Desalination 249 (2009) 949–959
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Desalination j o u r n a l h o m e p a g e : w w w. e l s ev i e r. c o m / l o c a t e / d e s a l
Theoretical and experimental investigation of humidification–dehumidification desalination unit E.H. Amer ⁎, H. Kotb, G.H. Mostafa, A.R. El-Ghalban Mechanical Power Engineering Department, Faculty of Engineering, Menoufiya University, Shebin El-Kom, Egypt
a r t i c l e
i n f o
Article history: Received 10 August 2008 Accepted 30 June 2009 Available online 13 October 2009 Keywords: Desalination Humidification Dehumidification Numerical simulation Experimental investigation Water productivity
a b s t r a c t A theoretical and experimental investigation of humidification–dehumidification desalination system is presented. The system is based on an open cycle for water and a closed cycle for the air stream. The air is circulated either by natural or forced circulation. The system modeling is based on various heat and mass balance equations and their numerical solution. The effect of operating parameters on the system characteristics has been investigated. An experimental test set-up has been fabricated and assembled. The set-up has been equipped with appropriate measuring and controlling devices. Detailed experiments have been carried out at various operating conditions and using several packing materials. The heat and mass transfer coefficients have been obtained experimentally and fitted in forms of empirical correlations. The results of the investigation have shown that the system productivity increases with the increase in the mass flow rate of water through the unit. Water temperature at condenser exit increases linearly with water temperature at humidifier inlet and it decreases as water flow rate increases. The higher water temperature at humidifier inlet or water flow rate, the higher is the air temperature and humidity ratio at condenser inlet and exit. A maximum productivity of 5.8 liter/h has been obtained using wooden slates packing and with forced air circulation. No significant improvement in the performance of the desalination unit has been achieved by forced circulation of air at high water temperatures. The average relative deviation of theoretical predictions from measurements is (− 0.9%) in the air temperature at condenser inlet, (3.8%) in the humidity ratio at condenser exit and (− 1%) in the water temperature at condenser exit. © 2009 Elsevier B.V. All rights reserved.
1. Introduction Water desalination using humidification–dehumidification (HD) processes has proven to be an efficient technique of obtaining fresh water from saline water. This technique presents several advantages such as flexibility in capacity, moderate installation and operating costs, simplicity, and possibility of using low temperature energy sources. The (HD) processes are appreciated when the fresh water demand is rather small. The current HD installations are in very compact units containing two exchangers: an evaporator where air is humidified and a condenser where distilled water is recovered. Compared to other distillation processes, the HD process functions at atmospheric pressure so that the components are not submitted to mechanical solicitations [1]. Several studies have been carried out to investigate the characteristics and performance of HD desalination systems. Nawayseh et al. [2] have constructed a simulation program in which the set of nonlinear equations describing the desalination unit were solved numerically. The results of the simulation have been found to agree with the experimental results of two different units constructed in Jordan and Malaysia. The air flow rate
⁎ Corresponding author. Tel.: +2 048 223 5695; fax: +2 048 222 0395. E-mail address:
[email protected] (E.H. Amer). 0011-9164/$ – see front matter © 2009 Elsevier B.V. All rights reserved. doi:10.1016/j.desal.2009.06.063
was found to have an insignificant effect on the productivity of desalinated water. Increasing the water flow rate was found to decrease productivity. In a later study, Nawayseh et al. [3] have used the simulation program to optimize the unit performance by studying the effect of condenser and humidifier areas as well as the feed water flow rate on the performance. Ben Bacha et al. [4] have modeled and simulated a solar multiple condensation evaporation system (SMCES). An experimental validation has been carried out to verify the accuracy of the model and a few requirements have been suggested for the best operation and production for the SMCEC desalination unit. Dai et al. [5] have derived a mathematical model for the unit to simulate the performance numerically. The results revealed that a large error occurred at high inlet temperature of feed water. The water productivity and the hourly water production behave differently with regard to the variation of mass flow rate of air. Optimum mass flow rate of air exists only for hourly water production. The optimum value of the mass flow rate increased with the decrease of the temperature of the inlet water feed to the humidifier. Farida et al. [6] have carried out a simulation study to investigate the performance of solar multi-effect humidification units based on the humidification–dehumidification principle. The study has focused on analyzing the effects of various components involved in the process along with the study of the effect of water flow rate on the desalination production. Nafey et al. [7] have presented a numerical investigation of
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the humidification–dehumidification–desalination (HDD) process in which water is heated using a solar concentrator and the air is heated using a flat-plate solar air heater. It has been reported that the productivity of the unit is strongly influenced by the air flow rate, cooling water flow rate and total solar energy incident through the day. Fath and Ghazy [8] have presented a numerical study of a single-stage humidification–dehumidification solar desalination process. The effect of solar intensity, ambient temperature, wind speed, the dehumidifier effectiveness, air circulation flow rate, feed water rate and temperature on the system productivity has been investigated. It has been reported that increasing the solar intensity and ambient temperature and decreasing wind velocity increase system productivity. Increasing the air flow rate up to 0.6 kg/s increases the productivity, after which it has no significant effect. The feed water flow rate has an insignificant influence on system productivity. The survey of literature reveals that, the published literature contains a few contradictions and mismatch in the drawn conclusions. At the time when most of the researches have agreed on the effect of water temperature at humidifier inlet on the unit productivity, the same is not true for air and water flow rates. Increasing the flow rate of water has been reported to increase the productivity [6], other publications reported the opposite [2]. In the case of the effect of air flow rate, the arguments are much more confusing. It has been reported that its effect is insignificant according to [2], and has a significant effect according to [7]. Other studies claim an optimum value for air flow rate which achieves best productivity, [5,8]. Thus, the objective of the current investigation is to carry out a detailed numerical simulation of a HDD unit to study the effect of operating parameters on the performance. Further, the study includes extensive experiments to investigate the effect of operating conditions on the unit parameters and productivity. The effect of using different packing materials in the humidifier has been experimentally investigated.
