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Energy Procedia 142 Energy Procedia 00(2017) (2017)1119–1126 000–000 www.elsevier.com/locate/procedia

9th International Conference on Applied Energy, ICAE2017, 21-24 August 2017, Cardiff, UK

Clarifying the bifurcation point on Design: A Comparative Analysis The 15th International Symposium on District Heating and Cooling between Solar-ORC and ORC-based Solar-CCHP feasibility ofJiaxin using heatMa demand-outdoor a a YingAssessing Zhanga, Shuaithe Deng , Li Zhaoa,*, Niathe , Minglu , Shan Lina, Zhengtao temperature function for a long-term Zhangadistrict heat demand forecast a aKey

Key Laboratory of Efficient Utilization of Low and Ministry of Education of China, Tianjin, a,b,c a Medium Gradea Energy (Tianjin University), b c c 300072, China

I. Andrić

a

*, A. Pina , P. Ferrão , J. Fournier ., B. Lacarrière , O. Le Corre

IN+ Center for Innovation, Technology and Policy Research - Instituto Superior Técnico, Av. Rovisco Pais 1, 1049-001 Lisbon, Portugal b Veolia Recherche & Innovation, 291 Avenue Dreyfous Daniel, 78520 Limay, France c Département Systèmes Énergétiques et Environnement - IMT Atlantique, 4 rue Alfred Kastler, 44300 Nantes, France

Abstract

An obvious knowledge gap exists in the design of ORC-based solar-CCHP, due to an abuse and indiscreet application of the existing research methodology specifically developed for solar-ORC. Such extended research methodology, which emphasizes onAbstract the promotion of power generation performance, does not provide a reasonable design result, because of a significant application difference between the ORC-based solar-CCHP and solar-ORC. Although the two systems have some similar technologies elements, a bifurcation pointaddressed on designin has been as clarified leadeffective to the lack of clear methodology District heating networks are commonly the not literature one ofwhich the most solutions for decreasing the bifurcation point, this focuses on the ORC-based framework or profile in this research field. To fill the gap and clarify the paper greenhouse gas emissions from the building sector. These systems require high investments which are returned through the heat solar-CCHP the energy utilization principle. Considered the medium-temperature working sales. Due system to the with changed climatecascade conditions and building renovation policies, heat demand in the futurefluids, couldincluding decrease, Benzene, Toluene, Decane, D4, MM, the whole system model is established which consists of a parabolic trough collector, ORC, prolonging the investment return period. heat exchanger for heating system, and a single-effect absorption chiller. An assessment framework for the system energy The main scope of this paper is to assess the feasibility of using the heat demand – outdoor temperature function for heat demand solar-CCHP conducted.ofThe efficiency proposed. And then a comparison study between the solar-ORC ORC-based forecast. is The district of Alvalade, located in Lisbon (Portugal), was used asand a case study. The district isis consisted 665 results showthat thatvary the configuration characteristics an important effect on thescenarios system energy for theand CCHP buildings in both construction period have and typology. Three weather (low, efficiency medium, high) threesystem, district which shouldscenarios be taken were into account prior to the performance the solar-ORC system. Specifically, the demand solar-ORC system, renovation developed (shallow, intermediate,ofdeep). To estimate the error, obtainedforheat values were the system energy efficiency increases as the increasing of the evaporating temperature. However, for the ORC-based solarcompared with results from a dynamic heat demand model, previously developed and validated by the authors. CCHP system, the behaviors of the thermal performance, such the overall efficiency of solar energyfor aresome totally changed The results showed that when onlysystem weather change is considered, theasmargin of error could be acceptable applications compared the solar-ORC system. optimal scenarios configuration parameters should after be a introducing tradeoff between the (the errorwith in annual demand was lowerConsequently, than 20% for the all weather considered). However, renovation thermal performance, size and initial investment cost. Finally, the optimum configuration design for the ORC-based solar-CCHP scenarios, the error value increased up to 59.5% (depending on the weather and renovation scenarios combination considered). system with of MM is achieved, where the overall efficiency of solar is 40.95%, theper cooling-to-power ratio of 4.95toand The value slope coefficient increased on average within the energy range of 3.8% up with to 8% decade, that corresponds the heat-to-power ratio of 4.2. decrease in the number of heating hours of 22-139h during the heating season (depending on the combination of weather and ©renovation 2017 The Authors. Published by Elsevier Ltd. hand, function intercept increased for 7.8-12.7% per decade (depending on the scenariosPublished considered). On the other © 2017 The Authors. by Elsevier Ltd. committee Peer-review under responsibility of the scientific of the 9th International Conference on Applied Energy. coupled scenarios). The values of suggested couldcommittee be used to modify the function Conference parameters on forApplied the scenarios Peer-review under responsibility the scientific of the 9th International Energy.considered, and improve the accuracy of heat demand estimations. © 2017 The Authors. Published by Elsevier Ltd. Peer-review under responsibility of the Scientific Committee of The 15th International Symposium on District Heating and * Corresponding author. Tel.: +0-000-000-0000 ; fax: +0-000-000-0000 . Cooling. E-mail address: [email protected]

