant wheel (DW). The exergy destruction in the air-heating coil could be lowered by increasing the efficiency of the air-heating coil. La et al., 2012 shows also that heat source has the largest exergy destruction coefficient. As shown in Fig. 6c, the air-heating coil effectiveness is very low, between 5% - 25% with respect to different reference temperatures. Air fans efficiency is low particularly the outdoor air fan (OAF) due to the high frictional losses which occurred for the supply air as it has more internal frictional loses from different flow directions than of the exit air (See Fig. 1). It also shows that when the regeneration temperature increases, the exergy destruction coefficient for the air-heating coil increases due to the increase of heat transfer irriversivility. It is the same for the cases of desiccant wheel and heat exchangers (HX1 and HX2) and evaporative cooler (EC). The explanations are presented in the different components analyses shown in Section 4.1. On the other hand, when the reference temperature increases, the exergy destruction coefficient for air-heating coil (HC) increases. For comparison,

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464 4.3. System performances

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Fig. 13a shows the thermal exergy coefficient of performance of the system. The thermal coefficient of performance (TEXCOP) varied with the reference temperature, from 0oC to 22oC, the thermal exergy performance of the system decreases, after which, the performance increases. However, based on the different regeneration temperatures, from 0oC to 22oC, the thermal exergy performance of the system decreases as the regeneration temperature increases and suddenly changed from 22oC to 30oC. On the other hand, for the electric exergetic performance presented in Fig. 13b, the trend of the electric exergetic coefficient of performance (EEXCOP) is the same as the thermal exergetic coefficient of performance shown in Fig. 13a in terms of reference temperatures. It means that increasing the regeneration temperature also affects the electric consumption of the system due to the increase of system internal frictional losses resulted to the decrease of the EEXCOP. On the other hand, the increase of the regeneration temperature resulting to higher increase of the system thermal exergy destruction due to components irreversibility and resulted to the decrease of the TEXCOP as explained in the different components presented in Section 4.1. Fig. 13c shows the over-all system exergetic coefficient of performance (SEXCOP). It shows the same trend based on regeneration temperature and reference temperatures. Based on the analysis of each of the three exergy coefficient of performances (TEXCOP, EEXCOP and SEXCOP), the system performance decreases as the regeneration temperature increases. Furthermore, there is a sudden decrease of the performance from 60oC to 65oC and 70oC to 75oC. Based on Fig. 2, the dehumidification and cooling of supply air (Point 5) is almost the same for regeneration temperatures of 65oC and 70oC. The sustainability index (SI) shows that as the regeneration temperature increases, the sustainability index decreases from 0oC to 22oC and changed from 22oC to 30oC (Fig. 13d). The sustainability index is related to exergy efficiency as shown in Eq. (9). Hence, from Fig. 13c, the over-all system exergy efficiency varies as the dead state temperature changes. This is the same trend for the case of the evaporative cooler shown by Caliskan et al., 2011.

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Fig. 14a shows the reversible coefficient of performance (COPRev). It shows that the COPRev has nothing to do with the reference temperature. As presented, the COPRev suddenly increases from 60oC regeneration temperature to 65oC. The COPRev then reduces at 70oC and increases a little at 75oC. This trend of COPRev is due to the system demumidification performance shown in Fig. 3. As shown in Fig. 3, there is a large 16

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increase of the system dehumidification performance from 60oC to 65oC which resulted to higher increase of COPRev and then from 65oC to 70oC, the dehumidification performance is not large even with this same increase of regeneration temperature increase of 5oC. In addition, from 70oC to 75oC, the increase is not as large as from 60oC to 65oC, this resulted to lower increase of COPRev. This trend little increase of the dehumidification performance of the desiccant wheel resulted from the heating of the desiccant material in the dehumidification side due to the heat-carry over from the regeneration side to dehumidification side of the desiccant wheel as observed from our previous studies (Enteria et al., 2010). The present system has higher COPRev of 4.20 at regeneration temperature of 65oC. Lavan et al., (1982) got 4.66 while Kanoglu et al., (2007) got a COPRev of 2.63 for double wheel + double evaporative coolers desiccant-evaporative air-conditioning system.

