Thermodynamic analyses and optimization of a novel CCHP system integrated organic Rankine cycle and solar thermal utilization

Thermodynamic analyses and optimization of a novel CCHP system integrated organic Rankine cycle and solar thermal utilization

Energy Conversion and Management 196 (2019) 453–466 Contents lists available at ScienceDirect Energy Conversion and Management journal homepage: www...

NAN Sizes 0 Downloads 50 Views

Energy Conversion and Management 196 (2019) 453–466

Contents lists available at ScienceDirect

Energy Conversion and Management journal homepage: www.elsevier.com/locate/enconman

Thermodynamic analyses and optimization of a novel CCHP system integrated organic Rankine cycle and solar thermal utilization ⁎

T



Di Wua, Jifeng Zuob, Zhijian Liua, , Zhonghe Hana, , Yulong Zhanga, Qiaomei Wanga, Peng Lia a b

Department of Power Engineering, School of Energy, Power and Mechanical Engineering, North China Electric Power University, Baoding 071003, Hebei, China School of Science, Agricultural University of Hebei, Baoding 071001, Hebei, China

A R T I C LE I N FO

A B S T R A C T

Keywords: CCHP system Solar thermal utilization Organic Rankine cycle Thermodynamic analyses Electric-thermal ratio

Combined cooling, heating and power systems (CCHP) have received wide attention recently because of its prominent role in energy utilization. It could both decrease the energy consumption and improve the overall energy efficiency. In this paper, a novel coupled CCHP system was proposed which is composed of a conventional CCHP system, a solar thermal utilization (ST) subsystem and an organic Rankine cycle (ORC) subsystem. The CCHP-ST-ORC system has been evaluated thermodynamically, and compared with CCHP system and CCHPST system, which is based on the GICE (rated internal combustion engine capacity) of 100 kW. Further, through the parametric study for CCHP-ST-ORC system, the effect of several operating parameters on the system thermodynamic performance is explored. Finally, a new operation strategy for CCHP-ST-ORC system was proposed and adopted in a residential building, which presents a better thermodynamic economy than CCHP-ST system. As the result, CCHP-ST-ORC system could generate an additional electricity of 5.1 kW compared with other two systems, which is at the expense of 39.3 kW heat. Further, it could expand the maximum power generation to 108 kW. The case study shows that the primary energy ratio of CCHP-ST-ORC system is 60.2% that is much higher than the 37.6% of CCHP-ST system. The results also indicate that the CCHP-ST-ORC system consumes 9.81 × 104 m3/year of natural gas, with a 12.4% reduction in energy consumption than CCHP-ST system. Therefore, these findings can promote the sustainable development of new energy utilization models, especially the coupling CCHP system.

1. Introduction With the increasing energy consumption, human is confronting serious energy crisis and environmental pollution [1]. Meanwhile, it is also exposed the shortcomings of conventional centralized energy supply system, which mainly reflects in the serious pollution, low overall efficiency and poor flexibility [2,3]. Therefore, the combined cooling, heating and power (CCHP) system was proposed and concerned, due to several advantages as follows [4]:

• simultaneous provision of heating, cooling and power • higher total primary energy efficiency • less pollution gas emissions • flexible adjustment for the ratio of electricity to heating In order to further improve the overall performance of CCHP system, optimal configurations and operation strategies were introduced and proposed [5,6]. Chahartaghi et al. [7] realized the energy,



environmental and economic evaluations for a CCHP system combined with Stirling engine and driven by hydrogen and helium. Afzali et al. [8] proposed a new operation strategy for CCHP system, according to the overall optimal partial loads of prime mover and novel performance curves. Mago et al. [9] evaluated the performance of a micro-CCHP system that operates with a novel following hybrid load (FHL) strategy. Further, the performance comparison was analyzed between the FHL strategy and following electric load (FEL) strategy, following thermal load (FTL) strategy. Hawkes et al. [10] explored the cost-effective operating strategies for three micro-CHP system in winter and summer. However, the mismatch for various loads demanded between the system and buildings is still the limitation of CCHP system, especially in peak load period [11]. Based on the strategy optimization, a novel integrated CCHP system is urgent to be proposed, which could realize a wide range of electric-thermal ratios. Many researches attempted to re-assemble the trigeneration system through combining CCHP system with other systems to fully utilize the renewable energies [12,13]. Remarkably, the solar thermal technology

Corresponding authors. E-mail addresses: [email protected] (Z. Liu), [email protected] (Z. Han).

https://doi.org/10.1016/j.enconman.2019.06.020 Received 5 April 2019; Received in revised form 8 June 2019; Accepted 11 June 2019 0196-8904/ © 2019 Elsevier Ltd. All rights reserved.

Energy Conversion and Management 196 (2019) 453–466

D. Wu, et al.

presented two typical configurations, namely, the parallel system and the sequential system. They revealed the higher overall economic exergy efficiency for sequential system. Chang et al. [30] explored a residential micro-CCHP system that coupled with fuel cell, ORC system and vapor compression cycle system, in which solar energy acts as one driving heat source. They indicated that the average efficiency of the combined CCHP system is 75.4% in summer and 85.0% in winter. However, they have not concerned the system provision range of electric-thermal ratios and do not propose an optimal strategy to further improve the system performance. Based on above considerations, this paper introduced a novel CCHPST-ORC system, which combines a conventional CCHP system, a ST subsystem and an ORC subsystem. In this paper, the key contribution is the comparison for the system thermodynamic performance with different scenarios, according to the magnitude of heat and electricity generated, exergy efficiency and primary energy ratio. On one hand, the comparison is between the CCHP-ST-ORC system, CCHP system and CCHP-ST system to exhibit the advantages of the introduced system. On the other hand, the comparison is based on the CCHP-ST-ORC system between different operating parameters that include the rated capacity of internal combustion engine (GICE), smoke temperature at WHR outlet (Tsmoke,WHR,out) and evaporation temperature of organic working fluid (Teva,org). Further, the originality of this paper is to propose a new operation strategy for CCHP-ST-ORC system, which indicates a better thermodynamic economy. These findings provide a comprehensive evaluation for distributed energy systems, and verified the superiority of CCHP-ST-ORC system based on the new operation strategy.

is usually adopted to increase the amount of heat generated. Yang et al. [14] proposed a solar hybrid CCHP system that hybridizes the CCHP system, solar collectors and photovoltaic panels. Wang et al. [15] combined a traditional CCHP subsystem, a compressed air energy storage subsystem and a solar energy utilization subsystem. Here, the solar energy is used to heat the high-pressure air. Su et al. [12] suggested that the biogas could be transferred into syngas through reforming reaction assisted by solar energy, and the syngas is used to be the primary energy in CCHP system. Because of the low solar-to-electrical efficiency, solar energy is more likely to be utilized to provide heat than for electricity generation [16,17]. Therefore, these hybrid systems can broaden the electric-thermal ratio through increasing the heat. The power only provided by prime mover, thus, the maximum electric energy produced is restricted by the rated capacity of prime mover. Therefore, in order to produce sufficient amount of energy, especially the power, and improve the flexibility of system, this paper intends to combine a conventional CCHP system, a solar energy utilization system and an Organic Rankine Cycle (ORC) subsystem. Herein, the ORC is a widespread technology, and its theories have gradually matured. Li et al. [18,19] explored the ORC system with different turbine models. They compared the system comprehensive performance with different working conditions and determined the optimal organic working fluid. For a standard and a regenerative ORC system, Braimakis et al. [20] researched the effect of system configuration and heat source temperatures on the total economic performance. Chaiyat et al. [21] combined the ORC and absorption systems, and found the integrated system has a higher total efficiency than normal ORC system. Uris et al. [22] investigated the technical and economic feasibility of a biomass-fueled ORC system in different scenarios. Further, there are some literature about the combination of CCHP system and ORC technology. Sadreddini et al. [23] combined the compressed air energy storage system, ORC system and ejector system into a CCHP system, which could provide heat, cooling and electricity. A thermodynamic evaluation was explored for the integrated CCHP system. On one hand, the effect of several parameters, such as inlet/ outlet pressure and outlet mass flow rate of cavern, as well as inlet temperature of gas turbine etc., on system performance was studied. On the other hand, the optimal performance was realized through adjusting the operating parameters, which could improve the round trip efficiency and exergy efficiency. Fang et al. [24] proposed a CCHP-ORC system, in which the electric chiller, ORC and gas boiler are severed as the adjusted devices to meet the user loads dynamically. Further, an optimal strategy for this system was presented with a wide adjustment range for user loads. Through comparing the typical CCHP system without ORC and electric chiller, the complementary CCHP-ORC system has a salient feature in economy, environment and energy consumption. Moreover, Yağlı et al. [25] used the subcritical and supercritical ORC technology to recover exhaust smoke heat in CCHP system. Rostamzadeh et al. [26] presented two novel micro-CCHP system that respectively combine the ORC cycle and Kalina cycle (KC). It is figured out that the optimum thermal efficiency is 76.54% for ORCbased system and 77.32% for KC-based system, and the optimum exergy efficiency is 48.37% for ORC-based system and 31.2% for KCbased system. Further, they point out the method to improve the system efficiency according to the parametric study. However, the study about the integrated system, combining the CCHP subsystem, the solar thermal utilization (ST) subsystem and the ORC subsystem, has not received attention widely. The main reason is the conversion efficiency limitation for ORC technology between the heat and electricity [27]. Maraver et al. [28] suggested that the ORC technology can be coupled to CCHP system, and it is limited to generate power ranging in 0–200 kW. However, the integrated CCHP system still has the potential to improve its performance and is higher than common CCHP systems. Typically, Zhao et al. [29] proposed a CCHP system coupled with ORC system and driven by solar energy, and