Energy balance for water: Equating the energy at segment inlet with the energy at segment outlet in addition to the various heat interactions between the segment and its surroundings leads to: ðm˙ w iw Þy + Δy = ðm˙ w iw Þy + Q w−a + Q evap
ð1Þ
Using Taylor series expansion: ðm˙ w iw Þy + Δy = ðm˙ w iw Þy +
∂ ðm˙ i Þdy ∂y w w
ð2Þ
Substituting into Eq. (1) and simplifying we get: ∂ ðm˙ i Þdy = hw Ah ðTwm −Tam Þ + Dh Ah hfg ðXs −Xam Þ ∂y w w
ð3Þ
Carrying out the differentiation in the left hand side of Eq. (3) leads to: ∂T ∂ m˙ w m˙ w Cpw w + Tw Cpw dy = hw Aw ðTwm −Tam Þ + Dh Ah hfg ðXs −Xam Þ ∂y ∂y
ð4Þ
Energy balance for humid air: The air exchanges heat by convection with water in addition to the heat from the vapor. This heat results in changing the enthalpy of the air stream. The air stream is composed of dry air accompanied by water vapor. The direction of air flow is opposite to the direction of water flow. ∂ ðm˙ i Þdy = ha Ah ðTwm −Tam Þ + Dh Ah hfg ðXs −Xam Þ ∂y a a
ð5Þ
The enthalpy of moist air is expressed as: ia = Cpa Tam + Xa ðCpv Tam + hfg Þ
ð6Þ
Differentiating the left hand side of Eq. (5) leads to: 2. Governing model A typical humidification–dehumidification–desalination unit consists usually of a humidifier, a condenser, and a heating source (either a solar collector or any other heat source). A similar configuration has been considered and a general energy balance has been applied to the different components in order to simulate the unit performance numerically. The mathematical model in the steady state regime allows determining the coupling equations between the water temperature, the humid air temperature and water content inside each component.
′ ∂T ′ ∂X m˙ a Cpa a + m˙ a hfg a dy = Ah ha ðTwm −Tam Þ + Dh Ah hfg ðXs −Xam Þ ∂y ∂y ð7Þ where: ′
′
ð8Þ
Cpa = Cpa + Xa Cpv and hfg = hfg + Cpv Ta Mass balance for water: ∂ m˙ w = Dh Ah ðXs −Xam Þ ∂y
ð9Þ
2.1. Humidifier modeling The mathematical model for the evaporation tower (humidifier) could be formulated by applying mass and energy balances on an elementary volume of height Δy whose surface area of is Ah as shown in Fig. 1.
Mass balance for humid air: The mass of air changes due to the amount of vapor that gets suspended in it as it flows inside the humidifier. Since the rate of dry air is constant, the mass balance is expressed as: m˙ a
∂Xa = Dh Ah ðXs −Xam Þ ∂y
ð10Þ
The vapor content of the air at the saturation condition Xs is calculated in terms of the saturation pressure Ps using the following equation [9]: Xs = 0:62198
Ps 1−Ps
ð11Þ
The saturation pressure for the water vapor Ps corresponding to a certain temperature is computed from [9] as: + 21:240964−2:71119 × 10 ½ −6096:938 T + 1:67395 × 10 T + 2:4335 lnðT Þ
Ps = Exp Fig. 1. An element of the humidifier.
a
−5 2 a
a
−2
Ta
ð12Þ
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where, Ta is the absolute temperature of air in Kelvin.The mass exchange coefficient (Dh) and the heat exchange coefficients (hw, ha) as functions of the air flow rate (ṁa) and water flow rate (ṁw) are taken from Ref. [9] as: 0:11515
2:09 m˙ a
Dh =
0:45
m˙ w
a
0:5894
; hw =
5900 m˙ a
0:169
m˙ w
a
and ha = Cpa Dh ð13Þ
2.2. Condensation tower modeling Energy and mass balances are applied to a segment of height Δy as shown in Fig. 2. The model of the condensation tower is given by the following set of equations: Energy balance for water: ∂ ðm˙ i Þdy = UAc ðTam −Twm Þ ∂y w w
ð14Þ
Energy balance for dry air: ′ ∂T ′ ∂X m˙ a Cpa a + m˙ a hfg a dy = hac Ac ðTam −Ti Þ + Dc Ac hfg ðXam −Xi Þ ð15Þ ∂y ∂y
m˙ a
∂Xa dy = Dc Ac ðXam −Xi Þ ∂y
ð16Þ
where, Ti is the temperature at the air–water interface in the condensation tower and Xi is the saturation humidity at the interface which can be obtained from Eq. (11) using the saturation pressure at the air–water interface Pi. This saturation pressure can, in turn, be calculated using Eq. (12) at the air–water interface temperature Ti. The overall heat transfer coefficient from the air to the cooling water inside the condenser is approximated by: U=
1 hac
+
1 hc
1 +
1 he
+
ð17Þ
δ Kp
The heat transfer coefficients hac, he, hc and the mass transfer coefficient Dc are given by [9]: hac = " hc =
0:779K 0:25 Gr dh1
#0:25 3 ρ2w ghfg Kw 4μw dh2 ðTa −Tw Þ
where, Tw is the absolute temperature of water in Kelvin. The unit productivity, the rate of water condensed inside the condensation tower, could be calculated from the following equation: m˙ distillate = m˙ a ðXaci −Xace Þ
he = and
Kw 0:8 0:33 ð0:023Re Pr Þ dh2
Dc =
hac Cpa
Fig. 2. An element the dehumidifier.
