ERJ Engineering Research Journal Faculty of Engineering Minoufiya University
A THEORETICAL AND EXPERIMENTAL STUDY FOR A HUMIDIFICATION - DEHUMIDIFICATION (HD) SOLAR DESALINATION UNIT
M.Abd ElKader, Aref., A., Gamal H. Moustafa, Yasser ElHenawy Department of Mechanical Power Engineering, Faculty of Engineering, Suez Canal University, Egypt
ABSTRACT A humidification-dehumidification (HD) solar desalination unit was designed. It seems to be suitable to provide drinking water for population or remote arid areas. Solar water and solar air collectors were designed to provide the hot water and air to the desalination chamber. The desalination chamber was divided into humidifier and dehumidifier towers. The circulation of air in the two towers was maintained by the forced convection. 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, heat and mass transfer inside the humidifier and dehumidifier. Such a technique was performed in order to increase the unit performance. Heat and mass balance was done and a set of governing equations was solved using the finite difference technique. The solar intensity was measured along the working day during the summer and winter months and a comparison between the theoretical and experimental results were performed. The average accumulative productivity of the system in November, December and January was ranged between 2 to 3.5 kg / m 2 day while the average summer productivity was found between 6 to 8 kg/m 2 day in June and 7.26 to 11 kg/m 2 day in July and August. دف ا ث إ ا ء ا ( ا ر أو ا ر ) ث د ظر ا رط ب وإ زاف ا رطو دام ا ط ا . ون ا ز ل ر ن وى ا ط ن ا ه وأ ر ن ا واء و ر ج رط ب و دل رارى زاف ا رطو . ث م ا ل دور ن ن ، ا ول ن ِ ء ا ذ َ وا ن ِ ا واء. و م ا دام وذج ظ ري ر ھ ذا ا ظ م د را ر و ت ا ز ا ا و ِط س و ا ل إ ا ز. وو د أن ا وذج ا ر ق ورة دة ا ا و م ره أ ر ن . و ن ا أ أن دل إ ج ا و دة ر وة دل ر ن ا واء، و دل ر ن ة ا ذ ، و دل ر ن ه ا ر د، و ذ ك دة ا ا ل ا وم. وو د أن دل إ ا و دة راوح ن2 : 3.5 م م/ 2 ل ور ل ا ء و ر ود ر و ر وزاد ھ ذا ا دل ون ن6 :8 م م/ 2 ًل ر و ن ل ا ف وأ زاد ھذا ا دل إ 7.26 : 11 م / م2 ل ور و وأ طس ن س ا ل. Keywords: Solar desalination, Solar collector, Humidification and Dehumidification, Solar intensity, Simulation; Finite difference, Water productivity
1. INTRODUCTION source, cost-free energy, and (ii) reduced operating Solar desalination is a suitable solution to supply costs and its simple structure so that it is suitable for remote regions in Egypt and other countries with a few families or small groups in remote areas. fresh water. The standard techniques, like Multi- Solar desalination can either be direct or Stage Flash (MSF), Multi-Effect (ME), Vapor indirect [8]. The direct solar humidification Compression (VC) and Reverse Osmosis (RO), are dehumidification desalination process (solar still) only reliable for large capacity ranges (100–50,000 was investigated by Nafey et al. [17–19]. The m3/day) of the fresh water production [7]. These indirect solar HDH process has the advantage of technologies are expensive for small amounts of separating the heating surface from the evaporation fresh water and also can not be used in locations zone, and therefore, the heating surface is relatively where there are limited maintenance facilities. In protected from corrosion or scale deposits. In fact, addition, the use of conventional energy sources to the humidification dehumidification desalination drive these technologies has a negative impact on the (HDD) is more costly than the conventional basin environment. Solar desalination processes is a future still solar desalination process [7]. Therefore, for the promising technology, has the following advantages simplicity of design, a modest level of manufacturing [8]. (i) Solar energy is environmental friendly energy technology is needed.