2. System description Three systems that could provide cooling heating and power simultaneously are introduced. Thereinto, the conventional CCHP system is the basic system, and the CCHP-ST is the system combined the solar thermal utilization technology, based on the basic system. While the CCHP-ST-ORC system integrates the basic system, solar thermal utilization technology and ORC subsystem. 2.1. Conventional CCHP system Conventional CCHP system includes an internal combustion engine (ICE), a waste heat recovery equipment (WHR), an air source heat pump (ASHP), an absorption heat pump (ABS) and heat exchangers (Fig. 1). ICE could provide electricity with full or partial load, and its exhaust smoke is transported into WHR. The circulating water could absorb heat from smoke in WHR, and the heated water is employed to provide heat load. Here, the heat exchanger 1 is a buffer to adjust the pressure and flow easily in engineering. For heat exchanger 2, it is used to recover the heat in jacket water to increase the heat produced by system. Meanwhile, ASHP and ABS could generate cooling for user, at the expense of heat and power respectively. If the energy generated by CCHP system could not meet the user loads, grid or gas boiler act as the supplement to provide electricity or heat, respectively. In this basic CCHP system, the cooling water of primary network is transported into heat exchanger 2 and WHR respectively, thus, it is heated by jacket water and smoke subsequently. And then, the heated cooling water returns to heat exchanger 1 to release the heat to circulating water in secondary network for heat provision. Finally, the cooling water is drawn out of the heat exchanger 1 to start a new heated cycle. 2.2. CCHP-ST system Apart from the components of conventional CCHP system, CCHP-ST system has evacuated tube collectors and a water tank (Fig. 2). Based on conventional CCHP system, the collector receives solar energy in daytime, to rise the water temperature. For the primary water flow, the water firstly leaves the heating exchanger 1, and flows into heat 454

Energy Conversion and Management 196 (2019) 453–466

D. Wu, et al.

Fig. 1. System schematic of conventional CCHP system.

exchanger 2, to absorb heat from jacket water. Then, the water flows into water tank to be heated by solar energy, and into WHR to recover the heat from exhaust smoke, subsequently. Finally, the water flows back to heat exchanger 1 to heat circulating water in secondary network.

difference is that the hot water heated by solar energy and smoke is used to drive the ORC system [31–33]. To be specific, one water stream is drawn out of tank and transported into WHR to absorb the heat from smoke. And then, the heated water flows into evaporator of ORC subsystem, releasing heat to working fluid. Finally, the stream of water flows back to tank and heated by solar energy. Moreover, in CCHP or CCHP-ST system, the smoke from ICE is drawn into WHR to recover the waste heat, and then released to environment. However, in CCHP-STORC system, the released smoke is fed into heat exchanger 1, to heat circulating water. For circulating water, it both flows into heat exchanger 1 and into heat exchanger 2, to absorb heat from both smoke

2.3. Novel CCHP-ORC-ST system CCHP-ORC-ST system integrates a CCHP subsystem, a ST subsystem and an ORC subsystem, aiming to increase electric quantity produced (Fig. 3). Its operation mode is similar with CCHP-ST system. The

Fig. 2. System schematic of CCHP-ST system. 455

Energy Conversion and Management 196 (2019) 453–466

D. Wu, et al.

Fig. 3. System schematic of CCHP-ST-ORC system.

Meanwhile, a lot of waste heat released from ICE is transferred to exhaust smoke, to jacket water, to environment, to oil cooler and aftercooler. The fraction of various heat stream to total waste heat has a relationship as follows.

and jacket water for meeting heat load. For ORC subsystem, it includes an evaporator, a condenser, a heat recovery device, a turbine and a pump (Fig. 3). R113 is selected as organic working fluid in this study. Because R113 is a dry fluid and its critical temperature is close to the water temperature of evaporator heat side, which makes more energy released. In evaporator, the heat is transferred from hot water to organic working fluid, which results in the organic working fluid becoming saturated or superheated vapor. Then the heated organic working fluid expands in turbine and drives turbine operating and producing electricity. In addition, heat exchange occurs between the outlet stream of turbine and inlet stream of evaporator, to reduce the energy loss. Subsequently, the organic working fluid goes through condenser, and the liquid of organic working fluid is available. Finally, the fluid is pumped back to evaporator, and start a new cycle. Relying on ORC technique, the novel system could generate power from both ICE and ORC subsystem simultaneously.

ftotal = fsmoke + f jacket + fenv + foilcooler + faftercooler

where ftotal is 1, representing the total waste heat from ICE; fsmoke, fjacket, fenv, foilcooler and faftercooler are respectively the fraction of the waste heat that goes to exhaust smoke, jacket water, environment, oil cooler coolant stream and aftercooler coolant stream. The required amount of natural gas can be calculated as follows.

EICE,mesh =

Egas =

3. Method

3.1.1. Internal combustion engine ICE is the cardinal prime mover of CCHP system to generate power through burning natural gas, whose performance is usually a function of partial load ratio (PLR). PLR is limited in the range of 0.0–1.0, which is expressed as follows.

EICE G ICE

Ereq ηmesh

EICE,mesh ηele

(3)

(4)

where Ereq is the power required, kW; ηmesh is mechanical efficiency of power generation equipment; EICE,mesh is the power generated when not considering the mechanical efficiency, kW; ηele is electrical efficiency of ICE; Egas is the natural gas input required, kW. According to a manufacturer, the ICE performance with several typical PLR values was obtained [34,35]. On the basis, the fitting curve for PLR-ηele, PLR-ηmesh, PLR- fjacket and PLR- fsmoke is shown in Fig. 4. Based on the fitting relationships, the approximate calculation is carried out with GICE below 200 kW. Correspondingly, the heat that goes to jacket water and exhaust smoke (Qjacket, kW, Qsmoke, kW) can be determined as follows.

3.1. Theoretical models

PLR =

(2)

(1)

where EICE is the actual electricity produced form ICE, kW; GICE is the rated ICE capacity, kW.

Qjacket = f jacket × (Egas − EICE,mesh ) 456

(5)

Energy Conversion and Management 196 (2019) 453–466

D. Wu, et al. 0.370

0.940 0.938

0.365

0.936 0.934 0.932

0.355

mesh

ele

0.360

0.350

0.930 0.928 0.926

0.345

0.924

0.340

0.922 0.920

0.335 0.4

0.5

0.6

0.7

0.8

0.9

0.4

1.0

0.5

0.6

0.7

0.8

0.9

1.0

PLR

PLR

(a) PLR-Șele

(b) PLR-Șmesh 0.534

0.315

0.532 0.314

0.528

0.312

0.526

fsmoke

fjacket

0.530 0.313

0.311 0.310

0.524 0.522 0.520

0.309

0.518

0.308

0.516 0.514

0.307

0.512

0.306 0.4

0.5

0.6

0.7

0.8

0.9

0.510

1.0

0.4

0.5

0.6

PLR

0.7

0.8

0.9

1.0

PLR

(a) PLR- fjacket

(b) PLR- fsmoke

Fig. 4. Function of PLR-ηele, PLR-ηmesh, PLR- fjacket and PLR- fsmoke.

