Thermodynamic and economic analysis of a combined plant for power and water production

Accepted Manuscript Thermodynamic and economic analysis of a combined plant for power and water production W.F. He, D. Han, T. Wen, H.X. Yang, J.J. Chen PII:

S0959-6526(19)31217-X

DOI:

https://doi.org/10.1016/j.jclepro.2019.04.140

Reference:

JCLP 16496

To appear in:

Journal of Cleaner Production

Received Date: 3 January 2018 Revised Date:

7 April 2019

Accepted Date: 13 April 2019

Please cite this article as: He WF, Han D, Wen T, Yang HX, Chen JJ, Thermodynamic and economic analysis of a combined plant for power and water production, Journal of Cleaner Production (2019), doi: https://doi.org/10.1016/j.jclepro.2019.04.140. This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting proof before it is published in its final form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.

ACCEPTED MANUSCRIPT

Thermodynamic and economic analysis of a combined plant for power and water production W.F. He1,2*, D. Han1, T. Wen2, H.X. Yang2, J.J. Chen1

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1. Nanjing University of Aeronautics and Astronautics, College of Energy and Power Engineering, Energy Conservation Research Group (ECRG), Nanjing, 210016, China

2. The Hong Kong Polytechnic University, Department of Building and Service Engineering, Renewable Energy

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Research Group (RERG), Hong Kong, China

Nanjing University of Aeronautics and Astronautics

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Corresponding author: W.F. He

No. 29 Yudao Street, Qinhuai District, Nanjing, Jiangsu Province, 210016, China. Tel./Fax numbers: +8602584893666

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E-mail address: [email protected], [email protected]

ABSTRACT: This paper proposes a novel combined system, driven by industrial waste heat, to satisfy the

simultaneous demand both for power and freshwater. The concept of organic Rankine cycle is applied to achieve

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the power generation, while a humidification dehumidification desalination unit is introduced to provide freshwater.

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For the purpose of efficient energy utilization, the internal energy of the discharged brine and preheated seawater is recovered. Based on the energetic and entropic analysis, the thermodynamic performance of the combined system is simulated, and the influence laws, mainly from the condensing temperature, terminal temperature difference of the recuperator and mass flow rate ratio of the feed seawater, are revealed. Finally, the economic viewpoint of the power and water combined system is also focused. The results show that maximum values of 13.1kW for the net output power, and 208kgh-1 for the water production can be acquired, when isobutane is applied as the working fluid. It is found that lower condensing temperature and terminal temperature difference of the recuperator indicate 1

ACCEPTED MANUSCRIPT a higher water production, while the total efficiency can be further determined in combination with the energy input. It is also discovered that the mass flow rate ratio of the feed seawater is a critical parameter to influence the system performance, although the final effect will be restricted by the entropy generation rate of the system components.

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Through the comparison among the combined systems, the advantages of the current type are proved by the unit area and cost of production. With respect to the economic performance, a fixed investment for the entire combined system, 49934€, the cost of production with 0.0032€L-1 for water and 0.063€kW-1h-1 for the electricity are obtained.

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It is also revealed that the production cost can be compressed with higher share of revenues, operation hours and

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design lifetime.

Keywords: combined system; organic Rankine cycle; humidification dehumidification desalination unit; energetic and entropic analysis; influence laws

Nomenclature

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Roman symbols

specific enthalpy (kJkg-1)

h

latent heat (kJkg-1)

hfg L

length (m)

mass flow rate (kgs-1)

m

pressure (MPa);

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p Q

heat load (kW) specific entropy (kJkg-1K-1);

s

concentration of seawater (gkg-1); entropy rate (kJs-1K-1)

T W

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S

temperature (K) power (kW)

Greek letters ε ω

effectiveness humidity ratio (gkg-1)

Subscripts a

air

b

brine; blower

c

condenser

d

dehumidifier

da

dry air

e

evaporator 2

generation; generator

h

hot; humidifier

m

mechanical; mixer

net

net

pin

pinch

pu

pump

r

recuperator

RH

relative humidity

s

saturation

sw

seawater

t

total

tur

turbine

w

water

we

waste exhaust

Abbreviation Annual cost

AMC

annual maintenance cost

ASV

annual salvage value

CT

condensing temperature

FAC

first annual cost

HCR

heat capacity ratio

HDH

humidification dehumidification

HRSG

heat recovery steam generator

MFRR

mass flow rate ratio

SOF TTD UAP UCP

pinch temperature difference

share of revenues

terminal temperature difference unit area production unit cost production

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1. Introduction

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PTD

organic Rankine cycle

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ORC

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AC

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g

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Owing to the contradiction between the economic development and environment, the shortage of some strategic resources, mainly including fossil fuel and freshwater, has attracted more and more attentions all over the world. As the leading secondary energy, electricity is necessary both in life and industry. Large amounts of energy, obtained from coal, oil and natural gas, will be consumed, and environment pollution with sulfide, nitride and carbon dioxide will arise. Accordingly, alternative energy sources should be applied to achieve power generation, relieving the serious energy and environmental crisis. 3

ACCEPTED MANUSCRIPT Waste heat is just a rational candidate, emerging as exhaust gas, steam and water in the industry. Simultaneously, new thermal cycles, such as organic Rankine cycle (ORC) [1-3], Kalina cycle [4] and carbon dioxide brayton cycle [5] were always involved. The organic Rankine cycle was especially appropriate to achieve

countries with large amount of waste heat, especially China.

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the utilization of waste heat or renewable energy [6, 7], and great potential of development is available for the

Sadeghi [8] used the zeotropic mixtures as the working fluid in the ORC to improve the thermodynamic

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performance. After the performance optimization, elevation magnitudes of the thermal efficiency, 27.8%, 25.0%

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and 24.8%, were obtained in the simple, parallel two-stage and series two-stage ORC systems, respectively. Furthermore, the use of zeotropic mixtures resulted in a more compact turbine, which declined the relevant investment. Jubori [9] delivered new performance methods for small-scale turbines in organic Rankine cycles, evaluating the influences of the layout for the single or two stage turbine on the performance of the ORC, driven by

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low grade waste heat. Different turbines, including axial, radial-inflow and radial-outflow type, were designed, and the performance of the ORC’s system with their single and two stage configurations was compared and optimized by varying their expansion ratio by numerical simulation. With R245fa as the cycling working fluid, it was found

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that the configurations with two-stage axial and radial-outflow turbines showed a higher performance, and the

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isentropic efficiencies were 85% and 83% with a output power of 15.8 kW and 14.3 kW. It was concluded that a two-stage scheme within a small-scale ORC driven by low grade waste heat showed great advantages. Organic Rankine cycles was regarded as a effective way to convert low grade waste heat into power [10], designed for unmanned operation with little maintenance. Hence, due to the excellent characteristics, several plants with basic ORC driven by waste heat were taken into reality, and the next step was to optimize the architectures for a lower cost [11]. White [12] introduced a mixed-integer non-linear programming to achieve the optimization of the ORC. After the applications for three industrial waste heat recovery occasions, the selection mechanisms for 4

ACCEPTED MANUSCRIPT different temperature of waste heat were concluded, and the influence from the working fluid type on the performance of the ORC was also demonstrated. Quoilin [13] focused on the thermodynamic and economic optimization for a small scale ORC with waste heat recovery. It was found that with the same fluid, different