Engineering Research Journal, Vol. 33, No. 2 April 2010, PP: 119-130 119 © Faculty of Engineering, Minoufiya University, Egypt M.Abd ElKader, Aref., A., Gamal H. Moustafa, Yasser ElHenawy, " A Theoretical and Experimental … "
Very limited configurations of the HDD process have falling film at the surface of a honeycomb wall in the been developed. Abdel-Monem et al. [1] humidifier. The structure sketch of the humidifier is experimentally and theoretically investigated the shown in Fig.2. It consists of a porous and durable main parameters effect on the productivity of the honeycomb wooden, which is characterized by a HDH process. Indoor tests with constant input energy large evaporation surface per unit volume of packed to the system were performed. The configuration material. The air is heated by a solar air collector and consisted of humidifier, dehumidifier and heating then is forced to pass through the humidifier where it equipments for air and water. The effect of weather becomes hot and humid due to the heat and mass conditions was not considered, and the unsteady state exchange between the seawater and air. Then, the performance of the system was not, therefore, humid air is cooled by the cold seawater when examined. Ettouney [6] made design and analysis of passing through the condenser at which the water various schemes for water desalination by vapor condenses and turns into fresh water. The humidification and dehumidification system (HDH). condenser is a helix-tube type one. The cold seawater A theoretical investigation of a humidification and flows in the tube channel and the fresh water is dehumidification desalination system configured by a produced from the condensate on the condenser double –pass flat plate solar air heater was examined surface. The remaining seawater drawn from the by Yamali and Solmus [21]. The effect of a cooling humidifier is cooled and collected at the bottom tower on a solar desalination system was studied by basin. Some of the water is fed to the solar collector Marmouch et. al [15]. Dai, et. al [4] reported the again since it is not very concentrated and is still parametric analysis to improve the performance of a warm. Thus, the fresh water can be continually solar desalination unit with a humidification and produced. The related parameters of the unit are as dehumidification. The performance analysis of a follows: The fan rated power is 500 W, the desalination unit of heat pump with humidification humidifier volume is 800 × 700× 600 mm 3, the solar and dehumidification was studied by Gao, et. al [10]. water collector area is 1.5 × 1 m 2, the solar air An experimental design and computer simulation of collector area is 2 × 1 m 2, and the dehumidifier heat multi-effect humidification (MEH) dehumidification transfer area is 0.8 2.Eeperments done the range of solar distillation was investigated by Garg et. al [11]. wind speed from 0.5:3.5m/sec. Fath and Ghazy [9] studied numerically a Various variables characterizing the present configuration that consists of a solar air heater, problem were measured. The water mass flow rate humidifier and dehumidifier. The feed water was not and unit productivity were measured by collecting heated, and hence, the effect of water temperature the amount of water in a graduated cylinder during a was not considered. Darwish [3] suggested a system prescribed time interval. The air mass flow rate was consisting of a solar pond, a humidifying column and determine by measuring the air velocity by rotating a dehumidifying stack. He reported that the air flow vanes anemometer ; with the range 0.2 : 40 m/s with rate has a strong effect on the system productivity; a resolution of 0.01 m/s. The air temperature was however, the feed water flow rate has a weak effect measured by a standard K - type thermocouple ; with on the productivity. On the other hand, Al-Hallaj and the range -50: 1000°C, with a resolution of 0.1 : Farid [2] reported that the feed water flow rate has a 199.9 °C and 1 °C above ).The Global solar radiation strong effect on the system productivity and the was measured by a silicon cell Pyranometer effect of the air flow rate is weak. In the present .accuracy with an of +/-5%.The relative humidity at article a theoretical study of humidification- the inlet and outlet of each component were dehumidification desalination unit which consists of measured by hygrometer; with the range 5 : 95 % RH flat plate water and air solar collectors, storage tank and resolution of 0.2 %). and dehumidifier is presented. The aim is to develop 3. MATHEMATICAL MODEL a computer program and to build an experimental apparatus that can be used to study the effect of A mathematical model of the unit is presented, in operating and design parameters as well as weather which the thermodynamic relations were used to conditions on the unit performance and fresh water study the flow, heat and mass transfer inside the productivity. A Comparison with experimental humidifier, dehumidifier, and water and air solar results are given. collectors. A set of governing one dimensional equations has been given and solved using a finite 2. SYSTEM DESCRIPTION difference technique. A computer program is A solar desalination unit with a humidification developed to evaluate the mathematical model. The and dehumidification cycle, which is configured following assumptions are considered: (i) the feed mainly by a flat plate solar water collector , V-groove water, cooling water and feed air are constant flow solar air collector, a humidifier and a condenser, is rate along the day. (ii) Thermal losses to the ambient shown in Fig. 1. Seawater is heated by a solar water are neglected. (iii) The energy stored in the collectors collector and then is sprayed by a sprinkler to form a is neglected.