ð19Þ
The system of Eqs. (4), (7), (9), (10), (14)–(16) are solved numerically to obtain the required values for air and water temperature, air humidity, and flow rates of air and water. Each tower is divided into (n) elements. The solution is obtained at the inlet and exit of each element for the two towers. Using an explicit forward finite difference scheme, the differential equations are expressed as a set of nonlinear algebraic equations. A computer program has been developed in Quick BASIC to enable the numerical solution of the model. The programming takes into account that the air flow is in a counter direction of water flow. The temperature of air and water as well as the air humidity ratio is obtained at each segment in order to calculate the temperature distribution along the flow direction. Iterative procedure is followed for each segment and the solution is assumed to converge when the difference between the values of all parameters calculated at the exit of each tower in two successive iterations is zero or equal to a small tolerance expressed as: J + 1
J
j yexit −yexit j J yexit
Mass balance for humid air:
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≤Tolerance
ð20Þ
The physical properties of water and air are obtained at each segment as a function of the mean temperature in any iteration. Two subprograms have been developed to perform the search technique to obtain the values of physical properties corresponding to temperature from the data tabulated in [10]. It has been considered while programming that the air conditions at humidifier exit are the same as those at condenser inlet. 3. Theoretical results Fig. 3 shows the variation of water temperature at condenser exit (Twce) with water temperature at humidifier inlet (Twhi) for different values of water flow rate (ṁw). It is seen from the figure that, Twce increases almost linearly with Twhi at a constant water flow rate. This is explained by the increase in the amount of heat gained by the air in the humidifier as Twhi increases. Consequently, air loses more heat to
ð18Þ
Fig. 3. Predictions of water temperature at condenser exit.
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the water in the condenser resulting in a higher water exit temperature. For the same value of Twhi, as ṁw increases, Twce decreases. This can be easily explained because higher flow rates lead to smaller temperature difference. Fig. 4 shows the variation of air temperature at both condenser inlet (Taci) and exit (Tace) with water temperature at humidifier inlet (Twhi) for different values for (ṁw). The figure shows that Taci increases linearly with Twhi. As Twhi increases, the air is heated as it leaves the humidifier leading to higher Tahe which is equal Taci. For the same value of Twhi, as ṁw increases Taci increases too. As ṁw increases, the heat supplied by water in humidifier is increased. As a result, the air becomes hotter at is leaves the humidifier and enters the condenser. The figure also shows that the higher Twhi is, the higher Tace is for a constant water flow rate. It has to be noted that Tace is always smaller than Taci for all cases as dictated by the energy balance. However, as hotter air enters the condenser, it gets cooled but leaves at a temperature higher than the value corresponding to lower Twhi since ṁw in the condenser is fixed. For the same value of Twhi, as ṁw increases Tace increases. Fig. 5 shows the variation of humidity ratio at condenser inlet (Xaci) and exit (Xace) with water temperature at humidifier inlet (Twhi) for different values for ṁw. The figure shows that Xaci increases with Twhi. As Twhi increases, water temperature becomes higher all over the humidifier which results in higher air temperature. When the air temperature increases, its ability to hold water vapor increases. Thus, the humidity ratio of air leaving the humidifier increases as well. It is noted also that the rate of increase in Xaci is higher at larger values for Twhi. For the same value of Twhi, the humidity ratio increases slightly as ṁw increases. This is because air reaches saturation condition as it leaves the humidifier, even at a small water flow rates. It can be also noticed that Xace increases with increase in Twhi since the temperature of air at condenser exit increases. At higher values of Twhi the rate of increase becomes steeper for a constant water flow rate. For the same value of Twhi the humidity ratio at condenser exit increases as ṁw increases. Fig. 6 shows the variation of the unit productivity with water temperature at humidifier inlet for different values for water flow rate (ṁw). The figure shows that, the unit productivity increases as Twhi increases. When Twhi increases, air accommodates more moisture in the humidifier. This moisture is condensed in the condensation tower giving higher unit productivity. For the same values of Twhi as ṁw increases the unit productivity increases.
Fig. 4. Predictions of air temperature at condenser inlet and exit.
Fig. 5. Predictions of air humidity ratio at condenser inlet and exit.
4. Experimental test set-up The experimental test set-up consists mainly of a desalination unit, the saline water loop and the heating source loop. The desalination unit is a cabinet containing both the humidification tower (evaporator) and dehumidification tower (condenser). The saline water loop consists of the storage tank, pump, filter, flow meter and the auxiliary valves and pipe connections. The heating source loop consists of a storage tank with a built in heat exchanger, pump, flow meter and the necessary hydraulic connections. A schematic diagram of the test setup is shown in Fig. 7. The desalination unit (dimensions 1.2 × 0.5 × 2 m) constitutes both the humidification and condensation towers. The unit is divided into two unequal compartments separated by a partition. The unit is insulated with 5 cm thick glass wool mates from all sides. The unit is kept at 20 cm above ground level to enable the collection of brine and distillate and also for easier handling of the unit. Aluminum angles form the frame of the desalination unit and the outer sides are made of galvanized steel sheets. Silicon sealing is applied at all sides to ensure that the unit is leak proof. In order to provide a closed loop for air circulation, a partition wall is made of a
Fig. 6. Predictions of unit productivity.
