Applied Thermal Engineering - Bu Benha/Mechanical... · desalination system operated with solar heat pump. ... and experimental investigation for various HDH solar ... researches

Applied Thermal Engineering 108 (2016) 1320–1335

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Applied Thermal Engineering

journal homepage: www.elsevier .com/locate /apthermeng

Research Paper

An integrated A/C and HDH water desalination system assisted by solarenergy: Transient analysis and economical study

http://dx.doi.org/10.1016/j.applthermaleng.2016.08.0261359-4311/� 2016 Elsevier Ltd. All rights reserved.

⇑ Corresponding author.E-mail address: [email protected] (A. Fouda).

A. Fouda a,⇑, S.A. Nada b, H.F. Elattar b

aDepartment of Mechanical Power Engineering, Faculty of Engineering, Mansoura University, 35516 El-Mansoura, EgyptbDepartment of Mechanical Engineering, Benha Faculty of Engineering, Benha University, Benha, 13511 Qalyubia, Egypt

h i g h l i g h t s

� Hybrid A/C and HDH desalination system solar assisted are proposed.� Transient analysis and parametrical study for system performance are presented.� Effects of operating and design parameters on the proposed system is investigated.� System capacity, consumption, performance and cost saving analysis are hourly and daily investigated and presented.� Proposed system operates more efficient in more hot and humid climatic conditions.

a r t i c l e i n f o

Article history:Received 1 June 2016Revised 31 July 2016Accepted 3 August 2016Available online 4 August 2016

Keywords:Integrated systemSolar energyA/CHDHHot and humid climates

a b s t r a c t

Theoretical investigation on the performance of a proposed innovated solar integrated system for air con-ditioning (A/C) and humidification dehumidification (HDH) water desalination in hot and humid regionsis presented. Transient analysis and parametrical study for the system are conducted under differentoperating and design conditions. System performance parameters (fresh water production rate, coolingcapacity, electrical power consumption, water recovery and system coefficient of performance) arehourly estimated and daily integrated for different operational and design conditions. The results revealthe increase of fresh water productivity, system water recovery, cooling capacity, electrical power con-sumption and COP of the system with the increase of air temperature, air humidity, and solar collectorsarea. For system assessment, the system performance parameters are compared with the performance ofa basic system under the same conditions. Cost saving analysis showed the proposed system operatesmore economically and efficiently that in hot and humid climatic areas. General numerical correlationsfor system performance parameters are proposed in terms of the design and operational conditions.

� 2016 Elsevier Ltd. All rights reserved.

1. Introduction

Air conditioning and fresh water are necessary for life perma-nency and nations development including agricultures and indus-tries activities. Fresh water and fossil fuel sources are dramaticallydepleting due to the rapid increase in world population and thedemands of recent life activities. Water desalination can be consid-ered as a solution of fresh water sources depletion but it consumesa considerable amount of energy. Moreover, air conditioning sys-tems consume considerable electrical energy specially in hot andhumid areas. The use of fossil fuel for electricity generation anddrive water desalination and air conditioning systems is not sus-tainable as they harm environment and accelerate fuel depletion.

Using renewable energy sources especially solar energy can beconsidered as a solution of these problems. Recently, integratedA/C and water desalination system using HDH approach derivedby mechanical vapor compression (MVC) and solar energy are pro-posed as an efficient solution for fossil fuel saving, fresh water pro-duction and environment protection. HDH desalination and A/Chybrid system is simple in design, construction and operationand appropriate for developing countries which suffer form energyand fresh water lacks problems. A broad survey on uses renewableenergy-driven technologies of integrated desalination systems waspresented by Ghaffour et al. [1]. A small scale solar HDH systemusing glass evacuated tubes solar air heater was studied and pre-sented by Li et al. [2]. Yıldırım and Solmus [3] theoretically inves-tigated the performance of HDH water desalination system drivenby solar energy for Antalya, Turkey under different design andoperational parameters. El-Agouz et al. [4] designed and

Nomenclature

A area, m2

A/C air conditioningADP apparatus dew pointCOP coefficient of performanceCp specific heat, kJ/kg KEES engineering equations solverE� total electric power, kWEps,day daily total electric energy, kW hFAR fresh air ratioFWST fresh water storage tankL latitude angle, �H hour angle, �HDH humidification-dehumidificationhfg water latent heat of evaporation, kJ/kgi specific enthalpy, kJ/kgIT total solar intensity, W/m2

m� mass flow rate, kg/smw,day daily fresh water production, kgQ�cc cooling coil capacity, kW

Q�R building cooling load, kW

QR,day daily building cooling energy, kW hR system water recoveryRSHF room sensible heat factorTOCS total operating cost savingPs saturation pressure, Pat temperature, �CW absolute humidity, kgv/kgaW�

c compressor power, kW

Greek symbolsb tilt angle, �d declination angle, �g efficiencyw arbitrary variable

Subscripta airatm atmosphereBS basic systemcc cooling coilcond condensatecw cold waterdeh dehumidifierh humidifieri inletinp inputm meano outletR room conditionRef refrigerationRe recirculateRH relative humidityPS proposed systemS supplied state of conditioned spaces saturation stateSC solar collectorv water vaporw water

A. Fouda et al. / Applied Thermal Engineering 108 (2016) 1320–1335 1321

constructed a pilot plant in a dry area using solar water collector toincrease fresh water production from salty water. Sharon andReddy [5] performed a review for various desalination units drivenby renewable energy particularly solar energy. Kabeel and El-Said[6] investigated and compared various configurations of hybriddesalination systems based on air HDH–single stage flashing andpowered by solar energy. Shatat et al. [7] presented theoreticalstudy on small scale HDH water desalination unit integrated withan evacuated tube solar collector.

Kumar et al. [8] developed thermal model to predict the perfor-mance of a single slope solar still integrated with evacuated tubecollector under New Delhi (India) climatic conditions. Liu et al.[9] presented thermal and economic analyses for water desalina-tion system using evacuated tube solar collectors. Al-Sulaimanet al. [10] presented thermodynamic analysis to evaluate the per-formance of HDH system integrated with parabolic trough solarcollector (PTSC). Hawlader [11] studied the influences of supplywater temperature and flashing on the performance of a newdesalination system operated with solar heat pump. Yuan et al.[12] proposed an integrated system for A/C and water desalinationusing humidification-dehumidification process and assisted byMVC heat pump. An open air MVC refrigeration unit for hybridA/C and water desalination system for ship was presented by Houaet al. [13]. Habeebullah [14] conducted experimentally the perfor-mance analysis for an integrated heat pump with a dehumidifica-tion technique for fresh water producing from atmospheric air. Theunit serve an office building with 250 m2 area and air supplycapacity of 1.586 m3/s, the building located on the sea in Jeddahcity, KSA, where the mean atmospheric temperature and relativehumidity are 34 �C and of 71%, respectively. Jain and Hindoliy[15] conducted performance study for two different pad materialsused in evaporative cooling (coconut and plash fibers) and com-

pared their performance with the traditional aspen and khus pads.Halima et al. [16] studied theoretically the performance of a simplesolar still integrated with a heat pump. Al-Enezi et al. [17] exam-ined and investigated the performance of water desalination sys-tem using HDH technique at low operational temperatures. Themaximum fresh water rate was observed at elevated hot waterand low cooling water temperatures and at high air and low hotwater mass flow rates. Nafey et al. [18] performed theoreticaland experimental studies for solar water desalination unit by uti-lizing flashing process on a small scale unit. The fresh water pro-ductivity was calculated by developing a mathematical modelunder various operating conditions. The mean summer water pro-ductivity obtained varied from 5.44 to 7 kg/day/m2 in July andAugust and from 4.2 to 5 kg/day/m2 in June. Mohamed andEl-Minshawy [19] presented simulation modeling to assess theperformance of seawater HDH water desalination system drivenby solar energy. The model was built with comparison study todisplay the influence of the various operating conditions on thesystem performance and fresh water production. Ghazal et al.[20] constructed an experimental setup to enhance the solar Humidification-Dehumidification Desalination (HDD) systems perfor-mance. The air and water solar heaters and the traditional HDDevaporator were substituted with compact system. It was observedthat a direct contact bubbling humidification is an efficient methodin HDD systems. Dayem and Fatouh [21] accomplished theoreticaland experimental investigation for various HDH solar assistedwater desalination systems to present the best efficient system.Three systems were suggested and compared for climatologicalenvironments of Cairo (30�N). Hermosillo et al. [22] presented the-oretical and experimental work on a novel suggested HDH desali-nation system. The evaporator was manufactured from treatedcellulose paper in which water was evaporated. He et al. [23]

