A Waste Heat-Driven Cooling System Based on Combined Organic Rankine and Vapour Compression Refrigeration Cycles

applied sciences

Article A Waste Heat-Driven Cooling System Based on Combined Organic Rankine and Vapour Compression Refrigeration Cycles

Youcai Liang, Zhibin Yu * and Wenguang Li School of Engineering, University of Glasgow, Glasgow G12 8QQ, UK; [email protected] (Y.L.); [email protected] (W.L.) * Correspondence: [email protected]; Tel.: +44-(0)-141-330-2530

Received: 11 September 2019; Accepted: 9 October 2019; Published: 11 October 2019

Abstract: In this paper, a heat driven cooling system that essentially integrated an organic Rankine cycle power plant with a vapour compression cycle refrigerator was investigated, aiming to provide an alternative to absorption refrigeration systems. The organic Rankine cycle (ORC) subsystem recovered energy from the exhaust gases of internal combustion engines to produce mechanical power. Through a transmission unit, the produced mechanical power was directly used to drive the compressor of the vapour compression cycle system to produce a refrigeration eﬀect. Unlike the bulky vapour absorption cooling system, both the ORC power plant and vapour compression refrigerator could be scaled down to a few kilowatts, opening the possibility for developing a small-scale waste heat-driven cooling system that can be widely applied for waste heat recovery from large internal combustion engines of refrigerated ships, lorries, and trains. In this paper, a model was ﬁrstly established to simulate the proposed concept, on the basis of which it was optimized to identify the optimum operation condition. The results showed that the proposed concept is very promising for the development of heat-driven cooling systems for recovering waste heat from internal combustion engines’ exhaust gas.

Keywords: organic Rankine cycle; vapour compression cycle; waste heat recovery; marine engine; cascade utilisation

1. Introduction Internal combustion (IC) engines have been the primary power source for automobiles, long-haul trucks, locomotives, and ships over the past century [1]. Over this time, periods of high fuel costs and concerns about foreign oil dependence have resulted in increasingly complex engine designs to reduce fuel consumption [2]. Although the most efficient mode of the modern large diesel engines is about 48–51% efficient in utilising the fuel energy [3], the remainder is still lost as waste to the environment through exhaust gas and jacket water. It appears that developing concepts for utilising the waste heat is important in the area of IC engine application. Current interest in reducing emissions and engine operating costs has led to the use of efficient waste heat recovery (WHR). For marine engines, the waste heat was originally used for direct heating services, including space heating in winter [4], heavy fuel oil (HFO) heating [5], ballast water heating [6], and hot water supply. However, the heating loads needed for shipboard service on a conventional vessel are usually much less than the available waste heat, resulting in a large amount of heat being unused. WHR technology for power production was then proposed to be coupled with the power propulsion plants or to meet the demand for auxiliary services without additional fuel costs and zero associated CO2 emissions. Steam-based WHR systems for both four- and two-stroke marine engines are available commercially, among others by MAN, ABB, Wärtsilä, Mitsubishi heavy industries, and

Appl. Sci. 2019, 9, 4242; doi:10.3390/app9204242 www.mdpi.com/journal/applsci Appl. Sci. 2019, 9, 4242 2 of 18