120 Engineering Research Journal, Minoufiya University, Vol. 33, No. 2, April 2010 M.Abd ElKader, Aref., A., Gamal H. Moustafa, Yasser ElHenawy, " A Theoretical and Experimental … "
values of heat transfer coefficients, at the top U t, at the bottom U b, and at side walls U si ,[12]. = + + U L U t U b U si (3) where 1 A U = ⋅ b b δ δ i , b b , b A cl + K i K b 1 A (4) U = ⋅ s si δ δ i , s b , s A cl + K i K b
where A s and A b are side and bottom collector areas 2 (m ). The top heat transfer coefficient U t is calculated from an empirical equation suggested by [12]: = [ − (ζ − )( − ε )] U t 1 45 .0 00259 .0 00144 p Fig.1 A Schematic diagram of desalination system. 1 1. Blower, 2. solar air collector , 3. humidifier, 4. N []()()344 T {}T −T /()N + f 31.0 +1 h dehumidifier, 5.distilled water , 6. seawater tank ,7. p p a win feed pump ,8.valve 9.solar water collector , 10. σ ( 4 − 4 ) ( − ) Tp Ta / T p Ta (5) sprinkler ,11. drain pump, 12. makeup of seawater + []ε + ()− ε −1 + []()+ − ε − p .0 0425 N 1 p 2N f 1 / g N where ζ is the collector tilt angle, N is the number of glass plate, εg is the emittance of glass, εp the emittance of absorber plate, T p is the average plate temperature and T a is the ambient temperature. σ is the Boltzmann constant. The term f is given as: = ( − + × −4 2 )( + ) f 1 04.0 hwin 5 10 hwin 1 .0 058 N (6)
where h win is the wind heat transfer coefficient and is = + given by, hwin 7.5 5.3 C win where, C win (m /sec) is the wind speed. The collector efficiency factor F cl is estimated by, Fig. 2 A Sketch of honey-comb packing material in (7) the humidifier = 1 Fcl 1 1 1 SU + + L U []D + ()S − D η C πD h 3.1. The Flat Plate Solar Water Collector L o o f bond i t f
The thermal energy balance of the solar collector is where S is the tube spacing (m), D i and D o are the written as follows, [16]: inner and outer diameters of the collector tubes (m). × = + + The collector fin efficiency ηf is calculated by, [12]: G t A cl Q loss Q u Q stg (1) δ ( − ) 2 tanh U K S D / 2 Where G t is the total solar intensity (W/ m ), A cl is L p p o (8) 2 η = the collector surface area (m ), Q loss is the heat loss f U K δ ()S − D / 2 from the collector and Q u is the useful heat L p p o transferred from the absorber to water. Q stg is the energy stored in the collector. The heat loss is For the flow with ( Re < 2300), the convective heat expressed by, [12]: transfer coefficient h tf and the Prandtl number, Pr are given by, [16]: = ( − ) (2) Q loss Acl U L T p T a 14.0 K 1 3 µ f ()1 3 D i f (9) Where T p is the average flat plate temperature (K) h = 86.1 Re Pr t f D L µ through the collector and T a is the ambient i f ,t temperature (K). The overall heat transfer coefficient 2 of the collector U L (W/ m ) is the summation of three
Engineering Research Journal, Minoufiya University, Vol. 33, No. 2, April 2010 121 M.Abd ElKader, Aref., A., Gamal H. Moustafa, Yasser ElHenawy, " A Theoretical and Experimental … "
The Reynolds number (Re) and Prandtl number (Pr) 3.2. Humidifier Modeling are given by, [12]: The humidifier consists mainly of honeycomb . µ wooden which is characterized by a large evaporation 4 m A C cl f p , f (10) Re = Pr = surface per unit volume of packed material. The π D N µ K i t f f mathematical model of the humidifier, which is 2 established on the basis of energy and mass balances, Where k f (W/m k), f (N sec/m ) and C p (J/kg k) the is presented. The heat and mass transfer between the thermal conductivity, dynamic viscosity and specific liquid film and the moist air is considered. The set of heat of the fluid are respectively. L is the tube length governing two dimensional equations is solved using (m). The absorbed solar radiation (q abs ) is determined the finite-difference method. Assumptions are taken by, [14]: into account are: (i) the flow of falling film is laminar = τα and no wavy movement takes place at the liquid-gas q abs G t ( ) avg (11) interface. (ii) Flow rate of falling film is believed τα constant because the evaporation rate of water out of Where ( ) is the average Transmittance- the liquid-gas interface is small enough . Figure 3 absorptance of cover plate. Also the heat removal shows the coordination set-up. Pleased along the factor ( F ) is calculated from, [12]: R interface line. . . m C p − U L Fcl / m C F = 1 − e p (12) R U L . where m is the mass flow rate per unit area of collector (kg/s. m 2). The useful energy from the collector is calculated from, [12]:
Q = A F q − U T − T (13) u cl R abs L f , i a The average fluid temperature is thus obtained as, [12]: Q F T = T + u 1 − R (14) f , a f , i A F U F cl R L cl The average plate temperature is obtained from, [12]: Fig.3 . Coordination set-up Q / N t (15) T = T + u p f , o −1 3.2.1. Governing equations of falling film 1 1 + If the liquid film falls slowly and steadily just π h tf D i L C bond L under the force of gravity, particularly in laminar flow, the mean velocity of the falling film, u s is given Where N t is the number of tubes and C bond is the conductive resistance of the bond. The fluid as [4]: temperature at the outlet from the collector is ρ g δ 2 u = s s (18) calculated from, [12]: s µ 3 s Q T = T + u (16) The thickness δs , Fig. 3 is given as [4]: f , o f , i . m C 1 p 3Γµ 3 δ = s The collector efficiency is determined by, [14]: s (19) gρ 2 Q s η = u (17) Γ c A G Where is the mass flow rate of saline water per cl t unit width of the wall (kg/s/m) and g is the acceleration of gravity, m/s 2. The energy balance for the falling film is
