Adjustable performance analysis of combined cooling heating and power system integrated with ground source heat pump

Adjustable performance analysis of combined cooling heating and power system integrated with ground source heat pump

Energy 163 (2018) 475e489 Contents lists available at ScienceDirect Energy journal homepage: www.elsevier.com/locate/energy Adjustable performance ...

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Energy 163 (2018) 475e489

Contents lists available at ScienceDirect

Energy journal homepage: www.elsevier.com/locate/energy

Adjustable performance analysis of combined cooling heating and power system integrated with ground source heat pump Jiangjiang Wang a, *, Yuzhu Chen a, Chao Dou a, Yuefen Gao a, Zheng Zhao b a b

School of Energy, Power and Mechanical Engineering, North China Electric Power University, Baoding, Hebei Province 071003, China School of Control and Computer Engineering, North China Electric Power University, Baoding, Hebei Province 071003, China

a r t i c l e i n f o

a b s t r a c t

Article history: Received 12 March 2018 Received in revised form 23 June 2018 Accepted 20 August 2018 Available online 21 August 2018

The difference in the electricity to cooling/heating ratio between a building and combined cooling, heating, and power (CCHP) system has a significant influence on the system configuration and performance. This paper designs a CCHP system coupled with a ground source heat pump (GSHP) to coordinate the match between system supplies and building demands. The energy flows in the cooling and heating work conditions are analyzed, and the thermodynamic models of components constructed. By means of a case study, the performances of the coupling CCHP system, under design and off-design working conditions, are evaluated and analyzed using energetic indicators, including the primary energy ratio and exergy efficiency, respectively. The adjustable areas expressed by electricity and cooling/heating, as well as the adjustable performance distributions, are obtained and discussed in order to guide operation regulation of the CCHP system integrated with the GSHP. Comparisons between the CCHP system with and without GSHP indicate that the coupling CCHP system in the specific case study can save averagely 40.6% and 39.5% of the primary energy in cooling and heating work conditions, respectively. © 2018 Elsevier Ltd. All rights reserved.

Keywords: Combined cooling heating and power system Ground source heat pump Primary energy ratio Exergy efficiency Adjustable area

1. Introduction Distributed energy systems have been considered to supplement the traditional central system, owing to their high overall efficiency and environmentally friendly performance [1], of which combined cooling, heating, and power system (CCHP) or combined heating and power system (CHP) is one of key forms [2]. The outstanding feature of the CCHP/CHP system is the cascading utilization of energy. Researches on distributed systems focusing on system configurations [3], performance evaluation [4], operational strategies [5], and optimization methods [6] aimed to establish optimal distributed systems in order to reduce energy consumption and greenhouse gas emissions. Gradually, renewable energy, including solar, wind, and geothermal energy, as well as bioenergy, has been introduced into distributed energy systems to amplify the benefits gained from energy, environment, and even economy [7]. According to the different technologies and characteristics, various integrated systems have been proposed and analyzed, such as CCHP systems based on biomass gasification with air [8] or steam [9], and CCHP

* Corresponding author. E-mail address: [email protected] (J. Wang). https://doi.org/10.1016/j.energy.2018.08.143 0360-5442/© 2018 Elsevier Ltd. All rights reserved.

systems coupled with solar collectors [10] or solar fuels [11]. However, one limitation in solar or wind integrated systems is their intermittent nature. Furthermore, hybrid biomass systems may be affected by low heat density and challenges in collection and transportation. Among the available renewable energy sources, geothermal energy has attracted significant attention due to its friendly performance and relative stability, particularly for shallow underground geothermal resources from groundwater or soil coupled with a heat pump system [12], namely a ground source heat pump (GSHP) system. Typical applications of GSHP systems consist of providing heating and/or cooling for users. Lucia et al. [12] reviewed GSHP systems that include both various GSHP technologies and thermodynamic models. The summaries demonstrated that systems with a GSHP can reduce the environmental impact of buildings. Zhai et al. [13] designed a mini-type GSHP system for a meeting room and discussed output capacities in the typical mode, which demonstrates the GSHP system applicability for building demands. Yuan et al. [14] proposed a control mode for the GSHP with a borehole-free cooling coupled system, and investigated the underground heat balance problem. This research concluded that the mode can improve energy efficiency and decrease energy consumption. In particular, the annual operational cost of the GSHP system is obviously reduced. To improve the performances of GSHP

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Nomenclature AHP CCHP CDE CDERR CHP COP FEL FTL GSHP HG HX ICE LG LX

absorption heat pump combined cooling heating and power carbon dioxide emission carbon dioxide emission reduction ratio combined heating and power coefficient of performance following electricity load following thermal load ground source heat pump high pressure generator heat exchanger internal combustion engine low pressure generator low temperature heat exchanger