Keywords: Heat demand; Forecast; Climate change 1876-6102 © 2017 The Authors. Published by Elsevier Ltd. Peer-review under responsibility of the scientific committee of the 9th International Conference on Applied Energy.

1876-6102 © 2017 The Authors. Published by Elsevier Ltd. Peer-review under responsibility of the Scientific Committee of The 15th International Symposium on District Heating and Cooling.

1876-6102 © 2017 The Authors. Published by Elsevier Ltd. Peer-review under responsibility of the scientific committee of the 9th International Conference on Applied Energy. 10.1016/j.egypro.2017.12.365

Ying Zhang et al. / Energy Procedia 142 (2017) 1119–1126 Author name / Energy Procedia 00 (2017) 000–000

1120 2

Keywords: distributed energy system; CCHP; working fluid; ORC; PTC

1. Introduction Distributed energy system (DES) plays a vital role in solving the contradiction between the energy crisis and environmental protection. It is known as the on-site energy systems with energy cycling technologies, realizing the conversion of low-grade thermal energy into electricity [1]. As one of the important technology, polygeneration systems have been put forward and developed rapidly in the last decades, which allows to supply multi-objectives energy products simultaneously [2]. More recently, the polygeneration systems driven by renewable energy, such as solar thermal energy has become a new challenging direction for DES research [3]. As one of prime movers for the polygeneration systems, organic Rankine cycle (ORC) represents an effective technology than internal combustion engines and Stirling engines to fulfil heat and electricity needs [2]. AlSulaiman proposed a combined cooling, heating and power (CCHP) system using parabolic trough solar collectors and an ORC, with a single-effect of absorption chiller (ACH) and a heating system by heat exchanger [4]. It was pointed out that this system has a promising prospect with a maximum exergy efficiency of 20% [5]. Freeman investigated the solar-powered ORC system for combined heating and power (CHP) in UK domestic applications, of which the hot water obtained heat from the condenser or solar collector in a storage cylinder [6]. Then the configuration of the CHP system was optimized in terms of two aspects, including the solar collector cycle optimization and working fluid selection [7]. Besides, Habka evaluated effects of the heating plant system on the ORC-CHP through series and parallel connection, which is driven by low-temperature and –caloric geothermal water [8]. It was suggested that the system performs better under the parallel configuration with mixtures as the working fluids, in terms of power productivity and heat source utilization in the given working conditions. Table 1 Interaction and restriction between the solar-ORC system and ORC-based solar-CCHP system. Outputs Performance criteria

Mutual restriction

Solar-ORC

ORC-based solar-CCHP

Power generation Collector thermal efficiency Electricity efficiency Net output of electric power High collector thermal efficiency leads to low electricity efficiency, as well as low net output of electric power

Cooling, heating and power Collector thermal efficiency Energetic efficiency Output of cooling, heating and power Condensing temperature of ORC and supply temperature of heating system Condensing temperature of ORC and supply temperature of cooling system

However, there is obvious interaction and restriction between the solar-ORC system and ORC-based solar-CCHP system, as shown in Table 1. And the goal of high overall efficiency of solar energy, exact and optimal configuration parameters in terms of the whole system instead of the solar-ORC subsystem has remained elusive in this field. Focused on this target, this paper comparatively analyzed the configuration characteristics and performance on an ORC-based solar-CCHP with parabolic trough collector (PTC), as well as on a solar-ORC system, over a wide range of evaporating temperature. The demonstration system is under construction and located in Binhai New District, Tianjin, China, with the capacity of the ORC power system of 200kW [9]. The system description and modeling is described in Section 2. Section 3 compared the performance under the assessment framework. Finally, Section 4 concludes the character of energy efficiency for the solar-CCHP system. Nomenclature ACH CHP CCHP DES ORC

absorption chiller combined heating and power combined cooling, heating and power distributed energy system organic Rankine cycle