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Fig. 14b shows that the actual coefficient of performance (COPA) is higher at regeneration temperature of 60oC and suddenly decreases as the regeneration increases to 75oC. This trend of decreasing COPA as the regeneration temperature increases is due to the not proportional increase of the system cooling capability as the regeneration energy supported to the system increases as the regeneration temperature increases. As the regeneration temperature increases, the heat carry-over from the regeneration side of the desiccant wheel to its dehumidification side increases (Enteria et al., 2010). The increase of the desiccant wheel temperature in the dehumidification side hampered the adsorption capacity as the sorption process is the difference of the water vapor pressure between the air and the desiccant surface. Also, the increase of the regeneration temperature increases the exergy of the supply air. As shown in Fig. 3, the increase of the air temperature after the desiccant wheel (state 3) is almost linear compared to the increase of the dehumidification capability as the regeneration temperature increases. Kodama et al., 2000 shows the same trend of result.

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Fig. 14c shows the Carnot coefficient of performance (COPC) which presents the increasing trend as the regeneration temperature increases. The increasing trend of the COPC is due to the increase of the regeneration temperature (T10). In addition, as shown in this research, the return temperature (T7) and the outdoor temperature (T7) are held constant. These values of the COPC are the possible Carnot coefficient of performance of the heat driven cooling system at closed cycle equivalent. However, open cycle desiccant air-conditioning system is not only for the temperature control but also for the humidity control. Therefore, the actual and the reversible coefficient of performances (COPA and COPRev) are different as explained in the above discussions for the COPA and COPRev. For comparison purposes, Kanoglu et al., 2007 presented a COPC of 8.74 based on ARI conditions. In this paper, the COPC at 75oC is 11.65.

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Fig. 14d shows the exergetic efficiency of first definition (Exergetic Efficiency I) which shows the decreasing trend as the regeneration temperature increases. The lowest value can be seen in the regeneration temperature of 70oC which has also a lower value of COPRev.. It shows that at 60oC, the exergetic efficiency I is 52.77%, while the lowest is 9.61% at 65oC. Fig. 14e shows the decreasing trend of the exergetic efficiency of the second definition (Exergetic Efficiency II). The negative value of the second efinition is due to the increase of the supply air exergy upon increase of the reference temperature. The exergetic efficiency of second definition decreases as the regeneration temperature increases due to the smaller increase of the cooling exergy capability of the system even with the increasing regeneration exergy supply to the system. This explanation is the same to the actual coefficient of performance (COPA). Moreover, the exergetic coefficient of performance of second definition decreases as the reference temperature increases. It means that at higher temperature, the system exergetic performance of second definition is not as efficient as in lower reference temperature due to the reduction of the system cooling exergy produced. Hence, it is expensive and inefficient to operate the system at higher outdoor air temperature than at lower outdoor air temperature.

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5. Conclusions

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This paper showed the exergy evaluation of the developed desiccant-evaporative air-conditioning system. The developed system was evaluated based on different regeneration temperatures and reference temperatures which the system will be subjected to during its actual operation. Air-heating coil (AHC), outdoor air fan (OAF) and then desiccant wheel (DW) contributed to high percentage of the total exergy destruction.

•

Based on the analysis of each of the three exergy coefficient of performances (TEXCOP, EEXCOP and SEXCOP), the system performance decreases as the regeneration temperature increases.

•

The sustainability index shows that as the regeneration temperature increases, the sustainability index decreases from 0oC to 22oC and changed from 22oC to 30oC (Fig. 13d).

•

The COPRev suddenly increases from 60oC regeneration temperature to 65oC. The COPRev then reduces at 70oC and then increases a little at 75oC. The trend of COPRev is due to the system dehumidification performance shown in Fig. 3.

•

The actual coefficient of performance (COPA) is higher at regeneration temperature of 60oC and decreases as the regeneration temperature increases to 75oC. This trend of decreasing COPA as the regeneration temperature increases is due to the not proportional increase of the system cooling capability as the regeneration energy supported to the system increases as the regeneration temperature increases.

•

The increasing trend of the COPC as the regeneration temperature increases is due to the increase of the regeneration temperature (T10). In addition, as shown in this research, the return temperature (T7) and the outdoor temperature (T7) are held constant. The exergetic efficiency of first definition (Exergetic Efficiency I) shows the decreasing trend as the regeneration temperature increases. The lowest value can be seen in the regeneration temperature of 70oC which has also a lower value of COPRev.

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The exergetic efficiency of second definition decreases as the regeneration temperature increases due to the little increase of the cooling exergy capability of the system even with the increasing regeneration exergy supply to the system.