Symbols C E Ex F f h m n Ne P

average specific capacity (kJ kg1 K1) electricity (kW) exergy (kW) area (m2) mass flow (kg s1) enthalpy (kJ kg1 K1) mass (kg) variable coefficient generation capacity (kW) pressure (kPa)

system, the exergy analysis was often used to discover the segments with more exergy destruction. Menberg et al. [15] proposed a water-based hybrid GSHP system and gas-fired boiler, and analyzed the exergy loss and destruction of the hybrid system by modeling each subsystem. Esen et al. [16,17] predicted the daily performance of GSHP in a fuzzy weighted pre-processing method with the limited experimental data and investigated the GSHP's energy and exergy performances as a function of depth trenches for heating season. The investigations shown that the energetic and exergetic efficiencies of the system increase when increasing the heat source (ground) temperature for heating season. Besides space heating/cooling provided by GSHP system, Balbay et al. [18,19] studied the using of GSHP systems for snow melting on bridge slabs and pavements, and researched the temperature distribution of slabs and pavements. Moreover, the GSHP have coupled with solar energy in recent researches [20]. Esen et al. [21] have analyzed a solar-assisted GSHP system, and obtained its coefficient of performance (COP) and other performances. Moreover, a GSHP system for heating and cooling is coupled with a CCHP system for electricity, heating, and cooling in order to supplement each other and relieve the limitation of the fixed heat to electricity ratio of the CCHP system. The CCHP and GSHP coupling system usually consists of a prime mover, waste heat utilization system, and GSHP. The electricity produced by the prime mover is fed to the building and GSHP. The cooling and heating demands can be met by the waste heat utilization system and GSHP. Typically, the alternatives of waste heat utilization system for producing chilled water includes absorption chiller/heat pump or adsorption chiller driven by heat sources with different heating temperature. The absorption chiller or heat pump is driven by the exhausted gas from power generation unit with high temperature. Wang et al. [8] employed the absorption chiller driven by three kinds of heat sources to recover the waste heat from the exhausted

Q R W PER T

energy (kW) constant for ideal gas practical power consumption (kW) efficiency primary energy ratio temperature (K)

Subscripts ac c ch co e ev ex g gr grid hw i in mix n ng out sys t 0

absorption heat pump compressor chilled water condenser electricity evaporator exergy ground source heat pump ground source water electricity grid hot water ICE inlet mixed hot water nominal natural gas outlet system throttle value standard reference state

h

gas to produce chilled water and hot water for space cooling and heating respectively and domestic hot water. Yang et al. [22] proposed and researched a new open-cycle absorption heat pump system to recover the waste heat from flue gas, and found that this system can get an excellent output even at high return water temperature. In addition, the adsorption chillers with the low temperature heat source could recover the waste heat with low grade energy level. Chorowski et al. [23] investigated an adsorption chiller which utilizes low-temperature heat from cogeneration and demonstrated that the adsorption chiller can be worked with a hot water of 65  C, a typical cogeneration heating temperature in distributed energy systems. The performance of adsorption chiller directly influences the operation parameters on cooling capacity [24] and then determines the recovery efficiency of waste heat in CCHP system. Although the recovery efficiency of adsorption chiller is lower than the absorption chiller, it provides one alternative to utilize the waste heat with the lower temperature. The researches on CCHP and GSHP coupling systems with the typical absorption chiller or heat pump have started to emerge in the last few years, only in performance analysis and configuration strategy. Kang et al. [25] proposed a hybrid system with four subsystems, including a power generation unit, GSHP unit, absorption chiller, and storage tank, and analyzed the energy and environmental performances in three basic modes. Similarly, Kang et al. [26] employed a comprehensive matrix approach to compare the configurations and performances of CCHP-organic Rankine cycle system with a GSHP in three energy management modes. Liu et al. [27] focused on hourly operation strategy of a CCHP system with GSHP and thermal energy storage under variable loads and discussed the economic and environmental benefits of the CCHP system with thermal energy storage. Moreover, aimed to obtain more benefits of CCHP and GSHP coupling system, optimization

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methodologies were applied to construct the system's configurations and determine the operation modes. Zeng et al. [28,29] employed the genetic algorithm to construct the optimization models based on energy, environment, and economy criteria in order to optimize the capacity and operation mode of the integrated system. The studies on CCHP and GSHP systems in these literatures [25e29] were only based on the general energy flows and the black-box models, and the specific configurations and flows were not considered. To reveal the mechanism to improve the performance, Kang et al. [30] present exergy analysis of a gasturbine CHP-GSHP coupling system and investigated the key operation parameters to affect the performances. Dou and Wang [31] proposed a novel coupled CCHP-GSHP system and analyzed the thermodynamic performances of heating supply modes. Differently to the reviewed literatures on CCHP and GSHP coupling system, this study is extended from our published literature [31] considering only heating mode and parameter's influences, and its originality lies in designing a hybrid CCHP and GSHP coupling system comprehensively integrated with the electricity output and the two forms of waste heat from an internal combustion engine (ICE) in cooling and heating operation modes, and focusing on its adjustable areas and performance distributions without adopting thermal energy storage to match the variable ratios of electricity to heating/cooling between supply and demand sides in different work conditions. Section 2 presents the flowchart and operational modes of the hybrid system; Section 3 discusses the construction of the thermodynamic models; Section 4 demonstrates the energy and exergy performances with the different adjustable ratio of electricity to heating/cooling of the hybrid system through a case study; and finally, Section 5 summarizes several conclusions. 2. System description The flowchart of the CCHP and GSHP coupling system is displayed in Fig. 1, and consists of three subsystems: the ICE, GSHP, and dual-source powered mixed-effect LiBreH2O absorption heat

