Humidification dehumidification desalination system using parabolic trough solar air collector

Descripción

Applied Thermal Engineering 75 (2015) 809e816

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Applied Thermal Engineering journal homepage: www.elsevier.com/locate/apthermeng

Research paper

Humidiﬁcation dehumidiﬁcation desalination system using parabolic trough solar air collector Fahad A. Al-Sulaiman a, b, *, M. Ifras Zubair a, Maimoon Atif a, Palanichamy Gandhidasan a, Salem A. Al-Dini a, Mohamed A. Antar a a b

Department of Mechanical Engineering, King Fahd University of Petroleum & Minerals (KFUPM), Dhahran 31261, Saudi Arabia Center of Research Excellence in Renewable Energy, King Fahd University of Petroleum & Minerals (KFUPM), Dhahran 31261, Saudi Arabia

h i g h l i g h t s Thermodynamic analysis of an HDH system driven by a parabolic trough solar collector was conducted. The ﬁrst conﬁguration reveals a GOR of 1.5 while the second conﬁguration reveals a GOR of 4.7. Effective heating of the HDH system was obtained through parabolic trough solar collector.

a r t i c l e i n f o

a b s t r a c t

Article history: Received 5 August 2014 Accepted 19 October 2014 Available online 28 October 2014

This paper deals with a detailed thermodynamic analysis to assess the performance of an HDH system with an integrated parabolic trough solar collector (PTSC). The HDH system considered is an open air, open water, air heated system that uses a PTSC as an air heater. Two different conﬁgurations were considered of the HDH system. In the ﬁrst conﬁguration, the solar air heater was placed before the humidiﬁer whereas in the second conﬁguration the solar air heater was placed between the humidiﬁer and the dehumidiﬁer. The current study revealed that PTSCs are well suited for air heated HDH systems for high radiation location, such as Dhahran, Saudi Arabia. The comparison between the two HDH conﬁgurations demonstrates that the gained output ratio (GOR) of the ﬁrst conﬁguration is, on average, about 1.5 whereas for the second conﬁguration the GOR increases up to an average value of 4.7. The study demonstrates that the HDH conﬁguration with the air heater placed between the humidiﬁer and the dehumidiﬁer has a better performance and a higher productivity. © 2014 Elsevier Ltd. All rights reserved.

Keywords: Humidiﬁcation dehumidiﬁcation (HDH) Thermal desalination technology Solar air-heated system Parabolic trough solar collector (PTSC) Gained output ratio (GOR)

1. Introduction Although the availability of water is abundant in the world, pure consumable water is a scarce resource. Most regions with scarcity of water are dry regions, where abundant solar energy is available. Therefore, desalination assisted by a solar energy technology is a viable option within such regions. Using a renewable energy resource for desalination is suitable for rural and remote areas and it is a preferred technology over fossil fuels. This allows the improvement of the living conditions with a very minimum impact on the environment. Solar energy technologies have had extraordinary development within the past two decades, where many * Corresponding author. Center of Research Excellence in Renewable Energy, King Fahd University of Petroleum & Minerals (KFUPM), Dhahran 31261, Saudi Arabia. Tel.: þ966 13 8604628. E-mail address: [email protected] (F.A. Al-Sulaiman). http://dx.doi.org/10.1016/j.applthermaleng.2014.10.072 1359-4311/© 2014 Elsevier Ltd. All rights reserved.

improvements such as concentrating technologies, among others, have been introduced [1]. Different methods are available for desalinating saline water through a thermal process. The basics of the thermal process, requires the saline water to be vaporized where the fresh water is collected. Examples of such technologies that use such process are multiple effects, multi-stage ﬂash and more recently humidiﬁcation dehumidiﬁcation process. One of the main problems faced with these technologies is relatively high demand of energy due to the vaporization phase change [2]. A number of studies were conducted to analyse the performance of water desalination using solar energy, e.g. Refs. [3e6]. Performance assessment of a small-scale multi-effect distiller for an optimized solar thermal desalination was carried out by Joo and Kwak [3]. They found that the performance ratio of their system was around 2.0. The performance assessment of v-trough solar concentrator for desalination applications was investigated by