Cummins [3]. Some work on marine WHRs has been reported in the past few years. The work conducted by Theotokatos and Livanos analysed a single pressure steam Rankine cycle (SRC) for a marine engine WHR [7], finding that the WHR is more attractive to combine with two-stroke IC engine than four-stroke IC engine from an economic perspective [8]. Altosole et al. [9] compared SRC-WHRs of a new design with a retrofitting one on board a passenger ship equipped with diesel–electric propulsion. Benvenuto et al. [10] studied a dual pressure SRC as a WHR system of a two-stroke diesel engine used for the propulsion of a tanker. As the exhaust gas temperature of marine engines is usually lower than 370 ◦C, it is economically unviable to operate a steam Rankine cycle system [11]. Organic Rankine cycle (ORC) has since been proposed as an alternative solution for two-stroke engines because of their low exhaust temperatures. Singh [12] conducted a review of different waste heat recovery systems for power generation, including Rankine cycle, Kalina cycle, exhaust gas turbine system, thermoelectric generation system, and the combination of these technologies, focusing on the utilisation of WHR for the supply of mechanical/electrical power to the ship. Baldi and Gabrielii [13] made comparisons between SRC, ORC, and Kalina cycle for marine WHR, and indicated that ORC produced about 7% additional power and that the SRC and Kalina cycles produced about 5% additional power. A comparison between ORC and conventional steam Rankine cycle was conducted by Andreasen et al. [14], which showed that ORC has better performance, as higher turbine efficiencies can be achieved for the ORC compared with the steam Rankine cycle system. These studies aimed to explore the ORC system’s potential as an alternative to SRC and demonstrated its potential to improve fuel efficiency at relatively low engine exhaust temperatures. Apart from heating and electricity, cooling is also required for food preservation and air conditioning, especially for cruise ships. Thermally powered cooling technologies have gained considerable interest. Liang et al. [15–17] proposed an electricity-cooling combined system (ECCS) for waste heat recovery of marine engines, in which the condensation heat of SRC is used to heat the generator of the absorption refrigeration cycle. Some scholars proposed to integrate power with cooling cycles, in which the ammonia mixture is used as the working medium [18–20]. In addition to the absorption refrigeration cycle, vapour compression cycle refrigerator is another technology that requires mechanical power/electricity to drive the compressor. Therefore, the concept of combining ORC and vapour compression cycle (VCC) was proposed as an alternative heat-driven refrigeration technology by Prigmore and Barber [21]. Compared to the thermally powered absorption cooling technologies, the ORC-VCC has some potential advantages in terms of performance and simplicity. Furthermore, the VCC powered by an ORC can make use of the heat source throughout the year [22] to provide either cooling or electricity when cooling is not required [23], increasing the operational flexibility and improving the economic profitability. Great efforts have been devoted to the development of ORC-VCC systems since such a concept was proposed. Wali [24,25] compared the performance of solar powered ORC-VCC systems for building cooling applications with five different working fluids. R113 and FC88 were considered as the best working fluids. To reduce the system complexity, Aphornratana and Sriveerakul [26] proposed an ORC-VCC concept, of which the compressor and expander are integrated in the same unit, use the same working fluid, and share the same condenser. Bu et al. [27–29] carried out a series of investigations on the working fluid of ORC-VCC ice makers and found that R600 is the most suitable working fluid. An experimental test on the ORC-VCC system was conducted by Wang [23], and reported a coefficient of performance (COP) of 0.48. Biancardi et al. [30] introduced the design, fabrication, and test of a solar-powered organic Rankine cycle heat pump and chiller system capable of delivering 63.3 kW of heating and cooling. When R11 was used as the working fluid, the heating COP of the system ranged from 1.6 to 2.25, while the cooling COP varied between 0.5 and 0.75. In addition to the above studies, several other researchers [31–35] have also reported the performance of the ORC-driven VCC for heating purposes. Despite numerous studies on individual WHR technologies for marine diesel engines, the effect of transmission ratio on the performance of the thermodynamic WHR has not been tackled thus far. Appl. Sci. 2019, 9, x FOR PEER REVIEW 3 of 20