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objective functions would lead to different optimal working conditions. For instance, the economical optimum resulted in an investment of 2136€kW-1, a net power of 4.2kW, and a thermal efficiency of 4%, while an overall efficiency of 5% was obtained under the thermodynamic optimum for the same fluid of n-butane. To overcome the

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problem from the single objective optimization, Wang [14] conducted a multi-objective optimization of the ORC

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using non-dominated sorting genetic algorithm-II (NSGA-II), considering thermodynamic and economic aspects simultaneously. With exergy efficiency and overall capital cost as the two objective functions, the optimum design solution with corresponding decision variables was selected from the obtained Pareto frontier. With respect to the freshwater, it was also a severe problem both for life and industrial production due to the

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existing pollution and population expansion. Accordingly, desalination methods, such as multi-stage flashing, multi-effect evaporation, vapor compression and so on [15-17], were proposed and implemented to achieve the water supply. Nevertheless, the methods above also belonged to the huge energy use and pollution version, and

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such systems were efficient only when the large scale water demand was existing due to the complicated

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configurations [18]. Owing to the necessarity of water and electricity both for life and industrial production, relevant devices to produce water and electricity at the same time, which have advantages of energy conservation and non-pollution, had great potential. In combination with the ORC power cycle for waste heat recovery, a compact and efficient desalination configuration is required to be coupled for water supply, and the humidification dehumidification desalination (HDH) method, which simulates the water vapor cycle in the nature environment, comes into the sight [19, 20]. Early in 1999, the method of computer simulation was introduced to the performance investigation of the 5

ACCEPTED MANUSCRIPT HDH desalination system by Nawayseh [21, 22]. After building the conservation equations, the thermodynamic performance of the HDH desalination system and the heat and mass transfer coefficients for the relevant components were predicted and analyzed. It was proved that the humidification dehumidification processes were

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appropriate for water production using solar energy. Campos [23] built the mathematical models of a HDH desalination unit with saturated air cycled. Through the experimental platform, critical thermal parameters were optimized for the internal temperature. After validating the theoretical models, a parametric study was completed to

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look for the relation between the water production and the input boundaries. After the simulation, it was revealed

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that the packing height in the humidifier, heat load from solar radiation and the mass flow rate of inflow seawater were the most critical factors to change the water supply, while the environment temperature was discovered to have a limited effect. Rajaseenivasan [24] accomplished a test investigation with solar collectors, which were applied both for the heating of water and air, in a HDH desalination system. The experimental data indicated the

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related parameters of the water producing capacity, including the mass flow rate and operation temperatures of the working fluid. An overall efficiency of 68% with the concave turbulators was obtained, and the highest water production for the configurations without turbulators, convex and concave turbulators were 12.4, 14.1 and 15.2

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kgm-2 one day, respectively.

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Actually, the combination between the ORC and HDH unit is not an original idea to accomplish power and water production. He [25, 26] arranged the HDH unit as the top cycle for desalination. The organic working fluid absorbed the carried heat from the reheated brine, and then it rotated the turbine with power generation. The research results showed it was possible for such configurations to provide power and water simultaneously. He [27] also conducted the performance analysis of the combined system with ORC as the top cycle, while an air-heated HDH desalination system was applied as the bottom frame. Ariyanfar [28] focused on the combined ORC and HDH systems, in which ORC was regarded as the top cycle and the HDH unit with natural air circulation mode was 6

ACCEPTED MANUSCRIPT coupled as the bottom cycle. Based on the macro framework above, thermodynamic and economic performance of four scenarios was calculated and compared. In the light of the literature survey, it can be concluded that attentions has been paid on the combined system

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with the ORC and HDH unit coupled. However, the relevant limitations are also evident: 1. The reheat layout in the combined systems with HDH as the top cycle will increase the complexity and relevant cost. 2. The discharging of the hot brine into the ambient will impact the thermal efficiency of the HDH desalination system. 3. The air-heated

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HDH desalination system to absorb the condensing heat will raise the heat transfer area and the corresponding cost.

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4. The water production will be influenced a lot, while the natural air circulation was applied in the HDH unit. In this paper, a novel system was proposed to solve such limitations for power and water production, in which the ORC is adopted as the top cycle, while water-heated HDH desalination cycle is used as the bottom cycle. Furthermore, humidification and dehumidification processes with forced air circulation are relatively independent,

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and the internal energy from the brine and preheated cooling seawater are further recovered. Based on the built mathematical models, the entropic analysis is achieved to look for the true operation conditions to prove the feasibility of the system, and the energetic analysis is conducted to investigate the influence laws from the internal

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parameters on the performance of the combined system, including the power output, water production and total

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efficiency. Finally, the detailed economic performance of the combined system is also calculated, to calculate the fixed investment, the cost of water and electricity production as well as the influence factors. The research results provide significant references for the design and further optimization for the power and water combined system. 2. Description of the combined plant For the currently investigated system, the ORC and HDH units are connected by the intermediate condenser. For the organic Rankine cycle, the saturated vapor is generated in the HRSG after absorbing the released heat from the waste exhaust, and then the organic vapor with high pressure and temperature expands in the turbine to drive 7

ACCEPTED MANUSCRIPT the rotor for power generation. The discharged vapor from the turbine is condensed by the feed seawater, which participates the seawater cycle in the humidifier, and the organic condensation is pressurized by the pump, entering the HRSG to close the thermal cycle of the organic working fluid.

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With respect to the HDH unit, it can be seen that the humidification and dehumidification processes, linked through the humid air stream, are relatively independent. For the feed seawater into the humidifier, it is heated in the recuperator and condenser for a top temperature. The hot seawater is sprayed into the humidifier, contacting the

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air with low temperature and humidity ratio. As a result, the cycled air is humidified and heated, while the seawater

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is concentrated. For the dehumidification process, the other stream of feed seawater flows into the dehumidifier to cool the hot saturated humid air. Freshwater is produced during air dehumidification, while the temperature of the feed seawater is raised, and the dehumidified air flows into the humidifier to close the air cycle. Finally, the preheated feed seawater joins the concentrated brine, and internal energy of the total stream is recovered in the

3. Methodology 3.1 Organic Rankine cycle

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recuperator.

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Detailed configurations of the power and water combined system is shown in Fig. 1. Apparently, it is observed

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that the organic working fluid, with isobutane, R245fa and R123 appointed, evaporates in the heat recovery steam generator (HRSG), recovering the waste heat from the hot exhaust, with thermal parameters in Table 1. Temperature profiles of the working fluids in the HRSG are presented in Fig. 2 (a), with TTDHRSG and PTDHRSG regarded as the terminal temperature difference (TTD) at the exhaust inlet of the evaporator and the corresponding pinch temperature difference (PTD), respectively. According to the energy equilibrium within the HRSG, the relevant heat transfer rate, QHRSG, can be calculated in Eq. (1).

QHRSG = mwe (hwe,i − hwe,o ) = mORC (htur − hp ) 8

(1)

ACCEPTED MANUSCRIPT In addition of the analysis based on the first law, the entropy situations within all the components are also focused. Hence, the entropy equation for the HRSG can be expressed as:

Sgen, HRSG = mwe (swe,o − swe,i ) + mORC (stur − s p )

(2)

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After the evaporation of the organic working fluid in the HRSG, the obtained steam enters the turbine, expanding with a power output of Wtur, and the specific expressions both for the energy and entropy equations can be obtained as:

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Wtur = mORC (htur − hc,i )

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Sgen,tur = mORC (sc,i − stur )

(3) (4)

After the discharge of the steam from the turbine, the exhaust is condensed under the cooling from the seawater with the detailed temperature profile in Fig. 2(b). The corresponding heat balance within the condenser can be calculated in Eq. (5).