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Riffat and Mayere [4]. They showed that a thermal efﬁciency of 38% was obtained when the heat transfer ﬂuid temperature reaches 100 C at the outlet of the concentrator. In addition, they concluded that the v-trough solar concentrator is a promising solution for small and medium water desalination applications. In a different study, performance evaluation of combining desalination systems with concentrated solar power plants was conducted by Palenzuela et al. [5]. They concluded that a low temperature multiple-effect distillation is more efﬁcient than a low temperature multipleeffect distillation integrated with a thermal vapour compression. In another study, a dynamic model of a solar system assisted a multiple-effect distillation system was investigated by Calle et al. [6]. The model can be used to optimize the operating control of such systems. Humidiﬁcation dehumidiﬁcation (HDH) desalination process is regarded as a favourable technique for small water capacity production plants. The main attraction towards this process is its capability to operate at low temperatures, the use of low-level technical features, and the possible integration of sustainable energy sources. In addition, one of the greatest advantages with this process of desalination is that it uses separate components for each thermal process, which in turn allows each individual element to be independently designed. Hence, it allows much greater ﬂexibility with the thermodynamic cycle for vaporizing water into air and consequently condensing the vapour [7]. An HDH desalination system when compared with solar stills has signiﬁcantly higher gained-output-ratio (GOR), which results in reducing the total area of the solar collector required for a given water demand. The GOR is deﬁned as the ratio of the collected fresh water multiplied by the water enthalpy of vaporization to the energy input of the system. HDH systems are more suitable for the developing world in terms of the capital investment and limited technical support. Subsequently, it involves relatively inexpensive and simple mechanisms that can also operate under a wide range of raw water quality without the need for complex maintenance operations. The foremost downside of the HDH systems is the relatively high thermal energy requirement in comparison to other technologies. Therefore, there is a need for further research to improve the performance of the HDH desalination systems. Solar stills generally integrate the functions of solar collection, water heating, evaporation, and condensation into a single system. Such systems result in a considerable thermal inefﬁciency. As a result, solar stills normally have a low GOR and require relatively large areas in order to produce fresh water. Alternatively, the HDH desalination systems overcome some of the shortcoming of the solar stills and consequently they have higher GOR.

The classiﬁcation of HDH systems is generally dependent on whether air or water is heated and if the air and/or the water ﬂow through an open or closed loop. There are various experimentations in using a combination of both air and water heaters, or the use of steam generators and water storage tanks, or the use of a combined humidiﬁer and dehumidiﬁer in place of two separate units. Table 1 shows a comparison of some of the experiments carried out on various HDH systems where the mode of heating is the main difference amongst them. Thermodynamic analysis of an HDH system is generally based on energy and mass balances of each individual component within the system. Narayan et al. [17] and Mistry et al. [18] showed that the top water temperature and mass ﬂow rate ratio of air and water streams play a major role in identifying the maximum GOR in an HDH system. Some of the existing HDH thermal desalination technologies were discussed by Narayan et al. [19]. Some of the HDH systems discussed in their study were multi-stage air heated cycle, mechanical compression driven cycle, HDH with thermodynamic balancing, HDH with common heat transfer wall, and hybrid HDH system with reverse osmosis, among others. Ettouney [20] evaluated the characteristics of four layouts of HDH systems. A common feature among these layouts was the air humidiﬁcation tower that had been used to increase the ambient air humidity to saturation at the desired design temperature. The main difference among the various layouts was the process of dehumidiﬁcation. One layout used a condenser to reduce the temperature of the humidiﬁed air and to condense the fresh water product. The other conﬁgurations considered were membrane air drying, desiccant air drying, and vapour compression. The main drawback stressed upon was the presence of air in bulk along with water vapour. The dehumidiﬁcation process efﬁciency had drastically reduced due to the above mentioned reason. Wang et al. [21] studied a photovoltaic (PV) driven humidiﬁcation dehumidiﬁcation desalination system. The main factors affecting the evaporation and condensation were considered in their study. The rate of evaporation of water and the condensation of the mass ﬂow rate increase along with the increase of evaporative raw water. In addition, the condensation rate was found to increase with lower cooling water temperatures. While the aforementioned parameters were set to their optimal levels, it was also found that the forced convection method had a higher yield of fresh water in comparison to the natural convection, A recent humidiﬁcation dehumidiﬁcation desalination technology known as humidiﬁcation compression was presented by Ghalavand et al. [22]. The main characteristic of this technology is the polytropic compressor where no heater is required.

