Energy and exergy analysis of novel combined cooling and power (CCP) cycles

Accepted Manuscript Research Paper Energy and Exergy Analysis of Novel Combined Cooling and Power (CCP) Cycles Hadi Rostamzadeh, Mohammad Ebadollahi, Hadi Ghaebi, Majid Amidpour, Reza Kheiri PII: DOI: Reference:

S1359-4311(16)34471-4 http://dx.doi.org/10.1016/j.applthermaleng.2017.06.011 ATE 10529

To appear in:

Applied Thermal Engineering

Received Date: Revised Date: Accepted Date:

28 December 2016 30 May 2017 4 June 2017

Please cite this article as: H. Rostamzadeh, M. Ebadollahi, H. Ghaebi, M. Amidpour, R. Kheiri, Energy and Exergy Analysis of Novel Combined Cooling and Power (CCP) Cycles, Applied Thermal Engineering (2017), doi: http:// dx.doi.org/10.1016/j.applthermaleng.2017.06.011

This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting proof before it is published in its final form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.

Energy and Exergy Analysis of Novel Combined Cooling and Power (CCP) Cycles Hadi Rostamzadeh1, Mohammad Ebadollahi2, Hadi Ghaebi1,1, Majid Amidpour2, Reza Kheiri1 Department of Mechanical Engineering, Faculty of Engineering, Mohaghegh Ardabili University, P.O.Box 179, Ardabil, Iran

Faculty of Mechanical Engineering, Department of Energy System Engineering, K.N. Toosi University of Technology, Tehran, Iran

Abstract: This paper presents energy and exergy analyses of the basic combined cooling and power (BCCP) cycle as well as three modified CCP cycles. These modified CCP cycles are brought together by an appropriate combination of the organic Rankine cycles (ORCs) and ejector refrigeration cycle (ERC) to produce power and refrigeration, simultaneously. The performance of different working fluids for each CCP cycle is investigated using isobutane as a fixed working fluid of the ERC and R123, R245fa, and isobutane as working fluids of the ORCs. Energy and exergy analyses are conducted showing that the thermal efficiency, primary energy saving ratio (PESR), exergy efficiency, and overall exergy destruction ratio can be improved by 24.5, 134, 72, and 32 % throughout the presented state-of-art modification as well as working fluid selection, respectively. Thus, selection of R123/isobutane as working fluid and CCP cycle incorporating both recuperation and turbine bleeding as cogeneration system are the most appropriate selection from thermodynamics and environment viewpoints. Moreover, exergy analysis demonstrated that the generator accounts for the major losses in the overall exergy destruction between all components. At the end, parametric study is conducted to examine the 1

Corresponding Author: E-mail: [email protected], Tel & Fax: (+98-45) 33512910

effects of different key thermodynamic parameters on performance of different cycles. It is shown that one can obtain a higher PESR by increasing of the generator pressure and ejector mass entrainment ratio or by decreasing of the evaporator pressure and condenser temperature. It is also found that increasing of the generator pressure and ejector mass entrainment ratio or decreasing of the evaporator pressure and condenser temperature will increase the thermal efficiency. Moreover, a higher exergy efficiency can also be obtained by increasing of the generator pressure as well as the ejector mass entrainment ratio.

Keywords: Combined cooling and power (CCP) cycles, Organic Rankine cycles (ORCs), Ejector refrigeration cycle (ERC), Energy analysis, Exergy analysis;

Nomenclature BCCP

basic combined cooling and power

in

inlet

CCP

combined cooling and power

e

evaporator

COP

coefficient of performance

ej

ejector

e

specific exergy

elec

electricity output

exergy rate

ERC

ejector refrigeration cycle

FFH

feed fluid hater

Ex

exergy

h

specific enthalpy

F

fuel

HE

heat exchanger

FFH

feed fluid heater

IHE

internal heat exchanger

g

generator

mass flow rate

HE

heat exchanger

ORC

organic Rankine cycle

IHE

internal heat exchanger

P

pressure

Int

intermediate

PESR

primary energy saving ratio

is

isentropic

pf

primary flow

KN

kinetical

heat transfer rate

L

loss

R

net power/refrigeration ratio

m

mixer

RCCP

recuperative CCP

n

nozzle

RTBCCP

CCP incorporating both recuperation and

net

net value

turbine bleeding s

specific entropy

ORC

organic Rankine cycle

sf

secondary flow

out

outlet

T

temperature

P

product

TBCCP

turbine bleeding CCP

pf

primary flow

Th.V

throttling valve

PH

physical

power

PT

potential

exergy destruction ratio (%)

pu

pump

ref

reference

room

cold room

y

Greek letters ??

efficiency (%)

s

superheated

μ

mass entrainment ratio

sf

secondary flow

t

turbine

Subscripts and superscripts

th

thermal

c

total

total value

v

vapor

CH

condenser chemical

cooling

cooling output

w

water

CV

control volume

0

dead state

D

destruction

1, 2, …

cycle locations

d

diffuser

1 Introduction In recent years most countries have been capitalized on their energy resources in order to meet their energy consumption needs. Such these actions lead to recording and publishing of energy data by International Energy Agency (IEA), periodically. So, improving these data and understanding of the worldwide energy consumption will come in handy in this rapid industrialization world. Therefore, using initiative of human in this industry will be as a matter of routine. On the other hand, world energy consumption is different from primary energy supply, drastically. This may be caused by loosing of energy in different process, such as energy refinement, transportation or conversion process. Among all well-known cycles, organic Rankine cycle (ORC) indicates much more satisfactory performances for lower and medium temperature sources for power production purposes. The utilized energy sources may consist of various types such as industrial waste heat, solar ponds, geothermal systems or biomass combustion. The ORC operates like ClasusiusRankine steam power plant, but uses organic fluids instead of water. This results in a high thermal efficiency as well as wide range of usability for the heat sources with different temperatures. The concept of thermodynamic analysis was introduced to characterize working condition of the organic Rankine cycle for different working fluids consideration [1, 2]. The analyses of the ORC were mainly based upon the classical laws of thermodynamics (first and second laws) for selecting more appropriate working fluids [3, 4]. In this respect, Yamamoto et al. [5] presented ORC as an environmentally-friendly system using organic substance R123 which combined circulated thermosyphon with a turbine system. They also presented a numerical simulation of the ORC for optimization of operating conditions. They concluded that R123 presents a higher turbine power than water which is a conventional working fluid for the

ORC. From the experimental analysis they made a conclusion that the ORC using R123 improves the cycle performance, considerably. Yari [6] compared performance of the different ORCs driven by high-temperature geothermal sources based on the second-law analysis. He found that the maximum first-law-efficiency of the ORC with an IHE, binary cycle with the regenerative ORC with an IHE, and ORC with flash-binary can be obtained 7.65, 15.35, and 11.81 %, respectively, using R123 as working fluid. Further efforts have been performed to introduce environmentally-friendly working fluids for the ORC in different applications [7, 8]. For example, Hettiarachchi et al. [9] presented a cost-effective optimum design criterion for the organic Rankine power cycle using low-temperature geothermal heat sources. They evaluated the optimum performance value for different working fluids including ammonia, R123, n-pentane, and PF5050. They also observed that for specific values of the evaporator and condenser temperatures, the optimization method will converge into a unique solution. Dai et al. [10] optimized the thermodynamic parameters of the ORC for different working fluids with regard to exergy efficiency by means of genetic algorithm. They demonstrated that organic fluids are much better than water in the ORC based on the low-grade waste heat sources. Safarian and Aramoun [11] proposed energy and exergy analyses of the basic as well as three modified ORCs using R113 as working fluid. They demonstrated that the ORC incorporating both turbine bleeding and regeneration has the highest thermal and exergy efficiencies for the proposed working fluid. On the other hand, ejector refrigeration cycle (ERC) is of high significance due to its simple structure, operation condition, and most importantly, running by low-grade sources. In recent studies, ejector has been introduced as an enhanced refrigeration driven device due to its specific purposes which it follows. With this respect, Chen et al. [12] studied working characteristics of

ejector using R141b, R245fa, and R600a as working fluids. They showed that different working fluids perform distinctively different in the ejector refrigeration system in which R141b presents the highest COP compared to the others. Meanwhile, Ersoy et al. [13] investigated the average performance of the solar-powered ejector cooling-system (SECS) using the evacuated-tube collector and R123 as refrigerant in the southern region of Turkey. They found that the maximum overall coefficient of performance and cooling capacity were 0.197 and 178.26 W/m2, respectively. In order to use the advantages of the ejector refrigeration cycle more sufficiently, selection of different suitable working fluids has been taken into account in recent studies. It is demonstrated that among all working fluids, R134a presents the highest COP for the ERC [1416]. However, isobutane due to its high COP and low fluid flammability is recommended as an interesting choice. Smierciew et al. [17] presented an experimental investigation on the ejector air-conditioning system driven by low-temperature solar heat sources. They demonstrated that ejector refrigeration systems are efficiently competitive with the absorption ones. They also suggested isobutane as an appropriate working fluid of the ERC owning COP of 0.19. More recently, Saleh [18] analyzed the ERC driven by low-temperature heat sources using different working fluids. A one-dimensional analysis is taken into account in this study leading to determination of the COP and ejector area ratio (EAR). He concluded that the maximum COP, mass entrainment ratio, and EAR can be evaluated 0.74, 0.99, and 16.23, respectively, using R245ca as working fluid.