122 Engineering Research Journal, Minoufiya University, Vol. 33, No. 2, April 2010 M.Abd ElKader, Aref., A., Gamal H. Moustafa, Yasser ElHenawy, " A Theoretical and Experimental … "
∂Τ ∂ 2Τ The coefficient of heat transfer h (W/m 2 K) is given s = α s by, [14]: us s 2 (20) ∂ ∂ β 2 o z y2 h = ξ V W/m C (26) 3.2.2. Governing equations of moist air The governing equations, including mass, The terms ξ and β are a function of the plate spacing. momentum, and energy conservations are listed as For a 6.7mm plate spacing the values of ξ and β are follows [4]: 54 and 0.7 respectively. As the plate spacing ∂u ∂v increases to 13.4 mm, the values of ξ and β are + = 0 (21) changed to be 49 and 0.6. ∂ ∂ x y Taking the phenomenon of evaporation into 2 account, introduced the concept of the wet bulb ∂u ∂u ∂ u 1 ∂p * 2 u + v = γ − (22) coefficient of heat transfer (h (W/m K) to correct the ∂ ∂ a ∂ 2 ρ ∂ x y y a x heat transfer process of a wet surface: 2 * = ( + λ ) ∂Τ ∂Τ ∂ Τ h h 1 e /C pa (27) u + v = α (23) ∂x ∂y a ∂y 2 Where h (W/m 2 oC) is the heat transfer coefficient and e is constant, given by: Boundary conditions: = (ω −ω ) ( − ) d ∂u e max min / Tmax Tmin (28) At y = − δ , u = u , = ,0 Τ = Τ s e ∂ e 2 y Here T max , T min , are the maximum and minimum temperatures at the liquid-gas interface. T is the where u e (m/sec) and T e (K) are the velocity and max temperature of incoming air. higher one of the temperatures of gas flow and water flow. T is the wet bulb temperature of the air. • At the water - air interface: min , ωmax and ωmin are the maximum and minimum u = ,0 Τ = Τ , and Τ = Τ humidity ratios corresponding to T max and T min.. n s n , 3.3. Dehumidifier modeling ∂Τ ∂Τ λW s j An energy balance between the cooling water α = = −α = + a ∂ y 0 s ∂ y2 0 ρ and the air/condensing vapor stream is made the y y2 sC ps energy balance is based on two following thermal equations [6]:
Where T n is the temperature at interface,T s is the . . ()()− = − temperature of saline water, W j is the evaporation mcw C pcw Tcwo Tcwi ma H o H i (29) rate of water, and λ is the latent heat of water. Solving the set flow equations (21-23) with boundary . ()− = (30) conditions, the distribution of the flow velocity and mcw C pcw Tcwo Tcwi U c Ac LMTD c temperature inside the humidifier were obtained. To where m cw ,and m a (kg/sec) are the mass flow rate of simplify the above solution, auxiliary equations are cooling water and incoming air. T cwo and T cwi (K) are needed. the temperature of cooling water at inlet and outlet. Hi and H o (Kj/kg) are the enthalpy of air at inlet and Auxiliary equations outlet. In Eq. (30) the logarithmic mean temperature difference and the overall heat transfer coefficient are The evaporation rate at the water surface estimated given by the following relations [6]: from, [13]:
1 M 2 ( − )− ( − ) =ε ()− (24) Tac Tcwo Tac Tcwi Wj 1 Ps Pv = (31) 2πRT LMTD c n ()T − T ln ac cwo Here M is the molecular weight of water and R is the ()− Tac Tcwi universal gas constant (8.314 kJ/kmol k). ε1 is the Knudsen coefficient of evaporation. It is calculated r from: ε = (2 ε)/(2- ε). The coefficient of evaporation, co 1 rco ln ε is calculated from: 1 1 r r 1 = co + + ci + (32) 1 Rf c 2 2πRT T U c hci rci kc hco ε = h * s n (25) ρ λ 2 M v The production of distilled water (m d ) is given by the following equation
Engineering Research Journal, Minoufiya University, Vol. 33, No. 2, April 2010 123 M.Abd ElKader, Aref., A., Gamal H. Moustafa, Yasser ElHenawy, " A Theoretical and Experimental … "
. = ()ω −ω md ma o i (33) ω ω Where o , and i are the absolute humidity (kg
H2O/kg dry air) of air at the inlet and outlet from the condenser. 3.4. V-groove solar air collector A V- shaped flat plate solar air collector is designed, Fig. 4. The calculation of the solar air collector are made with the following assumptions: (i) performance is steady state; (ii) There is no absorption of solar energy by a cover in so far, (iii) Fig.4 V-Groove air collector The radiation coefficient between the two air duct surfaces is found by assuming a mean radiant temperature equal to the mean fluid temperature. (iv) 4. THE SOLUTION PROCEDURE Loss through front and back are of the same The procedure is as follows: temperature .The energy balances for the absorber 1. The input the known data, such as the mass flow plate, back insulation and working fluid of the rates of air, feed seawater, cooling water, inlet collector are given by the following, [22]: temperature and humidity of the air are given. 2. The air and water temperature were obtained. In the single pass v-groove absorber: 3. The fresh water from the humidifier was collected. For the cover: 5. RESULTS AND DISCUSSION U t (T g-Ta) =h 1 (T p - T g) + h r (T p - T g) (34) Fig.5 shows the variations of solar intensity V-groove absorber: along one day of operation in summer season. The