3.1.2. Waste heat recovery equipment WHR could transfer the heat from exhaust smoke to circulating water, and its actual heat recovery efficiency (ηWHR) is determined as follows.

efficiency, and α0 is heat loss coefficient of evacuated tube collector. Tsolar and T0 are the water temperature at collector inlet and environment temperature, K; φrad is ambient solar radiation, W/m2. Other technical parameters are provided by manufacturer, and listed in Table 1. Therefore, the heat obtained from solar collectors (Qcollector, kW) is calculated as follows.

ηWHR = η0,WHR × (0.0951 + 1.525 × PLRWHR + 0.6249 × PLRWHR 2)

Qcollector = φrad × ηcollector × Acollector

Qsmoke = fsmoke × (Egas − EICE,mesh )

(6)

(7)

in which η0,WHR is the rated heat recovery efficiency of WHR. PLRWHR is partial load ratio of WHR, which is defined as the ratio of heat entering WHR (Qin,WHR, kW) to rated WHR capacity (GWHR, kW).

PLRWHR

Qin,WHR = G WHR

where Acollector is the effective heat collecting areas of a collector, m2. 3.1.4. Water tank Water tank, as a heat storage container, could store the hot water from collectors. The amount of heat provided from tank (Qtank,out, kW) is calculated as follows.

(8)

Qtank,out = Qtank,in × ηtank

3.1.3. Solar collector Evacuated tube collector has higher solar energy collection efficiency, which is used in CCHP system to fulfill the solar thermal utilization. Due to the solar collector only worked when solar energy existed, solar collectors run at 8:00–17:00 in this research. Its heat collecting efficiency (ηcollector) can be calculated as follows.

ηcollector = ηcollector,0 − α 0 × (Tsolar − T0)/ φrad

(10)

(11)

where Qtank,in is the heat input to water tank, kW, and ηtank is heat preservation coefficient of tank. 3.1.5. Heat exchanger Heat exchanger could transfer the heat from exhaust smoke or hot water to cooling water. Taking the heat loss into consideration, the relationship between the input heat (Qex,in, kW) and output heat (Qex,out, kW) is as follows.

(9)

in which ηcollector,0 is intercept of instantaneous heat collecting 457

Energy Conversion and Management 196 (2019) 453–466

D. Wu, et al.

Table 1 Technical parameters of system [35–38]. Items

Values

ICE Rated ICE capacity, GICE (kW) Inlet temperature of jacket water (°C) Mass flow rate of jacket water, mjacket,water (kg/s) Specific heat capacity of exhaust smoke (kJ/(m3·K))

100 15 0.2 1.1

WHR Rated WHR capacity, GWHR (kW) Rated heat recovery efficiency of WHR, η0,WHR

100 0.88

Solar collector Intercept of instantaneous heat collecting efficiency, ηcollector,0 Heat loss coefficient of collector, α0 (W/(m2·K)) Effective heat collecting areas of a collector, Acollector (m2) Heat preservation coefficient of water tank, ηtank Mass flow rate of circulating water in each solar collector (kg/s) Number of solar collectors Volume of tank (m3)

0.7 2.5 3.76 0.8 0.3 50 20

ORC Pinch point temperature difference (°C) Average temperature of cooling water (°C) Evaporation temperature of organic working fluid, Teva,org (°C) Relative internal efficiency of turbine, ηtur Pump efficiency in ORC subsystem, ηorg,pump Superheat of R113 in evaporator (°C) Mass flow rate of water in heat side of evaporator, mORC,water (kg/s)

5 15 70 0.85 0.85 0 2

Heat exchanger Heat transfer coefficient of heat exchanger, ηex Gas boiler Heat efficiency of gas boiler, ηboiler

Fig. 5. T-s diagram in ORC system.

where Qboiler is the heat that can be provided by gas boiler, kW; Ngas is the amount of gas natural consumed, kW; ηboiler is heat efficiency of boiler. 3.1.8. ORC subsystem Compared with traditional steam power generation, there is organic working fluid in ORC system, allowing a lower evaporating temperature. The temperature-entropy (T-s) diagram is presented in Fig. 5. For the ORC subsystem model, several simplifying assumptions is considered as follows:

0.9 0.85

Heat pump Cooling thermal coefficient of ASHP, COPASHP Cooling thermal coefficient of ABS, COPABS

4.0 0.7

Natural gas Inlet temperature of natural gas (°C) Lower calorific value of natural gas, QH,ng, (kJ/m3) Higher calorific value of natural gas, QL,ng, (kJ/m3) Air-fuel ratio Excess air coefficient

15 35,600 38,080 9.52:1 2

Others Mass flow rate of hot water to user (kg/s) Mass flow rate of water in primary network (kg/s) Backwater temperature from user (°C) Power consumed of pumps (kW)

2 2 40 0.5

• ORC system operates under a steady condition. • There are no pressure drops, friction loss and heat loss in components. • Organic working fluid of condenser outlet is saturated liquid. • The efficiency of pump and turbine is constant. Thermodynamic model of ORC system can be described as follows: ① Pumping process (4-5) The pump power consumed (Eorg,pump, kW) can be calculated as follows.

Qex,out = Qex,in × ηex

(12)

Eorg,pump = morg × (h5 − h4 ) = morg × (h5s − h4 )/ ηorg,pump

where ηex is the heat efficiency of heat exchanger. In addition, heat exchanger is arranged as counter flow, with terminal temperature difference of 5 °C.

where morg is the flow mass of organic working fluid in ORC, kg/s; h5 and h4 are the specific enthalpy of organic working fluid at outlet and inlet of pump respectively, J/kg; h5s is the specific enthalpy of organic working fluid at pump outlet for isentropic case, J/kg; ηorg,pump is the pump efficiency.

3.1.6. Air-source/Absorption heat pump ASHP and ABS could provide cooling for user through consuming electricity and heat respectively, for which the energy function is built by cooling thermal coefficient (COP).

Qcooling,ASHP = EASHP × COPASHP

(13)

Qcooling,ABS = Q heat,ABS × COPABS

(14)

② Heating process (6-1) The heat absorbed from evaporator heat side (Qeva, kW) can be determined as follows.

Qeva = morg × (h1 − h6)

where Qcooling,ASHP and Qcooling,ABS are the cooling quantity provided by ASHP and ABS, kW; COPASHP and COPABS are the COP of ABS and ASHP, respectively; EASHP is the electricity consumed by ASHP, kW; Qheat,ABS is the heat consumed by ABS, kW.

(17)

where h1 is the specific enthalpy of the organic working fluid at outlet of evaporator, J/kg; h6 is the specific enthalpy of the organic working fluid at inlet of evaporator, J/kg.

3.1.7. Gas boiler For gas boiler, the function between the heat provided and the natural gas consumed can be expressed as follows.

Q boiler = Ngas × ηboiler

(16)

③ Expansion process (1-2) The electricity produced by turbine (Etur, kW) is expressed as follows.

(15) 458

Energy Conversion and Management 196 (2019) 453–466

D. Wu, et al.

Etur = morg × (h1 − h2) = morg × (h1 − h2s)/ ηtur

Exfuel is the exergy of natural gas consumed, kW. The exergy of natural gas consumed (Exfuel, kW) can be approximately calculated by:

(18)

in which h1 and h2 are the specific enthalpy of organic working fluid at inlet and outlet of turbine respectively, J/kg; h2s is the specific enthalpy of organic working fluid at turbine outlet for isentropic case, J/kg; ηtur is the relative internal efficiency of turbine.

Ex fuel = 0.95 × QH,ng × Vng

where QH,ng is the higher calorific value of natural gas, kJ/m ; Vng is the volume flow rate of natural gas consumed by system, m3/s. For each device, the corresponding exergy efficiency is defined similarly.

