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Comprehensive study on vacuum humidification-dehumidification (VHDH) desalination
Applied Thermal Engineering 169 (2020) 114944
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Comprehensive study on vacuum humidification-dehumidification (VHDH) desalination
T
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Zohreh Rahimi-Ahar, Mohammad Sadegh Hatamipour , Younes Ghalavand, Alireza Palizvan Chemical Engineering Department, University of Isfahan, Isfahan, Iran
A R T I C LE I N FO
A B S T R A C T
Keywords: Vacuum humidification-dehumidification Desalination rate Exergy destruction Exergy efficiency
In this work, a solar-assisted pilot-scale vacuum humidification-dehumidification (VHDH) desalination plant is investigated experimentally as well as thermodynamically. In this process, the humidification performs at subatmospheric pressure, while the dehumidification occurs at over-atmospheric pressure. The annual experimental results show that the maximum desalination rate is obtained as 1200 mL h−1 m−2 during summer days at the best value of water to air mass flow rate ratio (w/a) and minimum humidifier pressure. It is resulted that despite achieving a high desalination rate at low pressure, the exergy efficiency decreases. This result reveals that the obtained value of humidifier pressure from energy-exergy analyses differs from the experimental one that is obtained regarding the desalination rate.
1. Introduction Increasing energy demand and freshwater shortage are the most severe difficulties worldwide due to the continuous increase in population and industrialization. To overcome the water shortage, it is recommended to produce desalinated water from saline water by using renewable energy sources or waste heat generated from a heat-producing unit [1]. As solar energy is accessible in most countries that challenge with water shortage, this energy source has attracted interest of the researches. Based on the ways of using solar energy, solar desalination processes are divided into direct and indirect processes. Due to the low vapor pressure and operating temperature in direct methods, the productivity is less than that of indirect counterparts. However, the lowcost process and applicability for small-scale applications encourage using direct methods in remote areas [2]. One of the direct techniques is humidification-dehumidification (HDH) desalination process. The low productivity of the conventional HDH systems encouraged the researchers to look for methods to improve the desalination rate of these systems. The proposed methods include using varied pressure humidification-dehumidification (VP-HDH) systems, coupling a heat pump to HDH system, and coupling the HDH with other desalination technologies (e.g., solar still or reverse osmosis). Unlike the conventional HDH processes that the dehumidifier and humidifier work at atmospheric pressure, in VP-HDH technology, the
humidifier and the dehumidifier work at different pressures. An expander or an LRVP could run the low-pressure humidification while; the high-pressure dehumidification should be run via a compressor. The LRVP has dual functionality: vapor condensation as well as vacuum creation [3]. By reducing the humidifier pressure, the humidity ratio of the outlet air stream from the humidifier increases, which improves productivity [4]. By high-pressure condensation, no extra heat source is required for water heater [5], and the rate of condensation as well as system productivity enhances. In the VP-HDH system, the performance ratio1 is lower than those of conventional HDH processes [6]. VP-HDH systems were investigated by Mistry et al. [7] and Narayan et al. [8] that experimented using a throttle valve instead of a mechanical expander. This replacement led to less energy consumption due to lower energy consumption and the irreversibility of the expander in comparison with a throttle valve. They also coupled a reverse osmosis (RO) module to the VP-HDH system to calculate the gained output ratio (GOR). It was resulted that the optimized VP-HDH-RO unit can compete with MED and MSF against GOR. It was recommended using a thermo-compressor due to its better performance than a mechanical-compressor. The energy and exergy analyses are established by following the first and second laws of thermodynamics. Exergy comprises the chemical and thermo-mechanical exergies. It is the maximum obtained work from a system whenever the concentration, temperature, and pressure of the system are brought into equilibrium from an initial
⁎
Corresponding author. E-mail addresses: [email protected] (Z. Rahimi-Ahar), [email protected] (M.S. Hatamipour), [email protected] (Y. Ghalavand). 1 Evaporation latent heat of the distillate to the overall energy input. https://doi.org/10.1016/j.applthermaleng.2020.114944 Received 11 July 2019; Received in revised form 12 December 2019; Accepted 12 January 2020 Available online 14 January 2020 1359-4311/ © 2020 Elsevier Ltd. All rights reserved.
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Nomenclature a Ac cp ex Ex FR F′ h H hfg I M ṁ P Q̇ R s Ss T UL Ẇ Y′
amb c da dest ha j k pw s sw v wf x
accuracy collector surface area (m2) specific heat (J kg−1 K−1) specific exergy (J kg−1) exergy flow rate (W) collector heat removal factor collector efficiency factor specific enthalpy (J kg−1) enthalpy (W) water latent heat of evaporation (J kg−1) solar intensity (W m−2) molecular weight mass flow rate (kg s−1) pressure (Pa) rate of heat (W) gas constant (J mol−1 K−1) specific entropy (J kg−1 K−1) salinity (kg kg−1) temperature (K) overall heat loss coefficient (W m−2 K−1) rate of work (W) absolute humidity (kg vapour kg dry air−1)
ambient collector dry air destruction humid air species counter pure water sun saline water vapor working fluid (saline water) mole fraction
Abbreviations Exp LRVP SAH Sim SWH VHDH w/a
experimental liquid ring vacuum pump solar air heater simulation solar water heater vacuum humidification- dehumidification water to air mass flow rate
Superscript Greek symbols
ε τα
CH PH *
relative error transmittance-absorptance product
chemical physical saturation state
Subscripts 0
dead state
difference of the PHE had little effect on the GOR. Ameri and Eshaghi [15] worked on the exergy analysis of a coupled RO-HDH desalination system and compared it with a five-element RO unit. The outlet saline water of the RO unit was conducted to the HDH system to eliminate the pump requirement. The results showed that the exergy destruction in the proposed configuration decreases by 20%, and the desalination rate increases. It was concluded that a higher desalination rate and exergy efficiency could be obtained in the coupled ROHDH configuration in comparison with the five-element RO unit. Exergy analysis of a hybrid HDH-RO desalination system was conducted by Al-Sulaiman et al. [16]. Total true specific exergy lost was introduced as a novel parameter, instead of exergy efficiency to evaluate the exergetic performance of their system. The investigation on the effect of the exergy destruction rate per desalination unit was clarified through the introduced parameter. The results showed that among the used components, the thermal vapor compressor contribution to the total exergy destruction is 50%. Therefore, it strongly affected system performance. Siddiqui et al. [17] designed a VP-HDH system. A sub-atmospheric pressure in the humidifier and over-atmospheric pressure in the dehumidifier were created via a throttle valve and a compressor, respectively. The effective parameters on the desalination rate and GOR were optimized. A maximum dehumidifier to humidifier pressure ratio of 2 was considered in the proposed process. The results of the exergy analysis strengthened the obtained optimum values of the variables. It was resulted that increasing the pressure ratio improves the exergy efficiency while maximum humidifier temperature up to 60 °C has a minor effect on irreversibility. Ayati et al. [18] compared the performance of different VP-HDH systems coupling to a heat pump (HP) with a conventional HDH system
state. The exergy analysis provides better insights into recognizing the inefficient sources and selecting the optimum value of the process parameters [9]. In the contact of two streams, heat and mass transfer occur at the expense of thermal and mass gradients, which leads to degradation of the energy and generation of entropy that is called irreversibility [10]. In this condition, the exergy efficiency is less than 100% and shows the degradation of the exergy [11]. The performance of the VP-HDH desalination system could be improved through the exergy analysis. An exergy analysis in a solar multi-effect HDH system was performed by Hou et al. [12]. They concluded that the proposed process has a low exergy efficiency as well as a high exergy loss. To improve the system productivity, the following recommendations were proposed: using the solar multi-effect HDH process to achieve a higher energy recovery rate and the GOR than a single-effect one and reusing the rejected water in such a system. A 3-stage HDH process was proposed by Kang et al. [13] to realize the cascade utilization of thermal energy, which decreased the process temperature range. The multi-stage HDH process improved the performance of the system enormously. A mathematical model based on mass/energy balances in each component of the system was developed. At a saline water temperature of 85 °C, the desalination rate and GOR reached 91 kg h−1 and 5.13, respectively. He et al. [14] proposed an HDH unit driven by waste heat. A plate heat exchanger (PHE) was used to recover the waste heat from the industrial gas exhaust. It was resulted that the dehumidifier is kept in balance condition (heat capacity ratio of 1) using the thermodynamic laws. It led to the maximum GOR and minimum total specific entropy generation. It was concluded that a low top temperature enhances system performance and heat recovery. The terminal temperature 2
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the same as that reported in our previous work [6]. The flow diagram and the apparatus photo are shown in Fig. 1. The air and water are heated by a flat plate solar air heater (SAH), and an evacuated tube solar water heater (SWH), respectively. The heated air in the SAH (stream 1) flows into the humidifier, which gains vapor from sprayed saline water (stream 7). The air gets humidified, and the brine (stream 8) drains from the tower to a metal tank. The outlet humid air from the humidifier (stream 2) exchanges its heat in direct contact with the sealing water streams of LRVP (streams 10 and 11), and the vapor condenses. The outlet air from the LRVP (stream 3) passes the first separator, and the air (stream 3) and water (stream 12) are separated. The exit air (stream 3) with reduced absolute humidity flows to the inner tube of the double pipe heat exchanger that is used as dehumidifier, and the feed water (stream 5) preheating occurs. The exit preheated water from the outer tube side of the dehumidifier (stream 6) is pumped to the SWH. The desalinated water (stream 9) and the air (stream 4) are separated from the exit air of the dehumidifier in the second separator. The solar radiation intensity, ambient temperature, relative humidity, air and water temperatures are measured at different points in the system, and data are recorded hourly during the tests. The system components used in this investigation are tabulated in Table 1. It should be noted that in the proposed system, most of condensation occurs in LRVP rather than dehumidifier, and from the experimental point of view, no dehumidifier is further required.
coupling to HP. It was resulted that coupling HP enhances the desalination rate and performance ratio (PR). The VP-HDH-HP system had superior performance over a humidification-compression, and conventional HDH coupled to HP (HC-HP and HDH-HP, respectively). It was due to the superiority of VP-HDH-HP processes over the HC-HP and HDH-HP systems. The parametric study showed that by increasing the dehumidifier to humidifier pressure ratio from 2.0 to 3.0, a 215% improvement in desalination rate results. Increasing w/a values from 1.5 to 2.0 increased the desalination rate by 15%, and changing w/a from 2 to 3 reduced it by 53%. At the optimum values of the parameters, the desalinated water cost of US $4.7/m3 was obtained for the VP-HDH-HP process as the more efficient system in comparison with other proposed processes. Exergy analysis of air and water heated HDH systems coupled with an HP was presented by Lawal et al. [19]. The air and water heated HDH-HP systems were exergetically compared with an HDH system containing an electric water heater (E-HDH). The product cost, exergy destruction, and exergy efficiency were evaluated for these systems. The analysis showed that the compressor and evaporator got the highest ranking in exergy destruction among the used components. The exergy efficiencies for the E-HDH, water and air heated HDH-HP units were 0.058%, 0.069%, and 1.097%, respectively. In this work, a VHDH desalination system is investigated under weather conditions in Isfahan (32.6546° N, 51.6680° E), Iran. This manuscript is a continuation of our previous work to complete the results obtained by the VHDH system by evaluating the system performance via SEEC and adding the exergy and cost analyses. Simultaneous theoretical and experimental studies lead to a decisive condition about selecting the desirable value of humidifier pressure. This study changes the best value of humidifier pressure that was selected as 50 kPa in our previous experimental study [6]. This work shows that the profound study via experimental and exergetical analyses can be a promising approach for better VHDH performance.