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Qc = mORC (hc,i − hc ) = msw1 (hsw1,2 − hsw1,1 )

(5)

Furthermore, the entropy generation rate of the condenser can be obtained in Eq. (6).

Sgen,c = mORC (sc − sc,i ) + msw1 (ssw1,2 − ssw1,1 )

(6)

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Afterwards, the condensate is pressurized in the pump, and the relevant power consumption, Wp and entropy

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generation rate, Sgen,p, can be attained in Eq. (7) and Eq. (8).

Wp = mORC (hp − hc )

(7)

Sgen, p = mORC (s p − sc )

(8)

In combination of the mechanical and generator efficiency with ηm=ηg=0.98, the final net power of the ORC unit, Wnet, can be acquired.

Wnet = Wturηmηg − Wp 3.2 Humidification dehumidification desalination unit

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(9)

ACCEPTED MANUSCRIPT The feed seawater is heated in the condenser, and then it is sprayed into the humidifier. After the direct contact between the hot seawater and the air, the temperature as well as the humidity ratio is raised while the seawater is concentrated. As a result, the relevant energy and energy equations can be obtained.

msw1 − mb = mda (ω2 − ω1 )

msw1hsw1,2 − mb hb = mda (ha 2 − ha1 )[29,30]

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(10) (11)

With respect to the entropic analysis, the relevant entropy generation rate during the humidification process

Sgen, h = mb sb − msw1ssw1,2 + mda (sa 2 − sa1 )

(12)

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can be calculated in Eq. (12).

(13)

After the humidified air is attained, it is condensed under the cooling effect of another stream of feed seawater. Consequently, freshwater is produced during the condensation, and the corresponding conservation equation can be expressed as:

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mw = mda (w2 − w1 )

mda ( ha 2 − ha1 ) = msw 2 ( hsw 2,1 − h0 ) + mw hw

Sgen, d = mw sw + msw2 (ssw2,1 − s0 ) + mda (sa1 − sa 2 )

(14)

(15)

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In order to recover the energy contained in the brine and the preheated seawater in the dehumidifier, a mixer

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and recuperator are installed. For the confluence of the brine and preheated feed seawater, the relevant mass, energy and entropy equations can be obtained.

msw = mb + msw2

(16)

msw hsw = mb hb + msw 2 hsw 2,1

(17)

Sgen, m = msw ssw − mb sb − msw2 ssw2,1

(18)

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ACCEPTED MANUSCRIPT After the mixing process, the obtained solution flows into the recuperator, releasing the energy to the feed seawater involved in the humidifier before entering the condenser. For the recuperator, the corresponding energy and entropy equations are given out in Eq. (19) and Eq. (20).

Qr = msw (hsw − hsw,o ) = msw1 (hsw1,1 − h0 )

3.3 Assessment of the power and water combined plant

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Sgen,r = msw (ssw,o − ssw ) + msw1 (ssw1,1 − s0 )

(19) (20)

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Based on the first law of thermodynamics, a definition of total efficiency is introduced to measure the energy conversion conditions within the power and water combined system, which is calculated as the summation of the

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evaporation heat and net generated power divided by the energy input.

ηt =

mw hfg + Wnet QHRSG

(21)

Furthermore, at the aspect of the feasibility as well as the irreversible loss for the thermal processes, the water

rate is calculated as:

sgen, HRSG + sgen,tur + sgen, p + sgen,c + sgen,h + sgen,d + sgen,r + sgen,m mw

(22)

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sgen,t =

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production is used as the standard to measure the entropy generation. Accordingly, the specific entropy generation

Thus, the mathematical models of the entire combined plant as well as the assessment criterions are built, and

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the corresponding performance can be calculated iteratively with the procedures, shown in Fig. 3. 4 Validation of the mathematical models for the power and water combined system Before the performance simulation, the established mathematical models must be validated through the comparison between the simulation results and the experimental data or published results. Since the ORC and HDH unit constitute the combined system, the mathematical models of the subsystems are validated, respectively. For the ORC subsystem, the net power generation is compared with the published data from Zhang [31], shown in Table 2, with a maximum deviation of 3.35%. With respect to the HDH desalination unit, the characteristics during 11

ACCEPTED MANUSCRIPT humidification and dehumidification are verified through the results from Narayan [32] at the consistent boundaries, shown in Table 3 and Table 4, respectively. It is found that the maximum error of the humidification performance emerges as 0.3% for the brine mass flow rate, while the corresponding value is 5.1% for the water production

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during dehumidification. Therefore, based on the limited difference between the simulation results and the published data, the accuracy of the established mathematical models is confirmed. 5. Result and discussion

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It is demonstrated that the power and water combined system is made up of the ORC and HDH units. The

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characteristics of the ORC are first calculated to determine the best performance as well as the relevant working conditions at the shown conditions in Table 5. With the organic working fluid in the condenser as the heat source, the thermodynamic performance of the HDH desalination unit is then calculated. It can be seen that, the pinch temperature difference (PTD) both for the HRSG and condenser are regarded as PTDHRSG=10K and PTDc=10K,

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with a condensation temperature of Tc=333K, while the terminal temperature difference of the recuperator is TTDr=5K. Moreover, based on the effectiveness definition [33], the effectiveness for the varied-humidity processes are considered at ε=0.85.

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5.1 Performance of the ORC unit

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In order to achieve the performance simulation for the entire combined plant, the specific characteristics of the top ORC cycle, should be first calculated, with three different types of dry organic working fluids as isobutane, R245fa and R123. Accordingly, the net output power and thermal efficiency are attained and presented in Fig. 4. In view of the temperature distribution along the flow path within the HRSG, it is observed that with the increase of the inlet temperature, as the saturated temperature corresponding to the inlet pressure of the turbine, the specific enthalpy difference will rise, while the relevant released heat from the waste exhaust will descend at the fixed PTD. As a result, due to the constant mass flow rate of the waste exhaust, mwe=3kgs-1, a reduction of the mass flow rate 12

ACCEPTED MANUSCRIPT for the organic working fluid is gained. Taking isobutane for instance, the mass flow rate declines from mORC=0.76kg-1 at Ttur=344K to mORC=0.37kg-1 at Ttur=396K. On the other hand, after the ascent of the temperature of the working fluid into the turbine, the specific enthalpy difference during the expansion process also rises, when

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the condensing temperature (CT) is fixed at Tc=333K. Therefore, based on the completely reverse trend of the specific enthalpy difference and the relevant mass flow rate for the organic working fluid, a peak net power of the ORC unit is obtained. The relevant maximum values of the net power for the three prescribed types of working

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fluids, isobutane, R245fa and R123, are Wnet=13.1kW, 12.5kW and 11.9kW, respectively.