E.H. Amer et al. / Desalination 249 (2009) 949–959
Fig. 7. Schematic diagram of the desalination system.
double layered wall with a thermal insulation between the two layers. The partition leaves gabs of 20 cm height at the top and at the bottom of the unit. An electric fan is supported at the top of the condensation tower to provide the forced circulation of air. The bottom of the desalination unit is shaped as a tray which is inclined at the horizontal plane by about 15º towards the outer side walls to collect the brine and distillate. The condensation tower occupies one compartment of the desalination unit. The dimensions of the condensation tower are 200 cm height, 40 cm length, and 50 cm width. A copper tube formed as a coil is used as a condenser of 15 m length and of 1.27 cm outer diameter. Fins are used to increase the surface area of the condenser. The total surface area of the condenser coil and its fins is approximately 6 m2. The cooling water flows inside the coil and the air flows over the finned coil inside the condenser cabinet in a counter direction. The other compartment of the desalination unit constitutes the humidifying tower. Its dimensions are 200 cm height, 80 cm length, and 50 cm width. A packing material is fixed inside the humidifier to have a surface area of approximately 6 m2. The packing is supported such that it does not block the air flow and remains continuously wet. The water is sprayed on the packing material using a hydraulic grid. A movable door is provided to facilitate the changing of packing material easily. Three different materials have been used in packing the humidifier. The differences in the packing materials are in both the actual wetted area and coefficient of mass transfer. The selected materials are either widely available in the local environment at almost no cost (gunny bag cloth), available in the local market at moderate price (plywood slates) or those used in the industry for cooling towers (PVC sheets). The reduced cost of backing material would certainly help in reducing the cost per liter of distilled water. The surface area of all materials is kept the same. A 20 W electric fan is used to circulate the air through the unit. The fan is supported on the top of the condenser duct. To circulate the water, a centrifugal pump of 0.5 hp is used. A filter is used in the suction side to separate any impurities coming from the elevated tank. A flow meter is connected on the delivery side of the pump to measure the water flow rate. A group of valves are used to separate each part of the saline water loop from the system when necessary. The elevated storage tank is made from galvanized steel sheets of 100 l in capacity. The tank is divided into two parts by a partition
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made from galvanized steel sheets to provide a constant suction head at the pump inlet. This helps to maintain a constant water flow rate. The tank is supported on a stand made from iron angles of 2 m height above the ground. The heating source loop consists of a storage tank with built in heat exchanger, and an electric heater. The tank is 100 l in capacity. The heat exchanger is made of copper tube of 15 m length and 1.27 cm diameter formed as a coil. The tank is supported on a stand made from iron angles. The storage tank as well as the connecting pipe lines is thermally insulated using glass wool to minimize the heat losses to the surroundings. A second 0.5 hp pump is used to circulate the heated water through the unit. In order to carry out experiments indoors, the storage tank is provided with a 6 kW electric heater powered through a temperature controller. During the course of experiments, several parameters have to be measured in order to evaluate the system performance. The quantities needed to be measured are: flow rate of air and water streams, temperature of air at the inlet and exit of each tower, temperature of water at the inlet and exit of each tower, temperature of water at inlet and exit of the storage tank, relative humidity of air at inlet and exit of each tower, and the productivity of the unit. A float type rotameter is used to measure the flow rate of water before it enters the condenser. The rotameter works in the range from (0.02 to 5) LPM. The accuracy of this apparatus is ±0.01 LPM. The flow meter has been calibrated experimentally by comparing its reading with the amount of water collected during a prescribed time interval. The air velocity has been measured using a hot wire anemometer (Model 490, KURZ INSTRUMENTS INC) working in the range from 0 to 2000 FPM (0–10 m/s) with an accuracy of ±0.01 FPM (±5 × 10– 5 m/s). Knowing the air velocity and the cross sectional area of the unit, the volume flow rate of air can easily be calculated. Manual control has been adopted to regulate the water flow rate using the valve placed before the flow meter. Temperatures are measured using Copper-Constantan (Type-T) thermocouples. From the calibration process, the accuracy of thermocouples has been found to be ±0.1 °C. Thermocouples are connected through (type-T) selector switch connected to 4½ digits readout device that provides cold junction compensation. The resolution of the readout device equals 0.1 °C and its accuracy is ±0.05 °C. The temperature of water at humidifier inlet has been controlled using a PID temperature controller (Digisence ColeParmer). The relative humidity of air streams is measured using a digital hygrometer (BK precision humidity meter model 720). The meter has 4 digits display and a resolution of 0.1% RH. The accuracy of the meter is ±2.5%. Using the values measured for relative humidity and dry bulb temperature, the humidity ratio is obtained from the psychometric chart. The unit productivity has been measured by collecting the amount of distillated water in a graduated cylinder during a prescribed time interval. This interval was measured using a digital stop watch. Based on the accuracy of each measuring instrument, an estimate of the uncertainty in measurements has been carried out following the procedure explained by Kline and McClintock [11]. It has been found out that the maximum uncertainty in the measurements is about 1.15%. 5. Experimental results 5.1. Experimental correlations for heat and mass transfer coefficients In order to evaluate the unit performance accurately, precise expressions have to be used for calculating the heat and mass transfer coefficients in the condenser and the humidifier. The unit has been operated under steady state conditions. Once equilibrium is reached; measurements for flow rates, water and air temperatures at inlet and