Table 1 Comparison of different desalination system. Heating mode

Production 2

2

Air heating (evacuated tubular solar water heater)

4 kg/m -day (increased to 10 kg/m -day including the water heating)

Air heating

4 L/m2-day (total 516 L/day)

Air heating

Up to 5 kg/h (GOR < 4) 2

Water heating

8 L/m -day (GOR < 2)

Water and air heating Water heating Water heating

9 L/day 13 L/m2-day (GOR ¼ 3e4.5) 12 L/m2-day (GOR < 4)

Water heating

3 L/m2-day (GOR < 0.5)

Comments

Reference

Single stage, double pass solar collector, pad humidiﬁer and ﬁnned tube dehumidiﬁer and 0.5 m3 water storage tank. No heat recovery. Water may be heated in the storage tank to increase the production signiﬁcantly. Five heating-dehumidiﬁcation stages; Forced air circulation; total collector area e 127 m2 Natural and forced air ﬂow, heat recovery in the condenser 2 m2 solar collector area, humidiﬁer and condenser speciﬁc areas are 14 and 8 m2 Packed bed humidiﬁer, air cooled dehumidiﬁer Thermal storage, natural air draft, 38 m2 collector area Forced circulation of air, multi-pass shell and tube condenser and wooden shaving packing in the humidiﬁer. Solar area 6 m2. No energy recovery.

Yamali and Solmus [8,9]

Houcine et al. [10] Nawayesh et al. [11] El-Hallaj et al. [12] Nafey st al. [13] Muller-Holst et al. [14] Farid et al. [15]

Ben Basha et al. [16]

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An experimental study of a two stage solar HDH system was carried out by Zamen et al. [23]. It was shown that two-stage process is appropriate in order to improve critical parameters such as the daily production output per unit area of the collector, the speciﬁc energy consumption, and the productivity. A study carried out by Iaquaniello et al. [24], on a system consists of a concentrating solar subsystem for power generation and thermal desalination processes, was concluded. They showed that such a system is effective to reduce the total cost of water production for large scale plants with respect to equipment lifetime and continuous operation capabilities. Mohamed and El-Minshawy [25] conducted a thermodynamic study on a solar air heated HDH system integrated with a PTSC. They had considered one conﬁguration in which they assumed the air to be heated before it enters the humidiﬁer. In addition, their proposed design assumed a common wall between the humidiﬁer and the dehumidiﬁer. From the literature review, it is observed that HDH desalination systems with no common wall between the humidiﬁer and the dehumidiﬁer and driven by parabolic trough solar air collectors have received no attention. In this study, we are comparing the performance of two different conﬁgurations of the HDH system incorporating a parabolic trough solar air collector. The ﬁrst one in which the PTSC is located before the humidiﬁer while the second one is in which the solar collector is located between the humidiﬁer and the dehumidiﬁer. In this study, detailed thermodynamic analysis was conducted to assess the performance of the systems examined. In this paper, the performance of the PTSC was considered for Dhahran, Saudi Arabia and the performance of the HDH system was examined as well as reported for the two conﬁgurations considered. 2. System description The two HDH conﬁgurations examined in this study are illustrated in Figs. 1 and 2. Each consists of a PTSC along with an openair open-water HDH system. The parabolic trough consists of an evacuated tube glass cover. The type of the HDH considered is multi-stage bubble generated system, which has a relatively high efﬁciency [26]. The basic design parameters and the operating parameters of the system considered are given in Table 2.

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Fig. 2. The second conﬁguration: open-water open-air humidiﬁcation dehumidiﬁcation desalination system.

In the ﬁrst conﬁguration, Fig. 1, the ambient air moves through the solar air heater and enters the humidiﬁer at high temperature where the air is humidiﬁed through the direct interaction with the water that comes from the dehumidiﬁer. On the other side, the humid air exits from the humidiﬁer and enters the dehumidiﬁer. The brackish water, at low temperature, enters the dehumidiﬁer to condense the humid air. Alternatively, in the second conﬁguration as shown in Fig. 2, the position of the parabolic trough solar air collector is placed between the humidiﬁer and the dehumidiﬁer. The later layout has been proposed [17] since it results in a higher system GOR. However, their study did not consider a solar heated system. As the humid air heats up in the solar collector, hot humid air is allowed to the dehumidiﬁer to have better condensation due to high temperature difference between hot humid air and cold inlet seawater (or brackish water). This results in better heat recovery such that water is heated and it leaves the dehumidiﬁer at a higher temperature. Hot water leaving the dehumidiﬁer is brought in direct contact with colder air in the humidiﬁer so that it both heats and humidiﬁes the air. This system is better compared with the original arrangement (Fig. 1) where hot air enters to the

Table 2 Design and operating parameters [1,26].