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Thermal efficiency is another important parameter for the ORC, which is calculated as the net power divided by the heat load of the HRSG. The increase of the evaporation temperature will result in the heat load declination due to the temperature profile described in Fig. 2(a). As a result, a continuous increasing trend of the thermal efficiency is obtained, with the values of ηORC=8.7%, 9.4% and 10.0% for isobutane, R245fa and R123,

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respectively.

5.2 Performance of the humidification dehumidification unit After the acquisition of the maximal net power from the turbine, the corresponding working conditions are

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applied to be integrated into the HDH unit. Before the performance analysis of the HDH unit, the entropy

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generation rates for all the thermal processes, including those within the ORC unit, are calculated to confirm the practicability of the operation conditions, shown in Fig. 5. In reality, during the performance simulation of the HDH desalination unit, the definition of effectiveness, currently with ε=0.85 for humidification and dehumidification, and heat capacity ratio (HCR) [33] are utilized to measure the characteristics of the heat and mass transfer processes. As a result, the thermodynamic performance of the combined system can be obtained. It is found that the entropy generation rates for all the involved thermal processes, except the humidification, are always positive under the appointed conditions. Taking isobutane for example, the range of the air mass flow rate, from mda=0.6kgs-1 to 13

ACCEPTED MANUSCRIPT mda=0.8kgs-1, is not workable to produce freshwater. Furthermore, it is also observed that the HRSG has the biggest irreversible loss during the running of the combined system for all the appointed working fluids. Water production and total thermal efficiency of the combined system are calculated and presented in Fig. 6,

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in which the impossible working conditions, with negative specific entropy generation rate, are marked with short dot. According to the definitions of effectiveness and heat capacity ratio (HCR) [33], it was proved that the balance conditions for the heat and mass transfer processes, containing the humidification and dehumidification, were

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significant for the system performance. For the freshwater production, it can be seen that the maximum values

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reach mw=208kgh-1, 199kgh-1 and 185kgh-1 for isobutane, R245fa and R123, respectively, at the balance condition of the dehumidifier, HCRd=1. It has been illustrated that the current HDH desalination unit is running at the conditions, in which the ORC unit has the largest output net power. Hence, the recovered energy in the HRSG is QHRSG=171kW, 164kW and 152kW, and the corresponding released heat in the condenser is 157kW, 150kW and

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140kW. Among the three cases, the conditions with the working fluid of isobutane have the maximum mass flow rates, both for the feed seawater and dry air into the humidifier, with values of msw1=1.52kgs-1 and mda=1.14kgs-1. Finally, a top value of the water production, mw=208kgh-1, is obtained. Due to the different trends of the output

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power, water production and relevant recovered energy in the HRSG, the top total efficiencies for the three

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designated types of organic working fluids are ηt=89.3%, 89.4% and 89.1%, respectively. Actually, based on the presentation of the specific entropy generation rates for all the components shown in Fig. 5, the specific entropy generation rate for the entire combined system can be summarized, and the relevant bottom values, with sgen,t=1.76kJkg-1K-1, 1.73kJkg-1K-1 and 1.82kJkg-1K-1 in sequence, are just corresponding to the maximal total efficiencies. It can be concluded the application of isobutane contributes to a maximal value both for the output power and water production, while the total efficiency is almost the same to those with the other types of organic working fluids. The specific performance of the combined system at the balance conditions is listed in Table 6. 14

ACCEPTED MANUSCRIPT 5.3 Influences from the critical parameters of the power and water combined system A. Condensing temperature of the ORC unit It has been proved that isobutane is the most significant option to be utilized in the current combined system,

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providing power and water. In the next analysis, isobutane is the only working fluid to accomplish the parametric simulation. Initially, the condensing temperatures, with three values of Tc=323K, 333K and 343K, are specified to investigate the corresponding influence laws on the performance of the combined system. For each condensing

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temperature, similar to the previous analysis, the top values for the net output power are first found with

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Wnet=17.0kW, 13.1kW and 9.71kW in sequence at the turbine inlet temperature of Ttur=375K, 380K and 384K. In view of the conditions with the aforementioned top output power, the water production and total efficiency from the HDH unit are calculated and presented in Fig. 7. It is seen that the variation of the condensing temperature will contribute to the performance alternation of the combined system. In response to the change of the evaporation

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temperature from Ttur=375K to Ttur=384K, the outlet temperature of the waste exhaust will also ascend from Twe,o=358K to Twe,o=374K based on the temperature profile described in Fig. 2(a). As a result, the case of Tc=323K has top values both for the heat load in the HRSG and condenser, with QHRSG=194kW and Qc=176kW. Therefore,

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the mass flow rate of the feed seawater to consume the condensation heat declines from msw1=2.13kgs-1 to

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msw1=1.10kgs-1. Furthermore, to realize the given effectiveness for the heat and mass transfer processes, the involved mass flow rate of the humid air also moves from mda=1.95kgs-1 to mda=0.67kgs-1. Due to the dominant effect to determine the water production, a top value with mw=212kgh-1 is obtained at the balance condition of the dehumidifier, when the condensing temperature keeps at Tc=323K. For the total efficiency, due to the smallest energy input for the case of Tc=343K, a maximum value of ηt=94% is obtained at the balance condition of the dehumidifier, HCRd=1. B. Terminal temperature difference of the recuperator 15

ACCEPTED MANUSCRIPT According to the stream diagram within the combined system, the concentrated brine and preheated seawater at the outlet of the dehumidifier will flow out. Consequently, it is a waste of energy if the carried thermal energy is not recovered, and the recuperator is available to preheat the feed seawater into the humidification process. The

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terminal temperature difference of the recuperator is also appointed to investigate the relevant influences on the performance of the combined system, with TTDr=2K, 5K and 8K, exhibited in Fig. 8. Obviously, the smaller the TTD value is, the better the energy recovery effect is achieved. It can be obtained that the actual produced water

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reaches mw=235kgh-1 at the case of HCRd=1, when the terminal temperature difference stays at TTDr=2K, which

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indicates an elevation magnitude of 26.1% will be attained from TTDr=8K to TTDr=2K. Similar to the situations of the condensing temperature alternation, a larger feed seawater mass flow rate will emerge to satisfy the heat load within the condenser for the case of TTDr=2K, because the feed seawater has been preheated much more completely before flowing into the condenser. As a result, a relevant mass flow rate of the humid air as

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mda=1.3kgs-1 also arise, and the maximum water production arrives at mw=235kgh-1. Furthermore, different from the previous conditions involved in the condensing temperature, the heat transfer rate between the organic working fluid and the seawater stream keeps the same at Qc=171kW, although the inlet

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temperature of the seawater varies. Therefore, the total efficiency of the combined system will be determined by the

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characteristics of the HDH desalination unit. Due to a much better energy utilization at TTDr=2K, the top total efficiency of the combined system reaches ηt=100%, an improvement of 19% compared to the case of TTDr=8K. C. Mass flow rate distribution between the streams of feed seawater As stated previously, the humidification and dehumidification processes with forced air circulation are relatively independent in the current power and water combined system. In combination with the significant impacts from the mass flow rate ratio (MFRR) of the working fluid on the system performance, the mass flow rate ratio between the two streams of feed seawater, in Eq. (23), is designated to analyze the corresponding correlations, 16

ACCEPTED MANUSCRIPT shown in Fig. 9.