477

pump (AHP) subsystem. Natural gas (state 1) as fuel drives the ICE to generate electricity for the building (state 4) or GSHP system (state 5). The waste heat, including exhaust gas (state 13) and jacket water (states 6/18), is used to drive the AHP or assist the GSHP. 2.1. Summer working condition In summer working conditions, the GSHP and AHP are relatively independent. The ICE provides electricity to the GSHP, while the waste heat from both the jacket water and exhaust gas is used to drive the AHP. The exhaust gas (state 13) releases heat in the highpressure generator (HG) of the AHP, then the exhaust gas (state 14) from the AHP is used to heat the hot water (state 16) through heat exchanger 2 (HX2) because the temperature in the state 14 is larger than that of the jacket water, and finally, it is emitted to the atmosphere (state 15). Thus, the waste heat includes the heat carried by the jacket water (states 6 and 18), the exhaust gas (states 13 and 14) and the hot water (states 16 and 17) from HX2 to continue recover the waste heat from the lower temperature exhaust gas. The mixture of jacket water (state 6) and hot water from HX2 (state 16) flows to low-pressure generator 1 (LG1) while the valves V3 and V7 are closed. The hot water (state 10) from LG1 is split into two streams that return to the jacket (state 18) and HX2 (state 17), respectively, for the next cycle. In the AHP cycle, the liquid refrigerant (H2O) in the LiBreH2O is converted into vapor by absorbing the heat in HG (state a7) and LG1 (state a4). Thereafter, the heat (state a9) released by HG evaporates the liquid refrigerant in LG2. The refrigerant vapor (states a12, a13) is fed to the condenser and condensed by the cooling water (state a17). Moreover, the low-pressure, low-temperature liquid refrigerant (state a14) evaporates by absorbing the heat from the chilled water (state a19) in the evaporator. The liquid refrigerant (state a15) is absorbed in the absorber and pumped to HG and LG1 for the next cycle. In the cycle, the strong, high-temperature solution is converted to a weak, low-temperature solution by absorbing the heat from the low-temperature heat exchanger (LX) and hightemperature heat exchanger (HX). Furthermore, the refrigerant

Fig. 1. Flowchart of CCHP and GSHP coupling system.

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vapor is cooled by the cooling tower (states d1, d2). In the GSHP cycle, the liquid refrigerant (R134a, state b4) is evaporated in the evaporator by absorbing the chilled water heat (states b5 and b6). Then, the refrigerant vapor (state b3) is compressed into vapor with a high temperature and pressure (state b2) by the compressor. Finally, the vapor flows into the condenser and releases the heat to the water from underground (state b9). Following throttling in the throttle value, the liquid refrigerant (state b1) falls into the evaporator for the next cycle. Overall, in the summer working conditions, the chilled water (states a20 and b10) is produced by both AHP and GSHP, so that the hybrid system has a higher cooling load regulation range. Moreover, the GSHP cooling output is driven by electricity generated by the ICE. Thus, the electricity to cooling ratio of the entire coupling system will be widened to match the building load. 2.2. Winter working condition In the winter working conditions, the exhaust gas is still fed to HG and HX2. As opposed to the summer working conditions, the mixed hot water (states 6 and 16), which is divided into three parts, flows into the HX1 (state 9), LG1 (state 4), and HX3 (state 8) when V3 is opened and V6 is closed. Furthermore, one part, which is fed to HX3, releases heat to evaporation in the AHP. Moreover, the other part of the mixed hot water preheats the ground source water inlet in the evaporator, a component of the GSHP. In the AHP subsystem, the internal cycle of LiBreH2O is similar to the summer working conditions while the external input sources differ. In winter conditions, the values containing V20 and V23 are opened, while V19, V21, V22, and V25 are closed. Furthermore, the hot water (state a18) used in the air conditioning is heated in the absorber and condenser. Similarly, in the GSHP, the condenser in summer is operated as an evaporator in winter; likewise, the evaporator in summer has a common function with the condenser in winter. The ground source water (state b7), which is preheated by heat from HX1, evaporates the liquid refrigerant from the throttle value (state b1). Then, the vapor (state b2) is compressed to vapor with a high temperature and pressure (state b3). Furthermore, the vapor (state b3) releases heat to the building in the condenser and produces a continuous heating output. The liquid refrigerant (state b4) is fed to the throttle value for the next cycle. Overall, the hot water (states a18 and b6) produced by the GSHP and AHP is used for the space heating system. The design parameters in the two working conditions are listed in Table 1.

Subsystem

Parameter

State

Summer, K

Winter, K

ICE

Exhaust gas Jacket water

13 6/16 17/18 14/15 8/11 a19/a20 c2/c1 a13/a18 a13/a14 13/14 7/10 a16/a17 a16/a17 9/12 b4/b3 b2/b1 b2/b1 b4/b3

751 358 343 443/393

751 358 343 443/393 443/393

HX2 HX3 Evaporator Condenser HG LG1 Absorber

GSHP

3.1. ICE For the ICE driven by natural gas, the power generation efficiency ðhi Þ can be estimated as follows [32]:

hi ¼ 0:2808  ðNen Þ0:0563 ;

(1)

where Nen is the nominal generation capacity of the ICE. The temperature of the gas exhausted from the ICE, Tex (K), is calculated as:

 2 Tex ¼ 2  105 Epgu  0:0707Epgu þ 758:33:

(2)

The heat ratio of the jacket water to exhausted gas differs with the various ICE types, and the ICE design parameters are listed in Table 2, according to a specific ICE from the Caterpillar company.