τg αp G t = h 1(T p-Tg) + h r (T p-Tg) +h 2(T p -Tf) (35) Figure shows that, there is a good agreement between Fluid medium: the theoretical model and the measured data. Fig.6 shows the variations of solar intensity along the three . m C dT days in June, July and August. h2 (T p - T f)+ U b (T a -Tf) = air air (36) W dx Figure 7 shows the characteristic efficiency and operating temperatures of the unit for different solving equations (5), (6) with (34), (35) and (36),the collector feed water .Efficiency curve is based on air, absorber plate temperature and heat transfer gross collector area and is determined using the coefficient were obtained. outdoor conditions. The thermal efficiency is a The pressure drop through the channel in the air function of the outdoor temperature, solar radiation, heater has been computed using the following and the temperature of the fluid at the inlet. expression [22]: Fig. 8.The figure shows, a comparison between the theoretical and measured accumulative . 2 3 productivity. It can be seen from the curve that, there ∆ = + D m air L P f o y (37) is as good agreement between the theoretical and L ρ D experimental results. The accumulative productivity on this day is about 10 kg/day/m 2 .In the real outdoor operation, a delay time was noticed between the start for laminar flow (Re < 2550) of the run time and the starting of fresh water production, as most of the energy received in early f o = 24/Re, y = 0.9 hours is used as sensible heat to warm up the test rig metal pants and the fluid flow. for transitional flow (2550 < Re < 10 4) Figures (9) and (10) show the measured values f 0.15 of the accumulative productivity at different days. It o = 0.0094, y=2.92 Re 4 5 can be seen from the curve that, the average for turbulent flow (10 < Re < 10 ) accumulative productivity of the system in winter (November, December and January) is ranged fo 0.2 = 0.059 Re , y = 0.73 between 2 to 3.5 kg / m 2.day and the average summer productivity is ranged between 7.26 to 11 kg/ 2 m .day( in June and August ).
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The effect of testing parameters in three operating days such as, feed water, air and cooling Figure (20) shows the effect of different honey- water flow rates are presented in Figures (11) and comb packing material in the humidifier on the (12). productivity. It seems from Fig. (13) That, for the same Fig.21 shows the effect of the humidifier inlet weather conditions (solar intensity, ambient water temperature on the unit productivity. It can be temperature and wind speed), the system productivity noticed that, the productivity decreases with of the unit increases with increasing the air flow rate increasing the water flow rate. This can be explained in three excessive days. This is due to the fact that by the fact that increasing the water flow rate leads to increasing the air flow rate leads to an increase in the a decrease of the enthalpy of air, hence, the he ability mass and heat transfer coefficients in the humidifier to extract vapor from the air is decreased. and hence entrained water vapor, which is lead to an increase in the system productivity. The effect of inlet air temperature on the productivity is shown in Fig. 22. It can be seen from The effect of cooling water flow rate on the the curve that, the ability of air to carry the water productivity is illustrated in Fig. (14). It is clear from vapor increases with the increase of the inlet air the curve that, Increasing the cooling water flow rate temperature. in the dehumidifier leads to a decrease in the surface temperature of the air cooler and this will lead to an Fig. 23 shows the effect of solar intensity on the increase in the condensation rate and hence the humidifier inlet temperature. It can be seen from the system productivity curve that, the increase of solar intensity increases the unit air temperature and hence increases the water Figure (15) shows the effect of solar intensity on productivity. the water productivity on 22 July 2009. The measured values of the average solar radiation and Figure 24 shows the effect different of inlet ambient temperature during 22 July 2009, in Port water temperature on the productivity. It can be seen Said-Egypt are given in Figure (16).The Average from the curve that, the productivity increases with relative humidity of the ambient air is 50–60% and decrease of inlet cooling water temperature the wind speed ranges between 1 and 3 m/s. The effect of the solar collector tilt angle Figure (17) represents the air relative humidity at variation on the solar desalination unit is theoretically the inlet and outlet of the humidifier for the indicated investigated and presented in Fig. 25. It is evident climatic conditions. One can notice that, the relative that, the productivity increases with the increase of humidity at the exit of the humidifier is close to the tilt angle slope in winter. However Fig. 26 show that, value of 95%. Due to the air heater device output, the the productivity increases with the decrease of tilt outlet relative humidity appears constant; angle slope in summer. independent on the inlet one. This result means that, 1000 the air leaving the humidifier is practically saturated. 900 Figure 18 shows the effect of the inlet temperature 22/7/2009 2 and relative humidity of air on the unit water 800 productivity. The results agree with the common knowledge in literature, that the higher inlet 700 temperature and humidity are beneficial for 600 producing more fresh water. The solar water productivity of the unit increases with the inlet 500 exp temperature and relative humidity. It is clear from the 400 curve also that the humidity has a strong influence on Solar radiation intensity ,W /m theo the system performance especially at the higher inlet 300 temperature of air. 200 9 10 11 12 13 14 15 16 17 18 Two different packing materials have been used Time ,hr in the test rig .The total surface area of the packing has been kept constant. However, the wetted area Fig. 5 The hourly variation of solar radiation during differs from one type to anther depending on the 22/7/2009 nature of each material. Figure 19 shows the variation of the daytime accumulative productivity for the two packing materials under same operating conditions.
