a theoretical and experimental study for a humidification

International Journal of Water Resources and Arid Environments 3(2): 108-120, 2014ISSN 2079-7079© PSIPW, 2014

Corresponding Author: M. Abd ElKader, Department of Mechanical Power Engineering, Faculty of Engineering, Suez Canal University, Port Said, Egypt.

108

A Theoretical and Experimental Study for aHumidification-Dehumidification (HD) Solar Desalination Unit

M.Abd ElKader, A. Aref, Gamal H. Moustafa and Yasser ElHenawy

Department of Mechanical Power Engineering, Faculty of Engineering, Suez Canal University, Port Said, Egypt

Abstract: A humidification-dehumidification (HD) solar desalination unit was designed. It seems to be suitableto provide drinking water for population or remote arid areas. Solar water and solar air collectors were designedto provide the hot water and air to the desalination chamber. The desalination chamber was divided intohumidifier and dehumidifier towers. The circulation of air in the two towers was maintained by the forcedconvection. Theoretical and experimental works were done at different environmental conditions.A mathematical model was formulated, in which the thermodynamic relations were used to study the flow, heatand mass transfer inside the humidifier and dehumidifier. Such a technique was performed in order to increasethe unit performance. Heat and mass balance was done and a set of governing equations was solved using thefinite difference technique. The solar intensity was measured along the working day during the summer andwinter months and a comparison between the theoretical and experimental results were performed. The averageaccumulative productivity of the system in November, December and January was ranged between 2 to 3.5 kg/ m day while the average summer productivity was found between 6 to 8 kg/m day in June and 7.26 to 112 2

kg/m day in July and August.2

Key words: Solar desalination Solar collector Humidification and Dehumidification Solar intensity Simulation Finite difference Water productivity

INTRODUCTION Fath [2] stated that solar desalination can either be

Solar desalination is a suitable solution to supply dehumidification desalination process (solar still) wasremote regions in Egypt and other countries with fresh investigated by Nafey et al. [3–4]. The indirect solar HDHwater. The standard techniques, like Multi-Stage Flash process has the advantage of separating the heating(MSF), Multi-Effect (ME), Vapor Compression (VC) and surface from the evaporation zone and therefore, theReverse Osmosis (RO), are only reliable for large capacity heating surface is relatively protected from corrosion orranges (100–50,000 m /day) of the fresh water production scale deposits. In fact, the humidification dehumidification3

[1]. These technologies are expensive for small amounts desalination (HDD) is more costly than the conventionalof fresh water and also can not be used in locations basin still solar desalination process [1]. Therefore, for thewhere there are limited maintenance facilities. In addition, simplicity of design, a modest level of manufacturingthe use of conventional energy sources to drive these technology is needed.technologies has a negative impact on the environment. Very limited configurations of the HDD process haveSolar desalination processes is a future promising been developed. Abdel-Monem et al. [5] experimentallytechnology, has the following advantages [2]: i) Solar and theoretically investigated the main parameters effectenergy is environmental friendly energy source, cost-free on the productivity of the HDH process. Indoor tests withenergy and ii) reduced operating costs and its simple constant input energy to the system were performed.structure so that it is suitable for a few families or small The configuration consisted of humidifier, dehumidifiergroups in remote areas. and h eating equipments for air and water. The effect of

direct or indirect. The direct solar humidification

Intl. J. Water Resources & Arid Environ., 3(2): 108-120, 2014

109

weather conditions was not considered and the unsteadystate performance of the system was not, therefore,examined. Ettouney [6] made design and analysis ofvarious schemes for water desalination by humidificationand dehumidification system (HDH). A theoreticalinvestigation of a humidification and dehumidificationdesalination system configured by a double –pass flatplate solar air heater was examined by Yamali and Solmus[7]. The effect of a cooling tower on a solar desalinationsystem was studied by [8]. Dai, et al. [9] reported theparametric analysis to improve the performance of a solardesalination unit with a humidification anddehumidification. The performance analysis of adesalination unit of heat pump with humidification and Fig. 1: A Schematic diagram of desalination system.dehumidification was studied by [10]. An experimentaldesign and computer simulation of multi-effecthumidification (MEH) dehumidification solar distillationwas investigated by [11]. Fath and Ghazy [12] studiednumerically a configuration that consists of a solar airheater, humidifier and dehumidifier. The feed water wasnot heated and hence, the effect of water temperature wasnot considered. Darwish [13] suggested a systemconsisting of a solar pond, a humidifying column and adehumidifying stack. He reported that the air flow rate hasa strong effect on the system productivity; however, thefeed water flow rate has a weak effect on the productivity.On the other hand, Al-Hallaj and Farid [14] reported that Fig. 2: A Sketch of honey-comb packing material in thethe feed water flow rate has a strong effect on the system humidifierproductivity and the effect of the air flow rate is weak.In the present article a theoretical study of wooden, which is characterized by a large evaporationhumidification-dehumidification desalination unit which surface per unit volume of packed material. The air isconsists of flat plate water and air solar collectors, storage heated by a solar air collector and then is forced to passtank and dehumidifier is presented. The aim is to develop through the humidifier where it becomes hot and humida computer program and to build an experimental due to the heat and mass exchange between the seawaterapparatus that can be used to study the effect of and air. Then, the humid air is cooled by the coldoperating and design parameters as well as weather seawater when passing through the condenser at whichconditions on the unit performance and fresh water the water vapor condenses and turns into fresh water.productivity. A Comparison with experimental results are The condenser is a helix-tube type one. The cold seawatergiven. flows in the tube channel and the fresh water is