2.2. Energy-exergy analyses of VHDH desalination system Theoretical studies are carried out based on energy-exergy analyses. Mass and energy conservation equations are solved through MATLAB. The exergetic model is implemented in this MATLAB code to explore the influences of humidifier pressure at different hours during test days, water salinity, and inlet w/a into the humidifier on the exergy efficiency.
2. Material and method 2.3. Error and uncertainty analyses The brackish water and seawater are synthesized by adding 10 and 35 g of sea salt to 1000 g drinkable water, respectively.
An error analysis is carried out to assign credible limits to the accuracy of the obtained values. In the present analysis, differences between the experimental and modeling data are taken into consideration to determine the error percent. All obtained data related to measurement instruments are
2.1. VHDH experimental setup and process description The experimental setup used for the VHDH desalination system is
Fig. 1. VHDH desalination system: (a) flow diagram, (b) setup photo. 3
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flow rate is kept at a fixed value of 10 m3 h−1. The effect of water mass flow rate on desalination rate reveals that at a specific value of 18 L h−1 the highest desalinated water obtains. It is concluded that at the values less than this, the wettability of packing and approaching to saturation state does not occur. Furthermore, at the higher water mass flow rates, the outlet water temperature from the SWH decreases and reduces the evaporation rate within the humidifier. In both conditions, the outlet air from the humidifier does not saturate, and the desalination rate reduces. The experimental results showing the hourly variation of desalinated water by changing the water mass flow rate is presented in Fig. 3. It is shown that the variation of water flow rate from 18 L h−1 to 10 L h−1, and 10 L h−1 to 30 L h−1 leads to a reduction in the productivity by about 35% and 17%, respectively. The humidifier pressure was decreased via the LRVP to approve the superiority of variable pressure operation in VHDH systems over conventional HDHs. It is observed that by the pressure variation from 90 kPa to 50 kPa and from 90 kPa to 70 kPa, the desalination rates change from 840 L h−1 m−2 to 1200 L h−1 m−2 and from 840 L h−1 m−2 to 1100 L h−1 m−2, respectively. It means that about 30% and 23.6% improvement in desalination rates are achieved by the pressure variation as mentioned above. It should be noted that in this situation, the inlet air and water temperatures into the humidifier are 75 °C and 96 °C, respectively. These temperatures are related to the maximum solar intensity during a summer day. The difference between enthalpies at the outlet and inlet of humidifier to their difference at saturation state is reinforced by humidifier pressure reduction. It is due to increasing the driving force in the humidifier that is calculated by Eq. (4) as follows [23]:
Table 1 Experimental set up components. Component
Capacity (unit)
SWH SAH Humidifier (packed bed) Dehumidifier (double pipe heat exchanger) Hot seawater tank Desalinated water tank LRVP water supply tank Circulation pumps Brine tank
1.5 (m2) 2 (m2) 2.5 (m) in length, 0.1 ID (m) 1.58 (m2) 200 (L) 5 (L) 100 (L) 32 (W) 50 (L)
considered to be distributed uniformly; hence, their uncertainty is of Type B and is calculated by Eq. (1) [20].
u=
a 3
(1)
The error analyses for the listed control and measuring devices and their uncertainty are tabulated in Table 2. As the exergy efficiency is a function of the number of input quantities η = f (x1, x1, x1, ...) , its uncertainty value is determined by Eq. (2) [21].
u (y ) = [(
∂y 2 2 ∂y 2 2 1 ) u (x1) + ( ) u (x2) + ...] 2 ∂x1 ∂x2
(2)
Regarding the uncertainty of (xi), u(xi) is tabulated in Table 2. The following equations are used to calculate the uncertainty error [22]:
ε (%) =
100 N
n
∑ i=1
|wexp (i) − w mod el (i)|
Driving − force =
wexp (i)
(h∗ − hout ) − (h∗ − hin ) ln
(3)
(h∗ − hout ) (h∗ − hin)
(4)
Fig. 4 shows the liquid temperature versus humid air enthalpy at the inlet and outlet streams of the humidifier by its pressure reduction from 90 kPa to 70 kPa and 50 kPa. At water mass flow rate of 18 L h−1, by pressure variation from 90 kPa to 50 kPa and from 90 kPa to 70 kPa, the driving force improves by about 31% and 23%, respectively. The variation of desalination rate at these pressures confirms the driving force variation. (30% and 23.6% improvement in desalination rate by pressure reduction from 90 kPa to 50 kPa and from 90 kPa to 70 kPa are concluded). The values of enthalpy are extracted by engineering equation solver (EES) by consideration of temperatures and relative humidity of streams at the inlet and outlet of the humidifier. The desalination rate has good agreement with the information obtains through the evaluation of the humidifier driving force. The parametric study indicates that reducing the humidifier pressure has the advantage that it intensifies the effects of this parameter over water mass flow rate on productivity. The hourly variation of the desalination rates during test days is reported in Fig. 5. The desalination rate follows the same trend as Fig. 3 in various seasons. The maximum desalination rates are 1140, 1200, 1050 and 800 mL h−1 m−2 of the SWH aperture area in June, August,
3. Results and discussion In the first step, the VHDH system is studied experimentally, and the average obtained desalination rate is reported during a one-year period. Then, the exergy analysis of the proposed system is carried out. 3.1. Experimental results The average ambient temperatures and radiation intensities from April 2017 to March 2018 are reported in Fig. 2. The ambient temperature varies between 7 and 37 °C while the solar insolation ranges between 0 and 1123 W m−2 during the days of experiments. The temperatures and radiation intensities increase until 14 PM in spring and summer days and until 12 PM in autumn and winter days and then starts to decrease. The solar intensity and the ambient temperature affect the process air and water temperatures as well as desalination rate. The main effective parameters are humidifier pressure and the mass flow rate of water. Due to the constant capacity of used LRVP, the air Table 2 The specification of the measurement instruments. Instrument
Range (Unit)
Accuracy
Standard uncertainty (Unit)
t100 thermometer resistor Hygrometer K type thermocouple Solar meter Digital multimeter
−200 to +600 (°C) 0–100 (%) −40 to +375 (°C) 0–2000 (W m−2) 0–600 (V) 0–10 (A) 0 to −1 (bar) 0 to 1000 (mL min−1) 0–16 (m3 h−1)
± 0.5 ±1 ± 1.5 ±1 ± (0.1%+1 dgt) ± (1.2%+2 dgt) ± 0.01 ± 20 ± 0.32
0.3 (°C) 0.6 (%) 0.9 (°C) 0.6 (W.m−2) 0.4 (V) 0.08 (A) 0.006 (bar) 11.8 (mL min−1) 0.18 (m3 h−1)
Vacuum gauge Rotameter Anemometer
4
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40
Ambient temperature (°C)
35 30 25 Temperature (Spring) 20
Temperature (Summer) Temperature (Autumn)
15
Temperature (Winter) 10 5 0 10
12
14
16
18
Time (hr) 1200 1100
Solar intensity (W/m2)
1000 900 800 700
Solar intensity (Spring)
600
Solar intensity (Summer)
500
Solar intensity (Autumn)
400
Solar intensity (Winter)
300 200 100 0 10
12
14
16
18
Time (hr) Fig. 2. The variation of the mean (a) ambient temperature, (b) solar intensity during test days (April 2017–March 2018). Water mass flow rate:30 L/h
Water mass flow rate:18 L/h
Water mass flow rate: 10 L/h
Desalination rate (mL.h-1.m-2)
1300 1200 1100 1000 900 800 700 600 500 9
11
13
15
17
19
21
Time (h) Fig. 3. The effect of water mass flow rate on desalination rate.
equation.
October, and January, respectively. The water production rate in summer is about 50% more than that obtained in winter. The highest desalination rates are related to the maximum solar intensity in each month. It should be noted that although solar intensity at 18 PM in the winter period is 0 W m−2, the system is still working. It is due to the existence of hot water in the storage tank of the SWH.
SEEC =
Wi̇ ṁ pw
(5)
The effect of humidifier pressure on SEEC is shown in Fig. 6. Depending on the input energy at different pressures and desalination rates, the SEEC is changed between 0.15 and 0.18 (kW h L−1) [24]. By comparing the obtained results in this study and other researches that are tabulated in Table 3, it can be claimed that the proposed system is more energy-efficient than the introduced investigations. This table
3.1.1. Performance analysis The HDH systems can be compared according to specific electrical energy consumption (SEEC) value, which is defined as the following 5
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Fig. 4. Humidification diagram at different humidifier pressures (Solar intensity of 1123 W m−2, water mass flow rate of 18 L h−1).
consumptions and desalination rates are different. This leads to different desalination costs for different humidifier pressures. A considerable difference between the desalinated water costs of 50 and 70 kPa does not result.
confirms that among conventional, variable pressure, and heat pump driven HDHs, the proposed system has the best performance based on SEEC.