3.2. AHP The two-stage Libr-H2O AHP is driven by the jacket hot water and exhausted gas from the ICE. Its coefficient of performance (COP) is strongly dependent on the heat source ratios. By combining the three-source chiller in a previous study [8] with the simulation from another study [10], a thermodynamic model was constructed according to material and energy balance equations. The detailed assumptions, and balance equations can be found in Ref. [33]. The thermophilic properties of the LiBreH2O solution can be directly obtained from the Engineering Equation Solver (EES) software. The COP of the AHP can be expressed as:

COP ¼

Qac;co þ Qac;a ðwinterÞ QHG þ QLG1 þ Qac;ev

(3)

COP ¼

Qac;ev ðsummerÞ; QHG þ QLG1

(4)

where QHG and QLG1 are the heat inputs from high and low-pressure generator, respectively, and Qac;co , Qac;a , and Qac;ev are the heat released from the condenser, absorber, and evaporator, respectively.

3.3. GSHP

Table 1 Design parameters.

AHP

3. Thermodynamic models and evaluation criteria

HX1 Evaporator Condenser

285/280 333/323

The GSHP [34] includes the compressor, throttle, evaporator, condenser, and heat exchanger HX1.

Table 2 ICE design parameters. Natural gas

Component analysis, mol %

291/302 751/443 358/343 291/302

315/318 751/443 358/343 311/315 358/343

280/285 280/292 302/291 318/311

Low heat value, MJ/Nm3 Capacity, kW Excess air factor Mass flow of jacket water, kg/s Mass flow of exhausted gas, kg/s Generating efficiency, % Heat recovery efficiency of exhausted gas, % Heat recovery efficiency of jacket water, %

CH4

C2H4

C2H6

N2

91.46

4

4.45

0.09

37.96 110 1.2 1.478 0.120 36.6 14.0 31.0

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3.3.3. Condenser

3.3.1. Compressor

mg;c;in ¼ mg;c;out

(5)

mg;co;in ¼ mg;co;out

(15)

mg;c;in hg;c;in  mg;c;out hg;c;out ¼ W

(6)

Qg;co ¼ mg;co;in hg;co;in  mg;co;out hg;co;out

(16)

(7)

  * ; Qg;co ¼ cg;co  mg;co  Tg;co;in  Tg;co;out ¼ hg;co  Fg;co  Tg;co



We

We ¼

hc

(17)

  n  mg;c  R  Tg;co  Tg;ev n1

(8)

.  Tg;ev ¼ Tgr;in  eðhg;ev Fg;ev Þ=ðcg;ev mg;ev Þ  Tgr;out    eðhg;ev Fg;ev Þ=ðcg;ev mg;ev Þ  1 .  Tg;co ¼ Tg;co;in  eðhg;co Fg;co Þ=ðcg;co mg;co Þ  Tg;co;out    eðhg;co Fg;co Þ=ðcg;co mg;co Þ  1 ;

(9)

where mg;co;in and mg;co;out are the inlet and outlet mass flows, respectively, of the GSHP condenser; Qg;co is the heat from the condenser; Tg;co;in and Tg;co;out are the inlet and outlet temperatures, * is the average condenser respectively, of the condenser; Tg;co temperature; mg;co;in and mg;co;out are the inlet and outlet mass flows, respectively, of the condenser; hg;co;in and hg;co;out are the inlet and outlet enthalpy, respectively, of the condenser; and hg;co is the condenser efficiency.

(10)

where W and We are the actual and nominal power consumptions, respectively, of the compressor; and mg;c;in and mg;c;out are the inlet and outlet mass flows, respectively, of the compressor. Furthermore, hg;c;in and hg;c;out are the inlet and outlet enthalpy, respectively, of the compressor, while hc is the compressor efficiency. Tg;ev is the evaporation temperature and Tg;co is the condensation temperature; hg;ev and hg;co are the heat transfer efficiencies of the evaporator and condenser, respectively; and Fg;ev and Fg;co are the heat transfer areas of the evaporator and condenser, respectively. Moreover, n is the variable coefficient; R is the ideal gas constant; Cgr is the average specific capacity for ground source water; Cch is the average specific capacity of chilled water; Tgr;in and Tgr;out are the inlet and outlet temperatures, respectively, of the ground source water; and Cg;co and Cg;ev are the average specific capacities of the refrigerant in the condenser and evaporator, respectively.

3.3.4. Throttle value

mg;t;in ¼ mg;t;out

(18)

Hg;t;in ¼ Hg;t;out ;

(19)

where mg;t;in and mg;t;out are the inlet and outlet mass flows, respectively, of the GSHP throttle value; and hg;t;in and hg;t;out are the inlet and outlet enthalpy, respectively, of the throttle value.

3.4. Heat exchanger The thermodynamic model for heat exchangers can be expressed by the following equation:

Qhx ¼ hhx mhx ðhin  hout Þ;

3.3.2. Evaporator

mg;ev;in ¼ mg;ev;out

(11)

Qg;ev ¼ mg;ev;out hg;ev;out  mg;ev;in hg;ev;in

(12)

  * Qg;ev ¼ cg;ev  mg;ev  Tgr;in  Tgr;out ¼ hg;ev  Fg;ev  Tg;ev

(20)

where Qhx is the heat exchanged in the heat exchanger, hhx is the heat exchanger efficiency, hin and hout are the inlet and outlet enthalpy, respectively, of the heat exchanger, and mhx is the mass flow.