Engineering Research Journal, Minoufiya University, Vol. 33, No. 2, April 2010 125 M.Abd ElKader, Aref., A., Gamal H. Moustafa, Yasser ElHenawy, " A Theoretical and Experimental … "
12 1100 . m cw = 400 kg/hr . 1000 m s = 300 kg/hr 10 . m a = 250 kg/hr 900
2 800 8
700 6 600
Solar Intensity, W/m 500 4
400 Accumulative productivity ,kg/day 12/8/2009 2 22/7/2009 300 15/7/2009 22/8/2009 15/6/2009 22/6/2009 200 0 9 10 11 12 13 14 15 16 17 18 9 10 11 12 13 14 15 16 17 18 Time, hr Time,hr
Fig. 6 The hourly variation of solar radiation along three operating days on summer Fig. 9 The variations of accumulative productivity 76 during June, July and August 4 75 . m cw = 400 kg/hr efficiency factor = 0.95 . 3.5 m s = 300 kg/hr 74 FR = 0.88 . m a = 250 kg/hr 73 3
72 2.5
71 2
70 1.5 FPC thremal efficiency,%69 1 Accumulative productivity ,kg/day 68 22/01/2009 0.5 22/12/2008 67 22/11/2008 0.01 0.012 0.014 0.016 0.018 0.02 0.022 0.024 0.026 0 (Ts - Ta )/G 9 10 11 12 13 14 15 16 17 18 Time,hr Fig.7 The instantaneous Efficiency of solar water collector Fig.10 The variations of accumulative productivity during November, December and January 12 22/7/2009 [Date20,23,24 /7/2009] Theo 2 ma250kg/hr(theo) 10 exp 1.8 ma200 kg/hr (theo)
1.6 8 ma100kg/hr (theo) 2 1.4 ma100kg/hr(exp)
6 1.2 ma200kg/hr(exp)
1 ma250kg/hr(exp) 4 0.8
. 0.6 W ater porductivity ,kg/m Accumulative productivity ,kg/day m = 400 kg/hr 2 cw . m s = 300 kg/hr 0.4 . . m a = 250 kg/hr m s = 300 kg/hr 0.2 . 0 m cw = 400 kg/hr 9 10 11 12 13 14 15 16 17 18 0 Time ,hr 9 10 11 12 13 14 15 16 17 18 Fig. 8 Hourly accumulative productivity 22/7/2009 Time,hr Fig.11 Effect of feed water mass flow rate on hourly system productivity
126 Engineering Research Journal, Minoufiya University, Vol. 33, No. 2, April 2010 M.Abd ElKader, Aref., A., Gamal H. Moustafa, Yasser ElHenawy, " A Theoretical and Experimental … "
12 2.5 ms200 kg/hr[Date,20-7-2009] 22/7/2009
10 ms300 kg/hr[Date,21-7-2009] 2 ms400 kg/hr[Date,22-7-2009] 8 2 1.5 6
1 4 Porductivity ,kg/hr.m . . m cw = 400 kg/hr Accumulative productivity2 ,kg/day m a = 300kg/hr 0.5 . . m s = 300 kg/hr m cw = 400kg/hr . m a = 250 kg/hr 0 0 9 10 11 12 13 14 15 16 17 18 0 500 1000 Time,hr 2 Solar radiation ,W/m Fig. 12 Effect of feed water mass flow rate on accumulative productivity Fig.15 Effect of solar radiation intensity on the productivity 12 1000 32.5 ma250 kg/hr[Date,20-7-2009] 22/7/2009 ma200 kg/hr[Date,23-7-2009] 10 900 32 ma150 kg/hr[Date,24-7-2009] 800 31.5 C o 8 2 700 31
6 600 30.5
500 30 4 Solar intensity,W /m Ambient temperature, 400 29.5 Accumulative productivity ,kg/day productivity Accumulative
2 . m s = 300 kg/hr G(exp) . 300 29 m cw = 400 kg/hr Ta
0 200 28.5 9 10 11 12 13 14 15 16 17 18 9 10 11 12 13 14 15 16 17 18 Time ,hr Time ,hr Fig.13 Effect of air mass flow rate on accumulative Fig.16 Variation of solar radiation on a horizontal productivity surface and ambient temperature during 22/7/2009 12 1000 100 mcw= 400kg/hr[Date,22/7/2009] 90 900 10 mcw= 300kg/hr[Date,23/7/2009] mcw= 200kg/hr[Date,24/7/2009] 80 800 22/7/2009
2 70 8 700 60 solar intensity 6 600 50 RH in(%) 40 500 RH out_exp (%) 4 Relative humidity ,% Solar Intensity ,W /m 30
400 . m cw = 400 kg/hr
Accumulative productivity,kg/day 20 2 m. = 300 kg/hr s m. = 300 kg/hr . 300 s m a = 250 kg/hr . 10 m a = 250 kg/hr 0 200 0 9 10 11 12 13 14 15 16 17 18 9 10 11 12 13 14 15 16 17 18 Time ,hr Time ,hr Fig.14 Effect of dehumidifier cooling water flow rate Fig.17 hourly variation of relative humidity of air at on accumulative productivity the inlet and outlet of the humidifier