System Description: A solar desalination unit with a The remaining seawater drawn from the humidifier ishumidification and dehumidification cycle, which is cooled and collected at the bottom basin. Some of theconfigured mainly by a flat plate solar water collector, water is fed to the solar collector again since it is not veryV-groove solar air collector, a humidifier and a condenser, concentrated and is still warm. Thus, the fresh water canis shown in Fig. 1. Seawater is heated by a solar water be continually produced. The related parameters of thecollector and then is sprayed by a sprinkler to form a unit are as follows: The fan rated power is 500 W, thefalling film at the surface of a honeycomb wall in the humidifier volume is 800 × 700× 600 mm , the solar waterhumidifier. The structure sketch of the humidifier is shown collector area is 1.5 × 1 m , the solar air collector area isin Fig.2. It consists of a porous and durable honeycomb 2 × 1 m and the dehumidifier heat transfer area is 0.8 m .

produced from the condensate on the condenser surface.

3

2

2 2

, , , ,

1 1b sb si

i b b b i s b scl cl

i b i b

A AU UA A

K K K K

= ⋅ = ⋅

+ +

( )( )

( ) ( ) ( ){ }0.31

1 45 0.00259 0.00144

1

344 / 1

t p

p p a win

U

N T T T N f h

= − − −

− + +

( ) ( )( ) ( )

4 4

1

/

0.0425 1 2 1 /

p a p a

p p g

T T T T

N N f N−

− −+ + − + + − −

( )

1

1 1 1cl

Lbond i t fL o o f

F

SUC D hU D S D

= + +

+ −


110

Various variables characterizing the present problem Q = A U (T – T ) (2)were measured. Thus, the temperature of water and air andthe relative humidity at the inlet and the outlet of each where T is the average flat plate temperature (K) throughcomponent and the water and air mass flow rates are the collector and T is the ambient temperature (K).measured. The wind and radiation are measured. The overall heat transfer coefficient of the collector

The mass flow rate is determined by measuring the air U (W/ m ) is the summation of three values of heatvelocity by rotating vanes anemometer (measuring range transfer coefficients, at the top U , at the bottom U and atof 0.2 m/s up to 40 m/s with a resolution of 0.01 m/s). side walls U , [16].

The air temperature is measured by standard K typethermocouple (Measuring range -50°C to 1000°C, U = U + U + U (3)resolution: 0.1°C up to +199.9°C and 1°C above).

The Global solar radiation is measured by silicon wherecell Pyranometer. Accuracy of instantaneous values is+/-5%, Then the calibration relation is given as: Indicated(mill volts) × 13.6 + 8.37 = (Solar radiation intensity) watts (4)/ meter .2

The relative at the inlet and the outlet of eachcomponent are measured by hygrometer (The humidity where A and A are side and bottom collector areasrange is 5 to 95% RH). (m ). The top heat transfer coefficient U is calculated from

Blower, 2. solar air collector, 3. humidifier, 4.dehumidifier, 5.distilled water, 6. seawater tank, 7.feed pump, 8.valve 9.solar water collector, 10.sprinkler, 11. drain pump, 12. makeup of seawater

Mathematical Model: A mathematical model of the unit ispresented, in which the thermodynamic relations wereused to study the flow, heat and mass transfer inside thehumidifier, dehumidifier and water and air solar collectors.A set of governing equations has been given and solvedusing a finite difference technique. A computer program (5)is developed to evaluate the mathematical model.The following assumptions are considered: (i) the feed where is the collector tilt angle, N is the number of glasswater, cooling water and feed air are constant flow rate plate, is the emittance of glass, the emittance ofalong the day, (ii) Thermal losses to the ambient are absorber plate, T is the average plate temperature andneglected and (iii) The energy stored in the collectors is T is the ambient temperature. is the Boltzmannneglected. constant. The term f is given as:

The Flat Plate Solar Water Collector: The thermal f = (1–0.04h + 5×10 h )(1 + 0.058N) (6)energy balance of the solar collector is written as follows,[15]: where h is the wind heat transfer coefficient and is