④ Heat recovery process (2-3 and 5-6) According to the principle of energy conservation, the following equation can be obtained.

① ICE

(19)

h3 − h2 = h6 − h5

The exergy efficiency of ICE (ηex,ICE) is defined as Eq. (26).

in where h3 and h2 are the specific enthalpy of organic working fluid on heat side at outlet and inlet of heat recovery device, J/kg. Similarly, h6 and h5 are the specific enthalpy of organic working fluid on cooling side at outlet and inlet of heat recovery device, J/kg.

ηex,ICE =

EICE Ex fuel − Qout,smoke × (1 −

Qcon = morg × (h3 − h4 )

The exergy efficiency of heat exchanger and WHR (ηex,ex/WHR) is defined as Eq. (27).

(20)

where h3 and h4 are the specific enthalpy of organic working fluid at inlet and outlet of condenser, J/kg.

ηex,ex/WHR

(

Qc,out,ex/WHR × 1 − 3.2. Performance indexes

=

(

Q h,in,ex/WHR × 1 − 3.2.1. Power provision The total power produced by CCHP-ST-ORC system (Esys, kW) is defined as follows.

Esys = EICE + EORC − Epump

ηex,ORC =

in which Qjacket and Qsmoke are the heat that is recovered and used to supply for meeting heat load, kW; α is the heating coefficient from jacket water to cooling water of secondary network; β is the heating coefficient from smoke to cooling water of secondary network; α and β are determined by system structure.

Ex fuel

(

T0 Tin,user

T h,out,ex/WHR

(

T0 Tin,ORC

)−Q

(

× 1− out,ORC

T0 Tout,ORC

)

(28)

3.2.4. Primary energy ratio The primary energy ratio of total system (PERsys) is defined by:

PERsys =

Esys + Qsys QL,ng × Vng

(29) 3

where QL,ng is the lower calorific value of natural gas, kJ/m .

3.2.3. Exergy efficiency The overall exergy efficiency of CCHP-ST-ORC system (ηex,sys) is defined as follows.

× 1−

) )

where Qout,ORC and Qin,ORC are the heat of water output from and input to evaporator, kW; Tout,ORC and Tin,ORC are the water temperature at outlet and inlet of evaporation, °C.

(23)

in,sys

h,out,ex/WHR

T0

EORC Qin,ORC × 1 −

3.2.2. Heat provision The total heat produced by system (Qsys, kW) is defined as follows.

)−Q

)−Q

T0 Tc,in,ex/WHR

The exergy efficiency of ORC subsystem (ηex,ORC) is defined as Eq. (28).

where Eorg,pump is the pump power consumed to convey organic working fluid, kW; ηtur,ele is electrical efficiency of turbine. For CCHP and CCHP-ST system, the power produced is defined similarly. The difference is that there is no EORC in such two systems.

T0 Tout,user

T h,in,ex/WHR

( × (1 −

× 1−

c,in,ex/WHR

③ ORC subsystem

(22)

Qsys = Qjacket × α + Qsmoke × β

T0

)−Q

(27)

(21)

EORC = Etur × ηtur,ele − E org,pump

T0 Tc,out,ex/WHR

where Qout,ex/WRH is the heat of water or smoke output from exchanger or WHR, kW; Qin,ex/WHR is the heat of water or smoke input to exchanger or WHR, kW; Tout,ex/WHR is the temperature of water or smoke at exchanger or WHR outlet, °C; Tin,ex/WHR is the temperature of water or smoke at exchanger or WHR inlet, °C. For the added footnote “h” and “c”, it represents the heat side and cooling side in heat exchanger or WHR.

where EICE is the power produced by ICE, kW; EORC is the net power produced by ORC subsystem, kW; Epump is the power consumed of water pumps, kW. For ORC system, the net power produced (EORC) is calculated as follows.

(

(26)

② Heat exchanger and WHR

The heat that rejected from organic working fluid to cooling water (Qcon, kW) in condenser is expressed as follows.

Esys + Qout,sys × 1 −

T0 ) Tout,smoke

where Qout,smoke is the heat of exhaust smoke from ICE, kW; Tout,smoke is the exhaust smoke temperature, °C.

⑤ Heat rejection process (3-4)

ηex,sys =

(25) 3

3.3. Parameter settings

)

The technical parameters of system are presented in Table 1, which refer to relevant literature and technical data.

(24)

4. Results and discussion

in which Qout,sys and Qin,sys are the heat supplied for users and the heat returned from users, kW. Correspondingly, Tout,user and Tin,user are the water temperatures supplied for users and returned from users, °C;

The system thermodynamic performance was compared between 459

Energy Conversion and Management 196 (2019) 453–466

D. Wu, et al.

Table 2 Average hourly performance of three types of CCHP system.

ST subsystem

CCHP-ST system

Only heat provision

Heat + Electricity

100 90

Items

CCHP

CCHP-ST

CCHP-ST-ORC

Power provision (kW) Heating provided (kW) Exergy efficiency (%) Primary energy ratio (%)

100 91.5 39.1 65.3

100 105.9 40.1 70.2

105.1 66.6 39.1 58.5

80

Heat (kW)

70

different CCHP systems and different operating parameters. Here, the systems operate continuously with full load state, and exclude the supplement, namely, the gas boiler and the gird. Moreover, the ASHP and ABS were not considered, because the cooling load can be met by converting the heat or power. Therefore, the total heat and electricity, primary energy ratio and exergy efficiency are evaluated in detail.

60 50

Heat 40 30

PLR=0%

PLR=40%-100%

20 10

Electricity

Heat

0

4.1. Performance comparison of three CCHP systems

0

10

20

30

40

50

60

70

80

90

100

Electricity (kW) The performance of three CCHP systems is compared. Thereinto, the Tsmoke,WHR,out is 300 °C; and the Teva,org is 70 °C. Moreover, the GICE is 100 kW, and the ICE operates with PLR of 100%. Because of the intermittent nature for solar energy, the system performance of CCHP-ST and CCHP-ST-ORC varies hourly. The average hourly system performance is presented in Table 2. As seen in Table 2, the total power generated is the same for CCHP and CCHP-ST systems, because they provide power only relying on the ICE. However, due to the solar thermal utilization, there is a heat provision difference of approximately 14 kW, with 91.5 kW for CCHP system and 105.9 kW for CCHP-ST system. For CCHP-ST-ORC system, the electricity increases about 5.1 kW compared with other two systems without ORC, at the expense of 39.3 kW heat. Moreover, these three systems have almost equal exergy efficiency around 40%. It indicates that the ST subsystem has smaller contribution for the improvement of system exergy efficiency. Further, the utilization of solar energy results in the primary energy ratio of CCHP-ST system being approximate 7.5% higher than that of CCHP system. However, the primary energy ratio of CCHP-ST-ORC system is only 58.5%, much lower than that of the other two systems. The reason is that, for CCHP-ST-ORC system, there is a large amount of heat was converted into a small amount of power through ORC subsystem, whereas the primary energy ratio index does not consider the energy quality improvement for the energy provision comparison. Figs. 6–8 show the range of electric-thermal provision for CCHP,

Fig. 7. Range of electric-thermal provision for CCHP-ST system. 100

80

Heat (kW)

Heat (kW)

CCHP-ST-ORC system Heat + Electricity

PLR=100%

Electricity

0 90

Electricity

Heat 10

20

30

40

50

60

70

80

90

100 110

CCHP-ST and CCHP-ST-ORC systems. As seen in Fig. 6, for CCHP system, the heat and electricity generated from system increase with an increasing PLR. The maximum for heat and electricity is 92 kW and 100 kW respectively. However, the minimum energy provision is 20 kW for heat and 40 kW for electricity, due to the lowest PLR limitation. For CCHP-ST system, solar energy heating could provide about 15 kW heat when PLR below 40% (Fig. 7). With PLR of 40%-100%, the amount of heat for CCHP-ST system is about 3 kW more than that for CCHP system, due to the solar thermal utilization. The maximum for heat is approximately 95 kW with rated ICE capacity operated. CCHP-ST-ORC system further widens the range of electric-thermal provision (Fig. 8). With PLR below 40%, there is only solar thermal utilization operated. With PLR of 40%-100%, the CCHP-ST-ORC system is reduced to CCHP-ST system. Further, under the PLR of 100%, the CCHP-ST-ORC system expand the maximum power generation to 108 kW, through changing the Tsmoke,WHR,out and mORC,water. The relationship between the Tsmoke,WHR,out, mORC,water and the amount of energy provision is presented in Fig. 9. It can be seen that the Tsmoke,WHR,out has significant effect on the amount of heat and electricity generated. With the Tsmoke,WHR,out decreases, heat for users has a sharp decline tendency, and meanwhile, the total electricity produced slightly increases. Because more smoke heat is transferred to ORC system. The different slope of heat and electricity generated indicates

PLR = 40%-100%

80

PLR=40%-100%

Fig. 8. Range of electric-thermal provision for CCHP-ST-ORC system.