3.1.2. Cost analysis The procedure of desalination cost estimation for the proposed system at different humidifier pressures is tabulated in Tables 4 and 5. The calculations are carried out for 10,000 L h−1 of feed water in average solar intensity of 1123 W m−2. Notably, the fixed capital cost for all humidifier pressures are the same, while the electrical energy
3.2. Energy-exergy analyses To calculate the temperatures and absolute humidities at inlet and outlet streams of each component (humidifier, dehumidifier, LRVP, SAH, and SWH), a time variation model is developed and is verified by
Fig. 5. The desalination rate during test days (April 2017-March 2018). 6
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Table 4 Fixed capital cost estimation of VHDH system for capacity of 10,000 (L h−1).
Fig. 6. Effect of humidifier pressure on mechanical energy consumption and SEEC.
Component
Cost (US $)
solar water heater vacuum pump Water pumps solar air heater Rotameters Humidifier Heat exchanger Pipes, valves and fittings Structural supports Total cost
17,750 35,100 565 2340 912 2925 5850 1755 877 68,074
Table 5 Cost estimation of VHDH system at different humidifier pressures [18,22].
experimental results to confirm the reliability of the results. The calculated maximum errors for mass flow rates, temperatures, and absolute humidities are 15%, 6.24%, and 9.1%, respectively. The error sources are due to improper insulation of the components and rubber hose connections, ignoring the transient-state condition in all components and other simplifying assumptions. The exergy analysis is carried out to determine the inefficiencies and to give recommendations for exergy efficiency improvement of the proposed system. The variation of the exergy efficiency of the components is affected by a reduction in humidifier pressure from 90 kPa to 50 kPa, variation in w/a, solar intensity, and water salinity. It is due to the changes in the temperatures and the corresponding absolute humidities. It should be noticed that the humidifier pressure domain is chosen by consideration of LRVP capacity. The exergy efficiency of the proposed system and other thermal desalination methods which operate at low temperatures (e.g., MED and MD) are tabulated in Table 6. The exergy efficiencies for humidifier pressure of 50 kPa, 70 kPa and 90 kPa are calculated as 21%, 27%, and 32%, respectively. Maximum uncertainty error values of exergy efficiency are determined by using Eq. (2) as ± 15.32%, ± 14.47%, and ± 14.34% for 50 kPa, 70 kPa and 90 kPa, respectively. Due to being away from the dead state, the lowest process exergy efficiency is related to the humidifier pressure of 50 kPa. This table confirms the comparable exergy efficiency of the VHDH process with MED and MD. In some cases, the exergy efficiency of the proposed system is more than that in MED and MD. It is due to the high exergy efficiencies of humidifier and dehumidifier in this process rather than other thermal desalination processes. The higher exergy efficiency of some processes in comparison with the VHDH system is due to their operation in a multi-stage or multi-effect process, which enhances the system performance.
Item (Formula, Unit)
Values in humidifier pressure 90 kPa
70 kPa
50 kPa
13614.8 0.18
13614.8 0.18
13614.8 0.18
0.057
0.057
0.057
First annual cost (FAC = CRF × P , $) Annual salvage (ASV = SFF × S, $) Electrical energy cost Ey = 300(WWaterpumps + WLRVP ) t
12048.02 775.83 28290.6
12048.02 775.83 31028.4
12048.012 775.83 38329.2
Annual RC (ARC = Ey × Z , $) Annual maintenance cost (AMC = 0.15 × FAC , $) Annual cost (AC = FAC + AMC + ARC − ASV , $) Annual yield (m3) Desalination cost ($.m−3)
763.8
837.8
1034.9
1807.20
1807.20
1807.20
13843.2
13917.2
14114.28
2,519,960 0.0055
3,299,947 0.0042
3,599,942 0.0039
Salvage (S = 0.2 × P , $) Capital recover factor (CRF = Sinking fund factor (SFF =
i (i + 1)n ) [(i + 1)n − 1
i ) [(i + 1)n − 1
for the determination of the desirable values of parameters. The results are summarized as follows:
• The maximum desalination rate is obtained in summer days by • •
4. Conclusion
•
A vacuum humidification-dehumidification (VHDH) solar desalination system is studied experimentally as well as theoretically. The annual desalination rate, system performance based on SEEC, and the process exergy efficiency are reported. It is confirmed that the results obtained through experimental, cost, and exergy analysis are reliable
•
1200 mL h−1 m−2 of the SWH aperture area at humidifier pressure of 50 kPa, the solar intensity of 1123 W m−2, and water mass flow rate of 18 L h−1. Experimental and theoretical analyses well determine the desirable value of humidifier pressure as the main parameter in VHDH systems. Unlike an enhancement in system performance at humidifier pressure of 50 kPa, the least exergy efficiency obtains. It shows that despite the beneficial influence of reducing the humidifier pressure, it imposes high irreversibility to the system. Simultaneous experimental and theoretical studies show that the exergy efficiency is not the sole deciding factor for the determination of desirable values of parameters. By considering productivity, SEEC, cost, and exergy efficiency, the best operating conditions for the humidifier are the pressure of 70 kPa, the solar intensity of 1123 W m−2, and w/a of 1.8.
Table 3 The SEEC results of different HDH desalination systems. System description
SEEC range (kW h L−1)
Reference
VHDH containing SWH, SAH, and LRVP Conventional HDH containing SAH, and an electrical water heater Conventional HDH containing SAH, and two electrical water heaters HC containing compressor, and SAH HDH-HP
0.15–0.18 0.12–0.49 0.31–0.74 0.58–0.74 0.25–0.55
Current study [25]
7
[5] [26]
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Table 6 The exergy efficiency of different low thermal desalination units. Desalination method Reverse Electrodialysis (RED) engine driven MED
24-effects 10-effects 22-effects with ideal membranes 4-effects With energy recovery Without energy recovery Multi-stage units in series pattern Multi-stage units in parallel pattern – With heat recovery Without heat recovery Humidifier pressure:50 kPa Humidifier pressure:70 kPa Humidifier pressure:90 kPa
RED-MED MED MED Direct contact membrane distillation (DCMD) Solar-powered vacuum membrane distillation (VMD) DCMD Proposed system
Maximum exergy efficiency (%)
Ref.
37 18.3 31 11.84 25.7 14.3 80 78 9.96 28.3 25.6 21 27 32
[27] [28] [29] [30] [31] [2] [32] –
Declaration of Competing Interest interests or personal relationships that could have appeared to influence the work reported in this paper.
The authors declare that they have no known competing financial Appendix A A.1. Energy equations
The energy analysis is performed by solving the equations of change for all components. The model is confirmed by experimental results (Table A.1). The results are related to 50 kPa humidifier pressure, w/a = 1.8, and radiation intensity of 1123 W m−2. The overall mass and energy balances, specific enthalpy of humid air, water vapor pressure, specific enthalpies of saline water and desalinated water are presented through Eqs. (1)–(7), respectively. The points (n) denotes the states in Fig. 1.
∑ ṁ in = ∑ ṁ out
(A.1)
∑ (ṁ . h)in − ∑ (ṁ . h)out + Q̇ + Ẇ
=0
(A.2)
hn = (c pda + c pv Yn′) Tn + hfg Yn′, n = 1, 2, 3, 4 Yn′ = 0.62198
(
PH∗ 2 O Patm − PH∗ 2 O
)
Ln PH∗ 2 O = 18.3036 −
(A.3)
, n = 1, 2, 3, 4
(A.4)
3816.44 −46.13 + Tn
(A.5)
hn = c psw Tn, n = 5, 6, 7, 8 hn = 141.355 + (4202.07Tn) −
(A.6)
(0.535Tn2)
+
(0.004Tn3),
n = 10, 11, 12
(A.7)
Energy balance equation for LRVP is calculated by Eq. (A.8).
ẆLRVP = ṁ da [(Y3′ − Y2′) h3 − h2] − ṁ 10 h10 + ṁ 11 h11 + ṁ 12 h12
(A.8)
As the SAH efficiency is affected by collector type that depends on optical efficiency, heat loss, and the environmental conditions, its energy
Table A.1 The model confirmation by experimental results. Point [Fig 1.a]
1 2 3 4 5 6 7 8 9 10 11 12
Mass flow rates (kg h−1)
Absolute humidity (kg vapour kg dry air−1)
Temperature (°C)
Experiment
Model
Error%
Experiment
Model
Error%
Experiment
Model
Error%
10 – – 10 20 20 20 – 0.2 960 1.90 960
10 12.32 10.23 10 20 20 20 17.68 0.23 960 2.09 960
– – – – – – – – 15.00 – 10.00 –
75 51 26 26 25 – 90 35 26 25 26 26
70.71 51.96 27.11 25.65 25 26.00 84.38 33.28 25.65 25 27.27 27.27
5.72 1.88 4.27 1.35 – – 6.24 4.91 1.35 – 4.88 4.88
0.005 0.22 – – – – – – – – – –
0.005 0.24 – – – – – – – – – –
0 9.1 – – – – – – – – – –
8
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Z. Rahimi-Ahar, et al.
balance equation is calculated by the following equations.
̇ QSAH = ṁ da (h1 − h4 ) = ṁ da (c pda + c pv Yn′)(T1 − T4 )= Ac FR [I (τα ) − UL (T4 − Tamb)] ṁ da c pda
FR =
Ac UL
[1 − exp(
(A.9)
Ac UL F ′ )] ṁ da c pda
(A.10)
The energy balance equation for SWH is presented through Eq. (A.11).
Mswtan k c psw
dT7 ̇ = Utan k Atan k (Tamb − T7) + ṁ 6 c psw (T6 − T7) + QSWH c dt
(A.11)
Eqs. (A.12) and (A.13) calculate the heat rates absorbed by the flowing saline water in heater and evacuated tube collectors, respectively.
̇ QSWH
wf
= ṁ sw (h7 − h6) = ṁ sw c psw (T7 − T6)
(A.12)
̇ QSWH c = I (ηA)c
(A.13)
The collector efficiency depends on the inlet and outlet water temperatures, ambient temperature, and solar intensity is calculated by Eq. (A.14).