Nowadays, combined cooling and power (CCP) cycles have been proposed by state-of-theart combination of thermodynamic cycles. Many explorations into the CCP cycles have been done, seeking their further applications [19-22]. For instance, Wang et al. [23] proposed a novel combination of the organic Rankine cycle and ejector refrigeration cycle for cogeneration

purposes. They added an extraction turbine between heat recovery vapor generator (HRVG) and ejector which could be driven by different resources. They analyzed the proposed cycle based on the first and second laws of thermodynamics, showing that HRVG has the biggest rate of exergy destruction. Similarly, Zheng and Weng [24] proposed a combined power and ejector refrigeration cycle based on a novel integration of the ORC and ERC using low-temperature heat sources. In their proposed cycle, the ejector was driven by the exhaust of the turbine to produce power and refrigeration, simultaneously. With R245fa as a working fluid, thermal and exergy efficiencies were obtained 34.1 and 56.8 %, respectively. Habibzadeh et al. [25] conducted simulation of the combined power and ejector refrigeration system using five different working fluids (i.e., R123, R141b, R245fa, R600a, and R601a) based on the first and second laws of thermodynamics. They showed that R601a has the highest thermal efficiency and lowest exergy destruction. They had also presented optimum values for key parameters of cycle, namely, the inlet pressure of the pump, the evaporator temperature, and etc. Xu et al. [26] combined a Brayton cycle and a transcritical ejector refrigeration cycle (ERC) to presented a modified CCHP system, using supercritical CO2. Energy and exergy analyses of the proposed system were conducted showing that the exergy efficiency of the modified system is (10.4-22.5) % higher than that of the basic one. Furthermore, the parametric study of the system demonstrated that an increase in the turbine inlet temperature increases the power output and exergy efficiency, considerably. Similarly, Wang et al. [27] employed an ejector expander instead of the expansion valve to improve the performance criteria of previous CCHP systems by reducing irreversibility of whole system. In the presented CCHP system by this group, solar energy is used as a renewable heat source to solve environmental problems. The results indicated that the thermal and exergy efficiencies can be decreased by increasing of the turbine inlet pressure, ejector inlet

temperature, and ejector back pressure. More recently, Megdouli et al. [28] combined vapour compression refrigeration cycle (VCRC) and ERC to introduce a new hybrid VCRC (HVCRC). Energy, exergy, and parametric studies were carried out using CO 2 as a refrigerant. The results of this group demonstrated that the COP of the HVCRC is increased 25 %, while input power reduced by 20 % compared to that of the VCRC for the same cooling capacity. Furthermore, Ghaebi et al. [29] enhanced the performance of the traditional combined power and ejector refrigeration (TCPER) cycle by employing of an ejector expander in place of expansion valve based on the first- and second-law-efficiency. They demonstrated that the use of an ejector in place of the expansion valve in the TCPER cycle can increase the cooling capacity up to 9.5 %, while the net produced work remains nearly constant.

In the above papers, no combination of the ORC and ERC based on the condenser and generator components integration has been mentioned. However, it is very important to take this factor into consideration when one would enhance the performance criteria of a system. This novel methodology is considered in this paper conjoining condenser and generator of ORC and ERC as integrated components, respectively. This study outlines a basic combined cooling and power (BCCP) cycle as well as three modified ones. The modified CCP cycles consist of recuperative CCP cycle, CCP cycle incorporating with turbine bleeding, and both of them. The proposed cycles can supply both power and refrigeration, simultaneously. Since power/refrigeration ratio (R) in some cases is larger than unit, the proposed cycles are also suitable for power-dominant production purposes. In order to have a better understanding from the proposed cycles, 1st and 2nd classical laws have been investigated using three different working fluids (i.e., R123, R245fa, and isobutane) and a fixed one (isobutane) for the ORC and ERC, respectively. In addition to different working fluids

consideration, comparison of different cycles has been also conducted. Finally, sensitivity analysis of key parameters of the cycles on the performance of the systems have been investigated, theoretically.

2 Systems evaluation The schematic of cycles and their corresponding P-h diagrams for the basic CCP cycle as well as three modified CCP cycles are shown in Figs. 1-4, respectively. These cycles are constructed from an appropriate combination of the ORCs and ERC which generate required cooling capacity as well as power, simultaneously. The applicable components of the proposed cycles are: generator, evaporator, condenser, pumps, turbine, ejector, heat exchanger (HE), internal heat exchanger (IHE), feed fluid heater (FFH), and throttling valve. Fig. 1(a) shows the schematic of basic CCP (BCCP) cycle, which is constructed from a basic ORC and a standard ERC. The detail process of power sub-cycle is as follows: 

Process (1-2): The saturated vapor leaves the generator and enters to the turbine. This vapor expands the turbine and generates turbine power by losing its pressure and temperature and also converting this saturated vapor into the superheated vapor at point 2.



Process (2-3): The superheated vapor transfers a specific amount of heating capacity of the ORC into the ERC through an isobar process. This transformed heat is in direct proportion to the efficiency of the heat exchanger. In other word, in practical, a major heat loss in this component may observe.



Process (3-4): The saturated liquid leaves the heat exchanger, and then enters to the pump-2 through an isentropic process by means of pump power.



Process (4-1): The subcooled liquid at point 4 enters into the generator and completes the ORC process. The utilized working fluid gains its heating capacity during this process by interaction with the circulated steam. As a matter of fact, this steam is considered as a main impulsion for this power sub-cycle.

On the other hand, the required transferred heat for the ejector refrigeration cycle is obtained by cold and hot working fluids interaction in the conjoining heat exchanger. Actually, this heat exchanger acts as a condenser of the ORC and generator of the ERC, simultaneously. The high pressure vapor from heat exchanger outlet (primary fluid-point 5) enters into the ejector and draws the lower pressure superheated vapor of the evaporator (secondary fluid- point 11) into the ejector. These two fluids are then mixed at mixing chamber (point 7) and then enter into the condenser, where the condensation process occurs by rejecting the condenser heat into the surroundings. The liquid from the condenser itself is divided into two parts. One part goes through the throttling valve (point 10) and then enters into the evaporator which produces cooling capacity (

) for the cooling users. The rest of the liquid is pumped back to the heat

exchanger by means of pump-1, completing the ejector refrigeration sub-cycle. Fig. 1(b) illustrates the P-h diagram which is corresponded to the BCCP cycle. Once notice it, it can be seen that there are two sub-cycles which interact with two working fluids, separately. These two sub-cycles are connected by a heat exchanger which acts as a condenser of the ORC and generator of the ERC, as mentioned earlier. Fig. 2(a) illustrates a schematic of the recuperative CCP (RCCP) cycle. In this proposed cycle, the turbine outlet temperature cooled down at two stages using an internal heat exchanger (IHE). This may cause part of hot working fluid ( exchanges heat capacity with cold working fluid (

)

) which will increase the inlet

temperature of generator. Fig. 2(b) depicts the schematic of P-h diagram corresponding to the

RCCP cycle. For further modification, a feed-fluid heater (FFH) is used instead of the IHE, incorporating with turbine by means of an extra extraction, called CCP cycle incorporating with turbine bleeding (TBCCP) (Fig. 3(a)). Ideally, the outlet of this feed-fluid heater is a saturated liquid at a constant pressure through of it. Furthermore, the schematic of P-h diagram of the TBCCP cycle is depicted in Fig. 3(b) corresponding to its cycle in Fig. 3(a). Finally, Fig. 4 illustrates the schematic of the modified CCP cycle incorporating both recuperation and turbine bleeding (RTBCCP) with its corresponding P-h diagram. Obviously, this cycle has a mixed behavior of the previous ones, causing to indicate much better performance compared to the previous ones based upon 1st and 2nd laws of thermodynamics.

3 Thermodynamic modelling of systems 3.1. Thermodynamic assumptions

For thermodynamic modeling of the proposed cycles, a thermodynamic code is arranged in Engineering Equation Solver (EES) software which is written based on specific thermodynamic assumptions. These thermodynamic assumptions are as follows: 

All processes included in the cycle are at steady state.



Kinetic energy at the inlet and outlet of all components are negligible.



Pressure and heat losses inside of the evaporator, generator, condenser, feed-fluid heater, pumps, internal heat exchanger, and ducts are negligible.



The reference state pressure and temperature are 0.101 MPa and 293 K, respectively.



The inlet and outlet pressures of the vapor and water of heat exchangers are set at reference pressure.



Isentropic efficiency of the turbine and pumps are assumed 90 and 100 %, respectively.



Coefficient accounting for losses in the conjoining heat exchanger is assumed to be 80%.



In the mixing chamber of the ejector, constant pressure assumption is taken into the consideration.



There is no heat transfer between the ejector and surroundings.



Flow through the expansion valve is assumed to be an isenthalpic flow.



Working fluid at the outlet of the generator and conjoining heat exchanger in the ERC side are assumed at saturated vapor state. In contrary, working fluid at the outlet of the condenser and conjoining heat exchanger in the ORC side are assumed to be in saturated liquid state.



Ejector nozzle, mixer, and diffuser efficiencies are 80, 95 and 80 %, respectively [29].

3.2. Energy analysis

Based upon the first law of thermodynamics, conservation of mass and energy laws can be applied to each component to specify the properties of all states. These governing equations for energy analysis of a cycle can be written as follow:

m

in

  m out  0

(1)

(2)

The primary energy saving ratio (PESR) is defined as the energy of the fuel saved from the cogeneration system with respect to the conventional one [30, 35]. The PESR can be used to compare the multigeneration systems against reference system with the same energy products. The PESR of the cogeneration system incorporating a turbine and evaporator for power and cooling production purposes, respectively, can be expressed as [30-35]:

where,

is reference electrical efficiency for separate electricity production,

is reference thermal efficiency for thermal bottoming cooling power generation mode, and

is reference coefficient of performance for cooling production purposes.

In general, two possible models for CCP configurations can be taken into account for the cooling generation: the separate (or parallel) cooling generation mode and the bottoming (or series) cooling generation mode [32]. In the separate cooling generation mode, the cooling side is decoupled from the power (topping) cycle. However, in the bottoming cooling generation mode, the cooling side is cascaded to the topping power unit. In this case, it is possible to consider an electrical bottoming cycle or a thermal bottoming cycle. In this study, the thermal bottoming cycle model is considered since fuel as primary energy is supplied in terms of thermal energy [32]. For the cooling sub-cycle, the single effect absorption chiller is considered as reference cooling group, where the cooling sub-cycle is fired by heat produced by the thermal bottoming [32]. Practical values for the electrical reference efficiency and reference COP are set to 0.25 and 0.5, respectively [31, 33]. On the other hand, the thermal reference efficiency considered 0.51 [35].