PTC

Yingname Zhang et al. / Procedia Energy Procedia 142 (2017) 1119–1126 Author / Energy 00 (2017) 000–000

1121 3

parabolic trough collector

2. System description and modeling 2.1. System description The schematic and state diagram of the system are dipecited in Fig. 1. The diathermic oil flows through vapor generator and recirculates in the PTC for heating by the oil pump. The ORC is chosen as a prime mover for the CCHP system, and its thermodynamic process is as follows. The working fluid (state 4) is pumped to be heated in the vapor generator; then the superheated gas (state 6) flows into the turbine; the exhaust gas (state 7) enters the heater to provide heat energy to users and is condensed into saturated vapor (state 8); then it condensed in the condenser and is pumped to continue the next cycle. Furthermore, the heat, which is gained over the condenser, will be applied to drive the single-effect absorption chiller (ACH) LiBr–H2O, providing cooling to users. Since the heat rejection temperature should be sufficiently high to drive the cooling and heating sub-system, totally five kinds of medium-temperature working fluid are researched in the paper: Benzene, Toluene, Decane, octamethylcyclotetrasiloxane (D4), hexamethyldisiloxane (MM). Meanwhile, the condensation condition is decided by the requirement of heat source temperature for the ACH of 95/75oC, while for the heating system of 95/70 oC in this paper. Besides, Syltherm 800 is chosen as heat transfer fluid in the PTC absorber tube due to the temperature application range. As a result, the configuration of the PTC-CCHP system is cascaded in series by coupling the ORC, cooling and heating cycle, which is defined as the sequential polygeneration system in this paper. The feature is the simultaneous production of power, cooling and heating from a single solar source based on the energy cascade utilization principle as shown in Fig. 1(b). 6

2

PTC field

Gen

VG

1 P_Oil

V1

7

10

5

3

P_ORC

T

Exp

Tout_oil

9

20

19 RHE

21 22 23

18 17 Abs 16 25

V4

Cond

Eva

͸ ͹

P_Water2

13

Cooling Water



V2

P_Water1 Gen

Tin_oil , moil Tpin

Tpin

V3 Fan coil

HE1

Cond_ORC 8 4 11 12

Irreversibility

24

ͷ Ͷ

Tin_sh_ACH

ͺ Tpin T

out_sh_ACH

Tout_sh_HS

Tin_sh_HS

14

Tamb

15 26

(a)

S

(b)

Fig. 1 The Sequential ORC-based solar-CCHP system: (a) Schematic of the system; (b) The symbolic T-S diagram of the system.

2.2. Mathematical models and performance criteria In order to reveal the mutual influence between the coupled cycle, the mathematical models of the main cycles are built, including the PTC model and CCHP model as follows. The model of PTC chosen for this study is the LS-3, one of the mature collector within the family of commercial PTCs, which allows the maximum temperature of 400 o C[10]. The expression for the collector thermal efficiency is defined by the ratio of the thermal power obtained by the fluid in the absorber ( QSTC-fluid ) to the solar power crossing the aperture area of the solar collector ( Gb  Aa ). Meanwhile, it can be tested as a function of the arithmetic mean of the diathermic oil temperature ( T abs ) at the inlet ( Tin ) and outlet ( Tout ) of the collector, the direct solar irradiance ( Gb ), and the incident angle of the direct solar irradiation according to the LS-3 model, as shown in Eq. (1) [10]:

Ying Zhang et al. / Energy Procedia 142 (2017) 1119–1126 Author name / Energy Procedia 00 (2017) 000–000