ηc = 0.695 − 1.357[
(T7 + T6) 2
− Tamb I
] − 0.1[
(
(T7 + T6) 2
− Tamb )2 I
]
(A.14)
A.2. Exergy equations The overall exergy balance equations are revealed as follows:
̇ ̇ , in − ∑ Ex mass ̇ , out = ∑ Ex dest ̇ − ∑ Ex ̇work + ∑ Ex mass ∑ Exheat
(A.15)
The exergy destruction due to the heat and mass transfer is calculated through Eq. (A.16).
̇ )in − (mex ̇ )out + ∑ [1 − ∑ (mex
T0 ̇ − ] Qout Tout
∑ Ẇ
=
̇ ∑ Exdest
(A.16)
The total exergy, extended equations related to chemical and physical exergies are presented as follows [33]:
ex = ex PH +
∑ ex CH j
(A.17)
ex PH = (h − h 0) − T0 (s − s0)
(A.18)
ex CH j
(A.19)
=
[x j ex ¯ CH j
+ RT0 x j ln(x j )]/ Mj
The exergy efficiency is calculated as follows:
ηex = 1 −
̇ Ex dest ̇ )in + ∑ [1 − ∑ (mex
T0 ̇ ] Qout Tout
− ∑ Ẇ
(A.20)
The exergy efficiency of SAH is calculated through the Eq. (A.21). The inlet exergy is the summation of exergies due to the inlet streams as well as sun exergy.
ηex , SAH = 1 −
̇ , SAH Ex dest ̇ )in + exs ∑ (mex
(A.21)
1 T 4 T ̇ exs = QSAH [1+ ( amb ) 4 − ( amb ) 3 Ts 3 Ts
(A.22)
The general equations of the exergy of humid air and water are defined as follows: T
exhan = (c pda + c pv Yn′)(Tn − T0) − T0 [(c pda + c pv Yn′) ln( Tn )]+ 0
1 + 1.6078Y ′
Y′
n′
0′
T0 [(Rda + Rv Yn′) ln( 1 + 1.6078Y0 )] + 1.6078Yn′ Rda ( Yn ), n = 1, 2, 3, 4
exswn = c pw (Tn − T0) − [T0 (c pw ) ln(
(A.23)
Tn )], n = 5, 6, 7, 8, 9, 10, 11, 12 T0
(A.24)
A.3. Exergy destruction ratio and overall exergy efficiency To compare the HDH systems it is recommended to calculate the overall exergy efficiency, which is defined by Eq. (A.25) [34]:
ηex , overall =
∑ Ex ̇out ̇ ∑ Ex in
(A.25)
9
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A.4. Hourly exergy efficiency variation at different humidifier pressures Fig. A.1 shows the hourly exergy efficiency variations at different humidifier pressures. By a pressure reduction in the humidifier, the absolute humidity of its outlet humid air (stream 2) increases. At the same time, a reduction in the humidified air and brine temperatures (stream 8) occurs. The temperature reduction affects the exergy contents of the outlet streams more than those resulting from absolute humidity increase; hence, the reduction in humidifier exergy efficiency occurs (Fig. A.1a). Despite decreasing the inlet humid air temperature into the LRVP (less inlet exergy content) by lowering the pressure, the similar outlet temperatures obtain at different pressures. It is due to the high flow rate and low temperature of the sealing water (stream 10) of the LRVP. It should be noted that the humidifier pressure variation from 90 kPa to 50 kPa imposes inlet work to the LRVP, as discussed in our previous study [4]. The effect of both factors leads to a decrease in the exergy efficiency of the LRVP by humidifier pressure reduction (Fig. A.1b). The temperature and absolute humidity variations by the pressure reduction are not remarkable in the dehumidifier, SWH, and SAH, and cannot influence their exergy efficiencies (Fig. A.1c and d). Fig. A.1 shows that the humidifier and LRVP are more sensitive to the pressure variation in comparison with other components. The highest exergy efficiency at the pressure of 90 kPa is due to proximity to the dead state temperature and pressure. It should be noted that the lowest amount of desalinated water produces at this pressure, which is not in our favor. In Fig. A.1, the time-variation exergy efficiencies during the summer days are presented, which change from the minimum value of 2.4% in the LRVP to the maximum value of 93% in the humidifier. The ascending trend of solar intensity from morning to 14 PM and its descending trend from 14 PM to 18 PM causes roughly similar trends in the exergy efficiency of all components. It is also revealed that LRVP is the most critical component of the VHDH system from the exergetic point of view. Fig. A.1d illustrates the exergy efficiency variation of the SWH and SAH during the summer days. In SWH, due to the existence of the storage tank, the energy saving is possible, and in the afternoon, the outlet water temperature from the SWH (stream 7), as well as exergy efficiency, do not change sensibly. While, in SAH, the outlet air temperature from the heater (stream 1) decreases by solar radiation decay after 14 PM; this causes a decrease in its exergy efficiency. The absorber plate and the tubes receive high energy at the maximum solar intensity; this results in an increase in the outlet fluid temperature, as well as the energy-exergy efficiencies. The maximum exergy efficiency values of roughly 6.2% and 10.7% corresponding to the maximum incident radiation for SAH and SWH are concluded, respectively. The high sun temperature degrades to the rather low temperatures of air and water; accordingly makes low exergy efficiency in SAH and SWH. The results suggest operating the process at high humidifier pressure (90 kPa) from the exergetic point of view; on the other hand, much higher desalination rates occur at lower pressures than those at atmospheric pressure operation. Therefore, the exergy efficiency can not be the sole criterion of decision making about the optimum values of the parameters, as also concluded by Hepbasli [11]. If it is targeted on reasonable productivity, cost as well as exergy efficiency, selecting the pressure of 70 kPa can be the best choice. The aim of exergy analysis is finding the causes of exergy losses in the components and improving the operating conditions as well as the geometry of the components. Modification of the absorber plate geometry (e.g., obstacle, baffle, finned, and corrugated surfaces) and using the strategies that decrease the destruction can improve the exergy efficiency. For this purpose, the heat pipe evacuated tube heaters with proper working fluid and the selective coating of the absorber tubes are proposed for SWH. The LRVP as the highest irreversible component in this process
Fig. A.1. Time-variation exergy efficiency of (a) Humidifier, (b) Liquid ring vacuum pump, (c) Dehumidifier, (d) Solar air and water heaters. 10
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Z. Rahimi-Ahar, et al. 90 Ss=10
Exergy Efficiency (%)
80
Ss=35
70 60 50 40 30 20 10 0
Dehumidifier
Humidifier
Vacuum Pump
Solar Air Heater
Solar Water Overal exergy Heater
Fig. A.2. Water salinity effect on exergy efficiency (humidifier pressure of 70 kPa). 80 (w/a)=1
Exergy Efficiency (%)
70
(w/a)=1.8
60
(w/a)=3
50 40 30 20 10 0 Dehumidifier
Humidifier
Vacuum Pump Solar Air Heater Solar Water Heater
Fig. A.3. The effect of the water to air mass flow rate ratio (w/a) on exergy efficiency (humidifier pressure of 70 kPa).
can be exergetically optimized by designing the most efficient case with less energy consumption, proper inlet water temperature, and mass flow rate. A.5. Effect of water salinity on exergy efficiency By raising the water salinity, an increase in viscosity and surface tension while a decrease in latent heat of vaporization is expected. The less impact of salinity change on exergy efficiency is due to keeping the humidifier temperature in values less than 68 °C, which prevents the intense variation in thermodynamics properties in the process. The effect of feed water salinity variation from 10 g kg−1 to 35 g kg−1 on exergy efficiency at P = 70 kPa is demonstrated in Fig. A.2. It is revealed that by changing the salinity from brackish water to seawater, a minor change in productivity, absolute humidities, and temperatures is observed, and a significant gap in exergy efficiency differences does not happen. As expected, the exergy efficiency of 35 g kg−1 salinity is lower than that at 10 g kg−1. It is due to the minor change in heat and mass transfer in both the humidifier and the dehumidifier at two mentioned salinities. A.6. Effect of the humidifier inlet water to air mass flow ratio (w/a) on exergy efficiency The effect of w/a variation from 3 to 1 on the exergy efficiency of the components is presented in Fig. A.3. The optimum value of w/a is determined as 1.8 [6], and the exergy efficiency in this ratio is at its maximum value. The general trend of changes in exergy efficiency due to w/a alteration is the same for all components except SWH. It is expected that due to the high Reynolds number at high w/a, the mass and heat transfer rates increase at constant water temperature, Hence, an enhancement in the exergy efficiency of the humidifier and dehumidifier concludes. While the remarkable temperature drops causes by w/a growth (for constant input heat), the humidifier temperature and the carried vapor by the air decrease. This phenomenon leads to the reduction of the outlet air temperature and its absolute humidity. Both effects cause high exergy destruction that leads to the reduction of exergy efficiency. On the other hand, at low values of w/a, the outlet humid air from the humidifier is not at its saturation state, so the low absolute humidity leads to the low exergy efficiency. It means that the exergy analyses agree with experimental results; when the optimum value of w/a is selected. At the w/a of 1 and 3, the humidified air with low absolute humidity enters into the LRVP and decreases the outlet absolute humidity, hence, the low LRVP exergy efficiency results. The temperature of the inlet air (stream 3) and outlet water (stream 6) from the dehumidifier change by w/a variation. In one hand, at w/a ratio of 1, the exergy content due to the high inlet air temperature into the dehumidifier is higher than that at the ratio of 3. This leads to an increase in exergy efficiency by w/a reduction. On the other hand, the provided absolute humidities at the optimum value of w/a improve the exergy content. Both phenomena lead to a reduction of exergy efficiency by w/a variation from 1.8 to 1 and 2. Due to the constant mass flow rate of the air stream, the meaningful change in the SAH exergy efficiency does not result. The decline in the exergy 11
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efficiency of SWH by increasing w/a value is explained by the exergy reduction due to lower outlet water temperature in comparison with the low w/ a case. The w/a increase leads to increasing thermal discharges related to exit water from the SWH and decreasing the exergy efficiency. The conclusion that derives from this section is the need to optimize the w/a value corresponding to relative high exergy efficiency. The theoretical results from the exergetic point of view have a good agreement with the experimental results concluded in previous literature [6].