Applying above mentioned equations for each control volume, the flow parameter can be calculated for each cycles. In this section, mathematical manipulation of BCCP cycle is presented in more detail. The energy balance equation for different elements of the proposed cycle can be expressed as follows:

For vapour generator: =

(14)

To calculate the heat exchanger duty in ORC, we have: =

(15)

while the heat exchanger duty in ERC can be calculated as: = The conjoining heat exchanger efficiency

(16)

can relate two above heat transfer rates as:

=

(17)

The isentropic efficiencies of the turbine and pumps can be expressed respectively as: (18) (19) (20) To calculate output power of the turbine as well as the input power of the pumps following relations can be respectively expressed: =

(21) =

(22)

=

(23)

For evaporator, the cooling output can be expressed as: =

(24)

For condenser: =

(25)

For the throttling valve: (26) Thus, the net power of the BCCP cycle can be expressed as: (27) The performance of the BCCP cycle can be evaluated in terms of thermal efficiency which is defined as the sum of output energies to the input ones: (28)

Considering all above calculated performance parameters, PESR of the BCCP cycle can be expressed as:

For more simplicity, some of the required thermodynamic relations in energy analysis of the RCCP, TBCCP, and RTBCCP cycles are listed in Table 1.

It is worth mentioning to introduce power/refrigeration ratio (R) which defines the combined cycle capability in producing of cooling and power in compared with each other:

(30) In the proposed combined ORC-ERC cycles, one of the main problem is ejector modelling. In general, two models are common: constant-pressure mixing model and constant-area mixing model [31, 36, 37]. In this study, a comprehensive thermodynamic modelling of ejector based on the constant-pressure model is presented in Appendix A [31]. In addition, a schematic diagram of the ejector with three main parts (i.e., nozzle, mixing chamber, and diffuser) are depicted in Fig. 5. The primary flow at state (pf) enters the ejector nozzle and expands to a mixture at state (n) with a nozzle efficiency of

. Meanwhile, the secondary flow enters to the ejector at state (sf),

and then two streams mix at constant pressure in the ejector to state (m). The mixed flow then flows through the ejector diffuser with a diffuser efficiency of to

, where it recovers its pressure

at outlet. The mass entrainment ratio of the ejector is another essential parameter in ejector

refrigeration cycles. This parameter is defined as the mass flow rate of the secondary flow ( m sf ) to that of the primary flow ( m pf ): (31) in which both m sf and m pf are in kg/s.

3.2. Exergy analysis

In analysis of the second law of thermodynamics it is necessary to discuss on an important parameter called exergy. Exergy of a system is defined as the maximum theoretical useful work

which can be obtained as the system interacts to equilibrium. Exergy of a system is not conserved. It can be destroyed, but not whole of it. If a system comes into an equilibrium state with its environment spontaneously, then in this case its exergy will be completely destroyed. Thus, it can be concluded that the value of exergy cannot be negative. The general form of the exergy balance equation for any component of the system can be expressed as: (32)

Exergy destruction is also a vital parameter in exergy analysis which indicates the main source of losses in the system by means of each component. In the absence of magnetic, electrical, nuclear, and surface tension effects, the total exergy rate of system ( components: physical exergy rate ( and chemical exergy rate (

), kinetic exergy rate (

) can be divided into four ), potential exergy rate (

),

) [38]:

E total  E PH  E KN  E PT  ECH

(33)

and the specific exergy can be expressed as:

etotal  e PH  e KN  e PT  eCH

(34)

in which

e

E m

(35)

which is more convenient to work with. The sum of the kinetic, potential, and physical exergies of a system is called thermo-mechanical exergy.

Because our system and its components are at rest relative to the environment, the kinetic and potential exergy rates can be set into zero. Also, because of rare chemical reactions happening and its negligible value compared with the physical exergy in organic materials, one can neglect the rate of chemical exergy. The specific physical exergy of a closed system for different working fluids can be calculated from the following equation:

e PH  h  h0 T 0 (s  s 0 )

(36)

in which h , s are specific enthalpy and entropy of the substance, respectively, and h0 , s 0 are those parameter at reference state (dead state) of known pressure and temperature of ( P0 ,T 0 ). In exergy analysis, two useful concepts by the name of fuel and product can be introduced. According to their definitions, product represents desired produced results and fuel represents the expended resources to generate the product. Both of these concepts can be expressed in terms of exergy. Considering that, let us express the exergy rate balance for the element i of a system as [38]:

E Fi  E Pi  E Di  E Li

(37)

in which E Pi and E Fi are the rate of generated product and supplied fuel of element i, respectively. On the other hand, E Li and E Di are the rate of exergy loss and exergy destruction of element i, respectively. It is crucial to notice that the fuel exergy rate is just a concept and does not have any relation with real fuels such as coal, gas, etc. Once assumed that all the outer surface of the system is at constant reference temperature, the rate of exergy loss can be neglected. This assumption is so reasonable in many cases [38]. The same equation for the total of a system can be written as:

E Ftotal  E Ptotal  E Dtotal  E Ltotal

(38)

whereas components are the corresponding ones in a system. i The exergy efficiency of element i (Ex ) is defined as the ratio of product exergy of element

i ( E Pi ) to the fuel exergy of the same element ( E Fi ): i Ex  E Pi E Fi

(39)

Obviously, a higher exergy efficiency result in a better performance. The total exergy efficiency of the system can be expressed as same as Eq. (39): total Ex  E Ptotal E Ftotal

(40)

total total in which E P and E F are the total exergy of product and fuel rate, respectively.

Another important parameter in related to the system inefficiencies is the exergy destruction ratio, which is defined as the ratio of exergy destruction of element i ( E Di ) to the overall exergy total

destruction of the system ( E D ):

y Di  E Di E Dtotal

(41)

This identified parameter has a deep impact on the aforementioned systems efficiency. Consider the BCCP cycle and employ Eqs. (32-41), then the exergy balance equations for each component of this cycle can be expressed as: For vapour generator: =

(42)

For heat exchanger: =

(43)

Fore turbine: =

(44)

For pumps: (45) (46) For evaporator, assuming

as cold room temperature, then: =

-

(47)

For condenser: =

(48)

For ejector: =

(49)

The total exergy destruction rate of the BCCP cycle is the sum of the exergy destruction rate in each component: (50) The performance of the BCCP cycle based on the exergy analysis can be defined as: (51) Table 2 expresses some of these exergy balance equations for different components of the RCCP, TBCCP, and RTBCCP cycles.

4 Model validation and accuracy Throughout this investigation, the presented theoretical work has been validated with the experimental one. For the case of simple ejector refrigeration cycle (ERC), the theoretical results were validated by previous published data from reference [17]. Under the same conditions and assumptions and using isobutane as a working fluid of the ERC, the present work shows a very good agreement with the results of Ref. [17]. This calculated accuracy is believed to be sufficient in most engineering applications. The results of this validation have been summarized in Table 3.

5 Working fluid selection Selection of an appropriate working fluid can be the number one issue in improving of the thermal efficiency, PESR, and exergy efficiency of the combined cooling and power cycles. An appropriate working fluid can be chosen based upon two important factors: having the highest efficiency and being eco-friendly working fluid. Therefore, a trade-off between these two factors needs to be reached.

In this paper, three appropriate working fluids (i.e., R123, R245fa, isobutane) for the ORCs and a fixed one (isobutane) for the ERC are suggested, theoretically. Some of these working fluids properties are listed in Table 4. Once notice them, it can be said that among all presented working fluids, isobutane/isobutane is the best choice from the environment viewpoint.

6 Results and discussion This section presents the obtained results from energy and exergy analyses of the basic CCP cycle and three modified ones using the aforesaid pairs of working fluids. But presenting all

obtained results for different working fluids seem to be useless and cumbersome. Therefore, some of the thermodynamic results for the best case of study (i.e., R123/isobutane as working fluid and RTBCCP cycle as the best system) have been presented. For this purpose, some input flow parameters are required which will come in handy for both energy and exergy analyses (Table 5). In addition, the input heat source value (generator duty) is fixed in order to have a realistic comparison between different cycles and working fluids.

Table 6 presents the thermodynamic properties for each state of the BCCP, RCCP, TBCCP, and RTBCCP cycles, using the most appropriate pair of working fluids, namely, R123/isobutane. These thermodynamic properties include the temperature, pressure, enthalpy, entropy, and mass flow rate which are essential for the next steps.

6.1. Energy evaluation of systems

Table 7 presents the results of energy analysis of different proposed cycles with different aforementioned working fluids used in the ORC and a fixed one (i.e, isobutane) in the ERC. Among all presented cycles, the BCCP cycle is the simplest cycle and hence indication of low value for some thermodynamic key parameters was predictable. It is found that among all applied working fluids, R123/isobutane and isobutane/isobutane have the highest and lowest thermal efficiency, respectively. Thus, R123/isobutane can be a suitable choice for all proposed CCP cycles, based on the first-law of thermodynamics. The maximum value of the thermal efficiency is corresponded to the RTBCCP cycle by 36.17 % using R123/isobutane, while the minimum value of the thermal efficiency is corresponded to the BCCP cycle by 29.05 % using isobutane/isobutane. As a result, appropriate working fluid selection as well as the presented state-of-art modification can improved the thermal efficiency up to 24.5%. Meanwhile, among

all applied working fluids, R123/isobutane and isobutane/isobutane also have the highest and lowest PESR, respectively. This indicates that R123/isobutane can be a suitable choice for all proposed CCP cycles, based on the energy saving viewpoint. The maximum value of the PESR is corresponded to the RTBCCP cycle by 0.3031 using R123/isobutane, while the minimum value of the PESR is corresponded to the BCCP cycle by 0.1295 using isobutane/isobutane. Thus, appropriate working fluid selection as well as the presented state-of-art modification can increase PESR nearly 134 %. Moreover, the minimum power efficiency is corresponded to the BCCP cycle by 12.09 % using isobutane/isobutane, while the maximum power efficiency is obtained for the RTBCCP cycle by 20.85 % using R123/isobutane. This also reveals that selection of R123/isobutane as working fluid and RTBCCP cycle as cogeneration system can improve the power efficiency over 73 % which is very considerable. However, since the working fluid of the ERC is fixed, therefore, the coefficient of performance (COP) for refrigeration purposes is fixed at 0.2419 for all cycles and working fluids.