Engineering Research Journal, Minoufiya University, Vol. 33, No. 2, April 2010 127 M.Abd ElKader, Aref., A., Gamal H. Moustafa, Yasser ElHenawy, " A Theoretical and Experimental … "
2.9 12
2.7 10 ms = 300 kg/hr RH 20% ms = 200kg/hr
2 2.5 RH 40% ms = 100 kg/hr 8 RH 60% 2.3 RH 80% 6
2.1 4 Water productivity,kg/day
W ater1.9 Productivity ,kg/hr.m 2 . . m a = 250 kg/hr m a = 250 kg/hr . . m cw = 400 kg/hr 1.7 m cw = 400 kg/hr . m s = 300 kg/hr 0 50 55 60 65 70 75 80 1.5 o Inlet water temperature, C 15 20 25 30 35 40 45 ◦ Ta, C Fig.18 Effect of relative humidity on water Fig.21 Effect of humidifier inlet water temperature productivity at different air temperature on the unit productivity 12 14 Wooden honey comb[Date 22-7-2009] . 12 m s = 300 kg/hr cloth[Date,23-7-2009] . 10 m cw = 400 kg/hr
10
8 8
6 6
4 WaterProductivity , kg/day 4 ma=100 kg/hr ma=200 kg/hr Accumulative productivity ,kg/day . m cw = 400 kg/hr 2 ma=300 kg/hr 2 . m s = 300 kg/hr ma=400 kg/hr . m a = 250 kg/hr 0 35 40 45 50 55 60 65 70 75 0 Inlet air temperature ,°C 9 10 11 12 13 14 15 16 17 18 Time ,hr Fig.19 Effect of two different packing materials on Fig.22 Effect of humidifier inlet air temperature on hourly accumulative productivity the unit productivity 2.5 1000 80 . 1.2 . m cw = 400 kg/hr 14,15,16/7/2009 m cw = 400 kg/hr . 75 m s = 300 kg/hr . 1.1 m s = 300 kg/hr . 900 m a = 250 kg/hr . 70 m a = 250 kg/hr 2 Tah,i 1
2 800 65 Tac,i 2 2 Tac,o 0.9 60 Ta 700 C 1.5 o md 55 0.8 600 50 0.7 1 500 Temperature , 45 0.6 Solar Intensity,W /m Water porductivity ,kg/hr.m W ater porductivity ,kg/m 400 40 0.5 0.5 Vh= 0.0396 Vh= 0.0264 35 300 0.4 Vh= 0.0132 G 30 0 200 25 0.3 9 10 11 12 13 14 15 16 17 18 300 400 500 600 700 800 900 1000 2 Time ,hr Solar intensity,W/m Fig.20 Effect of different volume of honey- comb Fig.23 Effect of solar intensity on humidifier inlet air packing materials on the productivity temperature and water productivity
128 Engineering Research Journal, Minoufiya University, Vol. 33, No. 2, April 2010 M.Abd ElKader, Aref., A., Gamal H. Moustafa, Yasser ElHenawy, " A Theoretical and Experimental … "
14 6. CONCLUSIONS mcw= 400kg/hr [Date,20-7-2009]
12 mcw= 300kg/hr [Date,21-7-2009] A humidification - dehumidification (HD) system has
mcw= 200kg/hr [Date,19-7-2009] been designed, installed and outdoor tested in the 10 faculty of Engineering, Suez Canal University, Port Said, Egypt. The system consists of an insulated 8 desalination chamber coupled with a solar Collector. 6 The desalination chamber divided into evaporator tower and dehumidifier tower. A parametric analysis Water productivity,kgWater /day 4 was conducted based on numerical simulation in order to optimize the system performance. The 2 . m s = 300 kg/hr . system was found to be suitable to provide drinking m a = 250 kg/hr 0 water for population or remote arid areas. It can be 15 20 25 30 35 40 seen from the above test results that: Tcwi, oC 1. A mathematical model is written for the all Fig. 24 Effect of different temperature of cooling system components(collectors, humidifier, inlet seawater on the water productivity dehumidifier) to predict the system productivity 1200 under different and wide range of operating conditions. 1000 21/01/2009 2. It is found that the accumulative productivity increases with the increase of inlet air 2 800 temperature and relative humidity. 3. Increasing the seawater flow rate leads to 600 decrease in water productivity to a certain value depending on the collector efficiency.