30

70

PLR=0%

Electricity (kW)

Heat

60

Heat

40

0

50

50

50

0

60

40

Heat

60

10

Heat + Electricity

10

Heat + Electricity

20

80

20

Only heat provision

30

90

40

CCHP-ST system

70

100

70

ST subsystem

90

100

Electricity (kW) Fig. 6. Range of electric-thermal provision for CCHP system. 460

Heat exchanger 1 WHR

Exergy efficiency of heat exchanger 2

0.85

mjacket water=4 kg/s

mjacket water=2 kg/s

0.80

ORC subsystem Heat exchanger 2

0.75 0.70 0.65 0.60 0.55 0.50 0.45 0.40 0.35 0.30 0.25 0.20 0.15 50

60

70

80

90

0.42 0.41 0.40 0.39 0.38 0.37 0.36 0.35 0.34 0.33 0.32 0.31 0.30 0.29 0.28 0.27 0.26 0.25 0.24 0.23 0.22

Exergy efficiency of other devices

Energy Conversion and Management 196 (2019) 453–466

D. Wu, et al.

100 110 120 130 140 150

GICE (kW) Fig. 11. Exergy efficiency of devices with different GICE. Fig. 9. Relationship between the Tsmoke,WHR,out, mORC,water and the amount of energy provision.

Fig. 11 shows the exergy efficiency of devices with different GICE. The change of GICE has a larger influence on the exergy efficiency of heat exchanger 2 and WHR, because these two devices are directly connected behind the ICE, devoting to recover the heat released. The heat exchanger 2 has a downward tendency for exergy efficiency as GICE increases. The reason is that the increased temperature difference between the heat side and cooling side of heat exchanger 2 leads to the irreversible loss increased. For heat exchanger 1, more heat is transferred from smoke to water, resulting in the input and output exergy increased. However, the increase of exergy output is slightly greater than that of energy input, which leads to the exergy efficiency risen. For the similar reason, the exergy efficiency of WHR first increases and then decreases with the maximum of 0.36. In addition, the exergy efficiency of ORC subsystem has a slight downward tendency, because the increase of heat consumption is greater than the increase of power produced.

that the power generation is at the expense of a large amount of lowerquality of heat. Similarly, with an increasing of mORC,water, the electricity increases slightly and the heat decreases.

4.2. CCHP-ST-ORC performance with different rated ICE capacity The system performance with various GICE is presented in Fig. 10. Here, the Tsmoke,WHR,out is 300 °C, and the Teva,org is 70 °C. In order to ensure an effective cooling for ICE, the mjacket,water is 2 kg/m3 with GICE below 100 kW, and 4 kg/m3 with GICE above 100 kW. Obviously, with an increased mjacket,water, the heat generated, total exergy efficiency and primary energy ratio decrease. The reason is that the heat transferred from jacket water to cooling water decreases. Further, based on constant mjacket,water, the electricity and heat increase with GICE increasing. As seen, the primary energy ratio obviously increases from 0.51 to 0.57 as GICE increases from 50 kW to 100 kW. Moreover, the exergy efficiency has a slight increased tendency in the range of 0.36–0.40. Therefore, increasing GICE has benefit for system performance improvement.

Electricity and heat (kW)

mjacket water=4 kg/s

mjacket water=2 kg/s

0.95

140

0.90

120

0.85 0.80

100

0.75

80

0.70

60

0.65 40

0.60

20

0.55

0

0.50

-20

0.45

-40

0.40

-60

0.35

-80

0.30 60

70

80

90

Exergy efficiency and primary energy ratio

1.00

180

50

As Tsmoke,WHR,out increases, the performance variation is presented in Fig. 12, based on the GICE of 100 kW and Teva,org of 70 °C. Due to the Tsmoke,WHR,out raised, the heat transferred from smoke to the stream of cooling water which is acted as the heat source for ORC system reduces, and meanwhile to the stream of cooling water which is acted as the heat source for heat load increases. Thus, the electricity decreases slightly, whereas the heat provided to users obviously increases. It indicates that increasing the Tsmoke,WHR,out could decrease the amount of power produced by ORC system, and significantly increase the amount of heat provided by system. Moreover, due to the dramatic increase of heat produced, the total primary energy ratio has an obvious upward tendency, ranging from 0.47 to 0.67. Meanwhile, the total exergy efficiency has a slight increase in the range of 0.38–0.40, due to the exergy loss of ORC system decreased. The exergy efficiency of devices is displayed in Fig. 13. As seen, the increased Tsmoke,WHR,out results in the exergy efficiency of WHR decreased. Additionally, due to the increased smoke temperature input to heat exchanger 1, its irreversible loss increased, which leads to a sharp drop tendency of exergy efficiency. Moreover, because the cooling water is heated by heat exchanger 1 and subsequently flows into heat exchanger 2, the increased water temperature at heat exchanger 2 inlet results in a reduced irreversible loss. Thus, the exergy efficiency of heat exchanger 2 has an upward tendency. While increasing the Tsmoke,WHR,out could result in a slight rise of exergy efficiency of ORC

Exergy efficiency Primary energy ratio

Electricity Heat 160

4.3. CCHP-ST-ORC performance with different smoke temperature at WHR outlet

100 110 120 130 140 150

GICE (kW) Fig. 10. System energy provision and performance with different GICE. 461

0.90 0.85 0.80 0.75 0.70 0.65 0.60 0.55 0.50 0.45 0.40 0.35 0.30

50

100

150

200

250

300

350

400

450

Electricity and heat (kW) and round trip efficiency of ORC subsystem

120 110 100 90 80 70 60 50 40 30 20 10 0 -10 -20 -30 -40 -50

Exergy efficiency and primary energy ratio

Electricity and heat (kW)

Exergy efficiency Primary energy ratio

Electricity Heat

Exergy efficiency Electricity Heat Primary energy ratio Round trip efficiency of ORC

120

0.70 0.68 0.66 0.64 0.62 0.60 0.58 0.56 0.54 0.52 0.50 0.48 0.46 0.44 0.42 0.40 0.38 0.36 0.34 0.32 0.30

110 100 90 80 70 60 50 40 30 20 10 0 50

55

60

65

500

70

75

80

85

90

95

Exergy efficiency and primary energy ratio

Energy Conversion and Management 196 (2019) 453–466

D. Wu, et al.

100

Teva,org

Tsmoke,WHR,out

Fig.14. System energy provision and performance with different Teva,org. Fig. 12. System energy provision and performance with different Tsmoke,WHR,out.

Heat exchanger 2 ORC subsystem

Heat exchanger 1

Heat exchanger 2

0.62

0.48

Exergy efficiency of devices

0.70 0.65

Exergy efficiency

WHR ORC subsystem

0.50

0.60 0.55 0.50 0.45 0.40 0.35 0.30 0.25

0.61

0.46

0.60

0.44

0.59

0.42

0.58

0.40

0.57

0.38

0.56

0.36

0.55

0.34

0.54

0.32

0.53

0.30

0.52

0.28

0.51

0.26

0.50

0.24

0.49

0.22

0.48

50

0.20 50

100

150

200

250

300

350

400

450

Exergy efficiency of heat exchanger 2

WHR Heat exchanger 1

0.75

55

60

65

500

70

75

80

85

90

95

100

Teva,org

Tsmoke,WHR,out

Fig. 15. Exergy efficiency of devices with different Teva,org. Exergy efficiency Primary energy ratio

Electricity Heat

subsystem.