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The exergetic evaluation is valuable in the evaluation of the desiccant-evaporative air-conditioning system we developed. It shows which components with large destruction of exergy that resulted to lowering of the system exergy efficiencies. Air heating coil (AHC), air fans (OAF and EAF) and desiccant wheel (DW) contributed to large percentage of system exergy destruction resulting to large impact to the reduction of the system performance. Hence, it is important to consider the components with high exergy destruction such as the air-heating coil, air fans and desiccant wheel for improvement.

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It was shown that the system presented in this paper which has one evaporative cooler and two sensible heat exchangers could have better performance compared to the typical double wheels and double evaporative coolers. For comparison purposes, the exergy efficiency of the liquid desiccant air-conditioning system is still high (6.8%, the lowest, for the basic system) (Xiong et al., 2010); however, handling of the liquid desiccant is more complex than solid desiccant. In the case of the solid desiccant, multi staging will increase the exergetic efficiency from 8.2% to 18.0% (La et al., 2012). However, multi-staging and making the air flow complex increases the fan power consumption that will affect the

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system performance when fan exergy input is considered as for our case. With respect to the closed cycle water-LiBr desiccant system, the open system has comparable COPA at lower regeneration temperature to half effect and single effect water-LiBr desiccant system shown by Gebreslassie et al. (2010) and Kaushik and Arora (2009). For example, in this paper, the COPA is 0.42 at 60oC while the water-LiBr desiccant system half effect has 0.458 (Gebreslassie et al., 2010). However, when converting the closed cycle water-LiBr system to cool and dehumidify the air, it is expected that the COPA will be lower. With respect to the closed cycle, solid desiccant system, the COPA is lower but could operate at much lower regeneration temperature. For example, at 60oC temperature, it could have a COP of less than 0.2 (Luo et al., 2006) which is less than the open-cycle counterpart. Askalany et al., (2013) presented some pairs of solid desiccant and refrigerant which could have higher COP. Hence, application of different types of desiccant-based air-conditioning system depends on the need and specific target. Wang et al., (2009) presented important guidelines of the desiccant-based air-conditioning system. For the case of open cycle desiccant-evaporative air-conditioning system it offers some advantages such as for chemical and biological treatments of the supply air to the indoor environment (Enteria and Mizutani, 2011). Further investigation through exergo-economic analysis is another step to deepen the evaluation of the system with respect to capital and operational cost in the light of thermodynamic point of view.

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Gebreslassie, B.H., Medrano, M., Boer, D., 2010. Exergy analysis of multi-effect water-LiBr absorption systems: From half to triple effect. Renew. Energ. 35, 1773-1782.

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Henning, H.M., Erpenbeck, T., Hindenburg, C., Santamaria, I.S., 2001. The potential of solar energy use in desiccant cooling cycles. Int. J. Refrig. 24, 220-229.

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Hepbasli, A., 2008. A key review on exergetic analysis and assessment of renewable energy resources for a sustainable future. Renew. Sust. Energ. Rev. 12, 593-661.

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Hurdogan, E., Buyukalaca, O., Hepbasli, A., Yilmaz, T., 2011. Exergetic modeling and experimental performance assessment of a novel desiccant cooling. Energ. Buildings 43, 1489-1498.

645 646

Hurdogan, E., Buyukalaca, O., Yilmaz, T., Hepbasli, A., 2010. Experimental investigation of a novel desiccant cooling system. Energ. Buildings 42, 2049-2060.

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Kanoglu, M., Carpinlioglu, M.O., Yildirim, M., 2004. Energy and exergy analyses of an experimental open-cycle desiccant cooling system. Appl. Therm. Eng. 24, 919-932.

649 650

Kanoglu, M., Bolatturk, A., Altuntop, N., 2007. Effect of ambient conditions on the first and second law performance of an open desiccant cooling process. Renew. Energ. 32, 931-946.

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Kaushik, S.C., Arora, A., 2009. Energy and exergy analysis of single effect and series flow double effect water-lithium bromide absorption refrigeration systems. Int. J. Refrig. 32, 1247-1258.

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Kaushik, S.C., Siva Reddy V., Tyagi, S.K., 2011. Energy and exergy analyses of thermal power plants: a review. Renew. Sust. Energ. Rev. 15, 1857-1872.

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Kodama, A., Jin, W., Goto, M., Hirose, T., Pons, M., 2000. Entropic analysis of adsorption open cycles for air conditioning. Part 2: interpretation of experimental data. Int. J. Energ. Res. 24, 263-278.