1322 A. Fouda et al. / Applied Thermal Engineering 108 (2016) 1320–1335

carried out a parametrical study on HDH water desalination sys-tem utilizing low grade heat sources for seawater heating priorthe humidifier. The desalination system performance and the lowgrade heat collector type were studied at various operating pres-sures. Chiranjeevi and Srinivas [24] carried out experimental andsimulation study to develop and analyze two stages HDH desalina-tion and cooling hybrid system to prove the probability of improv-ing the desalination yield with supplemented cooling support.Giwa et al. [25] investigated HDH desalination system poweredby photovoltaic system for fresh water and electricity productionin climatic conditions of Abu Dhabi, UAE. Buker et al. [26] devel-oped and tested pilot scale experimental set-up for to presenteffect of the different parameters such as weather condition, airflow and regeneration temperature on the proposed system perfor-mance. Using photovoltaic/thermal roof collector combined with aliquid desiccant. The results confirmed the potential of the exam-ined technology, and explained the specific conclusions for thepractice of such systems. Nada et al. [27] achieved an experimentalstudy to investigate the performance of an integrated HDH waterdesalination and A/C system using vapor compression refrigerationunit. Recently, Nada et al. [28] studied theoretically the perfor-mance of integrated A/C and water desalination systems usingHDH technique suggested for hot and dry weather areas. Four sys-tems with evaporative cooler and heat recovery units wereemployed, analyzed and assessed at various operating conditions.Elattar et al. [29] presented parametrical and economical studyof the performance of integrative A/C and HDH water desalinationsystem assisted by solar energy. The system performance wasinvestigated under steady-state operation utilizing solar thermalstorage and auxiliary heating units.

According to the above literature and the authors’ review,researches on combined solar-assisted systems driven by MVCchiller for A/C and HDH water desalination are still needed toinvestigate the effects of operational and design system parame-ters on the system performance. Energy saving, fresh water recov-ery, human thermal comfort issues and numerical correlations forsystem performance parameters are not addressed in the previousresearches. Therefore, the goal of the present work is a perfor-mance investigation of a combined A/C and HDH water desalina-

Fresh air,O

Mixed air,M

Return aR

Exhaust air,R

detalucric-eR

R,ria

ma,o•

ma,Re•

Coolingcoil

Air cooled chiller

Chemicaltreatment un

mw,fresh•

FWST

Freshwater

Fig. 1. Basic syste

tion system assisted by solar energy. A new configuration of thesystem is presented. The system performance parameters: systemcapacity, consumptions and performance parameters are instanta-neously and daily integrated investigated for different operationaland design system parameters (solar intensity, ambient air tem-perature, ambient absolute humidity, ambient air flow rate, freshair ratio, solar collectors’ areas, dehumidifier cooling water massflow rate, humidifier water mass flow rate and dehumidifier cool-ing water temperature). The system is tested under the conditionsof hot and humid climatic cities; namely under the meteorologicaldata of Jeddah city in KSA. Hourly and daily performance of thesystem are analyzed at various operational and design systemparameters. The modeling of the system performance is developedbased on heat and mass balances and solved by using EngineeringEquation Solver (EES) software and C++ programming language.For system assessment, the performance and productivity of thepresented system are compared with the basic air conditioningsystem. Cost saving analysis is also presented in the form of totaloperating cost saving (TOCS) due fresh water productivity andsaving in electric power consumption. General numerical correla-tions for hourly fresh water productivity, space cooling load andtotal electrical power consumption for the proposed system arecorrelated as a function in the system design and operationalparameters.

2. Systems description

Fig. 1 illustrates a schematic diagram for a basic air conditioningcentral system which is used as a reference system to assess theperformance of the proposed hybrid system. The basic system con-sists of mainly cooling/dehumidifying coil and air cooled chillerunit to support the air conditioning of a building. A water chemicaltreatment unit is used to treat the condensate water to be freshwater. Ambient air (O) is mixed with recirculated air (R). Themixed air (M) is then passed on a cooling coil where it is cooledand dehumidified (S). The condensate water passes through achemical water treatment unit where it is treated to be freshwater.

Supply air,S

ir,

ma,s•

unit

CompressorPower, Wc

•

Conditionedspace

it

m schematic.

A. Fouda et al. / Applied Thermal Engineering 108 (2016) 1320–1335 1323

The schematic diagram of the proposed hybrid integrated sys-tem is given in Fig. 2. The main system components are: air andwater solar collectors, humidifier, dehumidifier, air cooled chillerunit and chemical water treatment units. Air is drawn from ambi-

Humidifier

Circulatingpump

Wate

r sola

r coll

ector

Treatmentunit

Raw water(brackish, sea, etc.)

Chetreatm

Dehumidifier

mmakeup•

mw,fresh•

O

L H

mw,h•

mcw,deh•

ma,o•

mcond,deh•

FW

Freshwater

Blo

wdo

wn

wat

erAir s

olar c

ollec

tor

Water solar desalination system

Z

Fig. 2. Proposed system: (a) schem

ent (O) to flow into air solar collector, where it is sensibly heated(L). The air is then humidified in the humidifier (H) by sprayingthe hot water that received from the water solar collector. TheAir and water solar collectors are used to rise the air humidification

micalent unit

Coolingcoil

Air cooled chiller unit

M S

RR

Wc•

ma,Re•

K

Conditionedspace

ma,s•v2

Oma,o

•

mcond,cc•

R

ST

v1 v3

Air conditioning system

(a)

(b) Basic system Proposed system

atic, (b) psychrometric cycles.

6 8 10 12 14 16 18 20Day time (hour)

0

400

800

1200

1600

Sol

ar ra

diat

ion

inte

nsity

(w/m

2 )

Jeddah, KSA21 July

β= 0o

500

550

600

650

700

750

800

850

900

Dai

ly a

vera

ge s

olar

inte

nsity

(W/m

2 )

β= 0 o

β= 21.5 o

Jeddah city, KSA

(a)

1324 A. Fouda et al. / Applied Thermal Engineering 108 (2016) 1320–1335

capacity by elevating the air and sprayed water temperatures. Thehumidified air is then cooled and dehumidified (K) through thedehumidifier. The water condensate on the dehumidifier surfaceare drained to the water treatment section to produce fresh water.The dehumidified air is then mixed with the recirculated air (R).The mixed air (M) is then passes through the cooling coil of theair condition system to be cooled and dehumidified (S) producingadditional fresh water. Chilled water comes from air cooled chillerunit passes in the cooling coil unit to maintain coil apparatus dewpoint (ADP) of 7 �C. Finally, the cooled and dehumidified air (S) issupplied to the conditioned space to remove the cooling load andcool the space. Fig. 2b displays the psychrometric processes of aircycles for basic and proposed systems.

In the humidifier, brackish/sea water is treated in the treatmentunit and is then pumped and sprayed in the humidifier after heat-ing it in the solar water collector. Part of the sprayed water is evap-orated and carried up by the humidified air and the remainingamount is recirculated and mixed with make-up water (sea/brackish water). In the dehumidifier, a cooling water (sea/brackish)with temperature lower than dew point temperature of the air atpoint (H) is used to dehumidify the air. The condensate water isthen drawn and flowed into chemical treatment unit and finallykept in fresh water storage tank (FWST).

The suggested integrated system may be operated with threevarious manners; water desalination system only, A/C system only,and combined A/C and water desalination system. In case of usingit as water desalination system only, the system works under shut-ting valve V2 and opening valve V1 to exclude the A/C cycle forwhole system.While, in case of using it as A/C system only, the sys-tem is operated by closing the valve V2 and open the valve V3 toexclude the water desalination cycle from whole system. In caseof a combined A/C and water desalination system both A/C anddesalination cycles are operating in the same time by closingvalves V1 & V3 and opening valve V2. The current study is investi-gated under combined modes using fully solar utilization in desali-nation system during daylight hours.