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Comprehensive study on vacuum humidification-dehumidification (VHDH) desalination
T
⁎
Zohreh Rahimi-Ahar, Mohammad Sadegh Hatamipour , Younes Ghalavand, Alireza Palizvan Chemical Engineering Department, University of Isfahan, Isfahan, Iran
A R T I C LE I N FO
A B S T R A C T
Keywords: Vacuum humidification-dehumidification Desalination rate Exergy destruction Exergy efficiency
In this work, a solar-assisted pilot-scale vacuum humidification-dehumidification (VHDH) desalination plant is investigated experimentally as well as thermodynamically. In this process, the humidification performs at subatmospheric pressure, while the dehumidification occurs at over-atmospheric pressure. The annual experimental results show that the maximum desalination rate is obtained as 1200 mL h−1 m−2 during summer days at the best value of water to air mass flow rate ratio (w/a) and minimum humidifier pressure. It is resulted that despite achieving a high desalination rate at low pressure, the exergy efficiency decreases. This result reveals that the obtained value of humidifier pressure from energy-exergy analyses differs from the experimental one that is obtained regarding the desalination rate.
1. Introduction Increasing energy demand and freshwater shortage are the most severe difficulties worldwide due to the continuous increase in population and industrialization. To overcome the water shortage, it is recommended to produce desalinated water from saline water by using renewable energy sources or waste heat generated from a heat-producing unit [1]. As solar energy is accessible in most countries that challenge with water shortage, this energy source has attracted interest of the researches. Based on the ways of using solar energy, solar desalination processes are divided into direct and indirect processes. Due to the low vapor pressure and operating temperature in direct methods, the productivity is less than that of indirect counterparts. However, the lowcost process and applicability for small-scale applications encourage using direct methods in remote areas [2]. One of the direct techniques is humidification-dehumidification (HDH) desalination process. The low productivity of the conventional HDH systems encouraged the researchers to look for methods to improve the desalination rate of these systems. The proposed methods include using varied pressure humidification-dehumidification (VP-HDH) systems, coupling a heat pump to HDH system, and coupling the HDH with other desalination technologies (e.g., solar still or reverse osmosis). Unlike the conventional HDH processes that the dehumidifier and humidifier work at atmospheric pressure, in VP-HDH technology, the
humidifier and the dehumidifier work at different pressures. An expander or an LRVP could run the low-pressure humidification while; the high-pressure dehumidification should be run via a compressor. The LRVP has dual functionality: vapor condensation as well as vacuum creation [3]. By reducing the humidifier pressure, the humidity ratio of the outlet air stream from the humidifier increases, which improves productivity [4]. By high-pressure condensation, no extra heat source is required for water heater [5], and the rate of condensation as well as system productivity enhances. In the VP-HDH system, the performance ratio1 is lower than those of conventional HDH processes [6]. VP-HDH systems were investigated by Mistry et al. [7] and Narayan et al. [8] that experimented using a throttle valve instead of a mechanical expander. This replacement led to less energy consumption due to lower energy consumption and the irreversibility of the expander in comparison with a throttle valve. They also coupled a reverse osmosis (RO) module to the VP-HDH system to calculate the gained output ratio (GOR). It was resulted that the optimized VP-HDH-RO unit can compete with MED and MSF against GOR. It was recommended using a thermo-compressor due to its better performance than a mechanical-compressor. The energy and exergy analyses are established by following the first and second laws of thermodynamics. Exergy comprises the chemical and thermo-mechanical exergies. It is the maximum obtained work from a system whenever the concentration, temperature, and pressure of the system are brought into equilibrium from an initial
⁎
Corresponding author. E-mail addresses: [email protected] (Z. Rahimi-Ahar), [email protected] (M.S. Hatamipour), [email protected] (Y. Ghalavand). 1 Evaporation latent heat of the distillate to the overall energy input. https://doi.org/10.1016/j.applthermaleng.2020.114944 Received 11 July 2019; Received in revised form 12 December 2019; Accepted 12 January 2020 Available online 14 January 2020 1359-4311/ © 2020 Elsevier Ltd. All rights reserved.
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Nomenclature a Ac cp ex Ex FR F′ h H hfg I M ṁ P Q̇ R s Ss T UL Ẇ Y′
amb c da dest ha j k pw s sw v wf x
accuracy collector surface area (m2) specific heat (J kg−1 K−1) specific exergy (J kg−1) exergy flow rate (W) collector heat removal factor collector efficiency factor specific enthalpy (J kg−1) enthalpy (W) water latent heat of evaporation (J kg−1) solar intensity (W m−2) molecular weight mass flow rate (kg s−1) pressure (Pa) rate of heat (W) gas constant (J mol−1 K−1) specific entropy (J kg−1 K−1) salinity (kg kg−1) temperature (K) overall heat loss coefficient (W m−2 K−1) rate of work (W) absolute humidity (kg vapour kg dry air−1)
ambient collector dry air destruction humid air species counter pure water sun saline water vapor working fluid (saline water) mole fraction
Abbreviations Exp LRVP SAH Sim SWH VHDH w/a
experimental liquid ring vacuum pump solar air heater simulation solar water heater vacuum humidification- dehumidification water to air mass flow rate
Superscript Greek symbols
ε τα
CH PH *
relative error transmittance-absorptance product
chemical physical saturation state
Subscripts 0
dead state
difference of the PHE had little effect on the GOR. Ameri and Eshaghi [15] worked on the exergy analysis of a coupled RO-HDH desalination system and compared it with a five-element RO unit. The outlet saline water of the RO unit was conducted to the HDH system to eliminate the pump requirement. The results showed that the exergy destruction in the proposed configuration decreases by 20%, and the desalination rate increases. It was concluded that a higher desalination rate and exergy efficiency could be obtained in the coupled ROHDH configuration in comparison with the five-element RO unit. Exergy analysis of a hybrid HDH-RO desalination system was conducted by Al-Sulaiman et al. [16]. Total true specific exergy lost was introduced as a novel parameter, instead of exergy efficiency to evaluate the exergetic performance of their system. The investigation on the effect of the exergy destruction rate per desalination unit was clarified through the introduced parameter. The results showed that among the used components, the thermal vapor compressor contribution to the total exergy destruction is 50%. Therefore, it strongly affected system performance. Siddiqui et al. [17] designed a VP-HDH system. A sub-atmospheric pressure in the humidifier and over-atmospheric pressure in the dehumidifier were created via a throttle valve and a compressor, respectively. The effective parameters on the desalination rate and GOR were optimized. A maximum dehumidifier to humidifier pressure ratio of 2 was considered in the proposed process. The results of the exergy analysis strengthened the obtained optimum values of the variables. It was resulted that increasing the pressure ratio improves the exergy efficiency while maximum humidifier temperature up to 60 °C has a minor effect on irreversibility. Ayati et al. [18] compared the performance of different VP-HDH systems coupling to a heat pump (HP) with a conventional HDH system
state. The exergy analysis provides better insights into recognizing the inefficient sources and selecting the optimum value of the process parameters [9]. In the contact of two streams, heat and mass transfer occur at the expense of thermal and mass gradients, which leads to degradation of the energy and generation of entropy that is called irreversibility [10]. In this condition, the exergy efficiency is less than 100% and shows the degradation of the exergy [11]. The performance of the VP-HDH desalination system could be improved through the exergy analysis. An exergy analysis in a solar multi-effect HDH system was performed by Hou et al. [12]. They concluded that the proposed process has a low exergy efficiency as well as a high exergy loss. To improve the system productivity, the following recommendations were proposed: using the solar multi-effect HDH process to achieve a higher energy recovery rate and the GOR than a single-effect one and reusing the rejected water in such a system. A 3-stage HDH process was proposed by Kang et al. [13] to realize the cascade utilization of thermal energy, which decreased the process temperature range. The multi-stage HDH process improved the performance of the system enormously. A mathematical model based on mass/energy balances in each component of the system was developed. At a saline water temperature of 85 °C, the desalination rate and GOR reached 91 kg h−1 and 5.13, respectively. He et al. [14] proposed an HDH unit driven by waste heat. A plate heat exchanger (PHE) was used to recover the waste heat from the industrial gas exhaust. It was resulted that the dehumidifier is kept in balance condition (heat capacity ratio of 1) using the thermodynamic laws. It led to the maximum GOR and minimum total specific entropy generation. It was concluded that a low top temperature enhances system performance and heat recovery. The terminal temperature 2
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the same as that reported in our previous work [6]. The flow diagram and the apparatus photo are shown in Fig. 1. The air and water are heated by a flat plate solar air heater (SAH), and an evacuated tube solar water heater (SWH), respectively. The heated air in the SAH (stream 1) flows into the humidifier, which gains vapor from sprayed saline water (stream 7). The air gets humidified, and the brine (stream 8) drains from the tower to a metal tank. The outlet humid air from the humidifier (stream 2) exchanges its heat in direct contact with the sealing water streams of LRVP (streams 10 and 11), and the vapor condenses. The outlet air from the LRVP (stream 3) passes the first separator, and the air (stream 3) and water (stream 12) are separated. The exit air (stream 3) with reduced absolute humidity flows to the inner tube of the double pipe heat exchanger that is used as dehumidifier, and the feed water (stream 5) preheating occurs. The exit preheated water from the outer tube side of the dehumidifier (stream 6) is pumped to the SWH. The desalinated water (stream 9) and the air (stream 4) are separated from the exit air of the dehumidifier in the second separator. The solar radiation intensity, ambient temperature, relative humidity, air and water temperatures are measured at different points in the system, and data are recorded hourly during the tests. The system components used in this investigation are tabulated in Table 1. It should be noted that in the proposed system, most of condensation occurs in LRVP rather than dehumidifier, and from the experimental point of view, no dehumidifier is further required.
coupling to HP. It was resulted that coupling HP enhances the desalination rate and performance ratio (PR). The VP-HDH-HP system had superior performance over a humidification-compression, and conventional HDH coupled to HP (HC-HP and HDH-HP, respectively). It was due to the superiority of VP-HDH-HP processes over the HC-HP and HDH-HP systems. The parametric study showed that by increasing the dehumidifier to humidifier pressure ratio from 2.0 to 3.0, a 215% improvement in desalination rate results. Increasing w/a values from 1.5 to 2.0 increased the desalination rate by 15%, and changing w/a from 2 to 3 reduced it by 53%. At the optimum values of the parameters, the desalinated water cost of US $4.7/m3 was obtained for the VP-HDH-HP process as the more efficient system in comparison with other proposed processes. Exergy analysis of air and water heated HDH systems coupled with an HP was presented by Lawal et al. [19]. The air and water heated HDH-HP systems were exergetically compared with an HDH system containing an electric water heater (E-HDH). The product cost, exergy destruction, and exergy efficiency were evaluated for these systems. The analysis showed that the compressor and evaporator got the highest ranking in exergy destruction among the used components. The exergy efficiencies for the E-HDH, water and air heated HDH-HP units were 0.058%, 0.069%, and 1.097%, respectively. In this work, a VHDH desalination system is investigated under weather conditions in Isfahan (32.6546° N, 51.6680° E), Iran. This manuscript is a continuation of our previous work to complete the results obtained by the VHDH system by evaluating the system performance via SEEC and adding the exergy and cost analyses. Simultaneous theoretical and experimental studies lead to a decisive condition about selecting the desirable value of humidifier pressure. This study changes the best value of humidifier pressure that was selected as 50 kPa in our previous experimental study [6]. This work shows that the profound study via experimental and exergetical analyses can be a promising approach for better VHDH performance.