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exit of each tower, relative humidity of air at the inlet and exit of each tower, and the unit productivity are made. The following procedure has been followed to obtain the heat transfer coefficient. The energy lost by air and gained by water is obtained using the measured flow rate and temperature change for each fluid. The logarithmic mean temperature difference has been calculated and the overall heat transfer coefficient is obtained. Several experiments have been conducted to obtain a large data set for the regression analysis. A computer program has been constructed to perform the multiple nonlinear curve fitting using the method of least squares. The measured data have been given as an input to the program and a correlation has been obtained in the form: UAc =
β γ αTwhim˙ w
ð21Þ
where β, γ are dimensionless constants and α has dimensions such that both sides of the equations are of the same units. This form has been used in guidance of the previous publications. Calculations have been performed for all types of packing materials in the natural and forced modes of air flow. Table 1 shows the constants α, β, and γ used in Eq. (21) for the different materials. The table shows also the average deviation of measured values of (UA) from the fitted correlation. To obtain a correlation for the mass transfer coefficient Dh, the following procedures has been followed: • The rate of air flow has been calculated from the mass balance of the condenser using the measured rate of distilled water. This has been done to avoid the error in measuring the very low flow rate of air. DA • The mass transfer coefficient has been calculated from m˙ wp = Twhi Cpw dT where, if is the enthalpy of air at mean water temperature, ∫ if −ia T whe
and ia is the enthalpy of air at inlet or exit of humidifier. • The logarithmic mean difference has been used assuming a linear enthalpy temperature relation at saturation line and (if − ia)Lm is 2 where: Δi1 = (ia)Twm − (ia)hi and found from ðif −ia Þlm = Δi1 −Δi Δi ln Δi1 2 Δi1 = (ia)Twm − (ia)he. D Ap • The curve fitting program has been used to correlate h with the m ˙ w m˙ w in the form: ratio of m˙ a D h Ap m˙ w β =α ð22Þ m ˙ w m ˙ a
Table 2 Coefficients obtained for Eq. (22). Packing materials
Air circulation
α
β
Average deviation %
Gunny bag cloth
Forced Natural Forced Natural Forced Natural
7.4897 × 10− 2 0.6026 9.5789 × 10− 2 0.4300 7.4471 × 10− 2 0.3861
− 1.255 − 0.8307 − 1.147 − 0.6264 − 1.6461 − 0.6112
− 0.03 − 0.01 − 0.02 − 0.02 − 0.02 − 0.08
PVC Wooden slates
increases almost linearly with Twhi at a constant water flow rate. This tendency can be explained by the increase in the amount of heat gained by the air in the humidifier as Twhi increases. Consequently, air losses more heat to the water in the condenser resulting in a higher exit temperature. For the same value of Twhi, as ṁw increases Twce decreases. This can be easily explained because higher flow rates lead to smaller temperature difference. Fig. 9 shows the variation of air temperature at condenser inlet (Taci) with Twhi for different values for ṁw. The figure is plotted for natural air circulation and using Gunny bag as the humidifier packing. The figure shows that Taci increases linearly with Twhi. The air conditions at humidifier exit are the same as those at condenser inlet. As Twhi increases the air is heated as it leaves the humidifier leading to higher air temperature at humidifier exit (Tahe) which is equal to Taci. The figure also shows that Taci increases with the increase in ṁw for a given Twhi. As ṁw increases, the heat supplied by water in humidifier increases. As a result, the air becomes hotter as it leaves the humidifier and enters the condenser. Fig. 10 represents a sample of the results obtained for the variation of air temperature at condenser exit (Tace) with Twhi for forced air circulation using a Gunny bag cloth as packing material. It is seen that, the higher Twhi the higher Tace for a constant flow rate. It has to be noted that Tace is always smaller than Taci for all cases as dictated by the energy balance. However, as hotter air enters the condenser, it gets cooled but leaves the condenser at a temperature higher than the value corresponding to lower Twhi since ṁw in the condenser is fixed. For a constant Twhi, Tace increases with ṁw in both natural and forced circulation. Fig. 11 shows the variation of humidity ratio at both condenser inlet (Xaci) and exit (Xace) with Twhi for different values for ṁw. The figure is plotted for forced air circulation and using a Gunny bag cloth
The results obtained are summarized in Table 2 for the different packing materials. The correlations obtained in this section have been used to compare the theoretical predictions with actual measurements of the unit. 5.2. Measurements of unit performance Fig. 8 shows the variation of water temperature measured at condenser exit (Twce) with Twhi for different values of ṁw. This figure is plotted for naturally circulating air and a Gunny bag cloth as the packing material in the humidifier. It is seen from the figure that Twce Table 1 Coefficient obtained for Eq. (21). Packing materials
Air circulation
α
β
γ
Average deviation %
Gunny bag cloth
Forced Natural Forced Natural Forced Natural
3.4132 0.5598 4.8919 0.7267 19.2854 0.5915
0.8306 1.1719 0.7362 1.0746 0.4822 1.1691
0.1714 8.429 × 10− 2 0.2169 0.2199 − 4.887 × 10− 2 0.1266
−0.04 − 0.04 − 0.03 − 0.04 − 0.03 − 0.05
PVC Wooden slates
Fig. 8. Measurements of water temperature at condenser exit (Natural air circulationGunny bag cloth).