Electricity and heat (kW)

4.4. CCHP-ST-ORC performance with different evaporation temperature of organic working fluid Fig. 14 presents the impact of Teva,org for R113 in ORC subsystem on system performance. Here, GICE is 100 kW with full load operation, and Tsmoke,WHR,out is 300 °C. With an increased Teva,org, the electricity produced increases slightly, meanwhile, the round trip efficiency of ORC subsystem also increases. Specifically, for every 5 °C increase of Teva,org, power generation increases by 0.38 kW, and round trip efficiency increases by 0.9. Moreover, because the ORC subsystem and heating provision subsystem have different heat source, the Teva,org has no effect on the amount of heat produced. In addition, the overall exergy efficiency and primary energy ratio of system increase slightly, with the average of 0.39 and 0.59. For various Teva,org, the exergy efficiency of devices is displayed in Fig. 15. It can be seen that the exergy efficiency of ORC system improves from 0.26 to 0.36, when the Teva,org increased from 50 °C to 100 °C. In addition, increasing the Teva,org results in a higher water temperature at WHR inlet. The increased water temperature could lead to the reduced irreversible loss and increased exergy efficiency of WHR.

110 100 90 80 70 60 50 40 30 20 10 0 -10 -20 -30 -40 -50 -60

0.85 0.80 0.75 0.70 0.65 0.60 0.55 0.50 0.45 0.40 0.35 0.30 50

55

60

65

70

75

80

85

90

95

Exergy efficiency and primary energy ratio

Fig. 13. Exergy efficiency of devices with different Tsmoke,WHR,out.

100

PLR (%) Fig. 16. System energy provision and performance with different PLR of ICE.

462

Energy Conversion and Management 196 (2019) 453–466

D. Wu, et al.

ICE WHR Heat exchanger 2 Heat exchanger 1 ORC subsystem Round trip efficiency of ORC

0.80

12.0

0.75

11.0

Exergy efficiency

0.65 0.60

10.5 0.55 10.0

0.50 0.45

9.5 0.40 9.0

0.35 0.30

Round trip efficiency of ORC

11.5 0.70

8.5 0.25 8.0

0.20 50

55

60

65

70

75

80

85

90

95

100

PLR (%) Fig. 18. Ambient temperature and solar radiation of Weifang in typical year.

Fig. 17. Exergy efficiency of devices with different PLR of ICE.

8x104

4.5. CCHP-ST-ORC performance with different PLR of ICE

Electricity

Cooling

Heat

7x104

Building Loads (kW)

With GICE of 100 kW, Fig. 16 displays the system performance with different PLR of ICE. As the preconditions, the Tsmoke,WHR,out is 300 °C, and Teva,org is 70 °C. Since the PLR increases, the electricity and heat produced from system deservedly increase, at the expense of more fuel consumed. As seen, the primary energy ratio increases by 0.08, and exergy efficiency increases by 0.03 when the PLR increases from 50% to 100%, and further their increasing tendency slows down as PLR gets closer to 100%. Under various PLR of ICE, the exergy efficiency of devices is presented in Fig. 17. The ICE exergy efficiency remains around 0.43. And the exergy efficiency of ICE with partial loads is slightly lower than that with full load. The heat exchanger 2 has a declining tendency for exergy efficiency from 0.74 to 0.60. The reason is the improved jacket water temperature results in the higher irreversible loss. Moreover, the volume flow of smoke that fed into WHR increases, as the PLR goes up. The fact results in that the increase for energy output is greater than the increase for energy input, thus, the WHR exergy efficiency is rising.

6x104 5x104 4x104 3x104 2x104 1x104 0 Jan Feb Mar Apr May Jun Jul Aug Sep Oct Nov Dec

Month Fig. 19. Building loads in typical year.

5. Case study electric-thermal ratio through changing the Tsmoke,WHR,out and mORC,water to satisfy the user loads. In Fig. 20, Euser and Quser are the electricity and heat required by users, kW. The energy consumption for cooling load is contained in the Euser and Quser according to the conversion formulas (13) and (14). The max and min stand for the maximum and minimum for the corresponding variables, which is set in calculation. QICE is the heat provided for user from ICE, which includes the heat from jacket water and exhaust smoke, kW; whereas Qsolar is the heat provided for user from solar thermal utilization subsystem, kW. To be specific, the electrical load is first met by ICE. If the power produced by ICE could not satisfy the load, the ORC subsystem is running. And the mwater is increasing from the min to the max to increase the electricity produced. Then if the coupled system still could not meet the electrical load, more power is produced by decreasing the Tsmoke,WHR,out. With a decreasing Tsmoke,WHR,out, more heat is transferred to ORC system to increase the electricity. Finally, the power from grid is used as the supplement for providing the electricity that is still insufficient. For the heat provision, the heat from ICE is the preferred energy source. If there is not enough exhaust heat for meeting heat load, the hot water derived from tank is used to provide heat for users. Thus, the solar thermal utilization is the secondary heat source. At last, the gas boiler is the heating supplement.

To verify the thermodynamic economy for CCHP-ST-ORC system, the performance was analyzed based on one residential building. 5.1. Building information The building is located in Weifang city, China (latitude 119.15 °N, longitude 36.70 °E, altitude 35 m). The ambient feature in typical year is presented in Fig. 18. It indicates that the yearly average for global radiation is 159 W/m2 and for diffuse radiation is 93 W/m2. The average annual outdoor temperature is approximately 13.1 °C. The residential building has five storeys and its total building area is 12000 m2. The heat transfer coefficient for external walls is 0.622 W/ (m2·K) and for external windows is 3.1 W/(m2·K). Based on the building structure and ambient features, the building loads was obtained by energy consumption simulation software (Fig. 19). Moreover, the maximum hourly electricity is 113.7 kW. 5.2. Operation strategy To improve the thermodynamic economy, a detailed operation strategy is proposed for CCHP-ST-ORC system (Fig. 20). The new operation strategy is based on conventional FEL principle, and varies the 463

Energy Conversion and Management 196 (2019) 453–466

D. Wu, et al.

Fig. 20. Flow diagram of operation strategy for CCHP-ST-ORC system.

5.3. System performance

Table 3 Variable settings for systems.

In the case study, a CCHP-ST-ORC system is adopted in the residential building to satisfy the user loads. The ICE operates with various PLR according to the electricity required. Meanwhile, the system is running based on the new strategy. In addition, the collectors work at 8:00–17:00, resulting in large amount of heat entering the tank. For avoiding boiling, the water is replaced with oil acting as the working fluid in tank and collectors. The performance for CCHP-ST-ORC system based on the new strategy and CCHP-ST system based on the conventional FEL strategy is compared. Because the CCHP-ST system provides the electricity only relying on the ICE, its rated ICE capacity is the maximum hourly electrical load (Table 3). For the CCHP-ST-ORC system, the rated ICE capacity is selected as 105 kW, and ORC system bears the remaining load. The cooling load is generated by ABS, and its energy consumption is converted into heat load. Table 4 shows the performance for CCHP-ST system and CCHP-STORC system. Compared with Table 2, the exergy efficiency decreases for these two systems in this case. Because in the case study, the available energy is equal to the user loads that vary hourly rather than the total heat and power generated from coupled system. The extra

Items

Values

Specific heat capacity of oil (kJ/(kg·K)) GICE in CCHP-ST system GICE in CCHP-ST-ORC system The maximum value of mwater (kg/s) The minimum value of mwater (kg/s) The minimum value of Tsmoke,out (°C)

2.3 114 105 2 0.1 5 °C above the oil temperature at WHR inlet 0.7

COPABS

Table 4 Performance comparison between the CCHP-ST and CCHP-ST-ORC systems.