Despite numerous studies on individual WHR technologies for marine diesel engines, the effect of transmission ratio on the performance of the thermodynamic WHR has not been tackled thus far. Appl.In this Sci. paper,2019, 9, a 4242 marine engine waste heat recovery and cooling system with combined ORC and3 VCC of 18 is proposed.Appl. Sci. 2019 A comprehensive, 9, x FOR PEER REVIEW energy and exergy analysis was carried out and the potential3 of 20of the proposed thermodynamic system was demonstrated. In this paper,Despite a marine numerous engine studies waste on heatindividual recovery WHR and techno coolinglogies system for marine with diesel combined engines, ORC the effect and VCC of transmission ratio on the performance of the thermodynamic WHR has not been tackled thus far. is proposed. A comprehensive energy and exergy analysis was carried out and the potential of the 2. ThermodynamicsIn this paper, a marine and Operational engine waste heatConditions recovery ofand the cooling Proposed system System with combined ORC and VCC proposed thermodynamic system was demonstrated. is proposed. A comprehensive energy and exergy analysis was carried out and the potential of the 2.1. Systemproposed Components thermodynamic system was demonstrated. 2. Thermodynamics and Operational Conditions of the Proposed System Figure2. Thermodynamics 1 shows a schematic and Operational of the proposed Conditions heat of the driven Proposed cooling System system in the paper. As shown 2.1.in the System figure, Components the ORC expander was coupled with the VCC compressor by using a transmission unit. The Figurecombined2.1. System1 shows systemComponents a schematic aimed to of produce the proposed cooling heat for driven refrigeration cooling systemcabinets in by the waste paper. heat As shown recovery in thefrom ﬁgure, engines.Figure the The ORC 1 shows thermodynamic expander a schematic was statesof coupled the proposedof the with combined theheat VCC driven organic compressor cooling Rankine system by cycle-vapour usingin the apaper. transmission As compression shown unit. Thecycle combined in(ORC-VCC) the figure, system the are ORC aimedshown expander to in produce the was Temperature-Entropy coupled cooling with for the refrigeration VCC compressor (T-S) cabinets diagrams by byusing wastein a Figure transmission heat recovery2. Here, unit. a from 12- engines.cylinderThe ThecombinedWärtsilä thermodynamic systemRT-flex96C aimed states engineto produce of the with combined cooling 68640 for organickW refrigeration rated Rankine power cabinets cycle-vapour was by wasteselected. compression heat Therecovery exhaust cycle (ORC-VCC)temperaturefrom engines. areand shown heat The thermodynamicenergy in the Temperature-Entropycontained states in of the the exhaust combined (T-S) gas organic diagrams of the Rankine engine in Figure cycle-vapour is presented2. Here, compression ain 12-cylinder Figure 3. It Wärtsiläwas observedcycle RT-ﬂex96C (ORC-VCC) that the engine are temperature shown with in 68640 the peak Temperature-Entropy kW appeared rated power at was40% (T-S) selected.contract diagrams Themaximum in exhaust Figure continuous 2. temperature Here, a 12- rating and cylinder Wärtsilä RT-flex96C engine with 68640 kW rated power was selected. The exhaust heat(MCR) energy and that contained heat content in the exhaustincreased gas with of the the engine engine is load. presented The exhaust in Figure ga3s. temperature It was observed was that in a temperature and heat energy contained in the exhaust gas of the engine is presented in Figure 3. It therange temperature of 528.5 K to peak 597.9 appeared K, and the at 40%energy contract contained maximum in the gas continuous varied from rating 11,860 (MCR) to 35,787 and that kW. heat was observed that the temperature peak appeared at 40% contract maximum continuous rating content(MCR) increased and that with heat the content engine increased load. Thewith exhaustthe engine gas load. temperature The exhaust was gas intemperature a range of was 528.5 in a K to 597.9range K, and of the 528.5 energy K to 597.9 contained K, and inthe the energy gas variedcontained from in the 11,860 gas varied to 35,787 from kW. 11,860 to 35,787 kW.

Figure 1. Schematic of proposed organic Rankine cycle-vapour compression cycle (ORC-VCC) FigureFigure 1. Schematic 1. Schematic of proposed of proposed organic organic Rankine Rankine cycle-vapour cycle-vapour compression compression cycle cycle (ORC-VCC) (ORC-VCC) systems forsystems wastesystems for heat waste for recovery. waste heat heat recovery. recovery.