3.5. Evaluation criteria

(13) * ¼ Tg;ev

    Tgr;in  Tg;ev;in  Tgr;out  Tg;ev;out T

T

gr;in g;ev;in ln Tgr;out Tg;ev;out

;

(14)

where mg;ev;in and mg;ev;out are the inlet and outlet mass flows of the evaporator in the GSHP, respectively; Qg;ev is the heat absorbed by the evaporator; Tg;ev;in and Tg;ev;out are the inlet and outlet tem* is the average peratures, respectively, of the evaporator; Tg;ev

evaporator temperature; mg;ev;in and mg;ev;out are the inlet and outlet mass flow, respectively, of the evaporator; hg;ev;in and hg;ev;out are the inlet and outlet enthalpy, respectively, of the evaporator; and hg;ev is the evaporator efficiency.

The primary energy ratio (PER) and exergy efficiency are employed to evaluate performances from the first and second thermodynamic laws, respectively. The PER is defined as the ratio of the total output energy, including electricity, cooling, and heating, to the consumed natural gas and its value PERsys is expressed as:

PERsys ¼

8 > E þ Qg;ev þ Qac;ev  W > > ðsummerÞ > < Qng > E þ Qg;co þ Qac;a þ Qac;co  W > > > ðwinterÞ : Q

:

(21)

ng

The exergy efficiency is the ratio of the total exergy output to the hybrid system exergy input, and its value hex;sys is calculated as:

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hex;sys

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 T 8   0 Eþ  1  Qg;ev þ Qac;ev  W > > > T > ch > ðsummerÞ > < Exng ; ¼   >   T0 > > >  Q E þ 1  þ Q þ Q  W g;co ac;a ac;co > > Thw : ðwinterÞ Exng (22)

where E is the power output for the ICE, and Qac;a and Qac;ev are the heat absorbed from the absorber and evaporator, respectively, in the AHP. Furthermore, Thw is the average temperature of the hot water used in winter, while Tch is the average temperature of the chilled water in summer. The exergy was analyzed in the reference state, with 101.325 kPa and 298.15 K in summer, and 101.325 kPa and 278.15 K in winter.

4. Results and discussion 4.1. Validation of models There are three main models including ICE, AHP and GSHP during the simulation of thermodynamic performance in the Engineering Equation Solver (EES) software, in which the ICE model, especially the power generation efficiency, is fitted according to the data from a series of engines of the AB group [32]. Herein, the simulation models of AHP and GSHP are validated as follows: Because three are 20 states in the internal flows of the AHP, the average relative errors between our simulated parameters and the reference values in Ref. [35] at the same conditions are employed to evaluate its error, and each kinds of parameters are shown in Table 3 [33]. The average relative error of the mass flow rate appears to be somewhat larger, but considering its small magnitude, which varies from 0.0034 to 0.248 kg/s, the average relative error is acceptable, which validates the veracity of the absorption chiller model. By simulating, the COP of absorption chiller is obtained as 1.058 while slightly higher than the reference value which is 0.9402. The root-mean-square error is applied to analyze the simulation error of the GSHP model with the Ref. [36] because there are different work states to be verified. Table 4 listed the root-meansquare error of the GSHP's COP at the different inlet preheat temperature of the ground water into the GSHP. It can observed that the root-mean-square error is approximately 4.63%, which is acceptable in the thermodynamic analysis.

Table 3 Average relative error of the AHP simulation model with the Ref. [35]. Mass flow, kg/s

Enthalpy, kJ/kg

Concentration, %

Pressure, kPa

Temperature, K

12.71%

0.08%

0.80%

0.07%

0.57%

Values

Preheat temperature, K COP in the reference COP in this study Root-mean-square error

0.00 0.50 4.17 4.21 4.06 4.14 4.63%

1.00 4.27 4.22

1.50 4.33 4.32

2.00 4.38 4.43

2.50 4.42 4.48

3.00 4.45 4.495

A building located in Beijing was used as a case study for the integrated CCHP system, and the maximum and minimum loads are listed in Table 5. When the ICE operates following the lowest electricity load of 48 kW, the cooling and heating output from the AHP driven by exhausted gas and jacket water are 80 kW and 68 kW, respectively. In order to satisfy the maximum cooling and heating loads, the GSHP is introduced into the CCHP system, in which the cooling and heating capacities are 140 kW and 177 kW, respectively. At the design work conditions, the flow rates of main fluids are also summarized into Table 5. The mixed hot water, including jacket hot water and hot water from HX2, is employed to drive the AHP and assist the GSHP, and the allocation ratio has a significant influence on the performance of waste heat subsystems [31]. In summer, all of the mixed hot water is sent to the LG1 to drive the AHP, while in winter, it is allocated to the LG1, LX3, and LX1 at 30%, 10%, and 60%, respectively, in order to improve the integrated performance. And with the determined ratios, the whole system has the highest PER and exergy efficiency. When both the ICE and GSHP operate fully, the energy inputs and outputs are as displayed in Table 6. Under these working conditions, the cooling in summer and heating in winter of the GSHP are 140 kW and 177 kW, respectively, when the GSHP operates at a full load. Because some jacket water is used to assist the GSHP in winter working conditions, its COP increases from 3.75 to 4.43. Although its COP increases in winter, it is still lower than that in cooling working conditions at 5.45, and the consumed electricity generated by the ICE in summer is significantly lower than in winter. Similarly, the ICE operates at a full load and the waste heat, including jacket water and exhausted gas, is the same in summer and winter. However, the different jacket water allocation ratios and AHP COP in summer and winter working conditions result in varying cooling and heating outputs. Although the AHP COP in winter resulting from a larger ratio of HG to LG heat is larger than in summer, the cooling energy of the AHP, namely 141 kW, is larger than the heating energy in winter, namely 81 kW, because some jacket water is used to assist the GSHP. In total, the electricity and cooling in summer working conditions are 84 kW and 281 kW, respectively, while the electricity and heating in winter are 70 kW and 258 kW, respectively. Compared to the system performances in the cooling and heating working conditions in Table 6, it can be seen that both the PER and exergy efficiency in summer, at 122.0% and 31.0%, respectively, are larger than those in winter, at 109.7% and 26.0%, respectively. The main reason for the larger PER in summer is the GSHP contribution, where the COP is significantly larger than in winter. Compared to the exergy components, the electricity exergy in summer, at 84 kW, is higher than in winter, at 70 kW, and the cooling exergy exhibits no significant difference from the heating exergy. Consequently, the exergy efficiency in summer is also higher than in winter. 4.3. Adjustable ability under variable working conditions The outstanding feature of the coupling system is its wider adjustable area for satisfying the various electricity to cooling/ heating ratios of buildings. Controlling the variable operation of the GSHP can increase the system heating/cooling output. The following section discusses the performances and adjustable areas in off-design working conditions.