G + A = Q + Q + Q (1) wind speed. The collector efficiency factor F is estimatedt cl loss u stg

where G is the total solar intensity (W/ m ), A is thet cl2

collector surface area (m ), Q is the heat loss from the2loss

collector and Q is the useful heat transferred from theu

absorber to water. Q is the energy stored in the (7)stg

collector. The heat loss is expressed by, [16]:

loss cl L p a

p

a

L2

t b

si

L t b st

s b2

t

an empirical equation suggested by [16]:

g p

p

a

win win4 2

win

given by, h = 5.7 + 3.5C where, C (m /sec) is thewin win win

cl

by,

( )( )

tanh / 2

/ 2

U K S DL p p of

U K S DL p p o

− = −

( )0.141 3

1 3

,1.86 RePrf fi

t fi f t

K DhD L

=

.,4

Re Prcl f p f

t f

Cm AD N Ki f

= =

../

1 L cl pU F mCpmCF eR UL

− = −

.m

( ),Q A F q U T Tu cl R abs L f i a = − −

1, ,cl

Q Fu RT Tf a f i A F U Fcl R L

= + −

1

/,

1 1

t

tf i bond

Q NuT Tp f o

h D L C L

−= +

+

., ,p

QuT Tf o f imC

= +

uc

cl t

QA G

=


111

where S is the tube spacing (m), D and D are thei o

inner and outer diameters of the collector tubes (m). (14)The collector fin efficiency is calculated by, [16]:f

(8)

For the flow with (Re < 2300), the convective heattransfer coefficient h and the Prandtl number, Pr are given where N is the number of tubes and C is thetf

by, [16]: conductive resistance of the bond. The fluid temperature

(9)

The Reynolds number (Re) and Prandtl number (Pr) The collector efficiency is determined by, [17]:are given by, [16]:

(10)

where k (W/m k), µ (N sec/m ) and C (J/kg k) the thermal honeycomb wooden which is characterized by a largef f p2

conductivity, dynamic viscosity and specific heat of evaporation surface per unit volume of packed material.the fluid are respectively. L is the tube length (m). The mathematical model of the humidifier, which isThe absorbed solar radiation (q ) is determined by, [17]: established on the basis of energy and mass balances, isabs

q = G ( ) (11) film and the moist air is considered. The set of governingabs t

where ( ) is the average Transmittance-absorptance of Assumptions are taken into account are: (i) the flow ofcover plate. Also the heat removal factor (F ) is calculated falling film is laminar and no wavy movement takes placeR

from, [16]: at the liquid-gas interface and (ii) Flow rate of falling film

Fig. 3 shows the coordination set-up. Here, the zero(12) point for the vertical axis is set at the liquid-gas interface.

where is the mass flow rate per unit area of collector

(kg/s. m ). The useful energy from the collector is2

calculated from, [16]:

(13)

The average fluid temperature is thus obtained as,[16]:

The average plate temperature is obtained from, [16]:

(15)

t bond

at the outlet from the collector is calculated from, [16]:

(16)

(17)

Humidifier Modeling: The humidifier consists mainly of

presented. The heat and mass transfer between the liquid

equations is solved using the finite-difference method.

is believed constant because the evaporation rate of waterout of the liquid-gas interface is small enough.

Fig. 3: Coordination set-up

2

3s s

ss

gu =

13

23 s

ssg

Γ=

2

22

s ss su

z y∂Τ ∂ Τ

=∂ ∂

0u vx y∂ ∂

+ =∂ ∂

2

21

aa

u u u pu vx y xy∂ ∂ ∂ ∂

+ = −∂ ∂ ∂∂

2

2au vx y y

∂Τ ∂Τ ∂ Τ+ =

∂ ∂ ∂

At , , 0,2 s e ed uy u u

y∂

= − = = Τ = Τ∂

( )1

2

1 2s vn

MWj P PRT

= −

12

22* s n

v

RT ThM

=


112

Towards the falling film, axis y is set. Towards the gas u = 0, T = T and T = T2

phase, axis y is set. Axis x is defined in the direction of airflow and axis z is defined in the direction of falling film. where T is the temperature at interface,T is the

Governing Equations of Falling Film: If the liquid film water and is the latent heat of water. Solving the setfalls slowly and steadily just under the force of gravity, flow equations (21-23) with boundary conditions, theparticularly in laminar flow, the mean velocity of the falling distribution of the flow velocity and temperature insidefilm, u is given as [9]: the humidifier were obtained. To simplify the aboves

(18) Auxiliary Equations: The evaporation rate at the water

The thickness , Fig. 3 is given as [9]:s

(24)(19)

where is the mass flow rate of saline water per unit Knudsen coefficient of evaporation. It is calculated from:width of the wall (kg/s/m) and g is the acceleration of = (2 )/(2- ). The coefficient of evaporation, isgravity, m/s . The energy balance for the falling film is calculated from:2

(20) (25)

Governing Equations of Moist Air: The governing The coefficient of heat transfer h (W/m K) is givenequations, including mass, momentum and energy by, [17]:conservations are listed as follows [9]:

(21) The terms and are a function of the plate spacing.