Fig. 1. The ﬁrst conﬁguration: open-water open-air humidiﬁcation dehumidiﬁcation desalination system.

Parameter

Value

Trough length Trough width Receiver cover outer diameter Receiver cover inner diameter Absorber outer diameter Absorber inner diameter Dhahran-latitude Ambient relative humidity Atmospheric pressure (ta)n Intercept factor (Y) Reﬂectivity (r) Cover thermal conductivity Stainless steel conductivity Cover emissivity Absorber emissivity Seawater inlet temperature Collector outlet temperature Mass ratio Effectiveness of the humidiﬁer Effectiveness of the dehumidiﬁer

10 m 2.5 m 0.09 m 0.082 m 0.06 m 0.0 m 26.3 0.3 101.325 kPa 0.9 0.9 0.9 1.05 W/m C 16 W/m C 0.88 0.31 20 C 75 C 1.2 0.85 0.9

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humidiﬁer (coming from the solar collector), then it humidiﬁes and cools in the humidiﬁer due to interaction with colder water, resulting in less condensation potential in the dehumidiﬁer. Although the choice of the PTSC for an HDH Desalination system may increase the cost of the system due to the requirement of a tracking system, there are many advantages gained using PTSC, which may very well overweigh its disadvantages. The air used as the working ﬂuid in the HDH system achieves higher temperature as compared to ﬂat plate collectors of the same area, which would increase the overall thermodynamic efﬁciency of the system. The heat losses from a PTSC are extremely small as compared to a ﬂat plate collector and other similar types of collectors due to the relatively small absorber area. In addition, the receiver of the PTSC has an evacuated annulus and hence suppresses the convection and conduction heat losses. On the other hand, in this study, the air heated HDH system was selected in consideration of the system lifetime. Using a parabolic trough solar water collector, results in fouling built up in the inner side of the receiver and therefore this method of heating results in relatively short lifetime of the PTSC.

Qloss ¼ UL Ar ðTr Ta Þ

where UL is the heat transfer loss coefﬁcient and Ar is the receiver area. The wind heat transfer coefﬁcient, hw, is calculated using a Nusselt number correlation for the ﬂow of air along a tube in an outdoor environment as [1]

Nu ¼ 0:40 þ 0:54Re0:52 for 0:1 < Re < 1000 Nu ¼ 0:30Re0:6 for 1000 < Re < 50; 000

3.1. Modelling of the PTSC system The radiative heat transfer from the receiver to the covers inner surface is then given by Stuetzle et al. [27]

Qloss

4 pDo Ls Tr4 Tci ¼ 1εc Do 1 εc þ εc Dci

(1)

where Do is the outer diameter of the receiver, Dci is the cover inner diameter, L is the length of the collector, s is StefaneBoltzmann's constant, Tr is the receiver temperature, Tci is the cover inner temperature, and εc is the cover emissivity. The conductive heat losses through the cover thickness are given by

Qloss

2pkc LðTci Tco Þ ¼ ln DDcoci

1 Do Do lnðDo =Di Þ þ þ UL hfi Di 2kc

Uo ¼

!1 (6)

where hﬁ is the ﬂuid heat transfer coefﬁcient and Di is the inner diameter of the receiver. For a fully developed turbulent ﬂow, the following correlation was used [28]:

Nu ¼

n ðf =8ÞðRe 1000ÞPr m qﬃﬃﬃ 2 m f w 1:07 þ 12:7 8 Pr3 1

(7)

where Pr is the Prandtl Number, m is the dynamic viscosity of air, mw is the dynamic viscosity of water, and n is a positive integer. This correlation is only valid for 2300 < Re < 5 106 and 0.5 < Pr < 2000. Here n is equal to 0.11 for heating and equal to 0.25 for cooling and f is the Darcy friction factor given by

f ¼ ð0:79 lnRe 1:64Þ2

(8)

For gases, the viscosity ratio in the Nusselt number equation is given by (Tw/T)n. Where Tw is the water temperature and T is the air temperature. The useful energy gain in terms of the collector heat removal factor is given by Ref. [1]

Ar Qu ¼ FR Aa S UL Ti Ta Aa

(9)

where Aa is the aperture area and Ti is the ﬂuid inlet temperature. The heat removal factor, FR, is deﬁned in terms of the collector efﬁciency factor F0 and collector ﬂow factor F00 where 00