Table 4 Root-mean-square error of the GSHP simulation model with the Ref. [36]. Items

4.2. Design working conditions

3.50 4.46 4.48

4.3.1. Dynamic performance distributions 4.3.1.1. Adjustable areas and PER distributions. Figs. 2 and 3 display the cooling/heating and electricity adjustable areas, as well as PER distributions, in summer and winter working conditions,

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Table 5 Parameters of components. Items

Maximum loads Minimum loads GSHP capacity GSHP COP Flow rate (kg/s)

Ratios of mixed hot water

Summer

Winter

Power load, kW

Cooling load, kW

Heating load, kW

108 48

209 80 140 5.45 1.48 (states 6, 18) 0.12 (states 13, 14) 0.10 (states 16, 17) 6.71 (states a19, a20) 6.70 (states b5, b6) 1.00 (state a1) 1.55 (state b1) 3.60 (state b7) 100/0/0

240 93 177 3.75

Jacket water Exhaust gas Hot water from HX2 Outputs of AHP Outputs of GSHP LiBr solution R134a in GSHP Ground source water (LG1/HX3/HX1), %

Table 6 Simulation results under design working conditions. Items

Parameters

Summer

Winter

System input

Natural gas, kW Power to GSHP, kW GSHP, kW AHP, kW Electricity, kW COP of GSHP COP of AHP PER, % Exergy efficiency, %

299 26 140 141 84 5.45 0.99 122.0 31.0

299 40 177 81 70 4.43 0.77 109.9 26.0

System output

Performance

2.67 (states a19, a20) 6.04 (states b5, b6) 0.51 (state a1) 1.54 (state b1) 2.50 (state b7) 30/10/60

respectively. Analyzing the adjustable area in Fig. 2, it is observed to be a quadrilateral area, where no excess product exists and energy storage is not necessary. Line a1 indicates that no electricity output occurs and all generated electricity is used to drive the GSHP to produce cooling energy. Point P1 indicates that the GSHP operates fully and the total cooling energy, 174 kW, includes the GSHP and AHP outputs (at the ICE generated electricity of 26 kW). Line a2 represents that the GSHP does not operate and the constant cooling to electricity ratio, 1.24, is fixed due to the ICE and AHP construct, which is similar to the traditional CCHP system. Line a3 indicates that the GSHP power consumption increases from 0 (point P2) to

Fig. 2. PER distribution with variable load in summer.

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Fig. 3. PER distribution with variable loads in winter.

26 kW (point P3) when the ICE operates at a full load. Line a4 represents the state in which the GSHP operates at a full load of cooling mode. Furthermore, the coupling system PER changes with the variable working state. For lines a1 and a3, the increasing GSHP cooling output results in an increase in the system PER and a decrease in electricity to cooling ratio. Like line a3 shows, the PER and cooling to electricity ratio changes from 84.0% to 0.81e122.1% and 0.30, respectively. However, the increasing electricity output of the coupling system leads to a decreasing PER and an increasing electricity to cooling ratio, changes from 232.0% to 0e122.1% and 0.81, respectively, as indicated by line a4, when the GSHP operates at a full load. Moreover, this is explained by the point that the GSHP's COP is larger than that of the AHP, and the GSHP contribution is significantly higher. For line a4, the contribution of GSHP decreases from 81.4% to 50.7% with the increasing electricity outputs for building. The coupled CCHP system PER changes from 84.0% to 232.0% with the different states. Compared the adjustable areas in Figs. 2 and 3, the shapes are similar, while the adjustable capacity differs. Lines a1, a3, and a4 in Fig. 3 represent the same operational states as in Fig. 2; however, line a5 is different from a2 in Fig. 2. In the operational state in line a5, the GSHP also operates, and a 60% mixture of hot water from the jacket and HX2 is sent to HX1, whereas the GSHP does not operate in line a2 in Fig. 2. It is observed that the PER declines from 181.5% to 109.9% while the electricity to heating ratio increases from 0 to 1.26, on the line a3 in Fig. 3, with the increasing electricity for building when the heating produced by the GSHP is constant. And the higher heating produced by the GSHP, from 0 to 177 kW for the line a4 in Fig. 3, results in lower electricity for the building, 110 kWe70 kW, while the PER and heating to electricity ratio are changed from 64.5% to 115.4% and 1.26 to 0.27, respectively. In contrast, when the building demands the same electricity, the