ORCORC working working fluid fluid R134aR134a 450 450 5 5 380 360 400 400 360 3 3 340 8 8 6 6 350 320 8 350 320 8 2 300 11 2 4 300 0 11 4 0 300 280 13 12 300 1 280 13 12 12 260 1 10 7 Temperature [K] Temperature 12 260 250 0 240 10 7 Temperature [K] Temperature Temperature [K] 250 0 240

Temperature [K] 220 200 220200 200 200180 0.6 0.8 1.0 1.2 1.4 1.6 1.8 2.0 0.60.81.01.21.41.61.82.0 180 0.6 0.8 1.0Entropy 1.2 [kJ/kg 1.4⋅K] 1.6 1.8 2.0 Entropy [kJ/kg⋅K] 0.60.81.01.21.41.61.82.0 ⋅ ⋅ Entropy [kJ/kg K] Entropy [kJ/kg K] Figure 2. Temperature-Entropy (T-S) diagrams of both the ORC and VCC. Appl. Sci. 2019, 9, x FOR PEER REVIEW 4 of 20

Figure 2. Temperature-Entropy (T-S) diagrams of both the ORC and VCC. Appl. Sci. 2019, 9, 4242 4 of 18

40,000 600 Temperature 35,000 590 Heat content

580 30,000

570 25,000 560 20,000 550 Temeprature [K] Temeprature 15,000 540

530 10,000 Heat content of exhaust gas [kW] 520 5,000 20 30 40 50 60 70 80 90 100 110 Engine load [MCR%]

Figure 3. Temperature and heat content available in exhaust gas under diﬀerent loads. Figure 3. Temperature and heat content available in exhaust gas under different loads. The system was subject to two working conditions: (1) a belt drive was adopted, and the transmissionThe system ratio waswas adjustedsubject to bytwo changing working the conditions: diameter (1) of a the belt pulley drive whenwas adopted, the transmission and the ratio wastransmission not equal ratio to 1; was (2) adjusted the ORC by expanderchanging the and diameter the compressor of the pulley shared when the a commontransmission shaft ratio and the was not equal to 1; (2) the ORC expander and the compressor shared a common shaft and the transmission ratio was set to 1. transmission ratio was set to 1. 2.2. Thermodynamics Models and Equations 2.2. Thermodynamics Models and Equations ToTo establish establish a thermodynamicsa thermodynamics model model for for the the sy systemstem proposed, proposed, a number a number of assumptions of assumptions were were made,made, such such as as (1) (1) both both the the ambientambient temperature temperature an andd the the sea sea water water temperature temperature were were298 K; 298(2) the K; (2) the pinchpinch point point was was 30 30 K K in in the the evaporatorevaporator of of the the ORC ORC and and was was 5 K 5 in K the in theother other heat heatexchangers; exchangers; (3) the (3) the isentropicisentropic eﬃ efficiencyciency of of the the expander, expander, compressors, compressors, and pump was was 0.7, 0.7, 0.7, 0.7, and and 0.8, 0.8, respectively; respectively; (4) (4) the temperaturethe temperature of the of exhaust the exhaust gas exitinggas exiting boiler boiler was was higher higher than than the the acid acid dewdew pointpoint of of 393 393 K K [36]; [36 ];(5) (5) both heatboth and heat pressure and pressure loss in allloss heat in all exchangers heat exchangers and pipes and pipes were assumedwere assumed to be to negligible; be negligible; (6) the(6) combinedthe cyclescombined were operated cycles were under operated steady stateunder conditions; steady state (7) conditions; the mechanical (7) the loss mechanical of the expander loss of/compressor the expander/compressor coupling was negligible. coupling was negligible. 2.2.1. Components of the Exhaust Gas 2.2.1. Components of the Exhaust Gas Assuming that the molecular formula of hydrocarbons in the heavy fuel oil (HFO) supplied to Assuming that the molecular formula of hydrocarbons in the heavy fuel oil (HFO) supplied to the the main IC engine was C30H62 [37], the chemical equation of the combustion process can be expressed mainas ICthe enginefollowing was equation: C30H62 [37], the chemical equation of the combustion process can be expressed as the following equation: + + → + + + C30H62 a(O2 3.76N2) 30CO2 31H2O bO2 3.76aN2. (1) C30H62 + a(O2 + 3.76N2) 30CO2 + 31H2O + bO2 + 3.76aN2. (1) Considering that the oxygen concentration→ in the exhaust gas was 15.3% [38], the molar concentrationConsidering of thatthe exhaust the oxygen gas composition concentration was then in thecalculated: exhaust 75.37% gas wasfor nitrogen, 15.3% [5.37%38], the for molar carbon dioxide, 2.27% for steam (vapour), and 15.3% for oxygen. The properties of the exhaust gas concentration of the exhaust gas composition was then calculated: 75.37% for nitrogen, 5.37% can be evaluated on the basis of the mixing rule proposed by Poling [39]. for carbon dioxide, 2.27% for steam (vapour), and 15.3% for oxygen. The properties of the exhaust gas can2.2.2. be evaluated Thermodynamics on the basisEquations of the mixing rule proposed by Poling [39].