30 60 90 120 150 180 210 240 270 300 330 360Days of year

400

450β= 45 o

30 60 90 120 150 180 210 240 270 300 330 360Days of year

18

20

22

24

26

28

30

32

34

36

Dai

ly in

tegr

ated

sol

ar e

nerg

y (M

J/da

y)

β= 0 o

β= 21.5 o

β= 45 o

Jeddah city, KSA

(c)

(b)

Fig. 3. Solar radiation intensity: (a) solar intensity during the daylight, (b) dailyaverage solar intensity during a year, (c) daily integrated solar energy during a year.

3. Thermodynamic analysis and numerical modeling

A numerical model for the presented system is developed usingenergy and mass balance equations for the system components.The model is developed based on the following assumptions:

� Leakage (air/water) in the system components are neglected.� The water temperature leaving the humidifier is similar to airwet-bulb temperature leaving it.

� Humidifier efficiency is 100% (i.e. air is leaving the humidifier assaturated air).

� The process through the dehumidifiers is along the saturationcurve.

� The produced fresh water, cooling water and air dry-bulb tem-peratures are equal at the dehumidifier exits.

� Specific power of auxiliary components (power consumed byfans and pumps per unit amount of fresh water rate) is assumedas 0.029 kW/(m3

w,fesh/h) [30,31].� Water vapor at the humidifier outlet is free salt [32], thereforethe condensate water need a chemical treatment to be fresh/potable water with salinity 500 ppm according to the WHO(World Health Organization) as a recommended value.

� Thermal comfort conditions of the conditioned space areTdb,R = 24 �C and RH = 50%.

� Cooling coil surface temperature and RSHF are chosen as 7 �Cand 0.9, respectively as typical values for hot and humidclimates.

A. Fouda et al. / Applied Thermal Engineering 108 (2016) 1320–1335 1325

Fresh water productivity, system cooling capacity, system elec-trical power consumption, system water recovery and entire sys-tem coefficient of performance are expressed as a function in theoperating and design system parameters.

3.1. Solar collectors

Solar radiation is a main parameter in the proposed hybrid solarsystem. The hourly total solar radiation incident on the air andwater solar collectors are calculated according to the local geo-graphical coordinates of Jeddah city, KSA. The total solar radiationIT on a tilted surface is calculated as the sum of direct, diffuse andreflected radiations as follows:

IT ¼ Ib þ Id þ Ir ð1Þ

where Ib, Id and Ir are the direct, diffuse and reflected components ofthe solar radiation, respectively. IT is also presented by Duffie andBeckman [33] as follows.

IT ¼ IbRb þ Id1þ cosb

2þ ðIb þ IdÞqg

1� cosb2

ð2Þ

where ground reflectance qg equals 0.2 and Rb (tilt factor of beamradiation) for the specific case of a south-facing fixed surface witha tilt angle of b, is

6 8 10 12 14 16 18 20

Daylight time (hour)

6

8

10

12

14

16

18

20

22

24

26

m• w

,fres

h (kg

/h)

wa,o= 0.02 kgv/kga, FAR=0.25 tcw,deh,i= 20 oC, Aasc=10 m2, Awsc=10 m2

m•w,fresh (kg/h)

RPS

0.5

0.55

0.6

0.65

0.7

0.75

0.8

0.85

0.9

RP

S

ta,o= 30 oC

40 oC

50 oC

ta,o= 30 oC

40 oC

50 oC

6 8 10 12 14 16 18 20

Daylight time (hour)

0

2

4

6

8

10

E• P

S (k

W)

wa,o= 0.02 kgv/kga, FAR=0.25 tcw,deh,i= 20 oC, Aasc=10 m2, Awsc=10 m2

E•PS (kW)

COPPS

1

2

3

4

5

6

CO

PP

S

ta,o= 30 oC

40 oC

50 oC

ta,o= 30 oC

40 oC

50 oC

(a)

(c)

Fig. 4. Influences of ambient air dry bulb temperature on the proposed system perfo

Rb ¼ sinðL� bÞ sinðdÞ þ cosðL� bÞ cosðdÞ cosðhÞsinðdÞ sinðLÞ þ cosðdÞ cosðhÞ cosðLÞ ð3Þ

where L, d, and h are the latitude, declination and hour angles,respectively. Detailed calculation procedures for the IT and Rb aregiven by Duffie and Beckman [33]. According to the preceding equa-tions the solar radiation intensity, daily average and integratedsolar radiation intensity and energy are calculated and illustratedin Fig. 3a–c, respectively for Jeddah city. As shown in Fig. 3b andc, the optimum path of solar intensity during the year can beobtained from intersected lines of different tilt angles.

The calculated solar intensity is related to the temperature riseby the solar collector efficiency equation as suggested by Yuan andZhang [34].

gSC ¼ m�cpðtSC;o � tSC;iÞITASC

ð4Þ

The following empirical correlations for solar collector effi-ciency which was presented by [1,35] are used with Eq. (4) to cal-culate air and water temperatures downstream the solar collectors.

gw;SC ¼ a1 þ a2tSC;i � tw;o

ITð5Þ

ga;SC ¼ a1 þ a2tm � ta;o

ITð6Þ

where a1 = 0.37, a2 = �3.35 and tm ¼ tSC;oþtSC;i2 .

6 8 10 12 14 16 18 20

Daylight time (hour)

10

12

14

16

18

t a,S (

o C)

wa,o= 0.02 kgv/kga, FAR=0.25 tcw,deh,i= 20 oC, Aasc=10 m2, Awsc=10 m2

ta,S (oC)RHa,S

0.8

0.84

0.88

0.92

0.96

1

RH

a,S

ta,o= 30 oC

40 oC

50 oC

ta,o= 30 oC

40 oC

50 oC

6 8 10 12 14 16 18 20

Daylight time (hour)

5

6

7

8

9

10

Q• R

(kW

)

wa,o= 0.02 kgv/kga, FAR=0.25 tcw,deh,i= 20 oC, Aasc=10 m2, Awsc=10 m2

Q•R (kW)

Q•cc (kW)

8

10

12

14

16

18

20

Q• cc

(kW

)

ta,o= 30 oC

40 oC

50 oC

ta,o= 30 oC

40 oC

50 oC

(d)

(b)

rmance parameters: (a) m�w,fresh & R, (b) Q�

R & Q�cc, (c) E�ps & COPPS, (d) ta,s & RHa,s.

1326 A. Fouda et al. / Applied Thermal Engineering 108 (2016) 1320–1335

3.2. Humidifier

Energy balance of the humidifier is

m�a;oðia;h;o � ia;h;iÞ ¼ m�

w;h;icp;wtw;h;i �m�w;h;ocp;wtw;h;o ð7Þ

The absolute humidity of the inlet air to solar air collector isgiven by

ws ¼ 0:622RHPs

Patm � RHPsð8Þ

where Ps is the water vapor saturation pressure and is given by[36]:

Ps ¼ exp 23:196� 3816:44Ta � 46:13

� �ð9Þ

The air specific enthalpy is calculated by summing of the sensi-ble and latent parts as follows:

ia ¼ cp;ata þwð2500þ 1:84taÞ ð10ÞThe mass flow rate of the makeup water (sea or brackish) sup-

plied to the system is given by:

m�makeup ¼ m�

a;oðwa;h;o �wa;h;iÞ ð11Þ

6 8 10 12 14 16 18 20

Daylight time (hour)

0

10

20

30

40

m• w

,fres

h (kg

/h)

ta,o= 40 oC, FAR=0.25, tcw,deh,i= 20 oC, Aasc=10 m2, Awsc=10 m2

m•w,fresh (kg/h)

RPS

0.5

0.6

0.7

0.8

0.9

1

RP

S

0.05

wa,o= 0.01 kgv/kga

0.03 0.03

0.05

0.01

6 8 10 12 14 16 18 20Daylight time (hour)