F ¼

_ p mC FR A r UL F 0 1 exp ¼ _ p F0 Ar UL F 0 mC

(10)

where F0 is given by the following equation

(2)

where kc is the cover thermal conductivity, Tco is the cover outer temperature, and Dco is the cover outer diameter. The convection and radiation heat losses from the cover to the ambient and the sky are given by the following equation:

4 4 Qloss ¼ pDco Lhw ðTco Ta Þ þ εc pDco Ls Tco Tsky

(5)

where Nu is the Nusselt number and Re is Reynolds number. The overall heat transfer coefﬁcient between the environment and the ﬂuid was calculated based on the outer diameter of the absorber [1]

3. Mathematical modelling In order to achieve the objective of the study, a mathematical model is developed consisting of two main parts. The ﬁrst one is modelling the PTSC while the second one is modelling the HDH desalination system. The energy balances performed on this system assumes incompressible ﬂow and neglect pressure losses. The HDH desalination system is assumed to be well insulated with an evacuated glass tube as the cover of the absorber of the PTSC. The PTSC is designed to give a constant outlet temperature by controlling the mass ﬂow rate of the air.

(4)

(3)

where hw is the outside convection heat transfer coefﬁcient, Ta is the ambient temperature, and Tsky is the sky temperature. The overall heat transfer coefﬁcient considering the total heat loss is then

F0 ¼

Uo UL

The useful energy gain in terms of the collector inlet and outlet temperatures is given by

_ p Ti To Qu ¼ mC

(11)

where To is the outlet temperature. Taking the variation in the transmissivity, absorptivity, and the reﬂectivity into consideration the incidence angle modiﬁer is then given by Ref. [1]

Kta ¼ 1 6:74 105 q2 þ 1:64 106 q3 2:51 108 q4 (12)

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where q is the incidence angle modiﬁer. The absorbed radiation is deﬁned as

S ¼ Ib rðgtaÞn Kta

(13)

where Ib is the available beam radiation, r is the collector reﬂectivity, and ta is the transmittance absorptance product.

3.2. Modelling of the HDH system The energy balance equations for the humidiﬁer, dehumidiﬁer, and the solar air heater, respectively are presented next. The energy balance of the humidiﬁer is

m_ a ðh2a h3a Þ ¼ m_ w h3w m_ b h2w

(14)

where m_ a is the mass ﬂow rate of air, h2a is the enthalpy of air at the inlet of the humidiﬁer, h3a is the enthalpy of air at the exit of the humidiﬁer, m_ w is the mass ﬂow rate of seawater, h3w is the enthalpy of water entering the humidiﬁer, m_ b is the mass ﬂow rate of brine exiting the humidiﬁer, and h2w is the enthalpy of brine exiting the humidiﬁer. The energy balance of the dehumidiﬁer is

m_ a ðh3a h4a Þ ¼ m_ d hd m_ w ðh3w h4w Þ

(15)

where h4a is the enthalpy of air leaving the dehumidiﬁer, m_ d is the mass ﬂow rate of distilled water, hd is the enthalpy of distilled water, and h4w is the enthalpy of water at the inlet of the dehumidiﬁer. The detailed modelling of the PTSC was presented above in the previous subsection. The useful heat collected from the solar collector is

_ 2a h1a Þ Qu ¼ mðh

(16)

where h1a is the speciﬁc enthalpy of ambient air entering the PTSC. The effectiveness of the humidiﬁer and the dehumidiﬁer are given by

h¼

DH_ DH_ max

(17)

where DH_ is the change in the total enthalpy rate that can be obtained in an adiabatic heat and mass exchanger. The mass ﬂow rate of the distilled water is deﬁned in terms of the humidity ratio of the outlet and inlet of the humidiﬁer and can be written as