0.7 respectively. As the plate spacing increases to 13.4(22) mm, the values of and are changed to be 49 and 0.6.

introduced the concept of the wet bulb coefficient of heat(23) transfer (h (W/m K) to correct the heat transfer process

Boundary conditions:

where u (m/sec) and T (K) are the velocity ande e

temperature of incoming air. e = ( – )/(T – T ) (28)At the water - air interface:

n s n

n s

temperature of saline water, W is the evaporation rate ofj

solution, auxiliary equations are needed.

surface estimated from, [18]:

Here M is the molecular weight of water and R isthe universal gas constant (8.314 kJ/kmol k). is the1

1

2

h = V W/m °C (26)2

For a 6.7mm plate spacing the values of and are 54 and

Taking the phenomenon of evaporation into account,

* 2

of a wet surface:

h* = h(1 + e /C ) (27)pa

where h (W/m °C) is the heat transfer coefficient and e is2

constant, given by:

max min max min

Here T , T are the maximum and minimummax min,

( ) ( ). .

cw apcw cwo cwi o im C T T m H H− = −

( ).

cw pcw cwo cwi c c cm C T T U A LMTD− =

( ) ( )( )( )

ln

ac cwo ac cwic

ac cwo

ac cwi

T T T TLMTD

T TT T

− − −=

− −

ln1 1 1

coco

cicoc

c ci ci c co

rrrr Rf

U h r k h

= + + +

( ).

ad o im m= −

.

2 p f b a fh (T - T ) + U (T -T ) = air airmC dTW dx

2. 3air

oD m LP f yL D

∆ = +


113

temperatures at the liquid-gas interface. T is the highermax

one of the temperatures of gas flow and water flow.T is the wet bulb temperature of the air. and aremin, max min

the maximum and minimum humidity ratios correspondingto T and T .max min

Dehumidifier Modeling: An energy balance between thecooling water and the air/condensing vapor stream ismade the energy balance is based on two following Fig. 4: V-Groove air collectorthermal equations [6]:

so far, iii) The radiation coefficient between the two air(29) duct surfaces is found by assuming a mean radiant

Loss through front and back are of the same temperature.(30) The energy balances for the absorber plate, back

where m and m (kg/sec) are the mass flow rate of the following, [19]:cw a

cooling water and incoming air. T and T (K) are thecwo cwi

temperature of cooling water at inlet and outlet. H and H In the single pass v-groove absorber:i o

(Kj/kg) are the enthalpy of air at inlet and outlet. In Eq. For the cover:(30) the logarithmic mean temperature difference and theoverall heat transfer coefficient are given by the following U (T -T ) = h (T - T ) + h (T - T ) (34)relations [6]:

(31)

(32)

The production of distilled water (m ) is given by the The pressure drop through the channel in the aird

following equation heater has been computed using the following expression

(33)

where and are the absolute humidity (kg H O/kg dry0 i 2

air) of air at the inlet and outlet from the condenser.

V-Groove Solar Air Collector: A V- shaped flatplate solar air collector is designed, Fig. 4. The f = 24/Re, y = 0.9calculation of the solar air collector are made with thefollowing assumptions: I) performance is steady state, for transitional flow (2550 < Re < 10 )

ii) There is no absorption of solar energy by a cover in

temperature equal to the mean fluid temperature and iv)

insulation and working fluid of the collector are given by

t g a 1 p g r p g

V-groove absorber:

G = h (T -T ) + h (T -T ) +h (T -T ) (35)g p t 1 p g r p g 2 p f

Fluid medium:

(36)

Using equations (5), (6) with (34), (35) and (36)

[19]:

(37)

for laminar flow (Re < 2550)

0

4


114

f = 0.0094, y=2.92 Re received in early hours is used as sensible heat to warm00.15

for turbulent flow (10 < Re < 10 ) Figures (9) and (10) show the measured values of4 5

f = 0.059 Re , y = 0.73 It can be seen from the curve that, the average00.2

Numerical Procedure: The calculation procedure is as (November, December and January) is ranged between 2follows: to 3.5 kg / m .day and the average summer productivity is