E.H. Amer et al. / Desalination 249 (2009) 949–959
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Fig. 9. Measurements of air temperature at condenser inlet (Natural air circulationGunny bag cloth).
as the humidifier packing. The figure shows that Xaci increases with increasing Twhi. As Twhi increases the air is heated as it leaves the humidifier leading to higher Tahe. As the air temperature increases, its ability to hold water vapor increases. Thus, the humidity ratio of air leaving the humidifier increases as well. It is noted also that the rate of increase in Xaci is higher at larger values for Twhi. It can be noticed also that Xace increases with the increases in Twhi since the temperature of air at condenser exit increases. At higher values of Twhi the rate of increase becomes steeper for a constant water flow rate. The humidity ratio increases slowly as ṁw increases, then the rate of increase becomes more rapid at higher values of ṁw. This is due to the larger amounts of vapor carried away by the air stream. Higher humidity ratios result in larger unit productivity. For constant water flow rate, the productivity of the unit ṁdistillate increases as Twhi is increased as shown in Fig. 12. This is true for both natural and forced circulation of air. When Twhi increases, air
Fig. 10. Measurements of air temperature at condenser exit (Forced air circulationGunny bag cloth).
Fig. 11. Measured air humidity ratio at condenser inlet and exit (Forced air circulationGunny bag cloth).
accommodates more moisture in the humidifier. This moisture is condensed in the condensation tower giving higher unit productivity. Since the mass transfer coefficient in the humidifier is much higher in forced circulation than that in natural one, the unit gives more distillate in forced circulation operation. 5.3. Effect of using different packing materials Three different packing materials have been used in the investigation. The total surface area of the packing has been kept constant. However, the wetted area differs from one type to another depending on the nature of each material. The changes of air and water properties at different locations on the unit are presented in the following figures. Fig. 13 shows the variation of Twce with Twhi for the three packing materials under natural and forced modes of operation. A general conclusion remains true that forced circulation leads to higher value
Fig. 12. Measured unit productivity (Forced air circulation-Gunny bag cloth).
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Fig. 13. Variation of Twce with Twhi for different packing materials at water flow rate of 2.772 kg/min.
Fig. 15. Variation of Tace with Twhi for different packing materials at water flow rate of 2.772 kg/min.
for Twce which increases with Twhi. This is explained by the increase in ṁa which leads to a higher heat gain in the humidifier. Air loses this heat to the condenser water resulting in raising its temperature. The values of Twce in descending order of magnitude are for wooden slates, gunny bag cloth and PVC. Similar results are obtained for the air temperature at inlet and exit from the condenser as shown in Figs. 14 and 15. The air temperature at inlet and exit of condenser increases with the increases of Twhi. Gunny bag gives higher values of Tace compared to other packing materials and under both natural and forced flow condition. The effect of packing materials is more pronounced on air exit temperature than at inlet. Forced circulation leads to high temperatures at the condenser exit and lower values for Taci. This is attributed to the increase in ṁa for the same heat supplied by the humidifier which results in lower air temperature at humidifier exit (i.e., condenser inlet). It is also noticed from all measurements (though not presented) that the difference between Taci for natural and forced circulation increases at higher values of Twhi. The variation of humidity ratio with Twhi, when different packing materials are used, is shown in Figs. 16 and 17 at inlet and exit of the
condenser respectively. The humidity ratio is proportional to Twhi for the same flow rate. In case of forced circulation, the increase in air velocity (i.e., air flow rate) leads to better heat and mass transfer coefficients. Consequently, the productivity increases. Gunny bag cloth gives higher values for X than the other materials. But, the difference between X at inlet and exit of the humidifier is much larger in case of wooden slates. This in turn leads to a higher unit productivity using wooden slates as shown in Fig. 18. In order to show the changes of productivity, the values recorded during the extensive experiments are listed in Table 3 for the various operating conditions. The rate of distilled water in natural circulation is listed and the percentage increase in productivity when forced circulation is used. The table shows that, a maximum productivity of about 5.4 l/h is obtained using wooden slates and inverses to about 5.8 l/h using forced circulation for air. It is noted from the table that forced circulation of the air results in an enhancement in the unit productivity. However, the magnitude of productivity enhancement
Fig. 14. Variation of Taci with Twhi for different packing materials at water flow rate of 2.772 kg/min.
Fig. 16. Variation of Xaci with Twhi for different packing materials at water flow rate of 2.772 kg/min.
E.H. Amer et al. / Desalination 249 (2009) 949–959
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Table 3 Unit productivity (l/h) and percentage change due to forced circulation. Water temperature at humidifier inlet (°C)
Water mass flow rate ( kg/min)
PVC
85
2.772 2.264 1.308 2.772 2.264 1.308 2.772 2.264 1.308 2.772 2.264 1.308 2.772 2.264 1.308 2.772 2.264 1.308 2.772 2.264 1.308 2.772 2.264 1.308
3.96 3.63 2.25 3.5 3.15 1.93 3.05 2.7 1.67 2.66 2.34 1.42 2.26 1.98 1.2 1.94 1.67 0.99 1.63 1.38 0.81 1.38 1.17 0.62
80
75
70
65
60 Fig. 17. Variation of Xace with Twhi for different packing materials at water flow rate of 2.772 kg/min.
depends on the water flow rate and its temperature. The percentage enhancement ranges between 5% and 55%. The percentage enhancement is small at higher temperatures and it increases at lower water temperature. Better improvement in productivity is noticed for gunny bag cloth. The variation of the percentage enhancement in productivity differs from one packing material to another. The enhancement in productivity as a result of using forced circulation is much higher in case of gunny bag cloth. The productivity increases by 5%, 15%, and 33% as the flow rate is decreased from 2.8, 2.3, to 1.8 kg/min respectively. Further decrease in flow rate leads to a reduction in the magnitude of productivity enhancement. 5.4. Comparisons of results In order to verify the soundness of the theoretical model derivation and its numerical simulation, the theoretical predictions have been compared with the experimental measurements. Fig. 19
Fig. 18. Variation of unit productivity with Twhi for different packing materials at water flow rate of 2.772 kg/min.