0

2

4

6

8

10

E• P

S (k

W)

ta,o= 40 oC, FAR=0.25, tcw,deh,i= 20 oC, Aasc=10 m2, Awsc=10 m2

E•PS (kW)

COPPS

3.4

3.6

3.8

4

CO

P PS

wa,o= 0.01 kgv/kga

0.03

0.05

(a)

(c)

Fig. 5. Influences of the ambient humidity on the proposed system performanc

3.3. Dehumidifier

Energy balance of the dehumidifier can be expressed as.

m�a;oðia;deh;i � ia;deh;oÞ ¼ m�

cw;dehcp;wðtcw;deh;o � tcw;deh;iÞþm�

cond;dehcp;wtcond;deh ð12ÞThe condensed water flow rate leaving from the dehumidifier

can be calculated by

m�cond;deh ¼ m�

a;oðwa;deh;i �wa;deh;oÞ ð13Þ

3.4. Air conditioning system

The air conditioning system capacity (space cooling load andcooling coil load), room sensible heat factor, and condensate waterover the cooling coil can be calculated from

Q �R ¼ m�

a;sðia;R � ia;SÞ ð14Þ

Q�c:c ¼ m�

a;sðia;cc;i � ia;cc;oÞ ð15Þ

RSHF ¼ Cpaðta;R � ta;sÞðia;R � ia;sÞ ð16Þ

6 8 10 12 14 16 18 20

Daylight time (hour)

5

6

7

8

9

10

Q• R

(kW

)

ta,o= 40 oC, FAR=0.25, tcw,deh,i= 20 oC, Aasc=10 m2, Awsc=10 m2

Q•R (kW)

Q•cc (kW)

0

5

10

15

20

25

30

Q• cc

(kW

)

wa,o= 0.01 kgv/kga

0.03

0.05

6 8 10 12 14 16 18 20Daylight time (hour)

10

12

14

16

18

t a,S (o C

)

ta,o= 40 oC, FAR=0.25, tcw,deh,i= 20 oC, Aasc=10 m2, Awsc=10 m2

ta,S (oC)RHa,S

0.6

0.8

1

1.2

RH

a,S

wa,o= 0.01 kgv/kga

0.03

0.05

0.01

0.03

0.05

(d)

(b)

e parameters: (a) m�w,fresh & R, (b) Q�

R & Q�cc, (c) E�ps & COPPS, (d) ta,s & RHa,s.

A. Fouda et al. / Applied Thermal Engineering 108 (2016) 1320–1335 1327

m�cond;cc ¼ m�

a;sðwa;cc;i �wa;cc;oÞ ð17Þwhere m�

a;s ¼ m�a;o þm�

a;Re.Refrigeration unit performance coefficient in terms of ambient

air temperature can be calculated by [37]:

COPRef ¼ Q �c:c

W�c

ð18Þ

COPRef ¼ 11:1� 0:4ta;o þ 7:2� 10�3t2a;o � 5:18� 10�5t3a;o ð19ÞThe total fresh water rate and water recovery for basic and pro-

posed system are calculated from

m�w;fresh;BS ¼ m�

cond;cc ð20Þ

m�w;fresh;PS ¼ m�

cond;cc þm�cond;deh ð21Þ

RBS ¼m�

w;fresh;BS

m�a;owa;o

ð22Þ

RPS ¼m�

w;fresh;PS

m�makeup þm�

a;owa;oð23Þ

6 8 10 12 14 16 18 20Daylight time (hour)

5

10

15

20

25

30

35

m• w

,fres

h (kg

/h)

ta,o= 40 oC, wa,o= 0.02 kgv/kga, FAR=0.25, tcw,deh,i= 20 oC, Aasc=10 m2, Awsc=10 m2,

m•w,fresh (kg/h)

RPS

0.65

0.7

0.75

0.8

0.85

0.9

RP

S

m•a,o=0.1 kg/s

0.15

0.2

0.1

0.150.2

6 8 10 12 14 16 18 20

Daylight time (hour)

0

2

4

6

8

10

E• P

S (k

W)

ta,o= 40 oC, wa,o= 0.02 kgv/kga, FAR=0.25, tcw,deh,i= 20 oC, Aasc=10 m2, Awsc=10 m2,

E•PS (kW)

COPPS

3.2

3.6

4

4.4

4.8

CO

PP

S

m•a,o=0.1 kg/s

0.1

0.15

0.2

0.15

0.2

(a)

(c)

Fig. 6. Influences of the ambient air flow rate on the proposed system performa

Coefficient of Performance for basic and hybrid proposed sys-tems are defined by the following equations:

COPBS ¼Q �

R þm�cond;cchfg

E�Bs

ð24Þ

COPPS ¼Q �

R þm�w;freshhfg

E�ps

ð25Þ

The total power consumption for basic and proposed systemscan be given as follows:

E�Bs ¼ W�

c ð26Þ

E�ps ¼ W�

c þ E�aux ð27Þ

3.5. Operating and system design parameters

The system’s capacity, consumption and performance parame-ters are: m�

w,fresh, Q�R, Q�

cc, E�ps, R, ta,s, RHa,s and COP are determinedas a function in all the operational and design parameters usingEqs. (1)–(27). The operating and design parameters are varied inthe following range for the sake of the parametric study:

6 8 10 12 14 16 18 20

Daylight time (hour)

0

2

4

6

8

10

12

14

Q• R

(kW

)

ta,o= 40 oC, wa,o= 0.02 kgv/kga, FAR=0.25, tcw,deh,i= 20 oC, Aasc=10 m2, Awsc=10 m2,

Q•R (kW)

Q•cc (kW)

0

5

10

15

20

25

Q• cc

(kW

)

m•a,o=0.1 kg/s

0.15

0.2

0.1

0.15

0.2

6 8 10 12 14 16 18 20

Daylight time (hour)

12

13

14

15

16

t a,S (o C

)

ta,o= 40 oC, wa,o= 0.02 kgv/kga, FAR=0.25, tcw,deh,i= 20 oC, Aasc=10 m2, Awsc=10 m2,

ta,S (oC)RHa,S

0.8

0.84

0.88

0.92

0.96

1

RH

a,S

m•a,o=0.1 kg/s

0.15 0.2

0.1

0.15 0.2

(d)

(b)

nce parameters: (a) m�w,fresh & R, (b) Q�

R & Q�cc, (c) E�ps & COPPS, (d) ta,s & RHa,s.

1328 A. Fouda et al. / Applied Thermal Engineering 108 (2016) 1320–1335

� ta,o = 30–50 �C (ambient air temperature)� wa,o = 0.01–0.05 kgv/kga (ambient air absolute humidity)� m�

a,o = 0.1–0.2 kg/s (air mass flow rate)� Asc = 5–15 m2 (air and water solar collectors area)� FAR = 0.1, 0.25, and 0.5 (fresh air ratio)� m�

cw,deh = 0.1–0.2 kg/s (dehumidifier cooling water mass flowrate)

� m�w,h = 0.1–0.2 kg/s (humidifier water mass flow rate)

� tcw,deh,i = 15–25 �C (dehumidifier cooling water temperature)

The air specific humilities and enthalpies that illustrated in theprevious equations are calculated from air properties. The set ofequations from (1)–(27) are solved numerically in the same timeto conserve mass and energy for the hybrid presented system usingEES software and C++ programming language.

4. Results and discussions

Hourly analysis of the system performance is investigated topresent the effect of the operational and design parameters onthe system performance. Finally, cost saving analysis and perfor-mance comparison study for the presented hybrid system withthe basic system are conducted.