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Koroneos, C., Nanaki, E., Xydis, G., 2010. Solar air conditioning systems and their applicability – an exergy approach. Resour. Conserv. Recy. 55, 74-82.

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La, D., Li, Y., Dai, Y.J., Ge, T.S., Wang, R.Z., 2012. Development of a novel rotary desiccant cooling cycle with isothermal dehumidification and regenerative cooling using thermodynamic analysis method. Energy 44, 778-791.

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Luo, H.L., Dai, Y.J., Wang, R.Z., Wu, J.Y., Xu, Y.X., Shen, J.M., 2006. Experimental investigation of a solar adsorption chiller used for grain depot cooling. Appl. Therm. Eng. 26, 1218-1225.

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Marletta, L., 2010. Air conditioning systems from a 2nd law perspective. Entropy 12, 859-877.

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Pons, M., Kodama, A., 2000. Entropic analysis of adsorption open cycles for air conditioning. Part 1: first and second law analyses. Int. J. Energ. Res. 24, 251-262.

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Qureshi, B.A., Zubair, S.M., 2003. Application of exergy analysis to various psychrometric processes. Int. J. Energ. Res. 27, 1079-1094.

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Rosen, M.A., Dincer, I., Kanoglu, M., 2008. Role of exergy in increasing efficiency and sustainability and reducing environmental impact. Energ. Policy 36, 128-137.

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Rosen, M., 2001. The exergy of stratified thermal energy storages. Sol. Energy 71, 173-185.

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Figures and Tables

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Humidity Ratio (kg/kg)

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Fig. 3:

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1072

Table 1:

Regeneration Temperature

OA m³/h

EA kg/s

kg/s

75 C

: 200.33 0.0647 : 207.95 0.0645 : 94.38 0.0309 : 92.64 0.0305

SC M AN U

1078 1079 1080

TE D

1081 1082 1083

AC C

EP

1084

1089

m³/h

: 201.32 0.0650 : 206.55 0.0646 : 94.30 0.0310 : 92.59 0.0305 : 200.65 0.0648 : 207.01 0.0645 : 94.27 0.0309 : 92.71 0.0305

1077

1088

kg/s

65 C 70 C

1076

1087

m³/h

: 201.18 0.0650 : 205.57 0.0646 : 94.31 0.0310 : 92.78 0.0306

1075

1086

kg/s

RA

60
C

1074

1085

m³/h

SA

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1073

1090 1091

36

Table 2

Temperature, T (°C)

Humidity Ratio, ω (g/kg)

Specific Enthalpy, h (kJ/kg)

Specific Entropy, s (kJ/kg-K)

State 65°C

70°C

75°C

70°C

75°C

0.015

60°C :

65°C

70°C

75°C

63.37

60°C

Reference

:

1

:

30.1

30.1

30.2

30.1

:

0.0166

0.0165

0.0166

0.0165

:

72.7

72.5

:

2

:

32.0

32.1

32.2

32.3

:

0.0166

0.0165

0.0166

0.0165

:

74.7

74.5

74.8

74.8

:

3

:

49.3

52.0

54.4

57.6

:

0.0118

0.0111

0.0108

0.0103

:

80.2

81.1

82.8

84.8

4'

:

36.9

38.2

39.4

40.9

:

0.0118

0.0111

0.0108

0.0103

:

67.4

66.9

67.3

4

:

30.1

30.8

31.4

32.2

:

0.0118

0.0111

0.0108

0.0103

:

60.5

59.3

5'

:

22.9

22.9

23.0

23.1

:

0.0167

0.0167

0.0166

0.0167

:

65.5

5

:

26.0

26.4

26.7

27.1

:

0.0117

0.0110

0.0107

0.0103

:

6

:

26.8

27.2

27.6

28.1

:

0.0167

0.0167

0.0166

0.0167

7

:

26.3

26.4

26.7

26.9

:

0.0117

0.0116

0.0117

0.0119

8

:

26.8

27.1

27.4

27.7

:

0.0143

0.0143

0.0143

0.0143

9

:

42.1

44.0

45.7

47.8

:

0.0143

0.0143

0.0143

0.0143

10

:

60.7

64.6

68.5

74.2

:

0.0143

0.0143

0.0143

11

:

39.6

41.5

43.1

45.2

:

0.0189

0.0194

12

:

40.2

41.9

43.2

45.4

:

0.0189

Water

:

21.6

21.7

21.6

21.7

:

-

72.6

:

70°C

75°C

0.261

0.296

60°C

65°C

:

70°C

75°C

-

0.296

0.295

:

0.056

0.055

0.057

0.055

0.302

0.302

0.303

0.303

:

0.096

0.097

0.100

0.101

:

0.306

0.306

0.310

0.315

:

1.026

1.269

1.498

1.826

67.7

:

0.265

0.262

0.262

0.262

:

0.287

0.367

0.436

0.533

59.2

58.7

:

0.243

0.237

0.236

0.233

:

0.095

0.132

0.159

0.198

65.5

65.5

65.7

:

0.272

0.272

0.272

0.273

:

0.021

0.020

0.019

0.018

56.0

54.6

54.3

53.5

:

0.227

0.221

0.219

0.215

:

0.055

0.083

0.095

0.120

:

69.6

69.9

70.3

70.8

:

0.286

0.287

0.288

0.290

:

0.018

0.021

0.024

0.029

:

56.3

56.3

56.6

57.3

:

0.228

0.228

0.229

0.232

:

0.056

0.058

0.059

0.053

:

63.5

63.6

64.1

64.3

:

0.259

0.260

0.261

0.262

:

0.008

0.010

0.012

0.015

:

79.3

81.1

83.0

85.1

:

0.311

0.316

0.322

0.329

:

0.496

0.606

0.716

0.869

0.0143

:

98.4

102.3

106.5

112.3

:

0.370

0.381

0.394

0.410

:

2.060

2.518

3.016

3.823

0.0194

0.0204

:

88.6

91.8

93.6

98.2

:

0.354

0.365

0.371

0.388

:

0.429

0.545

0.637

0.808

0.0194

0.0194

0.0204

:

89.2

92.2

93.7

98.4

:

0.356

0.367

0.371

0.389

:

0.458

0.563

0.643

0.819

-

-

-

:

90.7

90.8

90.3

90.7

:

0.319

0.320

0.318

0.320

:

0.080

0.078

0.083

0.079

TE D

M AN U

72.7

65°C

Specific Exergy, ψ (kJ/kg)

0.295

EP

:

65°C

AC C

25

60°C

SC

60°C

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37

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Figures and Tables Caption

Schematic diagram of air-conditioning system.

the

developed

desiccant-evaporative

Fig. 2:

Actual view of the experimental set-up for the testing of the developed desiccant-evaporative air-conditioning system: (a) installed falling film

RI PT

Fig. 1:

SC

water evaporative cooler; (b) installed air fans for both outdoor air (OAF) and exit air (EAF); (c) complete system in the experimental chamber, and; (d) data gathering and monitoring system. States of air inside the developed desiccant-evaporative air-conditioning system in the psychrometric chart for different regeneration temperatures (RT).

Fig. 4:

Outdoor air fan (OAF) exergetic results: (a) inlet exergy rate with respect to different reference temperatures (RT) and regeneration temperatures (RegT); (b) exergy destruction rate with respect to different reference temperature and regeneration temperature (RegT); (c) exergetic efficiency with respect to different reference temperatures (RT) and regeneration temperatures (RegT), and; (d) improvement potential with respect to different reference temperatures (RT) and regeneration temperatures (RegT).

Fig. 5:

Exit air fan (EAF) exergetic results: (a) inlet exergy rate with respect to different reference temperatures (RT) and regeneration temperatures (RegT); (b) exergy destruction rate with respect to different reference temperatures (RT) and regeneration temperatures (RegT); (c) exergetic

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Fig. 3:

efficiency with respect to different reference temperatures (RT) and regeneration temperatures (RegT), and; (d) improvement potential with respect to different reference temperatures (RT) and regeneration temperatures (RegT).

Fig. 6:

Air heating coil (HC) exergetic results: (a) inlet exergy rate with respect to different reference temperatures (RT) and regeneration temperatures (RegT); (b) exergy destruction rate with respect to different reference temperatures (RT) and regeneration temperatures 38

ACCEPTED MANUSCRIPT

(RegT); (c) exergetic efficiency with respect to different reference temperatures (RT) and regeneration temperatures (RegT), and; (d)

Fig. 7:

RI PT

improvement potential with respect to different reference temperatures (RT) and regeneration temperatures (RegT). Desiccant wheel (DW) exergetic results: (a) inlet exergy rate with respect to different reference temperatures (RT) and regeneration temperatures (RegT); (b) exergy destruction rate with respect to different reference temperatures (RT) and regeneration temperatures

M AN U

SC

(RegT); (c) exergetic efficiency with respect to different reference temperatures (RT) and regeneration temperatures (RegT), and; (d) improvement potential with respect to different reference temperatures (RT) and regeneration temperatures (RegT).