2.2. Energy-exergy analyses of VHDH desalination system Theoretical studies are carried out based on energy-exergy analyses. Mass and energy conservation equations are solved through MATLAB. The exergetic model is implemented in this MATLAB code to explore the influences of humidifier pressure at different hours during test days, water salinity, and inlet w/a into the humidifier on the exergy efficiency.
2. Material and method 2.3. Error and uncertainty analyses The brackish water and seawater are synthesized by adding 10 and 35 g of sea salt to 1000 g drinkable water, respectively.
An error analysis is carried out to assign credible limits to the accuracy of the obtained values. In the present analysis, differences between the experimental and modeling data are taken into consideration to determine the error percent. All obtained data related to measurement instruments are
2.1. VHDH experimental setup and process description The experimental setup used for the VHDH desalination system is
Fig. 1. VHDH desalination system: (a) flow diagram, (b) setup photo. 3
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flow rate is kept at a fixed value of 10 m3 h−1. The effect of water mass flow rate on desalination rate reveals that at a specific value of 18 L h−1 the highest desalinated water obtains. It is concluded that at the values less than this, the wettability of packing and approaching to saturation state does not occur. Furthermore, at the higher water mass flow rates, the outlet water temperature from the SWH decreases and reduces the evaporation rate within the humidifier. In both conditions, the outlet air from the humidifier does not saturate, and the desalination rate reduces. The experimental results showing the hourly variation of desalinated water by changing the water mass flow rate is presented in Fig. 3. It is shown that the variation of water flow rate from 18 L h−1 to 10 L h−1, and 10 L h−1 to 30 L h−1 leads to a reduction in the productivity by about 35% and 17%, respectively. The humidifier pressure was decreased via the LRVP to approve the superiority of variable pressure operation in VHDH systems over conventional HDHs. It is observed that by the pressure variation from 90 kPa to 50 kPa and from 90 kPa to 70 kPa, the desalination rates change from 840 L h−1 m−2 to 1200 L h−1 m−2 and from 840 L h−1 m−2 to 1100 L h−1 m−2, respectively. It means that about 30% and 23.6% improvement in desalination rates are achieved by the pressure variation as mentioned above. It should be noted that in this situation, the inlet air and water temperatures into the humidifier are 75 °C and 96 °C, respectively. These temperatures are related to the maximum solar intensity during a summer day. The difference between enthalpies at the outlet and inlet of humidifier to their difference at saturation state is reinforced by humidifier pressure reduction. It is due to increasing the driving force in the humidifier that is calculated by Eq. (4) as follows [23]:
Table 1 Experimental set up components. Component
Capacity (unit)
SWH SAH Humidifier (packed bed) Dehumidifier (double pipe heat exchanger) Hot seawater tank Desalinated water tank LRVP water supply tank Circulation pumps Brine tank
1.5 (m2) 2 (m2) 2.5 (m) in length, 0.1 ID (m) 1.58 (m2) 200 (L) 5 (L) 100 (L) 32 (W) 50 (L)
considered to be distributed uniformly; hence, their uncertainty is of Type B and is calculated by Eq. (1) [20].
u=
a 3
(1)
The error analyses for the listed control and measuring devices and their uncertainty are tabulated in Table 2. As the exergy efficiency is a function of the number of input quantities η = f (x1, x1, x1, ...) , its uncertainty value is determined by Eq. (2) [21].
u (y ) = [(
∂y 2 2 ∂y 2 2 1 ) u (x1) + ( ) u (x2) + ...] 2 ∂x1 ∂x2
(2)
Regarding the uncertainty of (xi), u(xi) is tabulated in Table 2. The following equations are used to calculate the uncertainty error [22]:
ε (%) =
100 N
n
∑ i=1
|wexp (i) − w mod el (i)|
Driving − force =
wexp (i)
(h∗ − hout ) − (h∗ − hin ) ln
(3)
(h∗ − hout ) (h∗ − hin)
(4)
Fig. 4 shows the liquid temperature versus humid air enthalpy at the inlet and outlet streams of the humidifier by its pressure reduction from 90 kPa to 70 kPa and 50 kPa. At water mass flow rate of 18 L h−1, by pressure variation from 90 kPa to 50 kPa and from 90 kPa to 70 kPa, the driving force improves by about 31% and 23%, respectively. The variation of desalination rate at these pressures confirms the driving force variation. (30% and 23.6% improvement in desalination rate by pressure reduction from 90 kPa to 50 kPa and from 90 kPa to 70 kPa are concluded). The values of enthalpy are extracted by engineering equation solver (EES) by consideration of temperatures and relative humidity of streams at the inlet and outlet of the humidifier. The desalination rate has good agreement with the information obtains through the evaluation of the humidifier driving force. The parametric study indicates that reducing the humidifier pressure has the advantage that it intensifies the effects of this parameter over water mass flow rate on productivity. The hourly variation of the desalination rates during test days is reported in Fig. 5. The desalination rate follows the same trend as Fig. 3 in various seasons. The maximum desalination rates are 1140, 1200, 1050 and 800 mL h−1 m−2 of the SWH aperture area in June, August,
3. Results and discussion In the first step, the VHDH system is studied experimentally, and the average obtained desalination rate is reported during a one-year period. Then, the exergy analysis of the proposed system is carried out. 3.1. Experimental results The average ambient temperatures and radiation intensities from April 2017 to March 2018 are reported in Fig. 2. The ambient temperature varies between 7 and 37 °C while the solar insolation ranges between 0 and 1123 W m−2 during the days of experiments. The temperatures and radiation intensities increase until 14 PM in spring and summer days and until 12 PM in autumn and winter days and then starts to decrease. The solar intensity and the ambient temperature affect the process air and water temperatures as well as desalination rate. The main effective parameters are humidifier pressure and the mass flow rate of water. Due to the constant capacity of used LRVP, the air Table 2 The specification of the measurement instruments. Instrument
Range (Unit)
Accuracy
Standard uncertainty (Unit)
t100 thermometer resistor Hygrometer K type thermocouple Solar meter Digital multimeter
−200 to +600 (°C) 0–100 (%) −40 to +375 (°C) 0–2000 (W m−2) 0–600 (V) 0–10 (A) 0 to −1 (bar) 0 to 1000 (mL min−1) 0–16 (m3 h−1)
± 0.5 ±1 ± 1.5 ±1 ± (0.1%+1 dgt) ± (1.2%+2 dgt) ± 0.01 ± 20 ± 0.32
0.3 (°C) 0.6 (%) 0.9 (°C) 0.6 (W.m−2) 0.4 (V) 0.08 (A) 0.006 (bar) 11.8 (mL min−1) 0.18 (m3 h−1)
Vacuum gauge Rotameter Anemometer
4
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40
Ambient temperature (°C)
35 30 25 Temperature (Spring) 20
Temperature (Summer) Temperature (Autumn)
15
Temperature (Winter) 10 5 0 10
12
14
16
18
Time (hr) 1200 1100
Solar intensity (W/m2)
1000 900 800 700
Solar intensity (Spring)
600
Solar intensity (Summer)
500
Solar intensity (Autumn)
400
Solar intensity (Winter)
300 200 100 0 10
12
14
16
18
Time (hr) Fig. 2. The variation of the mean (a) ambient temperature, (b) solar intensity during test days (April 2017–March 2018). Water mass flow rate:30 L/h
Water mass flow rate:18 L/h
Water mass flow rate: 10 L/h
Desalination rate (mL.h-1.m-2)
1300 1200 1100 1000 900 800 700 600 500 9
11
13
15
17
19
21
Time (h) Fig. 3. The effect of water mass flow rate on desalination rate.
equation.
October, and January, respectively. The water production rate in summer is about 50% more than that obtained in winter. The highest desalination rates are related to the maximum solar intensity in each month. It should be noted that although solar intensity at 18 PM in the winter period is 0 W m−2, the system is still working. It is due to the existence of hot water in the storage tank of the SWH.
SEEC =
Wi̇ ṁ pw
(5)
The effect of humidifier pressure on SEEC is shown in Fig. 6. Depending on the input energy at different pressures and desalination rates, the SEEC is changed between 0.15 and 0.18 (kW h L−1) [24]. By comparing the obtained results in this study and other researches that are tabulated in Table 3, it can be claimed that the proposed system is more energy-efficient than the introduced investigations. This table
3.1.1. Performance analysis The HDH systems can be compared according to specific electrical energy consumption (SEEC) value, which is defined as the following 5
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Fig. 4. Humidification diagram at different humidifier pressures (Solar intensity of 1123 W m−2, water mass flow rate of 18 L h−1).
consumptions and desalination rates are different. This leads to different desalination costs for different humidifier pressures. A considerable difference between the desalinated water costs of 50 and 70 kPa does not result.
confirms that among conventional, variable pressure, and heat pump driven HDHs, the proposed system has the best performance based on SEEC.