One of the main drawback of isobutane as working fluid of the ORC is that even though it is an environmentally-friendly working fluid, but it produces the highest condenser load which is not desirable in cogeneration cycles. Therefore, increasing of net power and cooling capacity and decreasing of condenser duty can be the main solution for improvement of CCP cycles.

The power/refrigeration ratio (R) is another important parameter which is calculated for different working fluids and cycles, given in Table 7. It is demonstrated that among all suggested working fluids, only R123/isobutane produces more power output than the cooling output (R>1) in all proposed CCP cycles. Therefore, this pair of working fluids can be the best selection for powerdominant production purposes. On the other hand, isobutane/isobutane is the most appropriate

pair

of

working

fluid

for

refrigeration-dominant

production

purposes,

since

its

power/refrigeration is the lower than unit. Thus, the proposed cycles with their corresponding working fluids are appropriate for both power- and refrigeration-dominant productions purposes, depending on their arrangements. However, the proposed CCP cycle by the previous work (Ref. [24]) was more appropriate for refrigeration-dominant production, since power/refrigeration ratio had been set into 0.5. It is also shown that throughout this state-of-art modification, the rejecting heat by the condenser to the surroundings has been reduced, too.

6.2. Exergy evaluation of systems

Table 8 gives the exergy analysis results for the various CCP cycles and applied working fluids for different components as well as whole systems. The maximum exergy efficiency is calculated for the RTBCCP cycle by 51.71 % using R123/isobutane, while the minimum exergy efficiency is calculated for the BCCP cycle by 30.03 % using isobutane /isobutane. This demonstrates that presented modification and working fluids arrangement can increase the second-law-efficiency over 72 % which is also very considerable in cogeneration systems. Thus, selection of R123/isobutane as working fluid and RTBCCP cycle as cogeneration system are also the best selections from the second law of thermodynamics point of view. Moreover, the overall exergy destruction rate is increased nearly by 32 % throughout this successive modification and working fluid selection.

As it can be seen from Table 8, among all components, generator has the main contribution in the exergy losses for all proposed working fluids. Therefore, applying appropriate methods are necessary to reduce this component irreversibility. One method is to modify the proposed BCCP cycle based on the power sub-cycle, as presented in this study. Doing this so, the irreversibility

of this component is decreased by 32.4, 19, and 8.12 % for R123/isobutane, R245fa/isobutane, and isobutane/isobutane, respectively.

6.3. Sensitivity analysis of key thermodynamic parameters

As mentioned before, investigation of parametric study for different working fluids is a cumbersome and useless effort. Therefore, the effect of some key parameters on the BCCP cycle as well as three modified CCP cycles are investigated using R123/isobutane as the most appropriate working fluid.

6.3.1. The effect of generator pressure on the CCP cycles

Fig. 6(a) illustrates the variation of the thermal and exergy efficiencies with the generator pressure for different CCP cycles. As shown, an increase in the generator pressure increases inlet energy of the turbine, causing to extract more vapor from the turbine. This will result in a rise of the thermal efficiency, since the heat source capacity (generator duty) are remained constant. Meanwhile, the exergy efficiency is increased as the generator pressure increases. The logical reason behind this phenomenon is that an increase in the generator pressure has no effect on the supplied maximum theoretical work of the system, while increases the produced maximum theoretical work of the whole system, simultaneously. As a result, an increase in the generator pressure improves the performance of the proposed CCP cycles. Fig. 6(b) presents the variation of the ORC efficiency and PESR versus of the generator pressure for different cycles. As it is seen, increasing of the generator pressure results in a rise of the produced power and a fall in the refrigeration capacity, simultaneously. Therefore, the ORC

efficiency increases. Moreover, an increase in the generator pressure also increases the PESR, since the useful energy is saved more efficiently through the system operation. Fig. 6(c) has been plotted to show the effect of the generator pressure on the power/input heat and refrigeration/input heat ratios. As shown in this figure, an increase in the generator pressure increases the turbine expansion ratio, and hence the turbine produces more power. This results in increase of the net produced power. On the other hand, mass flow rate of the motive vapor (primary flow) is decreased by increasing of the generator pressure which will decrease the refrigeration/input heat ratio under constant external conditions. Fig. 6(d) shows the effect of the generator pressure on the overall exergy losses in different proposed cycles. As it can be figured out, throughout this state-off-the-art modification exergy destruction of cycles has been come down. This is because, the maximum produced work of the generator is decreased by increasing of the generator pressure under a constant maximum supplied work, and hence the overall exergy destruction rate is decreased, slightly. Using R123/isobutane as working fluid, the overall exergy destruction rate is fallen into the range of (18.45-23.78) kW, which is decreased by 22.41 % throughout this successive modification.

6.3.2. The effect of evaporator outlet pressure on the CCP cycles

Fig. 7(a) presents the variation of the thermal efficiency and exergy efficiency with various evaporator outlet pressures for four different aforementioned cycles, using R123/isobutane. As shown in this figure, the thermal efficiency is decreased with the evaporator outlet pressure for all different cycles, since the cooling capacity of evaporator decreases with constant net produced power and generator duty during cycle operation. Meanwhile, the maximum theoretical

produced work is decreased by a fall of the cooling capacity, and hence exergy efficiency is almost decreased by the evaporator outlet pressure, so slightly. Fig. 7(b) has been illustrated to show the effect of the evaporator outlet pressure on the ORC efficiency and PESR. Since the evaporator is not involved in the power sub-cycle, so the evaporator outlet pressure does not affect the ORC efficiency. At the same time, the PESR is decreased by an increase in the evaporator outlet pressure, and hence the system performance improves. Fig. 7(c) shows the effect of the evaporator outlet pressure on the power/input heat and refrigeration/input heat ratios, simultaneously. As it is mentioned earlier, increase in the evaporator outlet pressure decreases the refrigeration while has no effect on the produced power. As a result, the power/input heat ratio remains constant, while the refrigeration/input heat ratio is decreased under a constant heat source value. To investigate the effect of the evaporator outlet pressure on the overall exergy destruction rate for different cycles, Fig. 7(d) is presented. As illustrated in this figure, an increase in the evaporator outlet pressure increases the overall exergy destruction rate, so slightly, since the maximum theoretical produced work is decreased.

6.3.3. The effect of condenser temperature on the CCP cycles

Figs. 8(a) and 8(b) have been plotted to explain the condenser temperature variation with the thermal and exergy efficiencies for various proposed cycles. In this respect, one can obtain a higher thermal efficiency by decreasing of the condenser temperature. Actually, this is the matter of cogeneration cycles which must be taken into account by reducing heat losses during the condensation process. On the other hand, condenser temperature has no effect on the exergy

efficiency, since condenser does not contribute in exergy efficiency, theoretically. Since the condenser operates at the refrigeration sub-cycle, thus its temperature has no effect on the ORC efficiency which is shown in Fig. 8(b). Increasing of the condenser temperature also increases the inlet energy of the evaporator, slightly. This will result in a fall of refrigeration, and hence the PESR decreases for a fixed power production. Fig. 8(c) has been illustrated to show the effect of the condenser temperature on the power and refrigeration for different cycles. As it can be seen, an increase in the condenser temperature does not affect the produced power, since power is related to power sub-cycle. Thus, the power/input heat ratio is constant with the condenser temperature. As mentioned earlier, refrigeration is decreased when the condenser temperature increases. As a result, the refrigeration/input heat ratio is decreased under constant input heat supplementation. Fig. 8(d) is plotted to explain the effect of the condenser temperature on the overall exergy destruction rate for different cycles. As shown in the figure, increase in the condenser temperature decreases the overall exergy destruction rate, so slightly. The logical reason behind this behavior is that an increase in the condenser temperature decreases supplied maximum work of this component, and thus the overall exergy destruction decreases, gently.

6.3.4. The effect of ejector mass entrainment on the CCP cycles

Fig. 9(a) shows the effect of mass entrainment ratio on the thermal and exergy efficiencies for the BCCP cycle as well as three modified ones, namely: the RCCP cycle, the TBCCP cycle, and the RTBCCP cycle. An increase in the mass entrainment ratio increases mass flow rate of the secondary flow, and hence cooling capacity increases. As a result, the thermal efficiency is increased, considerably. However, the exergy efficiency is constant with the mass entrainment

ratio for all cycles. This is because, an increase in the mass entrainment ratio increases maximum supplied work which is independent from the exergy efficiency. Fig. 9(b) has been illustrated to show the effect of the ejector mass entrainment ratio on the ORC efficiency and PESR. Since the ejector is not involved in the power sub-cycle, so the ejector mass entrainment ratio does not affect the ORC efficiency. At the same time, the PESR is increased by an increase in the ejector mass entrainment ratio, and hence the system performance improves. Fig. 9(c) has been illustrated to show the effect of the ejector mass entrainment ratio on the power/refrigeration and refrigeration/input heat ratios, simultaneously. An increase in the ejector mass entrainment ratio increases refrigeration and does not vary the net produced power, totally. This will result in a rise of the refrigeration/input heat ratio and a fall in the power/refrigeration ratio. It is also worthwhile to mention here that the ejector mass entrainment ratio does not affect the overall exergy destruction rate so considerably, too.