m_ d ¼ m_ a ðu3 u4 Þ

813

4. Results and discussion Data of the weather conditions required for simulation are the hourly data of the direct solar irradiation (W/m2), wind speed (m/ s), and ambient temperature ( C). The city selected was Dhahran, Saudi Arabia. Such detailed data were only available for the month of May, whereas for the other months, only the hourly data of the solar radiation are available and the speed and ambient temperature values are given as monthly averages. For monthly calculations the average day of the month, speciﬁed by Dufﬁe and Beckman [1], are used along with the hourly data available for those speciﬁed days of the month. Considering the calculations carried out for the average day of the month of May, Figs. 3 and 4, discussed below, were considered as an example. The GOR of the HDH system is shown in Fig. 3, while operating on the 15th of May. The GOR value as expected is lower during morning and evening hours and reaches a maximum close to noon, which follows the trend of the availability of solar radiation. Fig. 4 shows the water temperature at the exit of the dehumidiﬁer. It can be observed that the temperature variation follows the pattern of solar radiation during the day and the maximum temperature values are well below the outlet temperature of the PTSC. The variation of direct solar irradiation during the four chosen months is shown in Fig. 5, which will be used as a reference to compare the PTSC efﬁciency and the productivity of the system considered. Four months, March, June, September, and December were chosen for monthly simulations. These speciﬁc months were chosen since the weather varies drastically between them and so do the metrological data. They represents the four seasons of the year. The irradiation during the morning hours in September is shown to have a greater value than that in March, as shown in Fig. 5. These trends are due to weather conditions where, for example, the speciﬁc day chosen in March was cloudy before noon time. Fig. 6 illustrates the variations in heat input, heat gain, heat loss and reﬂected thermal radiation for the June. Similar results were observed for the other months. These four values are function of the incidence solar radiation as PTCS modelling demonstrated. Hence, as it decreases these values decreases and consequently the collected fresh water decreases. After assessing the performance of the collectors for the selected four months, monthly averaged simulations were then performed for the whole 12 months of the years and the obtained results are discussed next. Fig. 7 depicts the monthly averaged solar irradiation

(18)

where u3 is the humidity ratio at the inlet the dehumidiﬁer and u4 is the humidity ratio at the exit of the dehumidiﬁer. The mass ﬂow rate of brine is then given by

m_ b ¼ m_ w m_ a ðu3 u4 Þ

(19)

The performance of an HDH system is deﬁned by the gainedoutput-ratio as

GOR ¼

m_ d hfg Ib Aa

(20)

where hfg is the water latent heat of condensation. The model presented was carried out using Engineering Equation Solver (EES), which solves the equations simultaneously through iterations.

Fig. 3. GOR variation with time (hourly basis) for the month for May 15, Dhahran, Saudi Arabia.

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Fig. 7. Monthly averaged solar irradiation versus month of the year. Fig. 4. Water temperature at the exit of the dehumidiﬁer versus time, for May 15, Dhahran, Saudi Arabia.

Fig. 5. Direct solar radiation versus time of day (hourly basis).

versus the month of the year where the data is shown to be valid since signiﬁcant variations occur between the months as well as the four main seasons of the year. Similar patterns were followed by the collector efﬁciency, average GOR, and the desalination water produced, as shown in Figs. 8e10 and will be discussed next. It can be observed that the collector efﬁciency does not vary signiﬁcantly except for the months of February and November

Fig. 6. Heat input, heat gain, heat losses, and reﬂected beam radiation for June.

where the solar radiation collected was low as shown in Fig. 7. In these two months, the data collected was during a cloudy day. Nevertheless, this data shows what the system performance could be if the solar radiation is low. The average collector efﬁciency versus the months of the year is presented in Fig. 8. The variation of the gained output ratios with the months of the year is illustrated in Fig. 9. It can be observed that for the second conﬁguration the GOR increases signiﬁcantly to reach, on average, 4.7. This demonstrates that the second conﬁguration has much higher GOR as compared to the ﬁrst conﬁguration as explained earlier. However, one needs to notice that in the second conﬁguration the air is saturated with a relative humidity close to one. Therefore, the receiver internal pipes should be made from an anticorrosive material. The average total production per day during each month of the year considering eight hours of production a day is shown in Fig. 10. In addition, comparing the data in Fig. 10, it is evident that the second conﬁguration drastically improves the fresh water productivity of the system throughout the year from an average value of about 293 kg of fresh water for the ﬁrst conﬁguration to around 954 kg for the second conﬁguration, which is a considerable improvement. 5. Conclusions Thermodynamic analysis of two conﬁgurations of the HDH desalination systems integrated with PTSC was conducted. The study demonstrates that the second conﬁguration corresponding to the modiﬁed cycle in Fig. 2 has signiﬁcantly higher GOR as compared to the ﬁrst conﬁguration. Therefore, it may be concluded that the use of an air heater between the humidiﬁer and the

Fig. 8. Average collector efﬁciency versus the month of the year.