2.2.2. ThermodynamicsFirst law modelling Equations First law modelling (1) Pump (1 2): → The power consumed by the pump can be calculated as

Wpump = m (h h ). (2) f 1 2 − 1 Appl. Sci. 2019, 9, 4242 5 of 18

The isentropic eﬃciency of the pump can be calculated as

h h η = 2is − 1 . (3) pump,is h h 2 − 1 (2) Evaporator-ORC (2 3, 5 6): → → The heat transferred in the evaporator-ORC can be calculated as

Q = m (h h ), (4) eva,ORC f 1 3 − 2

Q = mg(h h ). (5) eva,ORC 5 − 6 (3) Expander (3 4): → The power generated in the expander can be calculated as

Wexp = m (h h ). (6) f 1 3 − 4 The isentropic eﬃciency of the expander can be calculated as

h h η = 3 − 4 . (7) exp,is h h 3 − 4is (4) Condenser-ORC (4 1, 0 12): → → The heat transferred in condenser-ORC can be calculated as

Q = m (h h ), (8) cond,ORC f 1 4 − 1 Q = m (h h ). (9) cond,ORC w1 12 − 0 (5) Compressor (7 8): → The power consumed by compressor can be calculated as

Wcom = m (h h ), (10) f 2 8 − 7

Wcom = Wexp. (11) The isentropic eﬃciency of the compressor can be calculated as

h h η = 8is − 7 . (12) com,is h h 8 − 7 (6) Condenser-VCC (8 9, 0 11): → → The heat transferred in the condenser-VCC can be calculated as

Q = m (h h ), (13) cond,VCC f 2 8 − 9

Q = mw (h h ). (14) cond,VCC 2 11 − 0 (7) Expansion valve (9 10): → The working ﬂuid ﬂowing through the expansion valve can be taken as an isenthalpic throttling process h9 = h10. (15) Appl. Sci. 2019, 9, 4242 6 of 18

(8) Evaporator-VCC (10 7, 13 14): → → The heat transferred in the evaporator-ORC can be calculated as

Q = m (h h ), (16) eva,VCC f 2 7 − 10 Q = m (h h ). (17) eva,VCC air 13 − 14 Second law eﬃciency The exergy at state point i in the system can be deﬁned by