464

Items

CCHP-ST

CCHP-ST-ORC

Exergy efficiency Primary energy ratio (%) Natural gas consumption (m3/year)

33.4 37.6 1.12 × 105

33.8 60.2 9.81 × 104

Energy Conversion and Management 196 (2019) 453–466

D. Wu, et al.

Declaration of Competing Interest

energy produced is expelled into environment. For the same reason, the primary energy ratio of CCHP-ST system is only 37.6%. Whereas the primary energy ratio of CCHP-ST-ORC system is 60.2%, which is much higher than that of CCHP-ST system. Further, in CCHP-ST-ORC system, the heat can be met by solar collectors and ICE, without the gas boiler. While the CCHP-ST system also needs the gas boiler to replenish the heat required. Although the CCHP-ST-ORC system is equipped with a rated ICE capacity of only 105 kW, it presents excellent energy provision performance. In addition, the total natural gas consumption is 1.12 × 105 m3/year for CCHP-ST system, of which 1.15 × 104 m3/year is consumed by gas boiler. The CCHP-ST-ORC system consumes 9.81 × 104 m3/year of natural gas, with a 12.4% reduction in energy consumption. Therefore, compared with CCHP-ST system, the CCHPST-ORC system with new operation strategy could obviously improve the primary energy ratio and decrease the energy consumption.

None. Acknowledgments This work was supported by the Fundamental Research Funds for the Central Universities (No. 2018QN084). References [1] Abbasi Mohammad, Chahartaghi Mahmood, Hashemian Seyed Majid. Energy, exergy, and economic evaluations of a CCHP system by using the internal combustion engines and gas turbine as prime movers. Energy Convers Manage 2018;173:359–74https://linkinghub.elsevier.com/retrieve/pii/ S0196890418308367https://doi.org/10.1016/j.enconman.2018.07.095. [2] Cai Tianxing, Zhao Chuanyu, Xu Qiang. Energy network dispatch optimization under emergency of local energy shortage. Energy 2012;42(1):132–45https:// linkinghub.elsevier.com/retrieve/pii/S0360544212002800https://doi.org/10. 1016/j.energy.2012.04.001. [3] Liu Zhijian, Wu Di, He Bao-Jie, Wang Qiaomei, Yu Hancheng, Ma Wensheng, Jin Guangya. Evaluating potentials of passive solar heating renovation for the energy poverty alleviation of plateau areas in developing countries: a case study in rural Qinghai-Tibet Plateau, China. Sol Energy 2019;187:95–107https://linkinghub. elsevier.com/retrieve/pii/S0038092X19305122https://doi.org/10.1016/j.solener. 2019.05.049. [4] Chu Xiaolin, Yang Dong, Li Xiaohong, Zhou Rui. Evaluation of CCHP system performance based on operational cost considering carbon tax. Energy Procedia 2017;142:2930–5https://linkinghub.elsevier.com/retrieve/pii/ S1876610217361696https://doi.org/10.1016/j.egypro.2017.12.419. [5] Mago PJ, Chamra LM. Analysis and optimization of CCHP systems based on energy, economical, and environmental considerations. Energy Build 2009;41(10):1099–106https://linkinghub.elsevier.com/retrieve/pii/ S0378778809001157https://doi.org/10.1016/j.enbuild.2009.05.014. [6] Yang Cheng, Wang Xusheng, Huang Manman, Ding Su, Ma Xiaoqian. Design and simulation of gas turbine-based CCHP combined with solar and compressed air energy storage in a hotel building. Energy Build 2017;153:412–20https:// linkinghub.elsevier.com/retrieve/pii/S0378778816316474https://doi.org/10. 1016/j.enbuild.2017.08.035. [7] Chahartaghi M, Sheykhi M. Energy, environmental and economic evaluations of a CCHP system driven by Stirling engine with helium and hydrogen as working gases. Energy 2019;174:1251–66. [8] Afzali Sayyed Faridoddin, Mahalec Vladimir. Novel performance curves to determine optimal operation of CCHP systems. Appl Energy 2018;226:1009–36https://linkinghub.elsevier.com/retrieve/pii/ S0306261918308894https://doi.org/10.1016/j.apenergy.2018.06.024. [9] Mago PJ, Chamra LM, Ramsay J. Micro-combined cooling, heating and power systems hybrid electric-thermal load following operation. Appl Therm Eng 2010;30(8-9):800–6https://linkinghub.elsevier.com/retrieve/pii/ S135943110900355Xhttps://doi.org/10.1016/j.applthermaleng.2009.12.008. [10] Hawkes A, Leach M. Cost-effective operating strategy for residential micro-combined heat and power. Energy 2007;32(5):711–23https://linkinghub.elsevier.com/ retrieve/pii/S0360544206001277https://doi.org/10.1016/j.energy.2006.06.001. [11] Li B, Hu P, Zhu N, Lei F, Xing L. Performance analysis and optimization of a CCHPGSHP coupling system based on quantum genetic algorithm. Sustain Cities Soc 2019;46. 101408. [12] Su Bosheng, Han Wei, Chen Yi, Wang Zefeng, Qu Wanjun, Jin Hongguang. Performance optimization of a solar assisted CCHP based on biogas reforming. Energy Convers Manage 2018;171:604–17https://linkinghub.elsevier.com/ retrieve/pii/S019689041830582Xhttps://doi.org/10.1016/j.enconman.2018.05. 098. [13] Lu Shilei, Li Yuwei, Xia Hongwei. Study on the configuration and operation optimization of CCHP coupling multiple energy system. Energy Convers Manage 2018;177:773–91https://linkinghub.elsevier.com/retrieve/pii/ S0196890418311063https://doi.org/10.1016/j.enconman.2018.10.006. [14] Yang G, Zhai XQ. Optimal design and performance analysis of solar hybrid CCHP system considering influence of building type and climate condition. Energy 2019;174:647–63https://linkinghub.elsevier.com/retrieve/pii/ S0360544219304050https://doi.org/10.1016/j.energy.2019.03.001. [15] Wang Xusheng, Yang Cheng, Huang Manman, Ma Xiaoqian. Multi-objective optimization of a gas turbine-based CCHP combined with solar and compressed air energy storage system. Energy Convers Manage 2018;164:93–101https:// linkinghub.elsevier.com/retrieve/pii/S0196890418301985https://doi.org/10. 1016/j.enconman.2018.02.081. [16] Soheyli Saman, Shafiei Mayam Mohamad Hossein, Mehrjoo Mehri. Modeling a novel CCHP system including solar and wind renewable energy resources and sizing by a CC-MOPSO algorithm. Appl Energy 2016;184:375–95https://linkinghub. elsevier.com/retrieve/pii/S0306261916314192https://doi.org/10.1016/j. apenergy.2016.09.110. [17] Liu Zhijian, Liu Yuanwei, He Bao-Jie, Xu Wei, Jin Guangya, Zhang Xutao. Application and suitability analysis of the key technologies in nearly zero energy buildings in China. Renew Sustain Energy Rev 2019;101:329–45https://linkinghub.

6. Conclusions This paper proposed a novel CCHP-ST-ORC system that combines a conventional CCHP system, a ST subsystem and an ORC subsystem. Further, the thermodynamic performance was compared between the CCHP-ST-ORC system, CCHP system and CCHP-ST system, which is based on the GICE of 100 kW. Several conclusions have been reached as follows: (1) For CCHP-ST-ORC system, it could generate an additional electricity of 5.1 kW compared with other two systems, which is at the expense of 39.3 kW heat. Moreover, these three systems have almost equal exergy efficiency around 40%. The primary energy ratio is 65.3% for CCHP system and 70.2% for CCHP-ST system, while it is only 58.8% for CCHP-ST-ORC system. (2) The electric-thermal provision for CCHP system is limited by the PLR of ICE, in the range of 40 kW/20 kW through 100 kW/92 kW. For CCHP-ST system, solar energy heating could provide about 15 kW heat when PLR below 40%. With PLR of 40%-100%, there was 3 kW heat generated from CCHP-ST system more than that from CCHP system. Further, CCHP-ST-ORC system widens the range of electric-thermal provision. Under the PLR of 100%, the CCHP-STORC system expands the maximum power generation to 108 kW. (3) Increasing the GICE could improve the exergy efficiency and primary energy ratio, thus, which has benefit for system performance improvement. The total primary energy ratio has an obvious upward tendency, increasing from 0.47 to 0.67 with Tsmoke,WHR,out increased from 50 °C to 500 °C. For every 5 °C increase of Teva,org, power generation increases by 0.38 kW, and round trip efficiency of ORC subsystem increases by 0.9. With PLR increased from 50% to 100%, the primary energy ratio increases by 0.08, and exergy efficiency increases by 0.03, and further their increasing tendency slows down as PLR gets closer to 100%. (4) In the case study, the primary energy ratio of CCHP-ST system is only 37.6%. Whereas, the primary energy ratio of CCHP-ST-ORC system is 60.2%, which is much higher than that of CCHP-ST system. Moreover, in CCHP-ST-ORC system, the heat can be met by solar collectors and ICE, without the gas boiler. Furthermore, the CCHP-ST-ORC system consumes 9.81 × 104 m3/year of natural gas, with a 12.4% reduction in energy consumption than CCHP-ST system. In summary, the integrated CCHP-ST-ORC system has a wide range of electric-thermal provision, and realizes an obvious energy saving. Therefore, such energy utilization structure and operation mode provide a novel effective approach, which combines renewable energy with fossil energy, for satisfying the building loads. These findings have significant contribution to the development of coupling CCHP system, in terms of design, operation and application. 465