7 Conclusions A novel basic combined cooling and power (BCCP) cycle as well as three modified CCP cycles based on the first and second laws of thermodynamics were proposed in the present work. These novel cycles were achieved by an appropriate integration of organic Rankine cycles (ORCs) and ejector refrigeration cycle (ERC) for producing more power and cooling outputs, simultaneously. In defining of combined cooling and power cycle efficiencies, it is necessary to weight the refrigeration output to obtain meaningful values. Therefore, based upon this, definition of first and second laws of thermodynamics had been extended in this paper. In addition, sensitivity

analysis for some different key parameters of cycles was carried out. Some of the main concluded results are summarized here as follows: 

The maximum value of the thermal efficiency was corresponded to the RTBCCP cycle by 36.17 % using R123/isobutane, while the minimum value of the thermal efficiency was corresponded to the BCCP cycle by 29.05 % using isobutane/isobutane. Thus, appropriate working fluid selection as well as the presented state-of-art modification can improved the thermal efficiency up to 24.5 %.



The maximum value of the PESR was corresponded to the RTBCCP cycle by 0.3208 using R123/isobutane, while the minimum value of the PESR was corresponded to the BCCP cycle by 0.16 using isobutane/isobutane. Thus, appropriate working fluid selection as well as the presented state-of-art modification can double PESR.



The minimum power efficiency was corresponded to the BCCP cycle by 12.09 % using isobutane/isobutane, while the maximum power efficiency was obtained for the RTBCCP cycle by 20.85 % using R123/isobutane. As a result, the selection of R123/isobutane as working fluid and RTBCCP cycle as cogeneration system can improve the power efficiency over 73 %.



The maximum exergy efficiency was calculated for the RTBCCP cycle by 51.73 % using R123/isobutane, while the minimum exergy efficiency was calculated for the BCCP cycle by 30.03 % using isobutane/isobutane. This demonstrated that presented modification and working fluids arrangement can increase the second-law-efficiency over 72 %.



The overall exergy destruction rate was decreased nearly by 32 % throughout this successive modification and working fluid selection.



The coefficient of performance (COP) for all cycles and working fluids was constant at 0.2419, since the working fluid of the ERC is fixed.



R123/isobutane was introduced as the best selection for power-dominant production purposes, while isobutane/isobutane was the most appropriate pair of working fluid for refrigeration-dominant production purposes.



Among all components, generator had the main contribution in exergy losses for all working fluids.



Increasing of the generator pressure and ejector mass entrainment ratio or decreasing of the evaporator pressure and condenser temperature increased the thermal efficiency.



It was shown that one can obtain a higher PESR by increasing of the generator pressure and ejector mass entrainment ratio or decreasing of the evaporator outlet pressure and condenser temperature.



A higher exergy efficiency can be obtained by increasing of the generator pressure and ejector mass entrainment ratio.

Appendix (A): Ejector thermodynamic modelling As pointed out earlier, the ejector entrainment ratio ( ) is one of the key parameters in design of ejector for ERC which can be defined as the ejector mass flow rate of the secondary flow (state sf) divided by the motive mass flow rate of primary flow (state pf) (see Fig.(2)). For 1 kg of the mixture inside the ejector, the mass flow rate of secondary flow is

/(1+ ) kg and the mass flow

rate of primary flow is 1/(1+ ) kg. The drive steam enters the ejector and expands to suction pressure (

) with a nozzle efficiency defined as: (A.1)

in which

is corresponding isentropic state (n).

The energy balance between states (pf) and (n) is: (A.2)

Apply momentum conservation in the mixing section (n-m): (A.3)

The energy balance for the ejector as a control volume can be written as follow: (A.4)

The mixing efficiency is given as: (A.5)

where,

is the corrected form of

, in order to account for mixing section losses.

The energy balance equation between states (m) and (out) is: (A.6)

The mixture recovers pressure in the ejector diffuser with a given efficiency of: (A.7)

where

is corresponding isentropic enthalpy at the outlet of the ejector.

References [1] Hung, T.C., Shai, T.Y., and Wang, S.K. (1996) ‘A review of organic Rankine cycles (ORCs) for the recovery of low-grade waste heat’, Energy, Vol. 22, pp. 661-667. [2] Saleh, B., Koglbauer, G., Wendland, M., and Fischer, J. (2007) ‘Working fluids for lowtemperature organic Rankine cycles’, Energy, Vol. 32, pp. 1210-1221.

[3] Mago, P.J., Chamra, L.M., Srinivasan, K., and Somayaji, C. (2007) ‘An examination of regenerative organic Rankine cycles using dry fluids’, Appl. Therm. Eng., Vol. 28, pp. 9981007. [4] Maizza, V. and Maizza, A. (2001) ‘Unconventional working fluids in organic Rankine cycles for waste energy recovery systems’, Appl. Therm. Eng., Vol. 21, pp. 381-390. [5] Yamamoto, T., Furuhata, T., Arai, N., and Mori, K. (2001) ‘Design and testing of the organic Rankine cycle’, Energy, Vol. 26, pp. 239-251. [6] Yari, M. (2010) ‘Exergetic analysis of various types of geothermal power plants’, Renew. Energy, Vol. 35, pp. 112-121. [7] Tchanche, B.F., Papadakis, G., Lambrinos, G., and Frangoudakis, A. (2009) ‘Fluid selection for a low-temperature solar organic Rankine cycle’, Appl. Therm. Eng., Vol. 29, pp. 24682476. [8] Drescher, U. and Bruggemenn, D. (2007) ‘Fluid selection for the Organic Rankine Cycle (ORC) in biomass power and heat plants’, Appl. Therm. Eng., Vol. 27, pp. 223-228. [9] Hettiarachchi, H.D.M., Golubovic, M., Worek, W.M., and Ikegami, Y. (2007) ‘Optimum design criteria for an organic Rankine cycle using low-temperature geothermal heat sources’, Energy, Vol. 32, pp. 1698-1706. [10]

Dai, Y., Wang, J., and Gao, L. (2009) ‘Parametric optimization and comparative study of

organic Rankine cycle (ORC) for low-grade waste heat recovery’, Energy Convers. & Manage., Vol. 50, pp. 576-582. [11]

Safarian, S. and Aramoun, F. (2015) ‘Energy and exergy assessments of modified

Organic Rankine Cycles (ORCs). Energy Report, Vol. 1, pp. 1-7.

[12]

Chen, J., Havtun, H., and Palm, B. (2014) ‘Parametric analysis of ejector working

characteristics in the refrigeration system’, Appl. Therm. Eng., Vol. 69, pp. 130-142. [13]

Ersoy, H.K., Yalcin, S., and Yapici, R. (2007) ‘Performance of a solar ejector cooling-

system in the southern region of Turkey’, Appl. Energy, Vol. 84, pp. 971-983. [14]

Selvaraju, A. and Mani, A. (2004) ‘Analysis of a vapor ejector refrigeration system with

environment friendly refrigerants’, Int. J. Therm. Sci., Vol. 43, pp. 915-921. [15]

Varga, S., Lebre, P.S., and Oliveira, A.C. (2015) ‘Readdressing working fluid selection

with a view to designing a variable geometry ejector’, Int. J. Low-carbon Tech. Advance Access, Vol. 10, pp. 205-215. [16]

Wang, E.H., Zhang, H.G., Fan, B.Y., Ouyang, M.G., Zhao, Y., and Mu, Q.H. (2011)

‘Study of working fluid selection for organic Rankine cycle (ORC) for engine waste heat recovery’, Energy, Vol. 36, pp. 3406-3418. [17]

Smierciew, K., Gagan, J., Butrymowicz, D., and Karwacki, J. (2014) ‘Experimental

investigations of solar driven ejector air-conditioning system’, Energy & Buildings, Vol. 80, pp. 260-267. [18]

Saleh, B. (2016) ‘Performance analysis and working fluid selection for ejector

refrigeration cycle’, Appl. Therm. Eng., Vol. 107, pp. 114-124. [19]

Vijayaraghavan, S. and Goswami, D.Y. (2003) ‘On evaluating efficiency of a combined

power and cooling cycle’, J. Energy Resour. Tech., Vol. 125, pp. 221-227. [20]

Goswami, D.Y. and Xu, F. (1999) ‘Analysis of a new thermodynamic cycle for combined

power and cooling using low and mid temperature solar collectors’, J. Sol. Energy Eng., Vol. 121, pp. 91-97.

[21]

Xu, F., Goswami, D.Y., and Bhagawat, S.S. (2000) ‘A combined power/cooling cycle’,

Energy, Vol. 25, pp. 233-246. [22]

Hasan, A.A., Goswami, D.Y. and Vijayaraghavan, S. (2002) ‘First and second law

analysis of a new power and refrigeration thermodynamic cycle using a solar heat source’, Sol. Energy, Vol. 73, pp. 385-393. [23]

Wang, J, Dai, Y., and Sun, Z. (2009) ‘A theoretical study on a novel combined power and

ejector refrigeration cycle’, Int. J. Ref., Vol. 32, pp. 1186-1194. [24]

Zheng, B. and Weng, Y.W. (2010) ‘A combined power and ejector refrigeration cycle for

low temperature heat sources’, Sol. Energy, Vol. 84, pp. 784-791. [25]

Habibzadeh, A., Rashidi, M.M. and Galanis, N. (2013) ‘Analysis of a combined power

and ejector refrigeration cycle using low temperature heat’, Energy Convers. & Manage., Vol. 65, pp. 381-391. [26]

Xu, X.X., Liu, C., Fu, X., Gao, H., and Li, Y. (2015) ‘Energy and exergy analyses of a

modified combined cooling, heating, and power system using supercritical CO2’, Energy, Vol. 86, pp. 414-422. [27]

Wang, J.F., Zhao, P., Niu, X., and Dai, Y. (2012) ‘Parametric analysis of a new combined

cooling, heating, and power system with transcritical CO 2 driven by solar energy’, Appl. Energy, Vol. 94, pp. 58-64. [28]

Megdouli, K., Tashtoush, B.M., Nahdi, E., Elakhdar M., Mhimid, A., Kairouani, L.