Input the known data, such as the mass flow rates of August).air, feed seawater, cooling water, inlet temperature The effect of testing parameters in three operatingand humidity of the air. days such as, feed water, air and cooling water flow ratesCalculate the data within the solar water collector and are presented in Figures (11) and (12).obtain the temperature of feed water to the humidifier. It seems from Fig. (13) That, for the same weatherCalculate the data within the solar air collector and conditions (solar intensity, ambient temperature and windobtain the temperature and humidity of feed air to the speed), the system productivity of the unit increaseshumidifier. with increasing the air flow rate in three excessive days.Calculate the data within the humidifier and obtain This is due to the fact that increasing the air flow rateevaporation rate, outlet temperature of seawater, leads to an increase in the mass and heat transferoutlet temperature and humidity of the air. coefficients in the humidifier and hence entrained waterCalculate the data throughout the dehumidifier and vapor, which is lead to an increase in the systemobtain the outlet temperature and humidity of the air, productivity.outlet temperature of the cooling water and finally the The effect of cooling water flow rate on thefresh water production. productivity is illustrated in Fig. (14). It is clear from the

RESULTS AND DISCUSSION dehumidifier leads to a decrease in the surface

Fig.5 shows the variations of solar intensity along increase in the condensation rate and hence the systemone day of operation in summer season. The Figure productivity.shows that, there is a good agreement between the Figure (15) shows the effect of solar intensity on thetheoretical model and the measured data. Fig.6 shows the water productivity on 22 July 2009. The measured valuesvariations of solar intensity along the three days in June, of the average solar radiation and ambient temperatureJuly and August. during 22 July 2009, in Port Said-Egypt are given in Figure

Figure 7 shows the characteristic efficiency and (16).The Average relative humidity of the ambient air isoperating temperatures of the unit for different collector 50–60% and the wind speed ranges between 1 and 3 m/s.feed water. Efficiency curve is based on gross collector Figure (17) represents the air relative humidity at thearea and is determined using the outdoor conditions. The inlet and outlet of the humidifier for the indicated climaticthermal efficiency is a function of the outdoor conditions. One can notice that, the relative humidity attemperature, solar radiation and the temperature of the the exit of the humidifier is close to the value of 95%.fluid at the inlet. Due to the air heater device output, the outlet relative

Fig. 8. The figure shows, a comparison between the humidity appears constant; independent on the inlet one.theoretical and measured accumulative productivity. It This result means that, the air leaving the humidifier iscan be seen from the curve that, there is as good practically saturated.agreement between the theoretical and experimental Figure 18 shows the effect of the inlet temperatureresults. The accumulative productivity on this day is and relative humidity of air on the unit water productivity.about 10 kg/day/m . In the real outdoor operation, a delay The results agree with the common knowledge in2

time was noticed between the start of the run time and the literature, that the higher inlet temperature and humiditystarting of fresh water production, as most of the energy are beneficial for producing more fresh water.

up the test rig metal pants and the fluid flow.

the accumulative productivity at different days.

accumulative productivity of the system in winter

2

ranged between 7.26 to 11 kg/ m .day (in June and2

curve that, Increasing the cooling water flow rate in the

temperature of the air cooler and this will lead to an

22/7/2009

200

300

400

500

600

700

800

900

1000

9 10 11 12 13 14 15 16 17 18

Time ,hr

Sola

r ra

diat

ion

inte

nsity

,W/m

2

exp

theo

22/7/2009

0

2

4

6

8

10

12

9 10 11 12 13 14 15 16 17 18

Time ,hr

Acc

umul

ativ

e pr

oduc

tivity

,kg/

day

Theo

exp

m.cw = 400 kg/hr

m.s = 300 kg/hr

m.a = 250 kg/hr

200

300

400

500

600

700

800

900

1000

1100

9 10 11 12 13 14 15 16 17 18

Time, hr

Sola

r In

tens

ity, W

/m2

12/8/2009

15/7/2009

15/6/2009

efficiency factor = 0.95FR = 0.88

67

68

69

70

71

72

73

74

75

76

0.01 0.012 0.014 0.016 0.018 0.02 0.022 0.024 0.026

(Ts - Ta )/G

FPC

thre

mal

effi

cien

cy,%

0

2

4

6

8

10

12

9 10 11 12 13 14 15 16 17 18Time,hr

Acc

umul

ativ

e pr

oduc

tivity

,kg/

day

22/7/2009

22/8/2009

22/6/2009

m.cw = 400 kg/hr

m.s = 300 kg/hr

m.a = 250 kg/hr

0

0.5

1

1.5

2

2.5

3

3.5

4

9 10 11 12 13 14 15 16 17 18

Time,hr

Acc

umul

ativ

e pr

oduc

tivity

,kg/

day

22/01/2009

22/12/2008

22/11/2008

m.cw = 400 kg/hr

m.s = 300 kg/hr

m.a = 250 kg/hr


115

Fig. 5: The hourly variation of solar radiation during Fig. 8: Hourly accumulative productivity 22/7/200922/7/2009

Fig. 6: The hourly variation of solar radiation along Fig. 9: The variations of accumulative productivitythree operating days on summer during June, July and August

Fig. 7: The instantaneous Efficiency of solar water Fig. 10: The variations of accumulative productivitycollector during November, December and January

[Date20,23,24 /7/2009]