6 8 10 12 14 16 18 20

Daylight time (hour)

2

3

4

5

6

7

8

E• P

S (k

W)

ta,o= 40 oC, wa,o= 0.02 kgv/kga,FAR=0.25, tcw,deh,i= 20 oC

E•PS (kW)

COPPS

2

3

4

5

CO

P PS

Aasc=Awsc=5 m2

10 m2

15 m2

5 m2

10 m2

15 m2

6 8 10 12 14 16 18 20

Daylight time (hour)

6

9

12

15

18

21

24

27

30

m• w

,fres

h (kg

/h)

ta,o= 40 oC, wa,o= 0.02 kgv/kga,FAR=0.25, tcw,deh,i= 20 oC

m•w,fresh (kg/h)

RPS

0.5

0.55

0.6

0.65

0.7

0.75

0.8

0.85

0.9

0.95

RP

S

Aasc=Awsc=5 m2

10 m2

15 m2

5 m2

10 m2

15 m2

(a)

(c)

Fig. 7. Influences of solar collectors’ areas on the proposed system performanc

4.1. Hourly system performance

The hourly variation of proposed system performance parame-ters (m�

w,fresh, R, Q�R, Q�

cc, E�ps, ta,s, RHa,s and COPPS) are shown inFigs. 4–7 for various operational and design parameters. As shownin the figures the hourly variation of system performance followsthe hourly variation of the solar radiation intensity that is shownin Fig. 3a. The figures show that m�

w,fresh, R, Q�R, Q�

cc, E�ps, RHa,s andCOP increase with increasing the solar intensity until they reachmaximum values at noon where the solar intensity is maximumthen they decrease until the end of the day. The behavior of thehourly fresh water productivity and water recovery(Figs. 4a, 5a, 6a, 7a) can be attributed to the increase of the air andwater temperatures exits from the solar collectors with increasingthe solar intensity. Elevating air and water temperatures improvethe humidification capacity (ability of the air for carrying up watervapor in the humidifier) and subsequently increases the condensaterates in the dehumidifier and the cooling coil. The hourly variation ofthe system cooling capacity can be attributed to the psychometricanalysis which verifies the increase of the enthalpy of air leavingthe humidifier (H) and consequently the increase of the enthalpyof the mixing point (M) at the same fresh air ratio. As a result, thesupply air temperature drops and supply relative humidity rises(see Figs. 4d, 5d, 6d, 7d). Decreasing the supply air temperature with

6 8 10 12 14 16 18 20

Daylight time (hour)

5

6

7

8

9

10

Q• R

(kW

)

ta,o= 40 oC, wa,o= 0.02 kgv/kga,FAR=0.25, tcw,deh,i= 20 oC

Q•R (kW)

Q•cc (kW)

8

10

12

14

16

18

20

22

Q• cc

(kW

)

Aasc=Awsc=5 m210 m2

15 m2

6 8 10 12 14 16 18 20

Daylight time (hour)

10

12

14

16

18

t a,S (o C

)

ta,o= 40 oC, wa,o= 0.02 kgv/kga,FAR=0.25, tcw,deh,i= 20 oC

ta,S (oC)RHa,S

0.8

0.85

0.9

0.95

1

1.05

RH

a,S

Aasc=Awsc=5 m2

10 m2

15 m2

5 m2

10 m2

15 m2

(d)

(b)

e parameters: (a) m�w,fresh & R, (b) Q�

R & Q�cc, (c) E�ps & COPPS, (d) ta,s & RHa,s.

6 8 10 12 14 16 18 20

Daylight time (hour)

4

5

6

7

8

9

Q• R

(kW

)

ta,o= 40 oC, wa,o= 0.02 kgv/kga,FAR=0.25, tcw,deh,i= 20 oC,Aasc=10 m2, Awsc=10 m2,

Q•R (kW)

Q•cc (kW)

8

10

12

14

16

18

20

Q• cc

(kW

)Proposed system

Basic system

6 8 10 12 14 16 18 20

Daylight time (hour)

0

10

20

30

m• w

, fre

sh (k

g/h)

ta,o= 40 oC, wa,o= 0.02 kgv/kga,FAR=0.25, tcw,deh,i= 20 oC,Aasc=10 m2, Awsc=10 m2,

m•w, fresh(kg/h)

R

0.5

0.6

0.7

0.8

0.9

1

R

Proposed system

Basic system

6 8 10 12 14 16 18 20

Daylight time (hour)

2

3

4

5

6

7

8

E• (k

W)

ta,o= 40 oC, wa,o= 0.02 kgv/kga,FAR=0.25, tcw,deh,i= 20 oC,Aasc=10 m2, Awsc=10 m2,

E• (kW)COP

1

2

3

4

5

CO

P Proposed system

Basic system

6 8 10 12 14 16 18 20

Daylight time (hour)

8

12

16

20

24

t a,s

(o C)

ta,o= 40 oC, wa,o= 0.02 kgv/kga,FAR=0.25, tcw,deh,i= 20 oC,Aasc=10 m2, Awsc=10 m2,

ta,s (oC)RHa,s

0.6

0.7

0.8

0.9

1

RH

a,s

Proposed system

Basic system

6 8 10 12 14 16 18 20

Day time (hour)

5

10

15

20

25

E• (k

W)

m•w,h=0.15 kg/s,

m•cw,deh=0.15 kg/s, tcw,deh,i= 20 oC,

Aasc=5 m2, Awsc=5 m2, Basic systemProposed system

ta,o= 50 oC, wa,o= 0.05 kgv/kga

m•a,o=0.15 kg/s

m•a,o=0.2 kg/s

6 8 10 12 14 16 18 20

Day time (hour)

6

8

10

12

14

16

18

E• (k

W)

m•w,h=0.15 kg/s,

m•cw,deh=0.15 kg/s, m•

a,o=0.15 kg/s,Aasc=10 m2, Awsc=10 m2,

Basic systemProposed system

ta,o= 50 oC, wa,o= 0.05 kgv/kga

tcw,deh,i=15 oC

tcw,deh,i=25 oC

6 8 10 12 14 16 18 20

Day time (hour)

0

10

20

30

40

50

60

70

m• w

,fres

h (kg

/h)

m•w,h=0.15 kg/s,

m•cw,deh=0.15 kg/s, tcw,deh,i= 20 oC,

Aasc=5 m2, Awsc=5 m2, Basic systemProposed system

ta,o= 50 oC, wa,o= 0.05 kgv/kga

0.15 kg/s

m•a,o=0.2 kg/s

0.2 kg/s

0.15 kg/s

(a)

(d)(c)

(b)

(e) (f)

(g)

Fig. 8. Comparisons of the proposed with the basic system at different operating and design parameters.

A. Fouda et al. / Applied Thermal Engineering 108 (2016) 1320–1335 1329

0.1 0.12 0.14 0.16 0.18 0.2m•

a,o (kg/s)

-300

-225

-150

-75

0

75

150

225

TOC

S ($

/Yea

r)

ta,o= 30 oC, wa,o= 0.02 kgv/kga

m•w,h=0.15 kg/s, m•

cw,deh=0.15 kg/s,

Awsc= Aasc= 5 m2

10 m2

15 m2

15 oC

20 oC

25 oC

20 oC

10 m2

10 m2

tcw,deh,i= 20 oC

0.1 0.12 0.14 0.16 0.18 0.2m•

a,o (kg/s)

200

300

400

500

600

700

800

900

1000

1100

1200

TOC

S ($

/Yea

r)

ta,o= 50 oC, wa,o= 0.05 kgv/kga

m•w,h=0.15 kg/s, m•

cw,deh=0.15 kg/s,

tcw,deh,i= 25 oC

20 oC

15 oC

20 oC

20oC

Awsc= Aasc= 10 m2

10 m2

10 m2

5 m2

15 m2

(a)

(b)

Fig. 9. TOCS: (a) influences of solar collectors’ areas, (b) influences of dehumidifierinlet water temperature.

1330 A. Fouda et al. / Applied Thermal Engineering 108 (2016) 1320–1335

increasing relative humidity means the drop in the enthalpy of thesupply and then the increase in the building cooling load and thecooling coil capacity as shown in Figs. 4b, 5b, 6b and 7b. The increaseof hourly electrical power consumption and the system COPPS withthe increase of the solar intensity (see Figs. 4c, 5c, 6c, 7c) can beattributed to: (i) increasing of the cooling coil capacity leads tohigher electric power consumption at constant COPRef as describedin Eq. (18), (ii) increasing of the system cooling capacity and thefresh water rate overcomes the increase of the system electricalpower which leads to higher COPPS as given by Eq. (25).