MFRR =

msw2 msw1

(23)

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Regardless of the feasibility of the operation conditions, it is evident that the reduction of the mass flow rate ratio within the range of MFRR=0.75 to MFRR=1.25 will contribute to the performance elevation, including the water production and total efficiency. The maximum values at the balance condition of the dehumidifier are raised

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from mw=194kgh-1 and ηt=84% at MFRR=1.25 to mw=223kgh-1 and ηt=95% at MFRR=0.75, respectively. For the HDH desalination system, it was proved that the best performance always arises at the balance condition of the heat

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and mass transfer processes, including the humidification and dehumidification. As a result, it is useful to obtain a simultaneous balance condition both for the humidification and dehumidification, through adjusting the internal operation parameters. It can be observed that the balance conditions approach each other, with the declination of the mass flow rate ratio between the two streams of feed seawater, with the case of HCRd=1 at mda=0.79kgs-1 and

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HCRh=1 at mda=0.72kgs-1 at MFRR=0.75. However, the reality of the operation conditions must be taken into account. Unfortunately, the balance condition of the dehumidifier locates in the dashed region if the mass flow rate

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ratio of the feed seawater is reduced to MFRR=0.75. As a result, the actual improvement amplitude of the water production is compressed to 7.98% from mw=194kgh-1 at MFRR=1.25 to mw=210kgh-1 at MFRR=0.75,and the

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actual top value for the total efficiency arrives at ηt=90%. In the light of the correlations between the mass flow rate ratio and the system performance, it can be further calculated that the balance points will overlay at mda=0.74kgs-1, when the mass flow rate ratio is fixed at MFRR=0.71, and the corresponding water production and total efficiency appear at mw=225kgh-1 and ηt=95%. In combination with the obtained entropy generation rate, the true maximum values are obtained at mw=196kgh-1 and ηt=84%. 5.4 Comparison with the previous systems 17

ACCEPTED MANUSCRIPT In order to illustrate the advantages of the combined plant, the top performance comparison between the current combined system and previous ones [25-27] in Table 7, including the production, thermal efficiency and relevant cost of the heat transfer areas, is completed. It can be found that the combination scheme for the ORC and

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air-heated HDH (AH-HDH) desalination unit has maximum values of the production, with a value of 14.9kW for the power and 382kgh-1 for the produced water. Moreover, the corresponding thermal efficiency also arrives at a peak value of 184%.

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It can be concluded that the advantages of the current system are not found compared to the desalination

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system with AH-HDH configurations, in view of the thermodynamic analysis. As a result, total cost of the heat and mass transfer areas are calculated according to the given mathematical models [34, 35], with the geometric dimensions and materials [36]. Unit area of production (UAP) and unit cost of production (UCP), with water production and power generation both into consideration, are defined to characterize the unit consumption both for

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the area and cost aspects, respectively. Thereinto, the power generation is converted into the equivalent heat load based on the efficiency of ηe=40% [30], from thermal energy to electricity.

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UAP =

(24)

Ct mwhfg + Wnet / ηe

(25)

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UCP =

At mwh fg + Wnet / ηe

As demonstrated [25, 26], the combined system with HDH unit as the top cycle will increase the system complexity, with an extra reheat process for the brine discharged from the humidification. Accordingly, poor economic performance of the reheat prototypes is obtained. With respect to the combined configurations with ORC and AH-HDH desalination system, large heat transfer areas of the condenser and dehumidifier were obtained as 65.4m2 and 73.6m2 due to the participation of air, which are much larger than those of the current system with 6.12m2 and 27.4m2. It can be found that the unit area and cost of production has the lowest values for the current 18

ACCEPTED MANUSCRIPT combined system, with UAP=0.34 m2kW-1 and UCP=17.1€kW-1. Finally, according to the economic comparison among the four systems, the economic advantages of the current system is proved. 5.5 Economic performance of the power and water combined system

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In addition of the previous comparison focusing on the cost for the heat and mass transfer devices, the detailed economic performance of the entire combined system are also calculated. In combination with the areas of the heat and mass transfer devices (HMTD), the cost of the HMTD both in ORC and HDH subsystems, power machinery,

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including the blower, pumps and turbine, are calculated based on the mathematical models [11, 37, 38], shown in

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Table 8. Accumulating the cost for all the devices, the total capital of the combined system approaches 49934€. It is obvious that the cost of the ORC turbine has a maximum weight within all the components, with a value of 53.2%. Furthermore, after the acquisition of the fixed investment for the combined system, the detailed annual cost (AC), based on the values of first annual cost (FAC), annual maintenance cost (AMC) and annual salvage value

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(ASV) [39]. Furthermore, the water and electricity are produced simultaneously, and the cost of water and electricity can be obtained, when the share of revenues (SOF) is fixed. Therefore, the annual cost is calculated at SOF=1:1 (electricity vs water) and an operation time of 7200hy-1 for 10y, and then the cost of production can be

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obtained as 0.0032€L-1 for water and 0.063€kW-1h-1 for the electricity, presented in Table 9. Furthermore, the SOF,

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operation hours per year and lifetime, n, are prescribed to achieve the parametric analysis, to explore the relevant influence laws, shown in Table 10. It is found that higher values for SOF, operation time and lifetime are effective to reduce the production cost. The lowest cost of the production reach 0.0025€L-1 for water and 0.05€kW-1h-1 for the electricity, when the design lifetime extends to 20 years after its construction. 6. Conclusions and future work This paper focuses on a combined system to provide power and water. The thermodynamic and economic performance of the combined system at the given conditions is simulated. After the simulation and analysis, the 19

ACCEPTED MANUSCRIPT main conclusions drawn from the investigation are concluded as follows: 1)

Among the three appointed types of organic working fluids, it is proved a maximum value, 13.1kW, for the net output power is obtained, when isobutane is applied as the cycling fluid. At the prescribed range of the air mass flow rate, regions with negative values for the specific entropy

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2)

generation rate will arise. A maximum value, 208kgh-1, for the water production is gained at the balance condition of the dehumidifier, which is available due to the positive entropy generation for all the

A lower condensing temperature is effective to raise the output power and water production, and the

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3)

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components.

corresponding values arrive at 17.0kW and 212kgh-1, respectively, when the condensing temperature is reduced to 323K. At the aspect of the total efficiency, a maximum value of 94% is obtained, when the condensing temperature is fixed at 343K, because the relevant energy input within the HRSG is the

4)

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smallest.

Reducing the value of TTDr can elevate the water production as well as the total efficiency of the combined system. An elevation amplitude of 26.1% for the water production and 19% for energy

Within a reasonable range, adjusting the mass flow rate ratio between the two streams of feed seawater is

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5)

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utilization, can be achieved, respectively, once the TTDr is reduced from 8K to 2K.

an effective way to improve the performance of the combined system. Confined by the feasibility of the operation conditions, the actual elevation amplitude of the water production is 7.98%, when the mass flow rate ratio declines from 1.25 to 0.75, and the actual top efficiency reaches 90%. 6)

Based on the top performance comparison between the current and previous systems, the economic advantages of the current system are proved, with the values of UAP and UCP as 0.34m2kW-1 and 17.1€kW-1, respectively. 20

ACCEPTED MANUSCRIPT 7)

Taking the cost for all the components into account, the total fixed investment of the combined system approaches 49934€. Based on the annual cost calculation at SOF=1:1 (electricity vs water) and an operation time of 7200hy-1 for 10y, the cost of production approach 0.0032€L-1 for water and 0.063€kW-1h-1 for the

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electricity. Finally, after the parametric analysis, it is found that the production cost can be compressed to 0.0025€L-1 for water and 0.05€kW-1h-1 for the electricity, once the lifetime is appointed at 20 years during the design.