(2016) ‘Performance analysis of a combined vapour compression cycle and ejector cycle for refrigeration cogeneration’, Int. J. Refrig., Vol. 74, pp. 517-527.

[29]

Ghaebi, H., Rostamzadeh, H., and Matin, P.S. (2017) “Performance evaluation of ejector

expansion combined cooling and power cycles”, Heat Mass Transf, doi:10.1007/s00231-0172034-3. [30]

Freschi, F., Giaccone, L., Lazzeroni, P., and Repetto, M. (2010) “Multi-objective study of

a food-industry trigeneration”, 2nd International Conference on Engineering Optimization, September, Lisbon, Portugal. [31]

Maraver, D., Quoilin, S., and Royo, J. (2014) “Optimization of biomass-fueled combined

cooling, heating, and power (CCHP) systems integrated with subcritical or transcritical organic Rankine cycles (ORCs)”, Entropy, Vol. 16, pp. 2433-2453. [32]

Chicco, G. and Mancarella P. (2007) ‘Trigeneration primary energy saving evaluation for

energy planning and policy development’, Energy Policy, Vol. 35, pp. 6132-6144. [33]

Li, M., Jiang, X.Z., Zheng, D., Zeng, G., and Shi, L. (2016) ‘Thermodynamic boundaries

of energy saving in conventional CCHP (Combined Cooling, Heating and Power) systems’, Energy, Vol. 94, pp. 243-249. [34]

Frangopoulos, C.A. (2012) ‘A method to determine the power to heat ratio, the

cogenerated electricity and the primary energy savings of cogeneration systems after the European Directive’, Energy, Vol. 45, pp. 52-61. [35]

Cho, H., Mago, P.J., Luck, R., Chamra, L.M. (2009) ‘Evaluation of CCHP systems

performance based on operational cost, primary energy consumption, and carbon dioxide emission by utilizing an optimal operation scheme’, Appl. Energy, Vol. 86, pp. 2540-2549. [36]

Kheiri, R., Ghaebi, H., Ebadollahi, M., and Rostamzadeh, H. (2017) ‘Thermodynamic

modeling and performance analysis of four new integrated organic Rankine cycles (A comparative study)’, Appl. Therm. Eng., Vol. 122, pp. 103-117.

[37]

Rostamzadeh, H., Ghaebi, H., Behnam, F. (2017) ‘Energetic and exergetic analyses of

modified combined power and ejector refrigeration cycles, Therm. Sci. Eng. Progress, Vol. 2, pp. 119-139. [38]

Bejan, A., Tsataronis, G., and Moran, M. (1996) ‘Thermal design and optimization’, Jhon

Wiley & Sons, NY, USA.

Figures: Figure 1 Schematic diagram of the: (a) BCCP cycle and (b) corresponding P-h diagram. Figure 2 Schematic diagram of the: (a) RCCP cycle and (b) corresponding P-h diagram. Figure 3 Schematic diagram of the: (a) TBCCP cycle and (b) corresponding P-h diagram. Figure 4 Schematic diagram of the: (a) RTBCCP cycle and (b) corresponding P-h diagram. Figure 5 Schematic diagram of the ejector. Figure 6 The effect of generator pressure on the: (a) thermal and exergy efficiencies, (b) ORC efficiency and PESR, (c) power/input heat and refrigeration/input heat ratios, and (d) overall exergy destruction rate for different cycles. Figure 7 The effect of evaporator outlet pressure on the: (a) thermal and exergy efficiencies, (b) ORC efficiency and PESR, (c) power/input heat and refrigeration/input heat ratios, and (d) overall exergy destruction rate for different cycles. Figure 8 The effect of condenser temperature on the: (a) thermal and exergy efficiencies, (b) ORC efficiency and PESR, (c) power/input heat and refrigeration/input heat ratios, and (d) overall exergy destruction rate for different cycles. Figure 9 The effect of ejector mass entrainment ratio on the: (a) thermal and exergy efficiencies, (b) ORC efficiency and PESR, and (c) power/refrigeration and refrigeration/input heat ratios.

P [MPa]

(a)

4

1 5

12 8

9

3

11 7

10

6 2

0.2

0.4

0.6

0.8

h [kJ/kg] (b) Figure 1

P [MPa]

(a)

6

5

1

14

7

11

10

12 13

9

8 2

4 0.2

0.4

0.6

0.8

3

h [kJ/kg] (b) Figure 2

P [MPa]

(a)

7 1 6

5

2

15

8 11

12 14

13

10

4

0.2

0.4

9 3

0.8

0.6

h [kJ/kg] (b) Figure 3

P [MPa]

(a)

9 1 7

6

8

2

17

10 13

14 16

15

12

5

0.2

0.4

0.6

0.8

4

11 3

h [kJ/kg] (b) Figure 4

Figure 5

Solid Line: Thermal Efficiency

Dash Line: Exergy Efficiency

BCCP cycle

RCCP cycle

TBCCP cycle

RTBCCP cycle

BCCP cycle

RCCP cycle

TBCCP cycle

RTBCCP cycle 55

37

50 35

34 45

33

32

40

2

2.2

2.4

2.6

Generator Pressure (MPa)

(a)

2.8

3

Exergy Efficiency (%)

Thermal Efficiency (%)

36

Solid Line: ORC Efficiency

Dash Line: PESR

BCCP cycle

RCCP cycle

TBCCP cycle

RTBCCP cycle

BCCP cycle

RCCP cycle

TBCCP cycle

RTBCCP cycle

22

0.33 0.32

21

0.31

20

0.29 0.28

19 0.27 0.26

18

0.25 0.24

17

0.23 16

0.22 2

2.2

2.4 2.6 Generator Pressure (MPa)

(b)

2.8

3

PESR

ORC Efficiency (%)

0.3

Solid Line: Power/Input Heat Ratio

Dash Line: Refrigeration/Input Heat Ratio

BCCP cycle

RCCP cycle

TBCCP cycle

RTBCCP cycle

BCCP cycle

RCCP cycle

TBCCP cycle

RTBCCP cycle

22 16.1

15.9 20 15.7 19 15.5

18

15.3

17

16

15.1 2

2.2

2.4

2.6

Generator Pressure (MPa)

(c)

2.8

3

Refrigeration/Input Heat Ratio (%)

Power/Input Heat Ratio (%)

21

Pg=2 MPa

Pg=2.5 MPa

Pg=3 MPa

Overall Exergy Destruction Rate (kW)

25

20

15

10

5

0 BCCP cycle

RCCP cycle

TBCCP cycle

(d) Figure 6

RTBCCP cycle

Solid Line: Thermal Efficiency

Dash Line: Exergy Efficiency

BCCP cycle

RCCP cycle

TBCCP cycle

RTBCCP cycle

BCCP cycle

RCCP cycle

TBCCP cycle

RTBCCP cycle

37

55

50

35

45 34

33 0.15

0.16

0.17

0.18

0.19

0.2

0.21

Evaporator Outlet Pressure (MPa)

(a)

0.22

0.23

40 0.24

Exergy Efficiency (%)

Thermal Efficiency (%)

36

Solid Line: ORC Efficiency

Dash Line: PESR

BCCP cycle

RCCP cycle

TBCCP cycle

RTBCCP cycle

BCCP cycle

RCCP cycle

TBCCP cycle

RTBCCP cycle

22

0.31 0.3

21

0.28 19 0.27 18 0.26 17

16 0.15

0.25

0.16

0.17

0.18

0.19

0.2

0.21

Evaporator Outlet Pressure (MPa)

(b)

0.22

0.23

0.24 0.24

PESR

ORC Efficiency (%)

0.29 20

Solid Line: Power/Input Heat Ratio Dash Line: Refrigeration/Input Heat Ratio

BCCP cycle

RCCP cycle

TBCCP cycle

RTBCCP cycle

BCCP cycle

RCCP cycle

TBCCP cycle

RTBCCP cycle

16.1

Power/Input Heat Ratio (%)

20.5 20

15.9

19.5 15.7 19

18.5

15.5

18 15.3

17.5 17 0.15

Refrigeration/Input Heat Ratio (%)

21

15.1 0.17

0.19

0.21

0.23

Evaporator Outlet Pressure (MPa)

(c)

Overall Exergy Destruction Rate (kW)

Pe=0.15 MPa

Pe=0.2 MPa

Pe=0.25 MPa

25

20

15

10

5

0 BCCP cycle

RCCP cycle

TBCCP cycle

RTBCCP cycle

(d) Figure 7

Solid Line: Thermal Efficiency

Dash Line: Exergy Efficiency

BCCP cycle

RCCP cycle

TBCCP cycle

RTBCCP cycle

BCCP cycle

RCCP cycle

TBCCP cycle

RTBCCP cycle 54

37

52

50

35

48

46 34 44

33

42 295

296

297

298

299

300

301

302

Condenser Temperature (K)

(a)

303

304

305

Exergy Efficiency (%)

Thermal Efficiency (%)

36

Solid Line: ORC Efficiency

Dash Line: PESR

BCCP cycle

RCCP cycle

TBCCP cycle

RTBCCP cycle

BCCP cycle

RCCP cycle

TBCCP cycle

RTBCCP cycle

22

0.31

0.3

21

0.28 19 0.27 18 0.26 17

0.25

16

0.24 295

297

299

301

Condenser Temperature (K)

(b)

303

305

PESR

ORC Efficiency (%)

0.29 20

Solid Line: Power/Input Heat Ratio

Dash Line: Refrigeration/Input Heat Ratio

BCCP cycle

RCCP cycle

TBCCP cycle

RTBCCP cycle

BCCP cycle

RCCP cycle

TBCCP cycle

RTBCCP cycle

21

Power/Input Heat Ratio (%)