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STC =

QSTC-fluid U abs (T abs  Tabm ) =opt, 0  K ( )  Fe  L  Gb  Aa Cg Gb

(1)

where opt, 0 is the collector’s optical efficiency for a zero incident angle; Fe is the dirt degree of collector’s mirrors; C g is the geometric concentration ratio; and Tabm is the ambient temperature. Besides, K ( ) is the incident angle modifier [11]; U Labs is the thermal loss coefficient per unit area of the absorber tube [10]. To simplify the simulation of the system, an ideal ORC cycle is considered, which consists of four processes: isentropic compression (process: 4-5), isobaric evaporation (process: 5-6), isentropic expansion (process: 6-7), and isobaric condensation (process: 7-8-4). The ORC pump and expander are characterized by isentropic efficiencies, as well as the boundary conditions and other fixed parameters are shown in Table 2. The thermodynamic properties of the working fluid are acquired by linking MATLAB and the commercial software REFPROP 9.1. According to gross mechanical power output ( Wexp  200 ), the mass flow rate of the working fluid mwf is decided. The power consumed by the pump Wp, ORC is (process 5-6) obtained by Eq. (2). Besides, the surplus heat applied to drive the heating (process 7-8) and cooling (process 8-4) system is calculated by Eq. (3) and Eq. (4), respectively.

W p, ORC

1

p, ORC

(2)

mwf  (h5s  h4 )

(3)

Qheat  mwf  (h7  h8 )

(4)

Qcooling  mwf  (h8  h4 ) Table 2 Fixed parameters in the PTC-CCHP Item

Value

Gross mechanical power output, Wexp (kW)

200

Expander isentropic efficiency, ηexp (%)

0.75

Isentropic efficiency of the ORC pump, ηp_ORC (%)

0.75

Isentropic efficiency of the oil pump, ηo (%)

0.75

The inlet and outlet pressure of the oil pump, Pin/Pout (kPa)

100/600

The pinch point temperature difference, △Tmin (K)

10

Maximum evaporation temperature, Tev_max (ºC)

Tc – 10

Minimum evaporation temperature, Tev_min (ºC)

180

Minimum condensation pressure, Patm (kPa)

101

Minimum condensation temperature for the system in seriesa, Tcond_min (ºC)

105

Superheating temperature, △Tex (K)

10

If Tsat (Patm) > 105℃, then Tcond = Tsat (Patm) , for the medium-temperature working fluids to satisfy both the sufficient heat rejection temperature and atmospheric operation of the condensor. a

The system performance is evaluated from three aspects, including the thermodynamic performance, system size, as well as initial investment cost. For the thermodynamic performance, power consumption of the solar field circulating pump and the working fluid pump are considered. From the viewpoint of thermodynamics, the overall efficiency of solar energy is proposed, which represents the comprehensive utilization ratio in the multi-production process for the ORC-based solar-CCHP system:



Yingname Zhang et al. / Procedia Energy Procedia 142 (2017) 1119–1126 Author / Energy 00 (2017) 000–000

PRCCHP 

1123 5

Wexp  Wp,ORC  moil  ( Pout  Pin ) / oil p, oil   Qheat   Qcooling

(5)

Gb  Aa

where moil is the mass flow rate of the diathermic oil; Pin and Pout are the inlet and outlet pressure of the oil pump

respectively (set in Table 2); oil is the density of the diathermic oil; p, oil is the efficiency of the diathermic oil

 and  vary with the price fluctuation of cooling, heating and power. It is   0.8 [12]. Remarkably, the solar power crossing the aperture area of the solar

pump (set in Table 2). Besides, assumed that   0.5 and

collector ( Gb  Aa ) are adopted as the input energy to the whole plant. Table 3 Heat transfer coefficients for the different thermodynamic state. Item

Correlation

Tube side heat transfer coefficient (single phase oil side)

 htube, oil

Tube side heat transfer coefficient (single phase water side) Shell side heat transfer coefficient (subcooled and superheated regions) Shell side heat transfer coefficient (two-phase region during evaporating) Shell side heat transfer coefficient (two-phase region during condensing)

 d  f / 8  Re1000 Pr  112.7 t



i





htube, water  4200  1.35  0.02T water  ut 0.8



 

f 8  Pr 2/3  1  1   di l 

2/3



di 0.2

htube, water  0.36  s de  Re0.55  Pr1/3   s sw 

0.14

(Re=2000-106)

hshell, eva h0  F pf   q q0    Rp Rp 0 

0.133

nf



hshell, cond  0.95   l  l   l  g   g l 



1/3

Table 4 System costs model. Component

Dependent variable

Cost($)

PTC[6]

Aperture area,

Heat exchangers[17]

Heat exchange area,

Expander[17]

Volume flow rate,

Working fluid pump[17]

Electric power,

Wp, wf

HTF pump[17]

Electric power,

Wp, oil (W)

LiBr absorption chiller [18]