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The proposed configurations, with the ORC and HDH desalination subsystem coupled, is effective to provide

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water and electricity simultaneously. Based on the independent thermodynamic and economic analysis, it is proved that the proposed combined system is applicable for the seaside or brackish water areas, with various heat source, including the waste heat or renewable energy. However, in the future research, the multi-objective optimization should be completed to determine the best design and operation conditions.

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Acknowledge

The authors gratefully acknowledge the financial support by the National Natural Science Foundation of China (Grant No. 51406081), Postdoctoral Research Foundation of Jiangsu Province (Grant No. 1701144B), Hong Kong

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Reference

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Scholars Program (Grant No. XJ2017040) and The Hong Kong Polytechnic University.

[1] H. Wang, H.T. Wang. Power generation with low great waste heat. Beijing, Science Press, 2010. [2] J.F. Wang, Z.Q. Yan, M. Wang, S.L. Ma, Y.P. Dai. Thermodynamic analysis and optimization of an (organic Rankine cycle) ORC using low grade heat source. Energy, 49 (2013): 356-365. [3] O.A. Oyewunmi, C.J.W. Kirmse, A.M. Pantaleo, C.N. Markides. Performance of working-fluid mixtures in ORC-CHP systems for different heat-demand segments and heat-recovery temperature levels. Energy Conversion and Management, 148 (2017): 1508-1524. 21

ACCEPTED MANUSCRIPT [4] R. Bahrampoury, A. Behbahaninia. Thermodynamic optimization and thermo-economic analysis of four double pressure Kalina cycles driven from Kalina cycle system 11. Energy Conversion and Management, 152 (2017): 110-123.

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[5] J.F. Hinze, G.F. Nellis, M.H. Anderson. Cost comparison of printed circuit heat exchanger to low cost periodic flow regenerator for use as recuperator in a s-CO2 Brayton cycle. Applied Energy, 208 (2017): 1150-1161. [6] X.R. Wang, Y.P. Dai. Exergo-economic analysis of utilizing the transcritical CO2 cycle and the ORC for a

SC

recompression supercritical CO2 cycle waste heat recovery: A comparative study. Applied Energy, 170 (2016):

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193-207.

[7] C. Wieland, D. Meinel, S. Eyerer, H. spliethoff. Innovative CHP concept for ORC and its benefit compared to conventional concepts. Applied Energy, 183 (2016): 478-690.

[8] M. Sadeghi, A. Nemati, A. Ghavimi, M. Yari. Thermodynamic analysis and multi-objective optimization of

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various ORC (organic Rankine cycle) configurations using zeotropic mixtures. Energy, 109 (2016): 791-802. [9] A.M.A. Jubori, R. Al-Dadah, S. Mahmoud. New performance maps for selecting suitable small-scale turbine

931-946.

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configuration for low-power organic Rankine cycle applications. Journal of Cleaner Production, 161 (2017):

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[10] S. Lecompte, H. Huisseune, M.V.D. Broek, B. Vanslambrouck, M.D. Paepe. Review of organic Rankine cycle (ORC) architectures for waste heat recovery. Renewable and Sustainable Energy Reviews, 47 (2015): 448-461.

[11] M.T. White, O.A. Oyewunmi, M.A. Chatzopoulou, A.M. Pantaleo, A.J. Haslam, C.N. Markides. Computer-aided working fluid design, thermodynamic optimization and thermo-economic assessment of ORC systems for waste-heat recovery. Energy, 161 (2018): 1181-1198.

22

ACCEPTED MANUSCRIPT [12] M.T. White, O.A. Oyewunmi, A.J. Haslam, C.N. Markides. Industrial waste-heat recovery through integrated computer-aided working-fluid and ORC system optimization using SAFT-γ Mie. Energy Conversion and Management, 150 (2017): 851-869.

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[13] S. Quoilin, S. Declaye, B.F. Tchanche, V. Lemort. Thermo-economic optimization of waste heat recovery Organic Rankine Cycles. Applied Thermal Engineering, 31 (2011): 2885-2893.

[14] J.F. Wang, Z.Q. Yan, M. Wang, M.Q. Li, Y.P. Dai. Multi-objective optimization of an organic Rankine cycle

SC

(ORC) for low grade waste heat recovery using evolutionary algorithm. Energy Conversion and Management,

M AN U

71 (2013):146-158.

[15] M.W. Shahzad, K.C. Ng, K.T., B.B. Saha, W.G. Chun. Multi effect desalination and adsorption desalination (MEDAD): A hybrid desalination method. Applied Thermal Engineering, 72 (2014): 289-297.

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[16] H.F. Zheng. Solar desalination principle and technology. Beijing, Chemical and Industry Press, 2012. [17] P. Sharan, S. Bandyopadhyay. Solar assisted multiple-effect evaporator. Journal of Cleaner Production, 142 (2017): 2340-2351.

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[18] G.H. Zhao, Z.D. Tong. Technology of desalination engineering. Beijing, Chemical and Industry Press, 2011.

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[19] M.M. Farid, S. Parekh, J.R. Selman, S. Al-Hallaj. Solar desalination with a humidification-dehumidification cycle: mathematical modeling of the unit. Desalination, 151 (2002): 153-164. [20] G.P. Narayan, M.H. Sharqawy, E.K. Summers, J.H. Lienhard, S.M. Zubair, M.A. Antar. The potential of solar-driven humidification-dehumidification desalination for small-scale decentralized water production. Renewable and Sustainable Energy Reviews. 14 (2010): 1187-1201. [21] N.K. Nawayseh, M.M. Farid, A.A. Omar, A. Sabirin. Solar desalination based on humidification process-Computer simulation. Energy Conversion and Management. 40 (1999): 1441-1461. 23

ACCEPTED MANUSCRIPT [22] N.K. Nawayseh, M.M. Farid, S. Al-Hallaj,A.R. Al-Timimi. Solar desalination based on humidification process- Evaluating the heat and mass transfer coefficients. Energy Conversion and Management. 40 (1999): 1423-1439.

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[23] B.L.D.O. Campos, A.O.S.D. Costa,E.F.D.C. Junior. Mathematical modeling and sensibility analysis of a solar humidification dehumidification desalination system considering saturated air. Solar Energy, 157 (2017): 321-327.

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desalination system. Desalination, 404 (2017): 35-40.

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[24] T. Rajaseenivasan, K. Srithar. Potential of a dual purpose solar collector on humidification dehumidification

[25] W.F. He, X.K. Zhang, D. Han, L. Gao. Performance analysis of a water-power combined system with air-heated humidification dehumidification process. Energy, 130 (2017): 218-227. [26] W.F. He, L.N. Xu, D. Han, C. Yue, W.H. Pu. Performance investigation of a novel water-power cogeneration

110 (2016): 184-191.

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plant (WPCP) based on humidification dehumidification (HDH) method. Energy Conversion and Management,

[27] W.F. He, W.P. Zhu, D. Han, L. Huang. Performance simulation of a power-water combined plant driven by

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low grade waste heat. Energy Conversion and Management, 145 (2017): 107-116.