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ha hw hfg hﬁ DH_

Fig. 9. Monthly averaged GOR versus month.

Fig. 10. Averaged total distilled water production per month.

dehumidiﬁer, second conﬁguration, is much more advantageous than the conventional method of heating the air before the humidiﬁer. The ﬁrst conﬁguration has, on average, a GOR value of around 1.5 whereas the second conﬁguration has, on average, a GOR value of around 4.7 resulting in a higher productivity. The PTSC is used as an air heater and has an average collector efﬁciency of 0.71 throughout the year during daytime. The model developed can be used further to simulate the performance of the HDH systems proposed considering different locations. Thus, the use of PTSC is preferred for such systems due to its steady performance and effective heating almost during the whole year. Acknowledgements The authors acknowledge the support provided by King Fahd University of Petroleum & Minerals (KFUPM), Dhahran, Saudi Arabia. Nomenclature Aa Ar Cp Dci Dco Di Do F0 F00 FR GOR

aperture area, [m2] receiver area, [m2] speciﬁc heat at constant pressure, [kJ/kg C] cover internal diameter, [m] cover outer diameter, [m] receiver inner diameter, [m] receiver outer diameter, [m] collector efﬁciency factor collector ﬂow factor collector heat removal factor gained-output-ratio

Ib Ibtau kc KƬa L m_ a m_ b m_ d m_ w Qu Qloss S Ta Tci Tco Tr Tsky Tw UL Uo Vwind W

815

air speciﬁc enthalpy at ambient conditions, [kJ/kg] speciﬁc enthalpy of water, [kJ/kg] latent heat of condensation of water, [kJ/kg] ﬂuid heat transfer coefﬁcient, [W/m2 C] change in the total enthalpy rate that can be obtained in an adiabatic heat and mass exchanger, [W] available beam radiation, [W/m2] reﬂected beam radiation to the receiver, [W] thermal conductivity of the receiver cover, [W/m C] incidence angle modiﬁer length of PTSC, [m] mass ﬂow rate of air, [kg/s] mass ﬂow rate of brine, [kg/s] mass ﬂow rate of distillate, [kg/s] mass ﬂow rate of water, [kg/s] useful energy gain, [W] energy lost from the collector, [W] absorbed incident radiation, [W/m2] ambient temperature, [ C] cover inner temperature, [ C] outer cover temperature, [ C] receiver temperature, [ C] sky temperature, [ C] water temperature, [ C] heat transfer loss coefﬁcient, [W/m2 C] overall heat transfer loss coefﬁcient, [W/m2 C] wind velocity, [m/s] collector width, [m]

Greek symbols εc cover emissivity εr receiver emissivity ƞ effectiveness u humidity ratio r collector reﬂectivity s StefaneBoltzmann constant Τa transmittance absorptance product q incidence angle modiﬁer Subscripts a ambient b brine c cover d distillate i inner o outer r receiver w water References [1] J.A. Dufﬁe, W.A. Beckman, Solar Energy Thermal Processes, Wiley, 2013. [2] S. Veerapaneni, B. Long, S. Freeman, R. Bond, Reducing energy consumption for seawater desalination, J. Am. Water Works Assoc. 99 (2007) 95e106. [3] H.-J. Joo, H.-Y. Kwak, Performance evaluation of multi-effect distiller for optimized solar thermal desalination, Appl. Therm. Eng. 6 (2013) 491e499. [4] S. Riffat, A. Mayere, Performance evaluation of v-trough solar concentrator for water desalination applications, Appl. Therm. Eng. 50 (2013) 234e244. n-Padilla, J. Blanco, Evaluation of cooling [5] P. Palenzuela, G. Zaragoza, D.C. Alarco technologies of concentrated solar power plants and their combination with desalination in the mediterranean area, Appl. Therm. Eng. 50 (2013) 1514e1521. [6] A. Calle, J. Bonilla, L. Roca, P. Palenzuela, Dynamic modeling and performance of the ﬁrst cell of a multi-effect distillation plant, Appl. Therm. Eng. 70 (2014) 410e420. [7] G.P. Narayan, M.H. Sharqawy, E.K. Summers, J.H. Lienhard, S.M. Zubair, M. a. Antar, The potential of solar-driven humidiﬁcationedehumidiﬁcation desalination for small-scale decentralized water production, Renew. Sustain. Energy Rev. 14 (4) (May 2010) 1187e1201.

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Humidification dehumidification desalination system using parabolic trough solar air collector

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