55
50
(17.42%) (8.82%) (11.11%) (18.86%) (9.52%) (15%) (20%) (11.85%) (16.77%) (18.42%) (11.11%) (19.72%) (19.47%) (11.62%) (22.5%) (19.07%) (13.17%) (27.27%) (22.7%) (15.94%) (29.63%) (22.46%) (13.68%) (40.3%).
Gunny bag cloth
Wooden slates
5.1 (4.51%) 4 (15%) 2.62 (20.23%) 4.65 (4.95%) 3.45 (17.97%) 2.27 (20.26%) 4.13 (4.12%) 2.96 (18.24%) 1.95 (25.64%) 3.62 (6.08%) 2.49 (20.48%) 1.72 (27.91%) 3.08 (9.74%) 2.13 (23.94%) 1.48 (35.14%) 2.65 (11.7%) 1.8 (27.2%) 1.24 (41.13%) 2.25 (12.89%) 1.6 (23.75%) 1.08 (38.89%) 1.84 (15.2%) 1.38 (22.46%) 0.9 (44.44%)
5.4 (7.96%) 4.85 (9.28%) 3.2 (21.25%) 4.95 (7.88%) 4.4 (9.77%) 2.9 (20.69%) 4.55 (7.69%) 3.9 (12.82%) 2.65 (20.75%) 4.1 (8.54%) 3.45 (14.78%) 2.42 (19.83%) 3.65 (10.14%) 3.05 (15.41%) 2.21 (18.55%) 3.2 (14.69%) 2.7 (14.81%) 2 (17.5%) 2.8 (17.86%) 2.35 (17.02%) 1.8 (16.67%) 2.4 (25%) 2.05 (17.07%) 1.61 (14.9%)
shows a comparison between measurements and predictions for Taci. The figure shows that the theoretical results are in good agreement with measurements. The average and maximum relative deviation of predictions from measurements are (− 0.9%) and (−3%) respectively. Measurements of Xace are also compared with theoretical predictions and the comparison is presented in Fig. 20. The figure shows that the model is capable of predicting the humidity ratio with an average error of 3.8%. The maximum deviation of prediction from measurements is found to be 7%. Fig. 21 shows a comparison between measurements and predictions for Twce. The figure shows that the theoretical results are in good agreement with measurements. Here the average and maximum deviation of predictions from measurements are (− 1%) and (−3%) respectively. The operating conditions and the system dimensions of the unit reported in [4] have been given as an input to the simulation program. The output results have been compared with the corresponding data published in [4]. The comparison is depicted in Fig. 22 for the water temperature at condenser exit (Twce), and in Fig. 23 for the air
Fig. 19. Comparison between measurements and predictions for Taci (Gunny bag cloth, water flow rate of 2.264 kg/min).
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E.H. Amer et al. / Desalination 249 (2009) 949–959
Fig. 22. Comparison between theoretical results and reported results for Twce. Fig. 20. Comparison between measurements and predictions for Xace. (Gunny bag cloth, water flow rate of 2.772 kg/min).
nent. The deviation of actual measurements from the correlated data has not exceeded 5%. It can be concluded from the investigation that: temperature at condenser inlet (Taci). In these graphs, the solid line represents the predictions obtained from the current study, while the dashed line represents the theoretical results of Ref. [4]. The symbols show the experimental measurements reported in Ref. [4]. It can be noted that the deviation of theoretical values from measurements of the original data in Ref. [4] is not small. It is about 4 °C in water temperature and 5 °C in air temperature. Also, the current theoretical results are closer to measurements. The deviations are about 3 °C and 4 °C in water and air temperatures respectively.
6. Conclusions The humidification–dehumidification–desalination process has been extensively investigated both theoretically and experimentally. Three different packing materials have been used in experiments. The heat and mass transfer coefficients in the humidifier and condenser have been obtained experimentally. Empirical correlations have been obtained to express these coefficients in terms of the flow rates of working fluids and the temperature difference across each compo-
Fig. 21. Comparison between measurements and predictions for Twce (Gunny bag cloth, water flow rate of 1.308 kg/min).
• Increasing the temperature of water at humidifier inlet (Twhi) results in an increase in the values of all unit parameters (Twce, Taci, Tace, Xaci, Xace, ṁdistillate). • While the water temperature at condenser exit decreases with the increase in water flow rate, all the remaining parameters increase. • As air velocity and consequently air flow rate increases when forced circulation is used, the air temperature and its humidity ratio at condenser inlet are decreased, while all other unit parameters are increased. • Higher productivity of the unit is obtained using wooden slates and forced air circulation. The maximum productivity obtained is about 5.8 l/h at water mass flow rate of 2.8 kg/min and water temperature at humidifier inlet 85 °C. • The enhancement in productivity as a result of using forced circulation is much higher in the case of a gunny bag cloth. The productivity increases by 5%, 15%, and 50% as the flow rate is decreased from 2.8, 2.3, to 1.8 kg/min respectively. Further decrease in flow rate leads to a reduction in the magnitude of productivity enhancement. • At smaller values of water temperature at humidifier inlet (Twhi), the improvement in productivity is more pronounced when forced air circulation is used. • The average relative deviation of predictions from measurements are (−0.9%) in air temperature at condenser inlet, (3.8%) in humidity
Fig. 23. Comparison between theoretical results and reported results for Taci.