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[28] L. Ariyanfar, M. Yari, E. A. Aghdam. Proposal and performance assessment of novel combined ORC and HDD cogeneration systems. Applied Thermal Engineering, 108 (2016): 296-311. [29] M.H. Sharqawy, J.H. Lienhard, S.M. Zubair, Thermophysical properties of seawater: a review of existing correlations and data, Desalination and Water Treatment, 16 (2010): 354-380. [30] W.D. Shen, Z.M. Jiang, J.G. Tong, Thermodynamics. Beijing, Higher Education Press, 1982. [31] S.J. Zhang, H.X. Wang, T. Guo. Performance comparison and parametric optimization of subcritical organic Rankine cycle (ORC) and transcritical power cycle system for low-temperature geothermal power generation. 24

ACCEPTED MANUSCRIPT Applied Energy, 88 (2011): 2740-2754. [32] G.P. Narayan, J.H. Lienhard, S.M. Zubair. Entropy generation minimization of combined heat and mass transfer devices. International Journal of Thermal Sciences, 49 (2010): 2057-2066.

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[33] G.P. Narayan, K.H. Mistry, M.H. Sharqawy, S.M. Zubair, J.H. Lienhard, Energy effectiveness of simultaneous heat and mass exchange devices. Frontiers in Heat Mass Transfer, 023001 (2010): 1-13.

[34] A. Muley, R.M. Manglik, Experimental study of turbulent flow heat transfer and pressure drop in a plate heat

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exchanger with chevron plates, Journal of Heat Transfer, 121 (1999):110-117.

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[35] M. Mehrgoo, M. Amidpour. Constructal design and optimization of a direct contact humidification dehumidification desalination unit. Desalination, 293 (2012): 69-77.

[36] W.F. He, D. Han, W.P. Zhu, C. Ji. Thermo-economic analysis of a water heated humidification dehumidification desalination system with waste heat recovery. Energy Conversion and Management, 160

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(2018): 182-190.

[37] S. Quoilin, S. Declaye, B.F. Tchanche, V. Lemort. Thermo-economic optimization of waste heat recovery Organic Rankine Cycles. Applied Thermal Engineering, 31 (2011): 2885-2893.

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[38] R. Turton, R.C. Bailie, W.B. Whiting, J.A. Shaeiwitz. Analysis, synthesis, and design of chemical processes.

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Upper Saddle River, NJ: Prentice Hall PTR, 2009. [39] Y. Zhang, C.G. Zhu, H. Zhang, W.D. Zheng, S.J. You, Y.H. Zhen. Experimental study of a humidification dehumidification desalination system with heat pump unit. Desalination, 442 (2018): 108-117.

25

ACCEPTED MANUSCRIPT Table1 Thermal parameters of the waste exhaust Term

Unit

Value

Twe

K

423

mwe

kg/s

3

CO2

%

41.2

N2

%

58.8

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x

Table 2 Comparison of net power between the involved ORC and that from Zhang [31] R245fa (kW)

Term

R123 (kW)

345

347

349

351

353

345

347

349

351

353

Current

8.93

8.93

8.93

8.93

8.93

9.42

9.71

9.95

10.3

10.5

Zhang [31]

9.24

9.24

9.24

9.24

9.24

9.17

9.46

9.74

10.0

10.3

error(%)

3.35

3.06

2.87

2.80

2.73

2.65

2.57

2.11

2.34

2.09

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Ttur (K)

Table 3 Performance comparison during humidification between the involved HDH desalination unit and that from Narayan [32] ε

msw (kgs-1)

mda (kgs-1)

Tsw,2 (K)

Ta1 (K)

RH1

RH2

Ta2 (K)

mb (kgs-1)

Tb (K)

Current

0.9

0.15

0.1

336

307

1

1

324

0.14

310

Narayan [32]

0.9

0.15

0.1

336

307

1

1

325

0.14

310

error(%)

/

/

/

/

/

/

/

0.15

0.28

0.04

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Term

Table 4 Performance comparison during dehumidification between the involved HDH desalination unit and that

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from Narayan [32] ε

msw (kgs-1)

mda (kgs-1)

Tsw,0 (K)

Ta2 (K)

RH2

RH1

Ta1 (K)

mw (kgh-1)

Tsw,i (K)

Simulation

0.9

0.15

0.1

303

363

1

1

307

20.2

336

Narayan [32]

0.9

0.15

0.1

303

363

1

1

307

21.2

336

error(%)

/

/

/

/

/

/

/

0.06

5.1

0.1

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Term

Table 5 Detailed thermodynamic parameters of the power and water combined system [27] S (g/kg)

ε

RH1

RH2

T0(K)

Tc(K)

PTDHRSG (K)

PTDc (K)

TTDr(K)

35

0.85

1

0.9

288

333

15

10

5

Table 6 Specific performance of the power and water combined system at the balance conditions Fluid type

mORC (kgs-1)

Tt (K)

ptur (kPa)

Wnet (kW)

mw (kgh-1)

sgen,t (kJkg-1K-1)

ηt (%)

Isobutane

0.52

379

2218

13.1

208

1.76

89.3

26

ACCEPTED MANUSCRIPT R245fa

0.84

376

1351

12.5

199

1.73

89.4

R123

0.85

375

821

11.9

185

1.82

89.1

Table 7 Top performance comparison between the current system and the previous ones

ηt (%) At (m2) C (€) UAP (m2kW-1) UCP×10-2 (€kW-1)

system complexity

Wnet

mw (kgh-1)

He [25]

AH-HDH+ORC

6.04

19.5

95

14.0

873

0.49

31.0

He [26]

WH-HDH+ORC

3.75

103

76

40.7

2671

0.52

34.2

He [27]

ORC+AH-HDH

14.9

382

184

162

6274

0.55

21.4

Current

ORC+WH-HDH

13.1

208

89

58.7

2949

0.34

17.1

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Term

Seawater pump

Blower

Turbine

Refrigerant pump

HMTD (ORC)

HMTD (HDH)

Type

0.75kW40-125A

HG-1100

[11]

[11]

[37]

[37, 38]

Scale (kW or m2)

0.75kW

1.1kW

14.8kW

1.7kW

18.42m2

40.3m2

Price (€)

2×123.8

103.1

5900

10658

26561

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Components

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Table 8 Information of the components within the combined plant for power and water production

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Table 9 Economic performance of the combined plant for power and water production at the design conditions SOF

Operation time (h)

n (y)

i (%)

FAC (€)

5:5

7200

10

12

8838

AMC (€)

ASV (€)

AC (€)

Cw (€L-1)

Ce (€kW-1h-1)

1326

569.1

9594

0.0032

0.063

time and lifetime

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Table 9 Economic performance of the combined plant for power and water production at different SOF, operation

Cw (€L-1)

Ce (€kW-1h-1)

Operation time (h)

Cw (€L-1)

Ce (€kW-1h-1)

n (y)

Cw (€L-1)

Ce (€kW-1h-1)

4:6

0.0038

0.076

7200

0.0032

0.063

10

0.0032

0.063

5:5

0.0032

0.063

4800

0.0048

0.095

15

0.0027

0.054

6:4

0.0026

0.051

2400

0.0096

0.19

20

0.0025

0.05

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SOF

27

ACCEPTED MANUSCRIPT Fig. 1 Detailed configurations of the combined system for power and water production driven by waste exhaust Fig. 2 Temperature profiles along the flow passage of the involved heat exchangers within the power and