Primary heat exchanger (HX1) exergetic results: (a) inlet exergy rate with respect to different reference temperatures (RT) and regeneration temperatures (RegT); (b) exergy destruction rate with respect to different reference temperatures (RT) and regeneration temperatures (RegT); (c) exergetic efficiency with respect to different reference temperatures (RT) and regeneration temperatures (RegT), and; (d) improvement potential with respect to different reference temperatures (RT) and regeneration temperatures (RegT).

Fig. 9:

Secondary heat exchanger (HX2) exergetic results: (a) inlet exergy rate with respect to different reference temperatures (RT) and regeneration temperatures (RegT); (b) exergy destruction rate with respect to different reference temperatures (RT) and regeneration temperatures

EP

TE D

Fig. 8:

(RegT); (c) exergetic efficiency with respect to different reference

AC C

temperatures (RT) and regeneration temperatures (RegT), and; (d) improvement potential with respect to different reference temperatures (RT) and regeneration temperatures (RegT).

Fig. 10:

Direct evaporative cooler (EC) exergetic results: (a) inlet exergy rate with respect to different reference temperatures (RT) and regeneration temperatures (RegT); (b) exergy destruction rate with respect to different reference temperatures (RT) and regeneration temperatures (RegT); (c) exergetic efficiency with respect to different reference temperatures (RT) and regeneration temperatures (RegT), and; (d) 39

ACCEPTED MANUSCRIPT

improvement potential with respect to different reference temperatures (RT) and regeneration temperatures (RegT). Percentage contribution of exergy destruction for the developed desiccant-evaporative air-conditioning system components at different regeneration temperatures (RegT) and reference temperatures (RT): (a) 0oC RT; (b) 5oC RT; (c) 10oC RT; (d) 15oC RT; (e) 20oC RT; (f) 25oC RT, and; (g) 30oC RT.

Fig. 12:

Exergy destruction coefficient for the developed desiccant-evaporative air-conditioning system components at different regeneration temperatures (RegT) and reference temperatures (RT): (a) 0oC RT; (b)

SC

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Fig. 11:

M AN U

5oC RT; (c) 10oC RT; (d) 15oC RT; (e) 20oC RT; (f) 25oC RT, and; (a) 30oC RT. Performances of the developed desiccant-evaporative air-conditioning system at different regeneration temperatures (RegT) and reference temperatures (RT): (a) thermal exergetic COP with respect to different reference temperatures (RT) and regeneration temperatures (RegT); (b) electric exergetic COP with respect to different reference temperatures (RT) and regeneration temperatures (RegT); (c) system exergetic COP with respect to different reference temperatures (RT) and regeneration temperatures (RegT), and; (d) sustainability index with respect to different reference temperatures (RT) and regeneration temperatures (RegT).

Fig. 14:

Performances of the developed desiccant-evaporative air-conditioning system at different regeneration temperatures (RegT) and reference

EP

TE D

Fig. 13:

AC C

temperatures (RT): (a) reversible Carnot coefficient of performance (COP)Rev with respect to different reference temperatures (RT) and regeneration temperatures (RegT); (b) actual coefficient of performance (COP)Act with respect to different reference temperatures (RT) and

regeneration temperatures (RegT); (c) Carnot coefficient of performance (COP)C with respect to different reference temperatures (RT) and regeneration temperatures (RegT); (d) Exergetic efficiency I (First Definition), and; (d) Exergetic efficiency II (Second Definition).

40

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Table 1:

Measured air flow rates for different regeneration temperature (RT) during the experimental evaluation of the developed desiccant-evaporative air-conditioning system. Air states and equivalent specific enthalpy, specific entropy and specific exergy for different regeneration temperatures (RegT) for the case of 15oC reference temperatures (RT).

AC C

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Table 2:

41

ACCEPTED MANUSCRIPT International Journal of Refrigeration

Highlights

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Exergetic analyses of the developed desiccant-evaporative air-conditioning system
Exergetic analyses based on different regeneration temperatures
Exergetic performances for the system and components are shown

1