E = m [(h h ) T (s s )]. (18) i f i − 0 − 0 i − 0 The exergy destruction caused in each component can be calculated as X X I = E Eout. (19) i in − Expressions for exergy destruction are shown in following: Pump-ORC: I = Wpump + E E . (20) pump,ORC 1 − 2 Evaporator-ORC: I = E + E (E + E ). (21) eva,ORC 5 2 − 6 3 Expander-ORC: I = E E Wexp. (22) exp,ORC 3 − 4 − Condenser-ORC: I = E + E (E + E ). (23) cond,ORC 0 4 − 1 12 Compressor-VCC: I = Wcom + E E . (24) com,VCC 7 − 8 Condenser-VCC: I = E + E (E + E ). (25) cond,VCC 8 0 − 11 9 Expansion Valve: I = E E . (26) val,VCC 9 − 10 Evaporator-VCC: I = E + E (E + E ). (27) eva,VCC 10 13 − 7 14 The total exergy destruction caused in this system can be expressed as

Itotal = Ieva,ORC + Iexp,ORC + Ipump,ORC + Icond,ORC + Icond,VCC + Ival,VCC + Ieva,VCC + Icom,VCC. (28)

The exergy eﬃciency of the ORC-VCC waste heat recovery system can be calculated as the following: ηex WHR = Eout/Ein. (29) − In this case, the exergy input Ein included the exergy diﬀerence of exhaust gas stream, the power of the pump. Eout = (T0/Tcabin,in 1)Qeva VCC, (30) − − Ein = Eexh + WPump. (31) Appl. Sci. 2019, 9, 4242 7 of 18

The term EER (energy eﬃciency ratio) is a ratio of the useful cooling provided to the work required of air conditioners. Higher EER equates to lower operating costs.

Qeva,VCC EER = . (32) Wexp

2.3. Operational Conditions The heat transfer process should meet two requirements simultaneously. Firstly, the pinch point temperature difference (PPTD) should not be lower than 30 K in the modelling of ORC evaporator during the phase change process to ensure the feasible heat exchanger design in practical applications. Furthermore, the exhaust gas temperature should be higher than the acid dew point to avoid the possible corrosion in the evaporator (exhaust side), which is normally taken to be 393 K [40,41]. Figure4 shows the evaporation process based on the PPTD method, which has been explained in our previous study [35]. During the evaporation process, the evaporation temperature of working fluid rises as the pressure increases. For a working fluid with higher critical temperature, when ORC operates at T3a, the relative exhaust gas temperature T6a after heat transfer is above 393 K. In this case, the heat transferred in the evaporator will become smaller as the evaporation pressure increases, which would lead to a reduction of power output. However, for a working fluid with low critical temperature, the temperature during the evaporation process varies from T2 to T3b. As a result, the exhaust gas temperature would be lower than 393 K on the basis of the assumption that the PPTD is 30 K. To avoid a possible corrosion problem in the ORC evaporator, the outlet temperature of exhaust gas after heat transfer is set to be 393 K. In this case, the heat content in the ORC evaporator remains the same, even the evaporation pressure varies. On the other hand, as the evaporation pressure increases, the flow rate of working fluid decreases and the Appl. Sci. 2019, 9, x FOR PEER REVIEW 8 of 20 energy density becomes greater.

Figure 4. The temperature-entropy diagram on both exhaust gas and working fluid during the heat Figure 4. The temperature-entropy diagram on both exhaust gas and working fluid during the heat transfer process based on the pinch point temperature difference (PPTD) method. transfer process based on the pinch point temperature difference (PPTD) method. A program developed on the basis of the MATLAB platform was adopted to analyse the A program developed on the basis of the MATLAB platform was adopted to analyse the thermodynamic performance. Four environmentally friendly working fluids—R1233zd, R1234zez, thermodynamic performance. Four environmentally friendly working fluids—R1233zd, R1234zez, R245fa and R245ca—were considered as the ORC working fluids for the performance comparison, and R245fa and R245ca—were considered as the ORC working fluids for the performance comparison, the refrigerant R134a was used in the VCC. Their enthalpies and entropies of were obtained by using a and the refrigerant R134a was used in the VCC. Their enthalpies and entropies of were obtained by database Refprop 9.0. using a database Refprop 9.0.