Energy Conversion and Management 196 (2019) 453–466

D. Wu, et al.

[18]

[19]

[20]

[21]

[22]

[23]

[24] [25]

[26]

[27]

elsevier.com/retrieve/pii/S1364032118307792https://doi.org/10.1016/j.rser. 2018.11.023. Li Peng, Han Zhonghe, Jia Xiaoqiang, Mei Zhongkai, Han Xu, Wang Zhi. Analysis and comparison on thermodynamic and economic performances of an organic Rankine cycle with constant and one-dimensional dynamic turbine efficiency. Energy Convers Manage 2019;180:665–79https://linkinghub.elsevier.com/ retrieve/pii/S0196890418312615https://doi.org/10.1016/j.enconman.2018.11. 017. Li Peng, Han Zhonghe, Jia Xiaoqiang, Mei Zhongkai, Han Xu, Wang Zhi. Comparative analysis of an organic Rankine cycle with different turbine efficiency models based on multi-objective optimization. Energy Convers Manage 2019;185:130–42https://linkinghub.elsevier.com/retrieve/pii/ S0196890419301852https://doi.org/10.1016/j.enconman.2019.01.117. Braimakis Konstantinos, Karellas Sotirios. Integrated thermoeconomic optimization of standard and regenerative ORC for different heat source types and capacities. Energy 2017;121:570–98https://linkinghub.elsevier.com/retrieve/pii/ S0360544217300439https://doi.org/10.1016/j.energy.2017.01.042. Chaiyat Nattaporn, Kiatsiriroat Tanongkiat. Analysis of combined cooling heating and power generation from organic Rankine cycle and absorption system. Energy 2015;91:363–70https://linkinghub.elsevier.com/retrieve/pii/ S0360544215011330https://doi.org/10.1016/j.energy.2015.08.057. Uris María, Linares José Ignacio, Arenas Eva. Techno-economic feasibility assessment of a biomass cogeneration plant based on an Organic Rankine Cycle. Renew Energy 2014;66:707–13https://linkinghub.elsevier.com/retrieve/pii/ S0960148114000512https://doi.org/10.1016/j.renene.2014.01.022. Kang Ligai, Yang Junhong, An Qingsong, Deng Shuai, Zhao Jun, Li Zelin, Wang Yongzhen. Complementary configuration and performance comparison of CCHPORC system with a ground source heat pump under three energy management modes. Energy Convers Manage 2017;135:244–55https://linkinghub.elsevier.com/ retrieve/pii/S0196890416311293https://doi.org/10.1016/j.enconman.2016.12. 055. Fang F, Wei L, Liu J, Zhang J, Hou G. Complementary configuration and operation of a CCHP-ORC system. Energy 2012;46(1):211–20. Yağlı Hüseyin, Koç Yıldız, Koç Ali, Görgülü Adnan, Tandiroğlu Ahmet. Parametric optimization and exergetic analysis comparison of subcritical and supercritical organic Rankine cycle (ORC) for biogas fuelled combined heat and power (CHP) engine exhaust gas waste heat. Energy 2016;111:923–32https://linkinghub. elsevier.com/retrieve/pii/S0360544216307472https://doi.org/10.1016/j.energy. 2016.05.119. Rostamzadeh Hadi, Ebadollahi Mohammad, Ghaebi Hadi, Shokri Afshar. Comparative study of two novel micro-CCHP systems based on organic Rankine cycle and Kalina cycle. Energy Convers Manage 2019;183:210–29https:// linkinghub.elsevier.com/retrieve/pii/S0196890419300263https://doi.org/10. 1016/j.enconman.2019.01.003. Sadreddini Amirhassan, Fani Maryam, Ashjari Aghdam Muhammadali, Mohammadi Amin. Exergy analysis and optimization of a CCHP system composed of compressed air energy storage system and ORC cycle. Energy Convers Manage

[28]

[29]

[30]

[31]

[32]

[33]

[34] [35]

[36]

[37]

[38]

466

2018;157:111–22https://linkinghub.elsevier.com/retrieve/pii/ S019689041731107Xhttps://doi.org/10.1016/j.enconman.2017.11.055. Maraver Daniel, Sin Ana, Royo Javier, Sebastián Fernando. Assessment of CCHP systems based on biomass combustion for small-scale applications through a review of the technology and analysis of energy efficiency parameters. Appl Energy 2013;102:1303–13https://linkinghub.elsevier.com/retrieve/pii/ S0306261912005247https://doi.org/10.1016/j.apenergy.2012.07.012. Zhao L, et al. Solar driven ORC-based CCHP: Comparative performance analysis between sequential and parallel system configurations. Appl Therm Eng 2018;131:696–706. Chang Huawei, Wan Zhongmin, Zheng Yao, Chen Xi, Shu Shuiming, Tu Zhengkai, Chan Siew Hwa. Energy analysis of a hybrid PEMFC–solar energy residential microCCHP system combined with an organic Rankine cycle and vapor compression cycle. Energy Convers Manage 2017;142:374–84https://linkinghub.elsevier.com/ retrieve/pii/S0196890417302686https://doi.org/10.1016/j.enconman.2017.03. 057. Han Xu, Zeng Wei, Han Zhonghe. Investigation of the comprehensive performance of turbine stator cascades with heating endwall fences. Energy 2019;174:1188–99https://linkinghub.elsevier.com/retrieve/pii/ S0360544219304426https://doi.org/10.1016/j.energy.2019.03.038. Han Zhonghe, Mei Zhongkai, Li Peng. Multi-objective optimization and sensitivity analysis of an organic Rankine cycle coupled with a one-dimensional radial-inflow turbine efficiency prediction model. Energy Convers Manage 2018;166:37–47https://linkinghub.elsevier.com/retrieve/pii/ S0196890418303571https://doi.org/10.1016/j.enconman.2018.04.022. Han Zhonghe, Fan Wei, Zhao Ruocheng. Improved thermodynamic design of organic radial-inflow turbine and ORC system thermal performance analysis. Energy Convers Manage 2017;150:259–68https://linkinghub.elsevier.com/retrieve/pii/ S0196890417307379https://doi.org/10.1016/j.enconman.2017.08.025. Lin S, Li X, Chen B. Application Manual of Gas-fired Cooling, Thermoelectricity and Distributed Energy Technology. Beijing, China: China Electric Power Press; 2017. National Development and Reform Commission of China. Distributed Energy and Cogeneration of Heat, Power and Cooling. Beijing, China: China Electric Power Press; 2013. Han Z, Wu D, Liu Z, et al. Study on configuration optimization and economic feasibility analysis for combined cooling, heating and power system. Energy Convers Manage 2019;190:91–104. Li Fan, Sun Bo, Zhang Chenghui, Zhang Lizhi. Operation optimization for combined cooling, heating, and power system with condensation heat recovery. Appl Energy 2018;230:305–16. https://doi.org/10.1016/j.apenergy.2018.08.101. Sheykhi Mohammad, Chahartaghi Mahmood, Balakheli Mohammad Mahdi, Kharkeshi Behrad Alizadeh, Miri Seyyed Mahdi. Energy, exergy, environmental, and economic modeling of combined cooling, heating and power system with Stirling engine and absorption chiller. Energy Convers Manage 2019;180:183–95https://linkinghub.elsevier.com/retrieve/pii/ S0196890418312305https://doi.org/10.1016/j.enconman.2018.10.102.