20

15.9

19.5 15.7

19

18.5

15.5

18 15.3

17.5 17

15.1 295

297

299

301

Condenser Temperature (K)

(c)

303

305

Refrigeration/Input Heat Ratio (%)

16.1

20.5

Overall Exergy Destruction Rate (kW)

Tc=295 K

Tc=300 K

Tc=305 K

25

20

15

10

5

0 BCCP cycle

RCCP cycle

TBCCP cycle

(d) Figure 8

RTBCCP cycle

Solid Line: Thermal efficiency

Dash Line: Exergy Efficiency

BCCP cycle

RCCP cycle

TBCCP cycle

RTBCCP cycle

BCCP cycle

RCCP cycle

TBCCP cycle

RTBCCP cycle

42

55

41

39 50 38 37 36 45 35 34

33 32 0.25

0.27

0.29

0.31

Ejector Mass Entrainment Ratio

(a)

0.33

40 0.35

Exergy Efficiency (%)

Thermal Efficiency (%)

40

Solid Line: ORC Efficiency

Dash Line: PESR

BCCP cycle

RCCP cycle

TBCCP cycle

RTBCCP cycle

BCCP cycle

RCCP cycle

TBCCP cycle

RTBCCP cycle 0.4

22

0.38 21

20

0.34 0.32

19 0.3 18

0.28 0.26

17 0.24 16 0.25

0.27

0.29

0.31

Ejector Mass Entrainment Ratio

(b)

0.33

0.22 0.35

PESR

ORC Efficiency (%)

0.36

Dash Line: Refrigeration/Input Heat Ratio

BCCP cycle

RCCP cycle

TBCCP cycle

RTBCCP cycle

BCCP cycle

RCCP cycle

TBCCP cycle

RTBCCP cycle

1.5

21

1.4

20

1.3

19

1.2

18

1.1

17

1

16

0.9

15

0.8 0.25

0.27

0.29

0.31

Ejector Mass Entrainment ratio

(c) Figure 9

0.33

14 0.35

Refrigeration/Input Heat Ratio

Power/Refrigeration Ratio

Solid Line: Power/Refrigeration Ratio

Tables: Table 1 Thermodynamic equations of the RCCP, TBCCP, and RTBCCP cycles. RCCP cycle Equation = , =

Component Generator

TBCCP cycle Equation = , = , =

=

Turbine

= =

Heat exchanger

,

, ,

= =

= ,

Pump-3

-

Evaporator

= = =

Condenser Throttling valve Ejector

=

FFH

,

=

,

= =

,

,

=

,

=

,

=

,

=

,

= = =

See Appendix (A) = =

IHE

,

, ,

=

=

Pump-2

=

=

=

Pump-1

RTBCCP cycle Equation = , =

= = =

See Appendix (A)

See Appendix (A) = =

,

-

,

Net power of cycle ( )

Primary energy saving ratio (PESR)

ORC efficiency ( ) Coefficient of Performance ( ) Thermal efficiency ( )

=

=

=

= =

=

= =

=

Table 2 Some of the required equations for exergy analysis of the: (a) RCCP, (b) TBCCP, and (c) RTBCCP cycles. (a) RCCP cycle Component Pump 1 Pump 2 Generator

 i

Generated product rate E P =

= = -

 i

Fuel supplied rate E F

)

= = =

-

 i

Exergy destruction rate E D = = =

-

Evaporator Ejector

=

=

HE IHE Turbine Condenser Total system

-

+

 i

= =

Evaporator Ejector HE FFH Turbine Condenser Total system

= =

= = -

Generator

= =

=

= =

-

-

-

-

+

(c) RTBCCP cycle Component

 i

Generated product rate E P

Pump-1 Pump-2

=

Generator Evaporator Ejector HE

 i

Exergy destruction rate E D -

-

=

-

+

=

-

=

= = + = - = =

-

= = = = =

 i

Fuel supplied rate E F

 i

Exergy destruction rate E D

=

=

-

-

=

=

-

= = -

=

-

=

Pump-3

-

-

-

-

=

= =

= =

-

= =

=

=

=

 i

-

=

= = = = =

Fuel supplied rate E F

-

Pump-2 Pump-3

+

= = = = =

Generated product rate E P

Pump-1

-

=

= = = = = -

(b) TBCCP cycle Component

=

=

= =

=

-

= =

=

-

=

-

+

=

-

-

=

-

= -

=

IHE FFH Turbine Condenser Total system

=

=

= = =

-

-

-

+

= = + = - = =

= = = = =

Table 4 Thermodynamic data for the applied working fluids. Selected Working fluids

Chemical formula

R123 R245fa Isobutane

C2HCl2 F3 C3H3F5 CH(CH3)2CH3

Critical temperature ( c ) 183.68 153.86 134.66

Critical pressure (bar) 36.68 36.51 36.29

Boiling point ( c ) 27.85 15.3 -11.7

GWP

ASHARAE safety code

77 1030 3

B1 B1 A3

Table 3 Model validation between present work and experiment [17]. Parameter Present work Experiment Relative error (%) 8.713 9 3.18 Generator heat 1.706 1.75 2.51 Cooling capacity 9.826 11.28 modelling. 12.89 Table load 5 Input parameters for thermodynamic Condenser 0.01701 0.02 17.64 Pump power Parameter value Mass entrainment ratio 0.2422 0.24 0.91 2932.84 Reference state temperature Coefficient of performance 0.1954 0.19

Reference state pressure Generator pressure Evaporator outlet pressure Intermediate pressure Condenser temperature Evaporator superheated temperature Cold room temperature Mass flow rate of vapor Mass flow rate of water

Turbine isentropic efficiency Pump isentropic efficiency Conjoining heat exchanger efficiency Vapor inlet temperature Vapor outlet temperature Nozzle efficiency Mixer efficiency Diffuser efficiency

0.101 2.5 0.21 1 301 286 290 0.258 2 90 100 80 600 400 80 95 80

Table 6 Flow parameters for the BCCP, RCCP, TBCCP, and RTBCCP cycles, using R123/Isobutane. BCCP cycle

Point

RCCP cycle

1

433.2

2.5

468

1.711

0.4685

433.2

2.5

468

1.711

0.4977

2

341

0.181

427.5

1.724

0.4685

341

0.181

427.5

1.724

0.4977

3

318

0.181

247

1.159

0.4685

324

0.181

414.7

1.686

0.4977

4

318.9

2.5

248.7

1.159

0.4685

318

0.181

247

1.159

0.4977

5

314

0.5428

609.5

2.324

0.1973

318.9

2.5

248.7

1.159

0.4977

6

285.4

0.21

572.8

2.324

0.1973

331

2.5

261.6

1.198

0.4977

7

285.5

0.21

573

2.325

0.2506

314

0.5428

609.5

2.324

0.1947

8

303

0.3805

595.9

2.325

0.2506

285.4

0.21

572.8

2.324

0.1947

9

301

0.3805

266.5

1.23

0.2506

285.5

0.21

573

2.325

0.2473

10

281.7

0.21

266.5

1.237

0.05328

303

0.3805

595.9

2.325

0.2473

11

286

0.21

573.8

2.328

0.05328

301

0.3805

266.5

1.23

0.2473

12

301.1

0.5428

266.8

1.23

0.1973

281.7

0.21

266.5

1.237

0.05257

13

600

0.101

3128

8.303

0.258

286

0.21

573.8

2.328

0.05257

14

400

0.101

2730

7.496

0.258

301.1

0.5428

266.8

1.23

0.1947

15

286

0.101

54.02

0.1929

2

600

0.101

3128

8.303

0.258

16

295.9

0.101

95.3

0.3348

2

400

0.101

2730

7.496

0.258

17

-

-

-

-

-

286

0.101

54.02

0.1929

2

18

-

-

-

-

-

295.7

0.101

94.75

0.3329

2

TBCCP cycle

Point

RTBCCP cycle

1

433.2

2.5

468

1.711

0.7089

433.2

2.5

468

1.711

0.7089

2

390.8

1

454.7

1.715

0.254

390.8

1

454.7

1.715

0.2221

3

341

0.181

427.5

1.724

0.4549

341

0.181

427.5

1.724

0.4868

4

318

0.181

247

1.159

0.4549

323.1

0.181

414

1.683

0.4868

5

318.3

1

247.6

1.159

0.4549

318

0.181

247

1.159

0.4868

6

384.2

1

321.8

1.369

0.7089

318.3

1

247.6

1.159

0.4868

7

385.4

2.5

323.1

1.369

0.7089

331

1

261.2

1.2

0.4868

8

314

0.5428

609.5

2.324

0.1916

384.2

1

321.8

1.369

0.7089

9

285.4

0.21

572.8

2.324

0.1916

385.4

2.5

323.1

1.369

0.7089

10

285.5

0.21

573

2.325

0.2434

314

0.5428

609.5

2.324

0.1896

11

303

0.3805

595.9

2.325

0.2434

285.4

0.21

572.8

2.324

0.1896

12

301

0.3805

266.5

1.23

0.2434

285.5

0.21

573

2.325

0.2408

13

281.7

0.21

266.5

1.237

0.05174

303

0.3805

595.9

2.325

0.2408

14

286

0.21

573.8

2.328

0.05174

301

0.3805

266.5

1.23

0.2408

15

301.1

0.5428

266.8

1.23

0.1916

281.7

0.21

266.5

1.237

0.0512

16

600

0.101

3128

8.303

0.258

286

0.21

573.8

2.328

0.0512

17

400

0.101

2730

7.496

0.258

301.1

0.5428

266.8

1.23

0.1896

18

286

0.101

54.02

0.1929

2

600

0.101

3128

8.303

0.258

19

295.6

0.101

94.11

0.3308

2

400

0.101

2730

7.496

0.258

20

-

-

-

-

-

286

0.101

54.02

0.1929

2

21

-

-

-

-

-

295.5

0.101

93.69

0.3294

2

Table 7Output data obtained from energy analysis of the BCCP, RCCP, TBCCP, and RTBCCP cycles. Cycle