4.1.1. Effects of ambient air dry bulb temperatureThe effects of ambient air dry bulb temperature on the proposed

system performance parameters are shown in Fig. 4a–d for specificoperating and design parameters as an example. Fig. 4a shows thatthe hourly fresh water production rate and the system waterrecovery improve with increasing the ambient air temperature.This can be attributed to the increase of air temperature at thehumidifier inlet with increasing the ambient air temperaturewhich leads to improving humidification capacity, fresh water pro-ductivity and fresh water recovery.

Fig. 4b illustrates the hourly variation of the system coolingcapacity (space cooling load, Q�

R and cooling coil capacity, Q�cc) for

different ambient air temperatures. As displayed in the figure, thesystem capacity increases with increasing the ambient air tempera-ture. This can be attributed to the increase of air enthalpy at thedehumidifier outlet (k) and consequently at the mixing point (M)that lead to decreasing the supplied air enthalpy according to thepsychrometric analysis. Dropping the supplied air enthalpy causesa decrease of the supplied air temperature and an increase of relativehumidity to the space (see Fig. 4d) with the increase of the ambientair temperature. As explained in Section 4.1, decreasing the supplyair temperature and increasing relative humidity leads to thedecrease of the enthalpy of the supply air to the conditioned spaceand an increase in the system capacity for removing the space load.Fig. 4b also displays the improve in the cooling coil capacity with ris-ing in ambient air temperature. This can be attributed to the rising inthe system capacity and the drop in the air enthalpy downstream thecooling coil.

Fig. 4c shows the increase of the hourly electrical power con-sumption and the decrease of the system COPPS with increasingthe ambient air temperature. The increase of the electrical powerconsumption can be attributed to (i) the increase of the cooling coilcapacity as discussed in Fig. 4b that leads to higher electric powerconsumption and (ii) the decrease of the chiller COPRef with theincrease of the ambient air temperature (see Eqs. (18) and (19) thatleads to higher electric power consumptions (see Eq. (27)). Thedecrease of the system COPPS can be attributed to that the increasein the system cooling capacity and the fresh water rate can’t over-come the increase in electrical power (see Eq. (25)).

4.1.2. Effects of ambient air absolute humidityThe effects of the ambient air humidity on the hourly variation

of proposed system performance parameters are shown in Fig. 5a–d for specific operating and design parameters as an example.Fig. 5a shows that the hourly variation of m�

w,fresh and R increasewith increasing the humidity. The increase of the fresh water ratecan be attributed to the high amount of water vapor carried by airat the humidifier’s exit that leads to the increase of the drivingforce of mass transfer. The increase of the system water recoverywith the humidity can be attributed to the increase of the freshwater rate with decreasing the make-up water (see Eq. (23)).

Fig. 5b shows that Q�R, Q�

cc increase with the increase of wa,o. Thisis attributed to the increase of the latent heat capacities across thedehumidifier and cooling coil due to rising the air enthalpy down-stream the humidifier. Rising the air enthalpy downstream the

humidifier leads to an increase in the enthalpy of the mixed air(M) and a decrease in the supply air enthalpy (S) and subsequentlyan increase in the system capacity (Q�

R & Q�cc) to remove the space

cooling load. This decrease the supply air temperature and increasethe relative humidity as shown in Fig. 5d.

Fig. 5c displays the increase of system electric power consump-tion and COPPS with increasing the ambient air humidity. Increas-ing E�ps with the increase of the wa,o is due to the high cooling coilcapacity as discussed in Fig. 5b which results to higher electricalpower consumption. The increase of the system COPPS with theincreasing the air humidity is due to the increase of Q�

R and the freshwater productivity with a mount that overcomes on the increase insystem electrical power and this leads to higher COPPS.

(a)(b)

(c) (d)

(e)(f)

(g)

5060

7080

90100

110120

130

QR

,day (kWh)

wa,o = 0.02 kg

v /kga ,

FAR=0.25, tcw

,deh,i = 20 oC,

Aasc =10 m

2, Aw

sc =10 m2, m

•a,o =0.1-0.2 kg/sE

ps,day (kWh)

mw

,day (kg)

0 40 80 120

160

Daily energy consumption, Eps,day (kWh)

120

160

200

240

280

320

Daily fresh water productivity, mw,day (kg)

ta,o = 30 oC

40 oC

50 oC

5060

7080

90100

110120

130140

150

QR

,day (kWh)

ta,o = 40 oC,

FAR=0.25, tcw

,deh,i = 20 oC,

Aasc =10 m

2, Aw

sc =10 m2, m

•a,o =0.1-0.2 kg/sE

ps,day (kWh)

mw

,day (kg)

0 40 80 120

160

Daily energy consumption, Eps,day (kWh)

0

100

200

300

400

500

600

Daily fresh water productivity, mw,day (kg)

wa,o = 0.01 kg

v /kga

0.03

0.05

5060

7080

90100

110120

130

QR

,day (kWh)

ta,o = 40 oC, w

a,o = 0.02 kgv /kg

a , FAR

=0.25, tcw,deh,i =15 oC, m

•a,o =0.1-0.2 kg/s m

w,day (kg)

Eps,day (kW

h)

20 40 60 80 100

Daily energy consumption, Eps,day (kWh)

100

150

200

250

300

350

Daily fresh water productivity, mw,day (kg)

Aasc =A

wsc = 5 m

2

10 m2

15 m2

5 m2

10 m2

15 m2

5060

7080

90100

110120

130

QR

,day (kWh)

ta,o = 40 oC,

wa,o = 0.02 kg

v /kga , FAR

=0.25, A

asc =10 m2, A

wsc =10 m2, m

•a,o =0.1-0.2 kg/s m

w,day (kg)

Eps,day (kW

h)

20 40 60 80 100

120

Daily energy consumption, Eps,day (kWh)

100

150

200

250

300

Daily fresh water productivity, mw,day (kg)

tcw,deh,i = 15 oC

20 oC

25 oC

20 oC

25 oC

15 oC

5060

7080

90100

110120

130

QR

,day (kWh)

ta,o = 40 oC, w

a,o = 0.02 kgv /kg

a

FAR=0.25, tcw

,deh,i = 20 oC,

Aasc =10 m

2, Aw

sc =10 m2, m

•a,o =0.1-0.2 kg/s m

w,day (kg)

Eps,day (kW

h)

40 60 80 100

120

Daily energy consumption, Eps,day (kWh)

140

160

180

200

220

240

260

280

Daily fresh water productivity, mw,day (kg)

m•cw

,deh =0.1 kg/s0.15

0.2

0.10.15

0.2

5060

7080

90100

110120

130

QR

,day (kWh)

ta,o = 40 oC, w

a,o = 0.02 kgv /kg

a, tcw,deh,i = 20 oC

,FAR

=0.25, Aasc =10 m

2, Aw

sc =10 m2,

m•a,o =0.1-0.2 kg/s m

w,day (kg)

Eps,day (kW

h)

40 50 60 70 80 90 100

Daily energy consumption, Eps,day (kWh)

160

180

200

220

240

260

Daily fresh water productivity, mw,day (kg)

m•w

,h =0.1, 0.15,0.2 kg/s

3060

90120

150180

210

QR

,day (kWh)

ta,o = 40 oC, w

a,o = 0.02 kgv /kg

a , A

asc =10 m2, A

wsc =10 m

2, tcw,deh,i =20 oC

,m

•a,o =0.1-0.2 kg/s m

w,day (kg)

Eps,day (kW

h)

20 40 60 80 100

120

Daily energy consumption, Eps,day (kWh)

150

200

250

300

Daily fresh water productivity, mw,day (kg)

FAR= 0.1

0.25

0.5

Fig.10.Daily

performan

ceof

theproposed

systemat

differentoperatin

gan

ddesign

parameters.