0

0.2

0.4

0.6

0.8

1

1.2

1.4

1.6

1.8

2

9 10 11 12 13 14 15 16 17 18Time,hr

Wat

er p

ordu

ctiv

ity ,k

g/m

2

ma250kg/hr(theo)

ma200 kg/hr (theo)

ma100kg/hr (theo)

ma100kg/hr(exp)

ma200kg/hr(exp)

ma250kg/hr(exp)

m.s = 300 kg/hr

m.cw = 400 kg/hr

0

2

4

6

8

10

12

9 10 11 12 13 14 15 16 17 18

Time,hr

Acc

umul

ativ

e pr

oduc

tivity

,kg/

day

ms200 kg/hr[Date,20-7-2009]

ms300 kg/hr[Date,21-7-2009]

ms400 kg/hr[Date,22-7-2009]

m.a = 300kg/hr

m.cw = 400kg/hr

0

2

4

6

8

10

12

9 10 11 12 13 14 15 16 17 18

Time ,hr

Acc

umul

ativ

e pr

oduc

tivity

,kg/

day

ma250 kg/hr[Date,20-7-2009]

ma200 kg/hr[Date,23-7-2009]

ma150 kg/hr[Date,24-7-2009]

m.s = 300 kg/hr

m.cw = 400 kg/hr

0

2

4

6

8

10

12

9 10 11 12 13 14 15 16 17 18

Time ,hr

Acc

umul

ativ

e pr

oduc

tivity

,kg/

day

mcw= 400kg/hr[Date,22/7/2009]

mcw= 300kg/hr[Date,23/7/2009]

mcw= 200kg/hr[Date,24/7/2009]

m.s = 300 kg/hr

m.a = 250 kg/hr

22/7/2009

0

0.5

1

1.5

2

2.5

0 500 1000

Solar radiation ,W/m2

Pord

uctiv

ity ,k

g/hr

.m2

m.cw = 400 kg/hr

m.s = 300 kg/hr

m.a = 250 kg/hr

22/7/2009

200

300

400

500

600

700

800

900

1000

9 10 11 12 13 14 15 16 17 18

Time ,hr

Sola

r in

tens

ity,W

/m2

28.5

29

29.5

30

30.5

31

31.5

32

32.5

Am

bien

t tem

pera

ture

,o C

G(exp)Ta


116

Fig. 11: Effect of feed water mass flow rate on hourly Fig. 14: Effect of dehumidifier cooling water flow rate onsystem productivity accumulative productivity

Fig. 12: Effect of feed water mass flow rate on Fig. 15: Effect of solar radiation intensity on theaccumulative productivity productivity

Fig. 13: Effect of air mass flow rate on accumulative surface and ambient temperature duringproductivity 22/7/2009

Fig. 16: Variation of solar radiation on a horizontal

22/7/2009

200

300

400

500

600

700

800

900

1000

9 10 11 12 13 14 15 16 17 18

Time ,hr

Sola

r In

tens

ity ,W

/m2

0

10

20

30

40

50

60

70

80

90

100

Rel

ativ

e hu

mid

ity ,%

solar intensity

RHin(%)

RH out_exp (%)

m.cw = 400 kg/hr

m.s = 300 kg/hr

m.a = 250 kg/hr

1.5

1.7

1.9

2.1

2.3

2.5

2.7

2.9

15 20 25 30 35 40 45Ta,?C

Wat

er P

rodu

ctiv

ity ,k

g/hr

.m2

RH 20%

RH 40%

RH 60%

RH 80%

m.a = 250 kg/hr

m.cw = 400 kg/hr

m.s = 300 kg/hr

0

2

4

6

8

10

12

9 10 11 12 13 14 15 16 17 18Time ,hr

Acc

umul

ativ

e pr

oduc

tivity

,kg/

day

Wooden honey comb[Date 22-7-2009]

cloth[Date,23-7-2009]

m.cw = 400 kg/hr

m.s = 300 kg/hr

m.a = 250 kg/hr

14,15,16/7/2009

0

0.5

1

1.5

2

2.5

9 10 11 12 13 14 15 16 17 18

Time ,hr

Wat

er p

ordu

ctiv

ity ,k

g/m

2

200

300

400

500

600

700

800

900

1000

Sola

r In

tens

ity,W

/m2

Vh= 0.0396 Vh= 0.0264

Vh= 0.0132 G

m.cw = 400 kg/hr

m.s = 300 kg/hr

m.a = 250 kg/hr

0

2

4

6

8

10

12

50 55 60 65 70 75 80

Inlet water temperature,oC

Wat

er p

rodu

ctiv

ity,k

g/da

yms = 300 kg/hr

ms = 200kg/hr

ms = 100 kg/hr

m.a = 250 kg/hr

m.cw = 400 kg/hr

0

2

4

6

8

10

12

14

35 40 45 50 55 60 65 70 75Inlet air temperature ,°C

Wat

er P

rodu

ctiv

ity ,

kg/d

ay

ma=100 kg/hrma=200 kg/hrma=300 kg/hrma=400 kg/hr

m.s = 300 kg/hr

m.cw = 400 kg/hr


117

Fig. 17: Hourly variation of relative humidity of air at the Fig. 20: Effect of different volume of honey- combinlet and outlet of the humidifier packing materials on the productivity