3.1.2. Cost analysis The procedure of desalination cost estimation for the proposed system at different humidifier pressures is tabulated in Tables 4 and 5. The calculations are carried out for 10,000 L h−1 of feed water in average solar intensity of 1123 W m−2. Notably, the fixed capital cost for all humidifier pressures are the same, while the electrical energy
3.2. Energy-exergy analyses To calculate the temperatures and absolute humidities at inlet and outlet streams of each component (humidifier, dehumidifier, LRVP, SAH, and SWH), a time variation model is developed and is verified by
Fig. 5. The desalination rate during test days (April 2017-March 2018). 6
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Table 4 Fixed capital cost estimation of VHDH system for capacity of 10,000 (L h−1).
Fig. 6. Effect of humidifier pressure on mechanical energy consumption and SEEC.
Component
Cost (US $)
solar water heater vacuum pump Water pumps solar air heater Rotameters Humidifier Heat exchanger Pipes, valves and fittings Structural supports Total cost
17,750 35,100 565 2340 912 2925 5850 1755 877 68,074
Table 5 Cost estimation of VHDH system at different humidifier pressures [18,22].
experimental results to confirm the reliability of the results. The calculated maximum errors for mass flow rates, temperatures, and absolute humidities are 15%, 6.24%, and 9.1%, respectively. The error sources are due to improper insulation of the components and rubber hose connections, ignoring the transient-state condition in all components and other simplifying assumptions. The exergy analysis is carried out to determine the inefficiencies and to give recommendations for exergy efficiency improvement of the proposed system. The variation of the exergy efficiency of the components is affected by a reduction in humidifier pressure from 90 kPa to 50 kPa, variation in w/a, solar intensity, and water salinity. It is due to the changes in the temperatures and the corresponding absolute humidities. It should be noticed that the humidifier pressure domain is chosen by consideration of LRVP capacity. The exergy efficiency of the proposed system and other thermal desalination methods which operate at low temperatures (e.g., MED and MD) are tabulated in Table 6. The exergy efficiencies for humidifier pressure of 50 kPa, 70 kPa and 90 kPa are calculated as 21%, 27%, and 32%, respectively. Maximum uncertainty error values of exergy efficiency are determined by using Eq. (2) as ± 15.32%, ± 14.47%, and ± 14.34% for 50 kPa, 70 kPa and 90 kPa, respectively. Due to being away from the dead state, the lowest process exergy efficiency is related to the humidifier pressure of 50 kPa. This table confirms the comparable exergy efficiency of the VHDH process with MED and MD. In some cases, the exergy efficiency of the proposed system is more than that in MED and MD. It is due to the high exergy efficiencies of humidifier and dehumidifier in this process rather than other thermal desalination processes. The higher exergy efficiency of some processes in comparison with the VHDH system is due to their operation in a multi-stage or multi-effect process, which enhances the system performance.
Item (Formula, Unit)
Values in humidifier pressure 90 kPa
70 kPa
50 kPa
13614.8 0.18
13614.8 0.18
13614.8 0.18
0.057
0.057
0.057
First annual cost (FAC = CRF × P , $) Annual salvage (ASV = SFF × S, $) Electrical energy cost Ey = 300(WWaterpumps + WLRVP ) t
12048.02 775.83 28290.6
12048.02 775.83 31028.4
12048.012 775.83 38329.2
Annual RC (ARC = Ey × Z , $) Annual maintenance cost (AMC = 0.15 × FAC , $) Annual cost (AC = FAC + AMC + ARC − ASV , $) Annual yield (m3) Desalination cost ($.m−3)
763.8
837.8
1034.9
1807.20
1807.20
1807.20
13843.2
13917.2
14114.28
2,519,960 0.0055
3,299,947 0.0042
3,599,942 0.0039
Salvage (S = 0.2 × P , $) Capital recover factor (CRF = Sinking fund factor (SFF =
i (i + 1)n ) [(i + 1)n − 1
i ) [(i + 1)n − 1
for the determination of the desirable values of parameters. The results are summarized as follows:
• The maximum desalination rate is obtained in summer days by • •
4. Conclusion
•
A vacuum humidification-dehumidification (VHDH) solar desalination system is studied experimentally as well as theoretically. The annual desalination rate, system performance based on SEEC, and the process exergy efficiency are reported. It is confirmed that the results obtained through experimental, cost, and exergy analysis are reliable
•
1200 mL h−1 m−2 of the SWH aperture area at humidifier pressure of 50 kPa, the solar intensity of 1123 W m−2, and water mass flow rate of 18 L h−1. Experimental and theoretical analyses well determine the desirable value of humidifier pressure as the main parameter in VHDH systems. Unlike an enhancement in system performance at humidifier pressure of 50 kPa, the least exergy efficiency obtains. It shows that despite the beneficial influence of reducing the humidifier pressure, it imposes high irreversibility to the system. Simultaneous experimental and theoretical studies show that the exergy efficiency is not the sole deciding factor for the determination of desirable values of parameters. By considering productivity, SEEC, cost, and exergy efficiency, the best operating conditions for the humidifier are the pressure of 70 kPa, the solar intensity of 1123 W m−2, and w/a of 1.8.
Table 3 The SEEC results of different HDH desalination systems. System description
SEEC range (kW h L−1)
Reference
VHDH containing SWH, SAH, and LRVP Conventional HDH containing SAH, and an electrical water heater Conventional HDH containing SAH, and two electrical water heaters HC containing compressor, and SAH HDH-HP
0.15–0.18 0.12–0.49 0.31–0.74 0.58–0.74 0.25–0.55
Current study [25]
7
[5] [26]
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Table 6 The exergy efficiency of different low thermal desalination units. Desalination method Reverse Electrodialysis (RED) engine driven MED
24-effects 10-effects 22-effects with ideal membranes 4-effects With energy recovery Without energy recovery Multi-stage units in series pattern Multi-stage units in parallel pattern – With heat recovery Without heat recovery Humidifier pressure:50 kPa Humidifier pressure:70 kPa Humidifier pressure:90 kPa
RED-MED MED MED Direct contact membrane distillation (DCMD) Solar-powered vacuum membrane distillation (VMD) DCMD Proposed system
Maximum exergy efficiency (%)
Ref.
37 18.3 31 11.84 25.7 14.3 80 78 9.96 28.3 25.6 21 27 32
[27] [28] [29] [30] [31] [2] [32] –
Declaration of Competing Interest interests or personal relationships that could have appeared to influence the work reported in this paper.
The authors declare that they have no known competing financial Appendix A A.1. Energy equations
The energy analysis is performed by solving the equations of change for all components. The model is confirmed by experimental results (Table A.1). The results are related to 50 kPa humidifier pressure, w/a = 1.8, and radiation intensity of 1123 W m−2. The overall mass and energy balances, specific enthalpy of humid air, water vapor pressure, specific enthalpies of saline water and desalinated water are presented through Eqs. (1)–(7), respectively. The points (n) denotes the states in Fig. 1.
∑ ṁ in = ∑ ṁ out
(A.1)
∑ (ṁ . h)in − ∑ (ṁ . h)out + Q̇ + Ẇ
=0
(A.2)
hn = (c pda + c pv Yn′) Tn + hfg Yn′, n = 1, 2, 3, 4 Yn′ = 0.62198
(
PH∗ 2 O Patm − PH∗ 2 O
)
Ln PH∗ 2 O = 18.3036 −
(A.3)
, n = 1, 2, 3, 4
(A.4)
3816.44 −46.13 + Tn
(A.5)
hn = c psw Tn, n = 5, 6, 7, 8 hn = 141.355 + (4202.07Tn) −
(A.6)
(0.535Tn2)
+
(0.004Tn3),
n = 10, 11, 12
(A.7)
Energy balance equation for LRVP is calculated by Eq. (A.8).
ẆLRVP = ṁ da [(Y3′ − Y2′) h3 − h2] − ṁ 10 h10 + ṁ 11 h11 + ṁ 12 h12
(A.8)
As the SAH efficiency is affected by collector type that depends on optical efficiency, heat loss, and the environmental conditions, its energy
Table A.1 The model confirmation by experimental results. Point [Fig 1.a]
1 2 3 4 5 6 7 8 9 10 11 12
Mass flow rates (kg h−1)
Absolute humidity (kg vapour kg dry air−1)
Temperature (°C)
Experiment
Model
Error%
Experiment
Model
Error%
Experiment
Model
Error%
10 – – 10 20 20 20 – 0.2 960 1.90 960
10 12.32 10.23 10 20 20 20 17.68 0.23 960 2.09 960
– – – – – – – – 15.00 – 10.00 –
75 51 26 26 25 – 90 35 26 25 26 26
70.71 51.96 27.11 25.65 25 26.00 84.38 33.28 25.65 25 27.27 27.27
5.72 1.88 4.27 1.35 – – 6.24 4.91 1.35 – 4.88 4.88
0.005 0.22 – – – – – – – – – –
0.005 0.24 – – – – – – – – – –
0 9.1 – – – – – – – – – –
8
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balance equation is calculated by the following equations.
̇ QSAH = ṁ da (h1 − h4 ) = ṁ da (c pda + c pv Yn′)(T1 − T4 )= Ac FR [I (τα ) − UL (T4 − Tamb)] ṁ da c pda
FR =
Ac UL
[1 − exp(
(A.9)
Ac UL F ′ )] ṁ da c pda
(A.10)
The energy balance equation for SWH is presented through Eq. (A.11).
Mswtan k c psw
dT7 ̇ = Utan k Atan k (Tamb − T7) + ṁ 6 c psw (T6 − T7) + QSWH c dt
(A.11)
Eqs. (A.12) and (A.13) calculate the heat rates absorbed by the flowing saline water in heater and evacuated tube collectors, respectively.
̇ QSWH
wf
= ṁ sw (h7 − h6) = ṁ sw c psw (T7 − T6)
(A.12)
̇ QSWH c = I (ηA)c
(A.13)
The collector efficiency depends on the inlet and outlet water temperatures, ambient temperature, and solar intensity is calculated by Eq. (A.14).