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water combined system (a) HRSG (b) Condenser

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(c) Recuperator

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Fig. 3 Procedures during simulating the thermodynamic performance of the power and water combined system

Fig. 4 Net power and thermal efficiency of the ORC subsystem using Isobutane, R245fa and R123 as working fluids with the variation of the evaporation temperature

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Fig. 5 Specific entropy generation rates of the contained components using Isobutane, R245fa and R123 as working fluids

(b) R245fa

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(c) R123

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(a)Isobutane

Fig. 6 Water production and efficiency of the power and water combined system using Isobutane, R245fa and R123 as working fluid

Fig. 7 Water production and efficiency of the power and water combined system at different condensing temperatures of the ORC power subsystem Fig. 8 Water production and efficiency of the power and water combined system at different TTD values of the recuperator 28

ACCEPTED MANUSCRIPT Fig. 9 Water production and efficiency of the power and water combined system at different mass flow rate ratio for the two feed seawater streams

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Fig. 10 Expected levelized costs of the power and desalinated water production for the combined system

29

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Fig. 1 Detailed configurations of the combined system for power and water production driven by waste exhaust

30

ACCEPTED MANUSCRIPT

PTDc

TTDHRSG

PTDHRSG

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TTDc

(b) Condenser

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(a) HRSG

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TTDr

(c) Recuperator

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Fig. 2 Temperature profiles along the flow passage of the involved heat exchangers within the power and water combined system

31

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ACCEPTED MANUSCRIPT

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Fig. 3 Procedures during simulating the thermodynamic performance of the power and water combined system

32

ACCEPTED MANUSCRIPT 14

12

Isobutane, Wnet R123, Wnet

8

10 6 9

Isobutane, ηORC

8

4

R245fa, ηORC R123, ηORC

7

2

6

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11

Wnet(kW)

10

R245fa, Wnet

12

ηORC(%)

13

340

350

360

370

380

390

400

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Tt(K)

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0

5

Fig. 4 Net power and thermal efficiency of the ORC subsystem using Isobutane, R245fa and R123 as working fluids with the variation

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of the evaporation temperature

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ACCEPTED MANUSCRIPT 3.5

3.5

3.0

2.0 1.5 1.0

HRSG Turbine Pump Condenser Humidifier Dehumidifier Mixer Recuperator

2.5

sgen(kJ/kg-1K-1)

2.5

0.5

2.0 1.5 1.0 0.5

0.0

0.0

sgen=0

sgen=0

-0.5

-0.5 0.0

0.5

1.0

1.5

2.0

0.0

-1

1.0

1.5

2.0

-1

mda(kgs )

mda(kgs )

(b) R245fa

SC

(a)Isobutane

3.0

HRSG Turbine Pump Condenser Humidifier Dehumidifier Mixer Recuperator

2.5 2.0 1.5 1.0 0.5 0.0

sgen=0

M AN U

3.5

sgen(kJ/kg-1K-1)

0.5

RI PT

sgen(kJ/kg-1K-1)

3.0

HRSG Turbine Pump Condenser Humidifier Dehumidifier Mixer Recuperator

0.0

0.5

1.0

TE D

-0.5 1.5

-1

2.0

mda(kgs )

(c) R123

AC C

EP

Fig. 5 Specific entropy generation rates of the contained components using Isobutane, R245fa and R123 as working fluids

34

ACCEPTED MANUSCRIPT 240

100

210 80 180

90 60 30

Isobutane, sgen,h>0

Isobutane, sgen,h>0

Isobutane, sgen,h<0

Isobutane, sgen,h<0

R245fa, sgen,h>0

R245fa, sgen,h>0

R245fa, sgen,h<0

R245fa, sgen,h<0

R123, sgen,h>0

R123, sgen,h>0

R123, sgen,h<0

R123, sgen,h<0

40

20

0.0

0.4

0.8

1.2

1.6

-1

2.0

M AN U

mda(kgs )

SC

0

0

RI PT

120

60

ηt

mw

ηt(%)

-1

mw(kgh )

150

AC C

EP

TE D

Fig. 6 Water production and efficiency of the power and water combined system using Isobutane, R245fa and R123 as working fluid

35

ACCEPTED MANUSCRIPT 240

100

210 80

150

60 30 0 0.0

0.4

Tc=50K, sgen,h>0

Tc=50K, sgen,h<0

Tc=50K, sgen,h<0

Tc=60K, sgen,h>0

Tc=60K, sgen,h>0

Tc=60K, sgen,h<0

Tc=60K, sgen,h<0

Tc=70K, sgen,h>0

Tc=70K, sgen,h>0

Tc=70K, sgen,h<0

Tc=70K, sgen,h<0

0.8

1.2

40

20

0

1.6

-1

2.0

M AN U

mda(kgs )

RI PT

90

Tc=50K, sgen,h>0

SC

120

60

ηt

mw

ηt(%)

-1

mw(kgh )

180

Fig. 7 Water production and efficiency of the power and water combined system at different condensing temperatures of the ORC

AC C

EP

TE D

power subsystem

36

ACCEPTED MANUSCRIPT 240

100

210 80

90 60 30

60

TTDr=2K, sgen,h>0

TTDr=2K, sgen,h<0

TTDr=2K, sgen,h<0

TTDr=5K, sgen,h>0

TTDr=5K, sgen,h>0

TTDr=5K, sgen,h<0

TTDr=5K, sgen,h<0

TTDr=8K, sgen,h>0

TTDr=8K, sgen,h>0

TTDr=8K, sgen,h<0

TTDr=8K, sgen,h<0

20

0

0.0

0.4

0.8

1.2

1.6

-1

M AN U

mda(kgs )

2.0

SC

0

40

RI PT

120

ηt

mw TTDr=2K, sgen,h>0

ηt(%)

150

-1

mw(kgh )

180

AC C

EP

TE D

Fig. 8 Water production and efficiency of the power and water combined system at different TTD values of the recuperator

37

ACCEPTED MANUSCRIPT 240

100

210 80

ηt

120 90 60 30

60

MFRR=0.75, sgen,h>0

MFRR=0.75, sgen,h>0

MFRR=0.75, sgen,h<0

MFRR=0.75, sgen,h<0

MFRR=1, sgen,h>0

MFRR=1, sgen,h>0

MFRR=1, sgen,h<0

MFRR=1, sgen,h<0

MFRR=1.25, sgen,h>0

MFRR=1.25, sgen,h>0

MFRR=1.25, sgen,h<0

MFRR=1.25, sgen,h<0

20

0

0.0

0.4

0.8

1.2

1.6

-1

M AN U

mda(kgs )

2.0

SC

0

40

RI PT

mw

ηt(%)

150

-1

mw(kgh )

180

Fig. 9 Water production and efficiency of the power and water combined system at different mass flow rate ratio for the two feed

AC C

EP

TE D

seawater streams

38

ACCEPTED MANUSCRIPT 1. A power and water combined system, with ORC and HDH units coupled, is proposed. 2. Energetic and entropic analysis of the proposed system are achieved. 3. Influence laws from the critical parameters on the system performance are revealed.

AC C

EP

TE D

M AN U

SC

RI PT

4. Economic performance of the combined system is also focused.