Fig. 18: Effect of relative humidity on water productivity Fig. 21: Effect of humidifier inlet water temperature on theat different air temperature unit productivity

Fig. 19: Effect of two different packing materials on Fig. 22: Effect of humidifier inlet air temperature on thehourly accumulative productivity unit productivity

25

30

35

40

45

50

55

60

65

70

75

80

300 400 500 600 700 800 900 1000

Solar intensity,W/m2

Tem

pera

ture

,o C

0.3

0.4

0.5

0.6

0.7

0.8

0.9

1

1.1

1.2

Wat

er p

ordu

ctiv

ity ,k

g/hr

.m2

Tah,iTac,iTac,oTamd

m.cw = 400 kg/hr

m.s = 300 kg/hr

m.a = 250 kg/hr

0

2

4

6

8

10

12

14

15 20 25 30 35 40

Tcwi,oC

Wat

er p

rodu

ctiv

ity,k

g /d

ay

mcw= 400kg/hr [Date,20-7-2009]

mcw= 300kg/hr [Date,21-7-2009]

mcw= 200kg/hr [Date,19-7-2009]

m.s = 300 kg/hr

m.a = 250 kg/hr

21/01/2009

0

200

400

600

800

1000

1200

8 9 10 11 12 13 14 15 16 17 18

Time,hr

Sola

r Int

ensi

ty,W

/m2

5 10 15

25 35 4045 50 55

21/07/2009

0

200

400

600

800

1000

1200

6 7 8 9 10 11 12 13 14 15 16 17

Time ,hr

Sola

r R

adia

tion

Inte

nsity

,W/m

2

5 10 1525 35 4555


118

Fig. 23: Effect of solar intensity on humidifier inlet air Fig. 26: Effect of the collector tilt angle on the solartemperature and water productivity radiation unit in summer

Fig. 24: Effect of different temperature of cooling inlet Figure (20) shows the effect of different honey-seawater on the water productivity comb packing material in the humidifier on the

Fig. 25: Effect of the collector tilt angle on the solar unit air temperature and hence increases the waterradiation unit in winter productivity.

The solar water productivity of the unit increases with theinlet temperature and relative humidity. It is clear from thecurve also that the humidity has a strong influence on thesystem performance especially at the higher inlettemperature of air.

Two different packing materials have been used inthe test rig. The total surface area of the packing has beenkept constant. However, the wetted area differs from onetype to anther depending on the nature of each material.Figure 19 shows the variation of the daytime accumulativeproductivity for the two packing materials under sameoperating conditions.

productivity.Fig.21 shows the effect of the humidifier inlet water

temperature on the unit productivity. It can be noticedthat, the productivity decreases with increasing the waterflow rate. This can be explained by the fact that increasingthe water flow rate leads to a decrease of the enthalpy ofair, hence, the he ability to extract vapor from the air isdecreased.

The effect of inlet air temperature on the productivityis shown in Fig. 22. It can be seen from the curve that, theability of air to carry the water vapor increases with theincrease of the inlet air temperature.

Fig. 23 shows the effect of solar intensity on thehumidifier inlet temperature. It can be seen from thecurve that, the increase of solar intensity increases the


119

Figure 24 shows the effect different of inlet water Symbols:temperature on the productivity. It can be seen from the d Mean wall spacing, mcurve that, the productivity increases with decrease of K Thermal conductivity (W/m K)inlet cooling water temperature. Ps Saturated water vapor pressure, Pa

The effect of the solar collector tilt angle variation on Pv Water vapor pressure of the air, Pathe solar desalination unit is theoretically investigated Vh Volume of packing materials in the humidifier, m3and presented in Fig. 25. It is evident that, the u,v Velocity, m/sproductivity increases with the increase of tilt angle slope h Heat transfer coefficient, W/(m .s)in winter. However Fig. 26 show that, the productivity r Radius, mincreases with the decrease of tilt angle slope in summer. RH Relative humidity

CONCLUSIONS

A humidification - dehumidification (HD) system has Thickness (m)been designed, installed and outdoor tested in the faculty Thermal diffusitivity, m2/sof Engineering, Suez Canal University, Port Said, Egypt. Kinematic viscosity, m2/sThe system consists of an insulated desalination chamber Latent heat of water, kJ/kgcoupled with a solar Collector. The desalination chamber µ Dynamic viscosity, N/m2.sdivided into evaporator tower and dehumidifier tower. Density, kg/m3A parametric analysis was conducted based on numerical Coefficient of evaporationsimulation in order to optimize the system performance. V-groove angle of solar air heater, degreeThe system was found to be suitable to provide drinkingwater for population or remote arid areas. It can be seen Subscripts:from the above test results that: b Bottom, box