ηc = 0.695 − 1.357[
(T7 + T6) 2
− Tamb I
] − 0.1[
(
(T7 + T6) 2
− Tamb )2 I
]
(A.14)
A.2. Exergy equations The overall exergy balance equations are revealed as follows:
̇ ̇ , in − ∑ Ex mass ̇ , out = ∑ Ex dest ̇ − ∑ Ex ̇work + ∑ Ex mass ∑ Exheat
(A.15)
The exergy destruction due to the heat and mass transfer is calculated through Eq. (A.16).
̇ )in − (mex ̇ )out + ∑ [1 − ∑ (mex
T0 ̇ − ] Qout Tout
∑ Ẇ
=
̇ ∑ Exdest
(A.16)
The total exergy, extended equations related to chemical and physical exergies are presented as follows [33]:
ex = ex PH +
∑ ex CH j
(A.17)
ex PH = (h − h 0) − T0 (s − s0)
(A.18)
ex CH j
(A.19)
=
[x j ex ¯ CH j
+ RT0 x j ln(x j )]/ Mj
The exergy efficiency is calculated as follows:
ηex = 1 −
̇ Ex dest ̇ )in + ∑ [1 − ∑ (mex
T0 ̇ ] Qout Tout
− ∑ Ẇ
(A.20)
The exergy efficiency of SAH is calculated through the Eq. (A.21). The inlet exergy is the summation of exergies due to the inlet streams as well as sun exergy.
ηex , SAH = 1 −
̇ , SAH Ex dest ̇ )in + exs ∑ (mex
(A.21)
1 T 4 T ̇ exs = QSAH [1+ ( amb ) 4 − ( amb ) 3 Ts 3 Ts
(A.22)
The general equations of the exergy of humid air and water are defined as follows: T
exhan = (c pda + c pv Yn′)(Tn − T0) − T0 [(c pda + c pv Yn′) ln( Tn )]+ 0
1 + 1.6078Y ′
Y′
n′
0′
T0 [(Rda + Rv Yn′) ln( 1 + 1.6078Y0 )] + 1.6078Yn′ Rda ( Yn ), n = 1, 2, 3, 4
exswn = c pw (Tn − T0) − [T0 (c pw ) ln(
(A.23)
Tn )], n = 5, 6, 7, 8, 9, 10, 11, 12 T0
(A.24)
A.3. Exergy destruction ratio and overall exergy efficiency To compare the HDH systems it is recommended to calculate the overall exergy efficiency, which is defined by Eq. (A.25) [34]:
ηex , overall =
∑ Ex ̇out ̇ ∑ Ex in
(A.25)
9
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A.4. Hourly exergy efficiency variation at different humidifier pressures Fig. A.1 shows the hourly exergy efficiency variations at different humidifier pressures. By a pressure reduction in the humidifier, the absolute humidity of its outlet humid air (stream 2) increases. At the same time, a reduction in the humidified air and brine temperatures (stream 8) occurs. The temperature reduction affects the exergy contents of the outlet streams more than those resulting from absolute humidity increase; hence, the reduction in humidifier exergy efficiency occurs (Fig. A.1a). Despite decreasing the inlet humid air temperature into the LRVP (less inlet exergy content) by lowering the pressure, the similar outlet temperatures obtain at different pressures. It is due to the high flow rate and low temperature of the sealing water (stream 10) of the LRVP. It should be noted that the humidifier pressure variation from 90 kPa to 50 kPa imposes inlet work to the LRVP, as discussed in our previous study [4]. The effect of both factors leads to a decrease in the exergy efficiency of the LRVP by humidifier pressure reduction (Fig. A.1b). The temperature and absolute humidity variations by the pressure reduction are not remarkable in the dehumidifier, SWH, and SAH, and cannot influence their exergy efficiencies (Fig. A.1c and d). Fig. A.1 shows that the humidifier and LRVP are more sensitive to the pressure variation in comparison with other components. The highest exergy efficiency at the pressure of 90 kPa is due to proximity to the dead state temperature and pressure. It should be noted that the lowest amount of desalinated water produces at this pressure, which is not in our favor. In Fig. A.1, the time-variation exergy efficiencies during the summer days are presented, which change from the minimum value of 2.4% in the LRVP to the maximum value of 93% in the humidifier. The ascending trend of solar intensity from morning to 14 PM and its descending trend from 14 PM to 18 PM causes roughly similar trends in the exergy efficiency of all components. It is also revealed that LRVP is the most critical component of the VHDH system from the exergetic point of view. Fig. A.1d illustrates the exergy efficiency variation of the SWH and SAH during the summer days. In SWH, due to the existence of the storage tank, the energy saving is possible, and in the afternoon, the outlet water temperature from the SWH (stream 7), as well as exergy efficiency, do not change sensibly. While, in SAH, the outlet air temperature from the heater (stream 1) decreases by solar radiation decay after 14 PM; this causes a decrease in its exergy efficiency. The absorber plate and the tubes receive high energy at the maximum solar intensity; this results in an increase in the outlet fluid temperature, as well as the energy-exergy efficiencies. The maximum exergy efficiency values of roughly 6.2% and 10.7% corresponding to the maximum incident radiation for SAH and SWH are concluded, respectively. The high sun temperature degrades to the rather low temperatures of air and water; accordingly makes low exergy efficiency in SAH and SWH. The results suggest operating the process at high humidifier pressure (90 kPa) from the exergetic point of view; on the other hand, much higher desalination rates occur at lower pressures than those at atmospheric pressure operation. Therefore, the exergy efficiency can not be the sole criterion of decision making about the optimum values of the parameters, as also concluded by Hepbasli [11]. If it is targeted on reasonable productivity, cost as well as exergy efficiency, selecting the pressure of 70 kPa can be the best choice. The aim of exergy analysis is finding the causes of exergy losses in the components and improving the operating conditions as well as the geometry of the components. Modification of the absorber plate geometry (e.g., obstacle, baffle, finned, and corrugated surfaces) and using the strategies that decrease the destruction can improve the exergy efficiency. For this purpose, the heat pipe evacuated tube heaters with proper working fluid and the selective coating of the absorber tubes are proposed for SWH. The LRVP as the highest irreversible component in this process
Fig. A.1. Time-variation exergy efficiency of (a) Humidifier, (b) Liquid ring vacuum pump, (c) Dehumidifier, (d) Solar air and water heaters. 10
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Exergy Efficiency (%)
80
Ss=35
70 60 50 40 30 20 10 0
Dehumidifier
Humidifier
Vacuum Pump
Solar Air Heater
Solar Water Overal exergy Heater
Fig. A.2. Water salinity effect on exergy efficiency (humidifier pressure of 70 kPa). 80 (w/a)=1
Exergy Efficiency (%)
70
(w/a)=1.8
60
(w/a)=3
50 40 30 20 10 0 Dehumidifier
Humidifier
Vacuum Pump Solar Air Heater Solar Water Heater
Fig. A.3. The effect of the water to air mass flow rate ratio (w/a) on exergy efficiency (humidifier pressure of 70 kPa).
can be exergetically optimized by designing the most efficient case with less energy consumption, proper inlet water temperature, and mass flow rate. A.5. Effect of water salinity on exergy efficiency By raising the water salinity, an increase in viscosity and surface tension while a decrease in latent heat of vaporization is expected. The less impact of salinity change on exergy efficiency is due to keeping the humidifier temperature in values less than 68 °C, which prevents the intense variation in thermodynamics properties in the process. The effect of feed water salinity variation from 10 g kg−1 to 35 g kg−1 on exergy efficiency at P = 70 kPa is demonstrated in Fig. A.2. It is revealed that by changing the salinity from brackish water to seawater, a minor change in productivity, absolute humidities, and temperatures is observed, and a significant gap in exergy efficiency differences does not happen. As expected, the exergy efficiency of 35 g kg−1 salinity is lower than that at 10 g kg−1. It is due to the minor change in heat and mass transfer in both the humidifier and the dehumidifier at two mentioned salinities. A.6. Effect of the humidifier inlet water to air mass flow ratio (w/a) on exergy efficiency The effect of w/a variation from 3 to 1 on the exergy efficiency of the components is presented in Fig. A.3. The optimum value of w/a is determined as 1.8 [6], and the exergy efficiency in this ratio is at its maximum value. The general trend of changes in exergy efficiency due to w/a alteration is the same for all components except SWH. It is expected that due to the high Reynolds number at high w/a, the mass and heat transfer rates increase at constant water temperature, Hence, an enhancement in the exergy efficiency of the humidifier and dehumidifier concludes. While the remarkable temperature drops causes by w/a growth (for constant input heat), the humidifier temperature and the carried vapor by the air decrease. This phenomenon leads to the reduction of the outlet air temperature and its absolute humidity. Both effects cause high exergy destruction that leads to the reduction of exergy efficiency. On the other hand, at low values of w/a, the outlet humid air from the humidifier is not at its saturation state, so the low absolute humidity leads to the low exergy efficiency. It means that the exergy analyses agree with experimental results; when the optimum value of w/a is selected. At the w/a of 1 and 3, the humidified air with low absolute humidity enters into the LRVP and decreases the outlet absolute humidity, hence, the low LRVP exergy efficiency results. The temperature of the inlet air (stream 3) and outlet water (stream 6) from the dehumidifier change by w/a variation. In one hand, at w/a ratio of 1, the exergy content due to the high inlet air temperature into the dehumidifier is higher than that at the ratio of 3. This leads to an increase in exergy efficiency by w/a reduction. On the other hand, the provided absolute humidities at the optimum value of w/a improve the exergy content. Both phenomena lead to a reduction of exergy efficiency by w/a variation from 1.8 to 1 and 2. Due to the constant mass flow rate of the air stream, the meaningful change in the SAH exergy efficiency does not result. The decline in the exergy 11
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efficiency of SWH by increasing w/a value is explained by the exergy reduction due to lower outlet water temperature in comparison with the low w/ a case. The w/a increase leads to increasing thermal discharges related to exit water from the SWH and decreasing the exergy efficiency. The conclusion that derives from this section is the need to optimize the w/a value corresponding to relative high exergy efficiency. The theoretical results from the exergetic point of view have a good agreement with the experimental results concluded in previous literature [6].
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