Capacity,

A a (m2) A HE (m2)

Vwf (m3/s) (W)

Wcooling (kW)

1280  A a 200  327  A HE 1.58   225  170 Vwf  949.4  Wp, wf / 300 

0.25

527.45  Wp, oil / 300

0.25

316.47 Wcooling

For the system size, the PTC aperture area, heat transfer area and the expander size are all sensitive to the thermodynamic parameters. Considered two pinch point △Tpin =10K occurred at point A and state 7, the required PTC area Aa can be calculated according to Eq. (1), where the collector thermal efficiency STC , as well as

QSTC-fluid , is determined via the energy balance in the vapor generator in Fig. 1.The shell and tube heat exchangers are chosen for all the heat exchangers in this paper according to the engineering actual situation [9], and the prototype as well as their geometric dimensions can refer to [13]. The working fluid circulates on the shell side of the heat exchanger, and the diathermic oil as well as water circulates in the tube side. The appropriate heat transfer area is obtained based on the existing researches for the heat transfer correlations in [14, 15], as shown in Table 3. In addition, the expender size (SP) is evacuated  by SP Vwf 4 Hexp [16], where Vwf is the volume flow rate of

Ying Zhang et al. / Energy Procedia 142 (2017) 1119–1126 Author name / Energy Procedia 00 (2017) 000–000

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working fluid, which is the quotient of the mass flow rate of the working fluid and the density of the working fluid at the inlet of the expander; H exp is the enthalpy drop in the expander. This parameter characterizes the dimension of the expander or turbine via a proportional relationship. Thus the greater SP value corresponds to the bigger expander. For the initial investment cost, the equipment cost is calculated individually by the method in formal researches (Table 4). It should be noticed that the indirect project expense, including material, such as piping, auxiliary equipment, installation and commissioning, and labor cost, are not taken into account due to the uncertainty. 3. Results and discussion 3.1. PTC-ORC sub-system performance

ηel_D4

0.06

ηel_Benzene

0.05

12000

0.65

10000

0.60

ηel_Decane ηPTC_Toluene 0.55

0.04

ηPTC_MM

0.03

ηPTC_D4

0.02

ηPTC_Benzene ηPTC_Decane

0.01 0.00

0.70

0.50 0.45

(a)

Aa_D4

Aa_Benzene

8000

Aa_Decane

6000 4000

200

220 240 260 280 300 Evaporating temperature (oC)

SP_Toluene SP_MM

0.16

SP_Benzene

SP_D4

SP_Decane

0.12 0.08 0.04 200

AHE_MM

270

AHE_Benzene

AHE_D4

AHE_Decane

240 210 180 150

320

180

200

(b)

0.20

180

AHE_Toluene

300

120

180

0.24

Expander size (m)

Aa_MM

2000

0.40 180 200 220 240 260 280 300 320 340 360 380 400 Evaporating temperature (oC)

330

Aa_Toluene Heat exchanger area (m2)

ηel_MM

220 240 260 280 300 Evaporating temperature (oC)

320

Initial investment cost of PTC-ORC ($)

ηel_Toluene

0.07

PTC aperuture area (m2)

0.08

Efficiency of PTC (%)

Efficiency of the net electrical power (%)

Fig. 2 shows the thermal efficiency of the net electrical power, the system size, as well as the initial investment cost, for the different working fluids considered herein and over a range of evaporation temperature. It can be observed that the system energy efficiency increases as the increasing of the evaporating temperature. The maximum thermal efficiency of the net electrical power ηel is 7.5% for Benzene, and is achieved at a maximum evaporation temperature of 278 oC (Fig. 2(a)). It corresponds to the minimum PTC area of 2686 m2, HE area of 117 m2, SP of 0.038m, as well as initial investment cost of 5.53*105$. However, the PTC efficiency itself is affected by the temperatures in the vapor generator, which is declined with an opposite trend to that of the ηel (Fig. 2(a)). The reason is attributed to the increasing of the evaporation temperature which leads to higher collector ambient heat losses.

220 240 260 280 300 Evaporating temperature (oC)

320

(c)

2.4x106

Cel_Toluene

2.1x106

Cel_MM Cel_D4

1.8x106

Cel_Benzene Cel_Decane

6

1.5x10

1.2x106 9.0x105 6.0x105 3.0x105

(d)

180

200

220 240 260 280 300 Evaperating temperature (oC)

320

(e)

Fig. 2 System size and initial investment cost: (a) The efficiency of net electrical power and PTC efficiency; (b) PTC area requirement; (c) heat transfer area requirement; (d) the expander size requirement; (e) Initial investment cost of the PTC-ORC sub-system.