A mathematical model is written for the all system c w Cooling watercomponents (collectors, humidifier, dehumidifier) to I Inlet state, insulationpredict the system productivity under different and s Saline waterwide range of operating conditions. si SideIt is found that the accumulative productivity p Plateincreases with the increase of inlet temperature and n Liquids interfacerelative humidity. o OutletIncreasing the seawater flow rate to decrease v Water vaporproductivity of due to the lower evaporation causedby the lower temperature. However, reducing the REFERENCESseawater flow rate to extreme values was found tolower the water production due to the dramatic drop 1. Fath, H.E., 2000. Desalinationin the collector efficiency at high temperatures. Technology, The Role of Egypt in Region, IWT CThe water productivity of the unit increases with the 2000, Alexandria, Egypt.increase of packing material volume. 2. Fath, H.E., 1998. Solar DesalinationThe water productivity increases with the increase of Promising Alternative for Fresh Watercollector tilt angle in winter and decrease in summer. Production with Free Energy Simple TechnologyThe results of the proposed mathematical model are and Clean Environmental, Desalination,in good agreement with those of the experimental 116: 45-56.model of the system. 3. Abdelkader, M., A.S. Nafey A. Abdelmotalip andThe system daily productivity in summer is about A.A. Mabrouk, 2000. Parameters Affecting Solar Still7.26 to 11 kg/m day and about 2 to 3.5 kg/m day in Productivity, Energy Conversion and Management,2 2

winter. 411: 797-809.

2

Rf Fouling resistance, m oC/kW2

Greek:

c Dehumidifier


120

4. Abdelkader, M., A.S. Nafey, A. Abdelmotalip and 11. Garg, H.P., R.S. Adhikari and R. Kumar, 2002.A.A. Mabrouk, 2001. Enhancement of Solar Still Experimental Design and Computer Simulation ofProductivity Using Floating Perforated Black Plate, Multi-Effect Humidification (MEH)-DehumidificationEnergy Conversion and Management, 43: 1401-1408. Solar Distillation, Desalination, 153: 81-86.

5. Abdel-Monem, M., M. Abdel-Salam and Abdel- 12. Fath, H.S. and A. Ghazy, 2000. Solar DesalinationWahed, 1988. Theoretical and Experimental Studies of Using Humidification-Dehumidification, IWT C,Humidification-Dehumidification Desalination Alexandria, Egypt.System, Ph.D. Thesis, Department of Mechanical 13. Darwish, M.A., XXXX. Experimental and TheoreticalEngineering, Faculty of Engineering, Cairo Univ., Study of Humidification - Dehumidification DesaltingEgypt. System, Desalination, 94: 11-24.

6. Ettouney, H., 2005. Design and Analysis of 14. Al-Hallaj, S. and M. Farid, 1998. Solar DesalinationHumidification Dehumidification Desalination with a Humidification-Dehumidification Cycle:Process, Desalination, 183: 341-352. Performance Unit, Desalination, 120: 273-280.

7. Yamali, C. and I. Solmus, 2007. Theoretical 15. Moustafa M. Elsayed, Ibrahim S. Taha and Jaffer A.Investigation of a Humidification Dehumidification Sabbagh, 1994. Design of Solar Thermal Systems,Desalination System Configured by a Double Jeddah, Saudi Arabia.Pass Flat Plate Solar Air Heater, Desalination, 16. John, A., D. William and A. Beckman, 1980. Solar205: 163-177. Engineering of thermal processes, New York, USA:

8. Marmouch, H., J. Orfi and S. Ben Nasrallah, 2007. John Wiley.Effect of a Cooling Tower on a Desalination System, 17. MacLaine, I.L., Cross and P.J. Banks, 1981. J. HeatDesalination, 238: 281-289. Transfer, 103: 579-585.

9. Dai, Y.J., R.Z. Wang and H.F. Zhang, 2002. Parametric 18. Kettleborough, C.F. and C.S. Hsieh, 1983. J. HeatAnalysis to Improve the Performance of a Solar Transfer, 105: 366-375.Desalination Unit with Humidification and 19. Tchinda, R., 2009. A review of the mathematicalDehumidification, Desalination, 142: 107-118. models for predicting solar air heaters systems.

10. Gao, P., L. Zhang and H. Zhang, 2008. Performance Renewable and Sustainable Energy Reviews,Analysis of a New Type Desalination Unit of Heat 640: 1-26.Pump with Humidification and Dehumidification,Desalination, 202: 531-537.

a theoretical and experimental study for a humidification

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