Solar Intensity,W400 /m 4. The water productivity of the unit increases 5 10 15 with the increase of packing material volume. 200 25 35 40 5. The optimum tilt angle which gives maximum 45 50 55 solar radiation is found between 35:45 o in 0 winter and 15:20 o in the summer. 8 9 10 11 12 13 14 15 16 17 18 6. The results of the proposed mathematical model Time,hr are in good agreement with those of the Fig. 25 Effect of the collector tilt angle on the solar experimental model of the system. radiation unit in winter 7. The system daily productivity in summer is 1200 about 7.26 to 11 kg/m 2 day, and about 2 to 3.5 21/07/2009 kg/m 2 day in winter. 1000
2 Symbols
800 d Mean wall spacing, m K Thermal conductivity (W/m K) 600 Ps Saturated water vapor pressure, Pa Pv Water vapor pressure of the air, Pa 400 Vh Volume of packing materials in the 5 10 15 Solar Radiation Intensity,W /m 25 35 45 humidifier, m3 200 55 u,v Velocity, m/s
0 h Heat transfer coefficient, W/(m2.s) 6 7 8 9 10 11 12 13 14 15 16 17 r Radius ,m Time ,hr RH Relative humidity Fig.26 Effect of the collector tilt angle on the solar Rf Fouling resistance, m2 oC/kW radiation unit in summer Greek
δ Thickness (m)
α Thermal diffusitivity , m2/s
γ Kinematic viscosity, m2/s
λ Latent heat of water, kJ/kg
Engineering Research Journal, Minoufiya University, Vol. 33, No. 2, April 2010 129 M.Abd ElKader, Aref., A., Gamal H. Moustafa, Yasser ElHenawy, " A Theoretical and Experimental … "
Dynamic viscosity, N/m2.s [10] Gao, P. Zhang, L., and Zhang, H., ρ Density, kg/m3 "Performance Analysis of a New Type Desalination Unit of Heat Pump with ε Coefficient of evaporation Humidification and Dehumidification", υ V-groove angle of solar air heater , degree Desalination 202 (2008) 531-537. Subscripts [11] Garg, H.P., Adhikari, R.S., and Kumar, R., b Bottom, box "Experimental Design and Computer c Dehumidifier Simulation of Multi-Effect Humidification (MEH)–Dehumidification Solar Distillation", cw cooling water Desalination 153 (2002) 81-86. i Inlet state, insulation [12] John A., William, D. and Beckman, A., s Saline water "Solar Engineering of thermal processes", si Side New York, USA: John Wiley; 1980. p Plate [13] Kettleborough, C.F., and C.S. Hsieh, C.S., J. n Liquids interface Heat Transfer, 105 (1983) 366-375. o Outlet [14] MacLaine, I.L., Cross and Banks, P.J., J. Heat Transfer, 103 (1981) 579-585. v Water vapor [15] Marmouch, H., Orfi, J., and Ben Nasrallah, S., 7. REFERENCES "Effect of a Cooling Tower on a Desalination [1] Abdel-Monem M, Abdel-Salam M, and Abdel- System", Desalination 238 (2007) 281-289. wahed, "Theoretical and Experimental Studies [16] Moustafa M., Elsayed, Ibrahim S. Taha, and of Humidification–Dehumidification Jaffer A.Sabbagh., "Design of Solar Thermal Desalination System", Ph.D. Thesis, Department Systems", Jeddah ,Saudi Arabia ,1994. of Mechanical Engineering, Faculty of Engineering, Cairo Univ., Egypt, 1988. [17] Abdelkader M.,Nafey A.S., Abdelmotalip A., and Mabrouk A. A., "Parameters Affecting [2] Al-Hallaj S., and Farid, M., "Solar Desalination Solar Still Productivity", Energy Conversion with a Humidification–Dehumidification Cycle: and Management, 411(2000) 797–809. Performance Unit", Desalination 120(1998) 273–280. [18] Abdelkader M.,Nafey A.S, Abdelmotalip A., and Mabrouk A.A., "Solar Still Productivity [3] Darwish M.A.," Experimental and Theoretical Enhancement", Energy Conversion and Study of Humidification – Dehumidification Management 42 (2001) 1401–1408. Desalting System", Desalination 94, 11–24. [19] Abdelkader M.,Nafey A.S, Abdelmotalip A., [4] Dai, Y.J., Wang, R.Z., and Zhang, H.F., and Mabrouk A.A., "Enhancement of Solar "Parametric Analysis to Improve the Still Productivity Using Floating Perforated Performance of a Solar Desalination Unit with Black Plate", Energy Conversion and Humidification and Dehumidification", Management 43 (2001) 1401–1408. Desalination 142 (2002) 107-118. [20] Nafey, A.S., Mohamad, M.A., El-Helaby, S.O., [5] Eames, I.W., Marr, N.J., and Sabir, H., Int. J. and Sharaf, M.A., "Theoretical and Heat and Mass Transfer., 40 (1997) 2963-2973. Experimental Study of a Small Unit for Solar [6] Ettouney, H. "Design and Analysis of Desalination Using Flashing Process", Energy Humidification Dehumidification Desalination Conversion and Management 48 (2007) 528– Process", Desalination 183 (2005) 341–352 538. [7] Fath H.E. "Desalination Technology- The Role [21] Yamali, C., and Solmus, I., "Theoretical of Egypt in Region", IWT C 2000, Alexandria, Investigation of a Humidification Egypt. Dehumidification Desalination System [8] Fath H.E., "Solar Desalination Promising Configured by a Double Pass Flat Plate Solar Alternative for Fresh Water Production with Air Heater", Desalination 205 (2007) 163-177. Free Energy Simple Technology and Clean [22] Tchinda R."A review of the mathematical Environmental", Desalination 1998 (116) 45–56. models for predicting solar air heaters systems" [9] Fath H.S., and Ghazy A., "Solar Desalination Renewable and sustainable energy reviews 640 Using Humidification–Dehumidification", IWT (2009)1–26. C, Alexandria, Egypt 2000.
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