A.Fouda

etal./A

ppliedTherm

alEngineering

108(2016)

1320–1335

1331

Table 1Constants of predicted correlations and their errors.

w a b c d e f g h i k Error

m�w;fresh

m�w;fresh;max

1.282 �0.04 1.049 0.537 0.634 0.291 0.243 �0.005 �0.014 �0.014 Predicts 90% within error ±10%

Q�R

Q�R;max

0.117 �1.068 0.428 0.155 0.172 0.498 0.107 �0.0031 �0.118 0.265 Predicts 90% within error ±12%

E�PSE�PS;max

1.087 �0.263 0.662 1.643 0.385 �0.191 0.274 0.0264 �0.221 0.569 Predicts 90% within error ±15%

1332 A. Fouda et al. / Applied Thermal Engineering 108 (2016) 1320–1335

4.1.3. Effects of outdoor air flow rateThe effects of the outdoor air flow rate on proposed system per-

formance parameters are shown in Fig. 6a–d for specific operatingand design parameters as an example. Fig. 6a shows the increase ofm�

w,fresh and the drop of Rwith rising m�o,a. Increasing the fresh water

rate with increasing m�o,a is due to the high amount of water vapor

that passed through the humidifier and cooling coil which resultsin the increase of the amount of water vapor condensation on thedehumidifier/cooling coil surfaces. On the other side reducing thesystem water recovery with rising the ambient air flow rate can beattributed to that the percentage of increase of the make-up wateris higher than the percentage of increase in the dehumidificationcapacity, m�

w,fresh.Fig. 6b illustrates the increase of system cooling capacity with

m�a,o. This can be attributed to two opposite effects; (i) the increase

of supply air flow rate with respect, (ii) the decrease in the enthalpychange through the cooling coil and conditioned space as a result ofthe drop in air enthalpy downstream the air solar collector. Theincrease in the supply air flow rate overcomes the decrease in theenthalpy difference across the cooling coil and conditioned space.

Fig. 6c illustrates the increase in the electrical power consump-tion and the decrease in the system COPPS with decreasing the out-door air flow rate. Increasing the system electrical powerconsumption can be attributed to (i) the increase of the cooling coilcapacity as discussed in Fig. 6b which results to higher electricalpower consumption, and (ii) the increase of the power of auxiliarysystem components. The decrease of the system COPPS with theincrease of air flow rate can be attributed to the increase of thespace cooling load and the fresh water productivity with a percent-age lower than the percentage of the increase in the system electri-cal power.

4.1.4. Effects of solar collectors’ areasThe effects of the solar collectors’ areas on the proposed system

performance parameters are shown in Fig. 7a–d for specific operat-ing and design parameters as an example. Fig. 7a shows theenhancement of the fresh water production rate and the systemwater recovery with increasing the areas of air and water solar col-lectors. This can be attributed to the increase of the air and watertemperatures downstream the solar collectors which leads to theimprove of the air humidification capacity and consequently theincrease of the dehumidification capacity.

Fig. 7b shows that Q�R and Q�

cc increase with increasing the solarcollectors’ areas. This can be attributed to the increase of the sensibleheat capacities across the dehumidifier and cooling coil due to theincrease of the air enthalpy downstream the air solar collector andin consequence at the humidifier’s exit. Increasing of the air enthalpyat the humidifier’s exit leads to high air enthalpy at mixing point (M)and low enthalpy at the supply air point (S) with keeping constantADP and RSHF which leads to the increase of Q�

R and Q�cc.

Fig. 7c illustrates the increase both of the electrical power con-sumption and the system COPPS with increasing the solar collec-tors’ areas. Increasing the system electrical power consumptionis due to the increase of the cooling coil capacity as discussed inFig. 7b and the increase of power of auxiliary system components.The increase of the system COPPS with increasing the solar collec-

tors’ areas can be attributed to the increase of the space coolingload and the system water productivity which overcomes on theincrease of the system electrical power.

4.2. Comparisons and cost saving analysis

The comparison between the proposed system performanceparameters and the basic system is shown in Fig. 8a–g at theoperating and design parameters: FAR = 0.25, m�

a,o = 0.15 kg/s,m�

w,h = 0.15 kg/s, m�cw,deh = 0.15 kg/s, Awsc = Aasc = 10 m2 and

tcw = 20 �C. Fig. 8a and b shows an improvement of the proposed sys-tem fresh water production rate and the system water recovery ascompared with the basic system at ta,o = 40 �C, wa,o = 0.02 kgv/kga,and ta,o = 50 �C, wa,o = 0.05 kgv/kga, respectively. As shown inFig. 8b, the proposed system fresh water productivity is much higherthan the basic system. This is due to the existing of HDH system(humidifier/dehumidifier) with solar collectors that improve the airhumidification performance.

Fig. 8c shows that the proposed system cooling capacity (Q�R and

Q�cc) is greater than that of the basic system along the day. This can

be attributed to the increase of the air enthalpy change across thecooling coil due to elevating the air enthalpy downstream thehumidifier and decreasing the supplied air temperature with risingthe relative humidity (see Fig. 8g) due to existence of the HDH solarassisted system components.

Fig. 8d shows that the electrical power consumption and theCOP of the proposed system are greater than those of the basic sys-tem. This can be attributed to the higher Q�

cc of the proposed systemas compared to the basic system (see Fig. 8c) which leads to higherelectrical power consumption. The higher COP of the proposed sys-tem is attributed to the high Q�

R and fresh water production rate asdiscussed in Fig. 7c.

Fig. 8e and f illustrates that the proposed system electric powerconsumption is less than that consumed in basic system atta,o = 50 �C, wa,o = 0.05 kgv/kga. The possible explanation of that isattributed to the presence of dehumidifier in the system whichmakes the air enthalpy at the dehumidifier’s exit becomes lowerthan that of the outdoor air. This reduces the air enthalpy upstreamthe cooling coil which results in a reduction of the cooling coilcapacity and the electrical power consumption.

The previous comparison study verified that the suggested sys-tem has high fresh water productivity and space cooling capacitybut consumes low or high electrical power according to the oper-ating conditions. So, for fair comparison for proposed and basicsystems, the total cost savings due to fresh water productivityand electrical power consumption was selected as a judging crite-rion for comparison between the two systems. The TOCS is calcu-lated as follows:

TOCS ð$=YearÞ ¼ Dmw;fresh;Y ðkg=YearÞ �Water unit rate ð$=m3ÞþDEY ðkW h=YearÞ � Electricity unit rate ð$=kW hÞ

ð28Þwhere

Dmw;fresh;Y ¼ NY

Xsunsetsunrise

mw;fresh;PS �mw;fresh;BS� �

0 0.2 0.4 0.6 0.8 1

(FAR)-0.04(τ/Δτ)1.049(ta,o/ta,o,max)0.537(wa,o/wa,o,max)0.634(m.a,o/m

.a,o,max)0.291

(ASC/ASC,max)0.243(tcw,deh,i/tcw,deh,i,max)-0.005(m.cw,deh/m

.cw,deh,max

)-0.014(m.w,h/m

.w,h,max

)-0.014

0

0.2

0.4

0.6

0.8

1

1.2

m• w

,fres

h/m• w

,fres

h,m

ax

Numerical correlationNumerical results

0 0.4 0.8 1.2

m•w,fresh/m•

w,fresh,max(Cal. results)

0

0.4

0.8

1.2

m• w

,fres

h/m• w

,fres

h,m

ax (C

orr.

resu

lts)

+10%

-10%

0 2 4 6 8 10

(FAR)-1.068(τ/Δτ)0.428(ta,o/ta,o,max)0.155(wa,o/wa,o,max)0.172(m.a,o/m

.a,o,max)0.498

(ASC/ASC,max)0.107(tcw,deh,i/tcw,deh,i,max)-0.0031(m.cw,deh/m

.cw,deh,max

)-0.118(m.w,h/m

.w,h,max

)0.265

0

0.2

0.4

0.6

0.8

1

Q• R

/Q• R,

max

Numerical correlationNumerical results

0.2 0.4 0.6 0.8 1

(FAR)-0.263(τ/Δτ)0.662(ta,o/ta,o,max)1.643(wa,o/wa,o,max)0.385(m.a,o/m

.a,o,max)-0.191

(ASC/ASC,max)0.274(tcw,deh,i/tcw,deh,i,max)0.0264(m.cw,deh/m

.cw,deh,max

)-0.221(m.w,h/m

.w,h,max

)0.569

0.2

0.4

0.6

0.8

1

1.2

E• PS/E

• PS,

max

Numerical correlationNumerical results

0 0.4 0.8 1.2

E•PS/E•

PS,max (Cal. results)

0

0.4

0.8

1.2

E• PS/E

• PS,m

ax (C

orr.

resu

lts)

+15%

-15%

(a) (b)

(c) (d)

(e) (f)

0 0.2 0.4 0.6 0.8 1

Q•R/Q•

R,max (Cal. results)

0

0.2

0.4

0.6

0.8

1

Q• R

/Q• R

,max

(Cor

r. re

sults

)

+12%

-12%

Fig. 11. Proposed system numerical correlations predictions and errors: (a & b) m�w,fresh/m�

w,fresh,max, (c & d) Q�R/Q�

R,max, (e & f) E�ps/E�ps,max.