3.2. PTC-CCHP system performance in sequential configuration Fig. 3 shows the behavior of these fluids in the ORC-based solar-CCHP system. The varying regularity and magnitude order of the three indicators are totally changed with that of the solar-ORC sub-system. As the evaporation temperature increases, the total net power output, overall efficiency of solar energy and total initial investment cost decrease simultaneously (Fig. 3(a)-(c)). The essential reasons behind that are as follows. On the one



Ying Zhang et al. / Energy Procedia 142 (2017) 1119–1126 Author name / Energy Procedia 00 (2017) 000–000

1125 7

4000

Wnet_Toluene

3500

Wnet_MM

3000

Wnet_D4

2500

Wnet_Decane

Wnet_Benzene

2000 1500 1000 180

200

220 240 260 280 300 Evaporating temperature (oC)

320

(a)

Overall efficiency of solar energy (%)

Total net power output in CCHP mode (kW)

hand, the thermal properties of different working fluids lead to different heat rejection over the condenser. On the other hand, the configuration characteristic of the high integration within the system has a significant impact on the performance and capacity ratio. Furthermore, the amount of the cooling and heating output is larger than the power output though the grade of primary energy for the cooling and heating are different from that for the power. The results indicate that the Benzene and Toluene perform well. However, the MM is suggested to be the best candidate with the advantages of low toxicity and flammability, as well as good thermal stability, in practical applications [19]. Fig. 3(d) shows the Pareto frontier solutions for the sequential polygeneration system of the ORC-based solar-CCHP, which reveals the conflicts between the overall efficiency of solar energy, initial investment cost and system size. The design point C implies the optimum solution of multi-objective function where the system achieves the best possible values considering the above three performance indicators. Under this condition of the optimum solution, the overall efficiency of the solar energy is 40.95%, with the cooling-to-power ratio of 4.95 and heat-to-power ratio of 4.20. 0.48

PRCCHP_Toluene PRCCHP_MM

0.44

PRCCHP_D4

PRCCHP_Benzene

0.40

PRCCHP_Decane

0.36 0.32 0.28 180

200

220 240 260 280 300 Evaporating temperature (oC)

320

(b)

Total initial investment cost ($)

3.2x106 CCCHP_Toluene

2.8x106

CCCHP_MM CCCHP_D4

2.4x106

CCCHP_Benzene

Best in economic and size viewpoint

CCCHP_Decane

6

2.0x10

1.6x106

Best in thermodynamic viewpoint

Optimal point

1.2x106

Ideal point

8.0x105 180

200

220 240 260 280 300 Evaporating temperature (oC)

(c)

320

(d)

Fig. 3 Effects of evaporating temperature on the sequential polygeneration system of PTC-CCHP: (a) Total net power output; (b) Overall efficiency of solar energy (c) Total initial investment cost; (d) Results of multi-objective optimization for the system with MM.

4. Conclusion This paper compared the performance between the ORC-based solar-CCHP system and the solar-ORC system under five medium-temperature working fluids over a range of evaporation temperature. For the solar-ORC system, the system energy efficiency increases as the increasing of the evaporating temperature, which means that the optimal configuration parameters achieved at the maximum evaporating temperature. However, for the ORC-based solar-CCHP system, the behaviors of the system thermal performance, such as the overall efficiency of the solar energy are totally changed compared with the solar-ORC system, due to the effect of the configuration characteristics. Among the five working fluids, MM is suggested to be the best candidate taking into account of the security thermal stability. Furthermore, the Pareto frontier solutions are obtained to find the optimum system configuration under the assessment framework. The overall efficiency of the solar energy is 40.95%, with the cooling-to-power ratio of 4.95 and heat-to-power ratio of 4.2. To conclude, this investigation reveals the bifurcation point on the design of the ORC-based solar-CCHP and the solar-ORC system, which makes a contribution to the configuration optimization for the ORC-based solar-CCHP system, and can lead to useful directions for future researchers in this field.

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Ying Zhang et al. / Energy Procedia 142 (2017) 1119–1126 Author name / Energy Procedia 00 (2017) 000–000

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