Working fluid

BCCP

RCCP

TBCCP

RTBCCP

R123/Isobutane R245fa/Isobutane Isobutane/Isobutane R123/Isobutane R245fa/Isobutane Isobutane/Isobutane R123/Isobutane R245fa/Isobutane Isobutane/Isobutane R123/Isobutane R245fa/Isobutane Isobutane/Isobutane

(%) 33.58 30.97 29.05 34.47 31.89 29.18 35.5 32.5 30.02 36.17 33.13 30.1

0.2484 0.1843 0.1295 0.2682 0.2081 0.1332 0.2898 0.2232 0.1579 0.3031 0.2381 0.1602

(%) 17.69 14.47 12.09 18.74 15.55 12.18 20.03 16.3 13.22 20.85 17.08 13.32

R 0.2419 0.2419 0.2419 0.2419 0.2419 0.2419 0.2419 0.2419 0.2419 0.2419 0.2419 0.2419

18.12 14.8 12.36 19.26 15.97 12.51 20.58 16.75 13.59 21.43 17.55 13.69

18.95 15.64 13.42 20.13 16.87 13.58 21.78 17.88 14.84 22.65 18.7 14.95

0.882 0.826 0.987 0.882 0.826 0.987

0.769 0.774 0.996 0.817 0.835 1.009 0.264 0.243 0.206 0.282 0.264 0.211

0.058 0.060 0.062 0.057 0.060 0.062 0.056 0.059 0.061 0.056 0.059 0.061

82.58 85.82 88.2 81.47 84.67 88.05 80.18 83.91 87 79.35 83.14 86.9

16.38 17.02 17.49 16.16 16.79 17.46 15.9 16.64 17.25 15.74 16.49 17.23

102.7 102.7 102.7 102.7 102.7 102.7 102.7 102.7 102.7 102.7 102.7 102.7

1.107 0.8698 0.7064 1.192 0.9513 0.7166 1.294 1.007 0.7874 1.362 1.064 0.7942

Table 8 Exergy evaluation of the various cycle using R123/isobutane, R245fa/isobutane, and isobutane/isobutane. (a) BCCP cycle R123/Isobutane

R245fa/Isobutane

Isobutane/Isobutane

Component Pump-1 Pump-2 Generator Evaporator Ejector HE Turbine Condenser Total of system

(kW) 0.058 0.769 26.94 0.169 1.223 6.949 18.95 0.594 18.29

(kW) 0.058 0.769 41.75 0.654 2.726 6.949 20.76 2.196 41.75

(kW) 0 0 14.81 0.484 1.503 2.552 1.813 1.602 22.77

(%) 0 0 64.89 2.13 6.58 11.18 7.94 7.02 -

(%) 100 100 64.53 25.89 44.86 63.28 91.27 27.05 43.81

(kW) 0.060 0.774 23.45 0.176 1.271 4.57 15.64 0.560 14.98

(kW) 0.060 0.774 41.75 0.679 2.833 7.059 17.17 2.283 41.75

(kW) 0 0 18.3 0.503 1.562 2.489 1.531 1.722 26.11

(%) 0 0 70.09 1.93 5.98 9.53 5.86 6.59 -

(%) 100 100 56.17 25.89 44.86 64.74 91.08 24.75 35.87

(kW) 0.062 0.996 20.95 0.180 1.306 4.697 13.42 0.533 12.54

(kW) 0.062 0.996 41.75 0.517 2.912 7.196 14.75 2.346 41.75

(kW) 0 0 20.8 0.517 1.606 2.499 1.332 1.813 28.57

(%) 0 0 72.81 1.81 5.62 8.74 4.66 6.34 -

(%) 100 100 50.17 25.89 44.86 65.27 90.97 22.73 30.03

(b) RCCP cycle R123/Isobutane

R245fa/Isobutane

Isobutane/Isobutane

Component Pump-1 Pump-2 Generator Evaporator Ejector

(kW) 0.057 0.817 27.97 0.167 1.206

(kW) 0.057 0.817 41.75 0.645 2.69

(kW) 0 0 13.78 0.478 1.483

(%) 0 0 63.74 2.21 6.85

(%) 100 100 66.99 25.89 44.86

(kW) 0.060 0.835 24.5 0.173 1.254

(kW) 0.060 0.835 41.75 0.670 2.795

(kW) 0 0 17.25 0.497 1.541

(%) 0 0 69.19 1.99 6.18

(%) 100 100 58.69 25.89 44.86

(kW) 0.062 1.009 21.1 0.180 1.304

(kW) 0.062 1.009 41.75 0.697 2.907

(kW) 0 0 20.65 0.517 1.603

(%) 0 0 72.68 1.81 5.64

(%) 100 100 50.54 25.89 44.86

HE IHE Turbine Condenser Total of system

4.338 0.654 20.13 0.604 19.42

6.586 0.796 22.06 2.167 41.75

2.248 0.141 1.926 1.562 21.62

10.4 0.65 8.90 7.22 -

65.86 82.19 91.27 27.91 46.52

4.509 0.795 16.87 0.573 16.15

6.791 0.824 18.52 2.252 41.75

2.282 0.028 1.651 1.679 24.99

9.15 0.11 6.62 6.73 -

66.39 96.56 91.08 25.44 38.67

4.688 0.109 13.58 0.535 12.69

7.148 0.138 14.93 2.342 41.75

2.46 0.028 1.349 1.807 28.41

8.65 0.10 4.74 6.35 -

65.59 79.22 90.07 22.85 30.4

(c) TBCCP cycle R123/Isobutane

R245fa/Isobutane

Isobutane/Isobutane

Component Pump-1 Pump-2 Pump-3 Generator Evaporator Ejector HE FFH Turbine Condenser Total of system

(kW) 0.056 0.264 0.882 31.74 0.164 1.187 4.269 5.712 21.78 0.616 20.74

(kW) 0.056 0.264 0.882 41.75 0.635 2.647 6.747 8.027 23.82 2.133 41.75

(kW) 0 0 0 10.02 0.470 1.46 2.478 2.316 2.041 1.516 20.3

(%) 0 0 0 49.35 2.31 7.19 12.21 11.41 10.05 7.47 -

(%) 100 100 100 76.01 25.89 44.86 63.28 71.15 91.43 28.9 49.68

(kW) 0.059 0.243 0.826 26.95 0.172 1.243 4.468 3.931 17.88 0.580 16.92

(kW) 0.059 0.243 0.826 41.75 0.664 2.77 6.902 5.439 19.6 2.232 41.75

(kW) 0 0 0 14.81 0.492 1.528 2.434 1.508 1.726 1.651 24.15

(%) 0 0 0 61.32 2.04 6.32 10.08 6.24 7.14 6.83 -

(%) 100 100 100 64.54 25.89 44.86 64.74 72.27 91.2 26.03 40.53

(kW) 0.061 0.206 0.987 22.64 0.178 1.288 4.633 1.673 14.84 0.547 13.76

(kW) 0.061 0.206 0.987 41.75 0.689 2.872 7.098 2.104 16.31 2.314 41.75

(kW) 0 0 0 19.11 0.510 1.584 2.465 0.430 1.466 1.767 27.33

(%) 0 0 0 69.91 1.86 5.79 9.02 1.57 5.36 6.46 -

(%) 100 100 100 54.23 25.89 44.86 65.27 79.54 91.01 23.65 32.97

(d) RTBCCP cycle R123/Isobutane

R245fa/Isobutane

Isobutane/Isobutane

Component Pump-1 Pump-2 Pump-3 Generator Evaporator Ejector HE IHE FFH Turbine Condenser Total of system

(kW) 0.056 0.282 0.882 31.74 0.162 1.175 4.225 0.669 5.444 22.65 0.623 21.59

(kW) 0.056 0.282 0.882 41.75 0.628 2.62 7.018 0.812 7.018 24.78 2.111 41.75

(kW) 0 0 0 10.01 0.465 1.445 2.183 0.143 1.574 2.129 1.487 19.44

(%) 0 0 0 51.51 2.29 7.43 11.23 0.738 8.09 10.95 7.64 -

(%) 100 100 100 76.01 25.89 44.86 65.93 82.33 77.57 91.41 29.54 51.71

(kW) 0.059 0.264 0.826 26.95 0.170 1.231 4.427 0.822 3.458 18.7 0.588 17.72

(kW) 0.059 0.264 0.826 41.75 0.658 2.745 6.668 0.848 4.295 20.51 2.211 41.85

(kW) 0 0 0 14.81 0.488 1.514 2.241 0.026 0.836 1.811 1.623 23.35

(%) 0 0 0 63.42 2.09 6.48 9.59 0.11 3.58 7.75 6.95 -

(%) 100 100 100 64.54 25.89 44.86 66.39 96.92 80.52 91.17 26.62 42.43

(kW) 0.061 0.211 0.987 22.64 0.178 1.287 4.627 0.164 1.544 14.95 0.548 13.87

(kW) 0.061 0.211 0.987 41.85 0.688 2.869 7.04 0.207 1.877 16.42 2.312 41.75

(kW) 0 0 0 19.11 0.510 1.582 2.412 0.043 0.333 1.477 1.763 27.23

(%) 0 0 0 70.18 1.87 5.81 8.85 0.15 12.25 5.42 6.47 -

(%) 100 100 100 54.23 25.89 44.86 65.73 79.22 82.24 91.01 23.73 33.21

Highlights: 

A novel combined cooling and power (CCP) cycle is proposed.



Proposed cycle is modified based on the power sub-cycle concept.



First- and second-law analyses of the proposed cycles are carried out.



Several environmentally-friendly pairs of working fluids are selected for the proposed CCP cycles.



Sensitivity analysis of some key parameters of the proposed cycles is performed.