A. Fouda et al. / Applied Thermal Engineering 108 (2016) 1320–1335 1333

1334 A. Fouda et al. / Applied Thermal Engineering 108 (2016) 1320–1335

DEY ¼ NY

Xsunsetsunrise

½EBS � EPS�

NY: Number of days a year (NY = 365 day).In the current study the unit price of electrical energy and fresh

water are 0.05 $/kW h and 2.5 $/m3, respectively (typical values forGulf cities [38]).

Fig. 9a and b shows TOCS versus outdoor air flow rate atta,o = 30 �C, wa,o = 0.02 kgv/kga, and ta,o = 50 �C, wa,o = 0.05 kgv/kga,respectively with different operating and design parameters(Asc & Tcw,deh,i). As shown in the figures, the TOCS reduces and riseswith increasing m�

a,o at ta,o = 30 �C, wa,o = 0.02 kgv/kga, andta,o = 50 �C, wa,o = 0.05 kgv/kga, respectively. Reducing TOCS with ris-ing m�

a,o can be attributed to the higher electrical energy consump-tion of the proposed system and vice versa for more hot andhumid conditions (ta,o = 50 �C, wa,o = 0.05 kgv/kga). In addition, atlow ta,o and wa,o it is observed that the proposed system operateseconomic deficiently relative to the basic system (see Fig. 9a). Onthe other side the proposed system works best economically in hotand humid climatic conditions (see Fig. 9b) and the maximum valueof TOCS can be obtained throughout all the studied parameter rangesis 1079 $/Year at Awsc = Aasc = 10 m2 and Tcw,deh,i = 15 �C.

4.3. Daily system performance

The hourly system outputs and consumptions are integrated allover the day time to give the daily system output and consump-tions. Fig. 10 shows the variations of the daily system outputsand consumptions for different systems capacities serving differentbuildings of different cooling loads. The figures are plotted for var-ious operational and design system parameters to show theireffects on the daily fresh water productivity and electrical powerconsumptions. Fig. 10a–g displays an improvement of the dailyfresh water productivity and electrical power consumption withthe increase of the integral cooling load. This can be attributed tothe increase of the cooling coil capacity and the cooling load withincreasing the outdoor air flow rates which results to the increaseof the hourly and then daily fresh water productivity and systempower electrical consumption. Fig. 10a–d illustrates enhancementof the daily fresh water productivity and electrical power con-sumption with increasing the ambient air temperature, the ambi-ent relative humidity, solar collectors’ areas and outdoor air flowrat, respectively. This can be attributed to the same reasons ofincreasing the hourly values which are discussed in Section 4.2.Fig. 10e–f shows the increase of the electrical power consumptionand the decrease of the fresh water production rate with increasingboth of inlet cooling water temperature and flow rates to the dehu-midifier. The possible explanation can be attributed to the risingthe air enthalpy downstream the dehumidifier with increasingtcw,deh,I and m�

cw,deh. This leads to an increase the cooling coil capac-ity and E�ps. The reduction in the fresh water productivity can beattributed to the reduction in the latent heat capacity across thedehumidifier with increasing tcw,deh,I and m�

cw,deh. Fig. 10g showsthat m�

w,h has no effect on the electrical power consumption andfresh water productivity. The possible explanation is that the airenthalpy at the humidifier’s exit remains constant with changingof m�

w,h as the outlet water solar collector temperature decreaseswith increasing m�

w,h at constant solar intensity. The maximum dailyfresh water productivity, integrated cooling load and energyconsumption are 501 kg, 146.7 kW h and 140 kW h, respectively atwa,o = 0.05 kgv/kga (see Fig. 10b).

4.4. Numerical correlations prediction

To simplify and present the hourly values of m�w,fresh,PS, Q�

R andE�PS for proposed system, the numerical results are regressed to

predict general correlations in terms of s (time), ta,o,wa,o, m�

a,o, m�cw,deh, m�

w,h, tcw,deh,i, Asc, and FAR. The developed corre-lations with their errors and constants are presented in Eq. (29) andTable 1 are illustrated in Fig. 11.

w ¼ a� FARb sDs

� �c ta;ota;o;max

� dwa;o

wa;o;max

� e m�a;o

m�a;o;max

� fASc

ASc;max

� g

tcw;deh;itcw;deh;i;max

� h m�cw;deh

m�cw;deh;max

� �im�

w;hm�

w;h;max

� �k ð29Þ

Eq. (29) is valid for the following ranges 30 �C 6 ta,o 6 50 �C,0.01 kgv/kga 6wa,o 6 0.05 kgv/kga, 0.1 kg/s 6 m�

a,o 6 0.2 kg/s,0.1 kg/s 6 m�

cw,deh 6 0.2 kg/s, 0.1 kg/s 6 m�w,h 6 0.2 kg/s, 15 �C 6

tcw,deh,i 6 25 �C, 5 m2 6 Asc 6 15 m2, 0.1 6 FAR 6 0.5 and7:00 6 s 6 13:00 (similar and mirrored results with 13:00 6 s 619:00). Where the values of ta,o,max, wa,o,max, m�

a,o,max, m�cw,deh,max,

m�w,h,max, Tcw,deh,i,max, Asc,max, Ds, m�

w,fresh,PS,max, Q�R,max and E�PS,max

are 50 �C, 0.05 kgv/kga, 0.2 kg/s, 0.2 kg/s, 0.2 kg/s, 25 �C, 15 m2, 12,36.13 kg/h, 22.24 kW and 9.092 kW, respectively.

5. Conclusions

A proposed innovated solar integrated system for A/C and HDHwater desalination is proposed for climate conditions of hot andhumid regions. Hourly analysis of system performance is studiedusing a developed mathematical model with EES and C++. Systemcapacity parameters (m�

w,fresh, Q�R and Q�

cc) and system consumptionsand performance parameters (E�ps, R and COP) are hourly and dailyestimated for various operational and design system parameters(IT, ta,o, wa,o, tcw,deh,i and m�

cw,deh, m�w,h, m�

a,o, FAR and ASC). The resultsshow that the main influencing parameters affecting the system’scapacity, consumption and performance are ta,o and wa,o. Increasingof ta,o, wa,o and ASC cause an increase in fresh water productivity,fresh water recovery, cooling capacity, electrical power consumptionand COP of the system. TOCS parameter decreases with increasingoutdoor air flow rate at ta,o = 30 �C, wa,o = 0.02 kgv/kga and increasesat ta,o = 50 �C, wa,o = 0.05 kgv/kga. It was found that the proposed sys-tem works economically and more efficiently in hot and humid cli-matic conditions. The maximum TOCS can be attained from theproposed system within all the studied parameter ranges is 1079$/Year at wa,o = 0.05 kgv/kga, Awsc = Aasc = 10 m2, ta,o = 50 �C,tcw,deh,i = 15 �C.

Moreover, the proposed system can provide maximum dailyfresh water productivity and integrated cooling load with 501 kg,146.7 kW h, respectively and consume energy with 140 kW h atwa,o = 0.05 kgv/kga, Awsc = Aasc = 10 m2, ta,o = 40 �C, tcw,deh,i = 20 �C.General numerical correlations for proposed system performanceparameters (m�

w,fresh/m�w,fresh,max, (c) & (d) Q�

R/Q�R,max, (e) &

(f) E�ps/E�ps,max) are developed in terms of all studied design andoperational parameters.

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