coupling system for satisfying the larger heating demand will exhibit a higher PER. Moreover, the state with the lowest PER in winter, namely 64.5%, is the state P2 when the electricity output is 110 kW for the building and no heating output is produced from the GSHP. On the contrary, when the GSHP is operating at a full load and no electricity demand exists for the building, the system exhibits the highest PER, namely 181.5%. The adjustable areas and PER distributions of the CCHP system without GSHP, when the CCHP system operates following the building electricity load, are displayed in Fig. 4. When the electricity to cooling ratio of the building is equal to that of the CCHP system, 0.81, no excess cooling energy exists, as indicated the line a0. For other states in the adjustable area, there is an excess cooling output, and the storage unit is necessary for improving efficiency. If no storage unit exists, the system PER will decrease. Compared to the CCHP system PER with GSHP in Fig. 2, the CCHP system PER exhibits a significantly lower energy utilization efficiency, which changes from 36.0% to 88.0%. Moreover, compared to the adjustable areas, the CCHP system with the GSHP has a wider area, and the GSHP both widens the adjustable area and improves the PER. Additionally, making a comparison between the hybrid system and a CCHP system coupled with two types TES in Ref. [37]. The adjustable area of Ref. [37] with no surplus cooling exists is similar to area displayed in Fig. 2. However, the PER in reference is lower than PER of CCHP with GSHP coupling system. This can be explained that the reference only takes waste heat from ICE into consideration, but the hybrid system takes the geothermal energy and waste heat into utilization. Moreover, when the ICE and GSHP works at the full load in hybrid system, the cooling to electricity ratio, 3.33, is larger than the ratio in reference, 2.24 in strategy A. And a better cooling to electricity ratio will be achieved when the electricity demand drops with the constant output of GSHP.

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Fig. 4. PER distribution of CCHP system without GSHP.

Fig. 5. Exergy efficiency distributions with variable load in summer.

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Fig. 6. Exergy efficiency distributions with variable loads in winter.

4.3.2. Operation strategies under variable working conditions In order to analyze the operational GSHP states in the coupled CCHP system, the building loads during a typical summer/winter day using a case study are illustrated in Fig. 7.

250

Cooling demand Electricity demand

Demand, kW

200 150 100 50 0 1

3

5

7

9

11 13 Time, h (a)

15

250

17

19

21

23

Heating demand Electricity demand

200 Demand, kW

4.3.1.2. Exergy efficiency distributions. Figs. 5 and 6 display the exergy efficiency distributions in summer and winter working conditions, respectively. Compared the PER distributions in Figs. 2 and 3, the adjustable areas are similar; however, the exergy efficiency distributions are contrary to PER. The increased cooling output leads to a declining exergy efficiency, while the power output remains constant. In contrast, the exergy efficiency will be improved, while the electricity output increases, with the constant cooling output from the GSHP. The state with the highest exergy efficiency in summer, namely 36.5%, is the P2 state when the electricity output is 110 kW for the building and no cooling output from the GSHP occurs. However, the state with the lowest exergy efficiency, namely 6.6%, is the P1 state when the GSHP operates at the full load and no electricity outputs to the building. The reason for the opposite trends is that the electricity exergy is significantly higher than the cooling exergy at the same 1 kW energy output, and the greater electricity output leads to the higher exergy efficiency. Comparing Figs. 5 and 6, the exergy efficiency distributions in winter are similar to the summer working conditions. The changing exergy efficiency range is from 7.8% to 35.3%, and the state P2 also exhibits the highest exergy efficiency, while state P1 has the lowest exergy efficiency.

150 100 50

4.3.2.1. Cooling working conditions. From Fig. 7 (a), it can be observed that the cooling demand varies from 108 kW to 210 kW, while the power demand changes from 49 kW to 100 kW. Moreover, the electricity to cooling ratio ranges from 0.38 to 0.54. In order to satisfy the demands, the operational states of the ICE, AHP, and GSHP are shown in Fig. 8. It can be seen that the ICE does not

0 1

3

5

7

9

11 13 Time, h (b)

15

17

19

21

Fig. 7. Building demands during a typical summer (a) and winter day (b).

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485

Fig. 8. Operational states during the typical summer day.

operate following only the building electricity demands, which satisfies the building electricity and GSHP consumption for producing cooling energy. In terms of the GSHP, the lowest and highest power consumptions are 4 kW and 14 kW, respectively. Consequently, the ICE generates electricity from 55 kW to 106 kW. Approximately 10.8% of electricity from the ICE during the entire day is sent to the GSHP to produce cooling energy. The waste heat from the ICE following electricity generation is related to its generated electricity, so the cooling energy produced by the AHP changes congruously with the electricity generated by the ICE. The cooling demand shortage of the building is supplemented by the GSHP. During the typical day, the GSHP provides approximately 29.0% of the building cooling demand. The GSHP contributes to

improve the outputs of the traditional configurations of ICE and AHP and also increase the ratio of cooling to electricity. Furthermore, a conventional CCHP system without GSHP integration is used as a reference in order to analyze the benefits achieved by the GSHP. The reference system consists of the ICE, AHP, and nature gas boiler to supplement the heat shortage (thermal efficiency is assumed to be 80%), and an electricity grid to supplement the electricity shortage (total efficiency of natural-gasfired power plant and grid is assumed to be 40%). The ICE operates in two modes: following the thermal load (FTL) and following the electrical load (FEL) [38]. Additionally, the carbon dioxide emissions (CDE) of the three systems mentioned above can calculated by the correlation:

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CDE ¼ kng Qng þ kgrid Egrid

(23)

and the carbon dioxide emission reduction ratio (CDERR) can be expressed to:

CDERR ¼ 1 

CDEGSHP CDEref

(24)

where kng and kgrid are the CDE factor of nature gas and electricity from grid respectively, Qng and Egrid are the consumptions of natural gas and electricity respectively. CDEGSHP and CDEref are the CDEs of the hybrid system and reference system, respectively. {Gaganis, 2017 #4} The hourly primary energy consumptions of the CCHP systems with and without the GSHP in cooling working conditions are displayed in Fig. 9. It can clearly be observed that the CCHP system with the GSHP consumes the least natural gas. Compared to the coupling system, the reference CCHP system without the GSHP in the FTL and FEL operational modes consumes an additional 45.5% and 36.6% of natural gas, respectively. Two main points contribute to the benefits achieved by the coupling system: no surplus product exists in the coupling system while the reference CCHP system may produce excess electricity or heat; and the energy consumed in the coupling system is accorded with the principle of energy cascade utilization while the supplementary electricity or natural gas in the reference system results in some energy loss due to energy noncascade-utilization. Moreover, in comparisons to the reference systems with FEL and FTL, the CDERR of the hybrid system can reach 26.8% and 36.3%, respectively. These phenomena can be attributed to the utilizing of the renewable energy, which named geothermal energy.

4.3.2.2. Heating working conditions. It can be observed from Fig. 7 (b) that the heating demand varies from 108 kW to 210 kW, while

the power demand changes from 57 kW to 90 kW. And the electricity to heating ratio varies from 0.28 to 0.63. Similarly, the operational states of the ICE, AHP, and GSHP are illustrated in Fig. 10. The lowest and highest electricity consumptions of the GSHP are 11 kW and 31 kW, respectively, to provide 48 kW and 131 kW hot water for heating. In total, the electricity consumed by the GSHP accounts for approximately 21.9% of the generated electricity of the ICE during the typical winter day, and provides 51.6% of the total building heating demand. Fig. 11 displays the hourly natural gas consumptions of the CCHP systems with and without GSHP in the heating working conditions. The CCHP system with the GSHP still consumes the least natural gas while the reference CCHP system without the GSHP in the FTL and FEL operational modes consumes an additional 34.9% and 20.2% of natural gas over the coupling system, respectively. The reference CCHP system in the FTL mode consumes the greatest primary energy because there is significant surplus electricity with the higher difference between the heating and electricity demands. And compared to the system without GSHP in the FTL and FEL modes, the CDERR of the hybrid system are 53.7% and 25.7%, respectively.

5. Conclusions In this study, a hybrid CCHP system integrated with a GSHP was designed to be applied into different cooling or heating work conditions, the thermodynamic models were constructed and validated, and the performance analyses were performed under design and off-design working conditions. The analysis focusing on the adjustable ability and thermodynamic performance distribution under variable work conditions achieved the following conclusions. The waste heat of the jacket hot water and hot water from HX2 is allocated to HX1 in GSHP, LG1 in AHP and HX3 to match the working conditions with the ratios, 0%, 100% and 0% in summer, 60%, 30% and 10% in winter, respectively, and the comprehensive

Fig. 9. Natural gas consumption during the typical summer day.

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Fig. 10. Operational states during the typical winter day.

integration design improves the system performance. Under the design working conditions, the PER and exergy efficiency of the coupling system are 122.1% and 31.0%, respectively, in the cooling mode, and 109.9% and 26.0%, respectively, in the heating mode. The outstanding feature of the hybrid CCHP system is that the integration of GSHP into the CCHP system varies the electricity to cooling/heating ratio and widens the adjustable areas of electricity and cooling/heating. The GSHP has contrasting effects on the PER and exergy efficiency of the integrated CCHP system. For the hybrid system, the PER changes from 84.0% to 232.0% in summer mode, and from 64.5% to 181.5% in winter mode. And the exergy efficiency varies from 6.6% to 36.5% in cooling work condition, 7.8%e35.3% in heating work condition. Additionally, the increased cooling/heating output of the GSHP results in an increase in the systemic PER, while

leading to a decline of the exergy efficiency. In the coupling CCHP system, the ICE cannot simply operate following building loads, and the electricity to cooling/heating ratio of the building should be taken into consideration in the ICE operational strategy. For the specific case study, the GSHP consumes 10.8% and 21.9% electricity from ICE, and produces 29.0% and 51.6% cooling/heating output, respectively. A further advantage of the introduction of the GSHP is that no excess product is output or wasted, and it improves the system performance and decreases primary energy consumption. Compared to the conventional CCHP system without the GSHP, the coupling CCHP system in the specific case study can save averagely 40.6% and 39.5% of primary energy in cooling and heating working conditions, respectively.

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Fig. 11. Natural gas consumption during the typical winter day.

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