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ratio at condenser exit (Xace), and (−1%) in water temperature at condenser exit (Twce). References [1] Kh. Naser, Mohammed Mehdi Nawayseh, SaidAl-Hallaj Farid, AbdulRahman AlTimimi, Solar desalination based on humidification process-I. Evaluating the heat and mass transfer coefficients, Energy Conversion & Management 40 (1999) 1423–1439. [2] Kh. Naser, Mohammed Mehdi Nawayseh, AbdulAziz Farid, AhmadSabirinc Omar, Solar desalination based on humidification process-II. Computer simulation, Energy Conversion & Management 40 (1999) 1441–1461. [3] Naser K.h Nawayseh, Farid Mohammed Mehdi, Omar Abdul Aziz, Said Moh. AlHallaj, Tamimi Abdul Rahman, A simulation study to improve the performance of a solar humidification–dehumidification desalination unit constructed in Jordan, Desalination 109 (1997) 277–284. [4] H. Ben Bacha, M. Bouzguenda, M.S. Abid, A.Y. Maalej, Modeling and simulation of a water desalination station with solar multiple condensation evaporation cycle technique, Renewable Energy 18 (1999) 349–365. [5] Y.J. Dai, R.Z. Wang, H.F. Zhang, Parametric analysis to improve the performance of a solar desalination unit with humidification and dehumidification, Desalination 142 (2002) 107–118. [6] M.M. Farida, J.R. Sandeep Parekhb, SaidAl-Hallajb Selmanb, Solar desalination with a humidification–dehumidification cycle: mathematical modeling of the unit, Desalination 15 (1) (2002) 153–164. [7] A.S. Nafey, H.E.S. Fath, S.O. El-Helaby, A.M. Soliman, Solar desalination using humidification–dehumidification processes. Part I. A numerical investigation, Energy Conversion and Management 45 (2004) 1243–1261. [8] Hassan E.S. Fath, Ahmad Ghazy Solar desalination using humidification– dehumidification technology, Desalination 142 (2002) 119–133. [9] AliR. EL-Ghalban, Simulation of a closed humidification–dehumidification desalination cycle, Engineering research journal, Faculty of Engineering-Shebin Elkom, Egypt 26 (1) (January 2003) 99–110. [10] J.P. Holman, Heat Transfer, Ninth ed., Mc Graw-Hill, Singapore, 2002. [11] S.J. Kline, F.A. McClintock, Describing uncertainties in single sample experiments, Mechanical Engineering 75 (1953) 3–8.
Glossary A: Humidifier surface area per unit volume (m2) Ac: Surface area of condenser (m2) Ah: Surface area of humidifier (m2) CPa: Specific heat at constant pressure for air (J/kg K)
CPv: Specific heat at constant pressure for water vapor (J/kg K) CPw: Specific heat of water (J/kg K) Dc: Mass transfer coefficient in condenser (kg/m2 s) Dh: Mass transfer coefficient in humidifier (kg/m2 s) dh1: Hydraulic diameter of air flow (m) dh2: Hydraulic diameter of water flow (m) g: Gravitational acceleration (m/s2) Gr: Grashof number ha: Air heat transfer coefficient in the evaporation tower (W/m2 K) hac: Air heat transfer coefficient in the condensation tower (W/m2 K) hc: Heat transfer coefficient for the condensate film (W/m2 K) he: Heat transfer coefficient at the water-condenser inside wall (W/m2 K) hfg: Latent heat of evaporation of water at zero °C (J/kg) hw: Water heat transfer coefficient at the air–water interface (W/m2 K) ia: Enthalpy of air at inlet or exit temperature (J/kg) iw: Enthalpy of water at inlet or exit temperature (J/kg) k: Thermal conductivity of air (W/m K) kw: Thermal conductivity of water (W/m K) Pr: Prandtl number Ps: Saturation pressure (Pa) Q a–i: Rate of heat transfer from air to interface in condensation tower (W) Q a–w: Rate of heat transfer from air to water in condensation tower (W) Q evap: Rate of heat transfer to evaporate water in humidifier (W) Q w–a: Rate of heat transfer from water to air in condensation tower (W) Re: Reynolds number Tam: Mean temperature of air (K) Tace: Air temperature at condenser exit (oC) Taci: Air temperature at condenser inlet (oC) Tahe: Air temperature at humidifier exit (oC) Ti: Temperature at the air–water interface in the condensation tower (K) Twm: Mean temperature of water (K) Twce: Water temperature at condenser exit (oC) Twhe: Water temperature at humidifier exit (oC) Twhi: Water temperature at humidifier inlet (oC) U: Overall heat transfer coefficient in condenser (W/m2 K) Xa: Mean humidity ratio of air (kgv/kga) Xace: Humidity ratio at condenser exit (kgv/kga) Xaci: Humidity ratio at condenser inlet (kgv/kga) Xi: Saturation humidity at the interface in the condenser tower (kgv/kga) XS: Saturation humidity in the humidifier tower (kgv/kga) ṁ a: Mass flow rate of air (kg/s) ṁ w: Mass flow rate of water (kg/s) ṁ distillate: Unit productivity (kg/s) δ: Tube wall thickness (m) ρw: Density of water (kg/m3) µw: Dynamic viscosity of water (Pa s) Δy: Height of element (m)
959