3. Results

3.1. Belt Drive as the Transmission Unit When a belt drive was used as the transmission unit for the coupling of expander and compressor, both the ORC and VCC could be operated at their optimum conditions simultaneously. In this case, the transmission ratio could be adjusted by changing the diameters of two pulleys. As the refrigeration cycle was powered by the ORC expander, the variation of the ORC power output had a great effect on the performance of the refrigeration cycle. Figure 5 presents the comparison of ORC power output among the four different working fluids. The figure shows that for R1233zd and R1234zez, the power output firstly increases and then decreases after reaching its peak value roughly at 34oo kPa. This is because enthalpy change across the expander increased as the evaporation pressure increases. However, the mass flow rate of working fluid decreases as the ORC evaporation pressure increases. As a result, the different variation trends of these two parameters lead to an optimal power output when the ORC was operated nearly at the critical pressure. For the working fluids with lower critical temperatures, such as R245fa and R245ca, the maximum cooling capacity appeares at a higher evaporation pressure, whereas it appeares at a lower evaporation pressure for the working fluids with lower critical temperatures, namely, R1233zd and R1234zez. Furthermore, it can be observed that R1233zd and R245ca have a higher power output in comparison with R1234zez and R245fa.

Appl. Sci. 2019, 9, 4242 8 of 18

3. Results

3.1. Belt Drive as the Transmission Unit When a belt drive was used as the transmission unit for the coupling of expander and compressor, both the ORC and VCC could be operated at their optimum conditions simultaneously. In this case, the transmission ratio could be adjusted by changing the diameters of two pulleys. As the refrigeration cycle was powered by the ORC expander, the variation of the ORC power output had a great effect on the performance of the refrigeration cycle. Figure5 presents the comparison of ORC power output among the four different working fluids. The figure shows that for R1233zd and R1234zez, the power output firstly increases and then decreases after reaching its peak value roughly at 34oo kPa. This is because enthalpy change across the expander increased as the evaporation pressure increases. However, the mass flow rate of working fluid decreases as the ORC evaporation pressure increases. As a result, the different variation trends of these two parameters lead to an optimal power output when the ORC was operated nearly at the critical pressure. For the working fluids with lower critical temperatures, such as R245fa and R245ca, the maximum cooling capacity appeares at a higher evaporation pressure, whereas it appeares at a lower evaporation pressure for the working fluids with lowerAppl. critical Sci. 2019 temperatures,, 9, x FOR PEER REVIEW namely, R1233zd and R1234zez. Furthermore, it can be observed9 of that 20 R1233zd and R245ca have a higher power output in comparison with R1234zez and R245fa.

3800 R1233zd 3600 R1234zez 3400 R245fa R245ca 3200

3000

2800

2600

Power output [kW] Power 2400

2200

2000 750 1250 1750 2250 2750 3250 3750 ORC evaporation pressure [kPa]

Figure 5. Effects of ORC evaporation pressure on ORC power output for different working fluids. Figure 5. Effects of ORC evaporation pressure on ORC power output for different working fluids. In this system, the power generated in the ORC expander was consumed by the VCC compressor. In this system, the power generated in the ORC expander was consumed by the VCC Hence, the sole useful energy output of the whole system was the cooling generated in the VCC. compressor. Hence, the sole useful energy output of the whole system was the cooling generated in Figure6 shows the variation of cooling capacity and EER with evaporation pressure. The cooling the VCC. Figure 6 shows the variation of cooling capacity and EER with evaporation pressure. The capacitycooling has capacity a similar has variation a similar trend variation with thattrend of with the ORCthat of power the ORC output, power as output, shown as in shown Figure 5in. ThisFigure is because5. This the is power because consumed the power by consumed the VCC by compressor the VCC compressor is equal to thatis equal generated to that generated by the ORC by expander.the ORC For aexpander. given condensation For a given temperature, condensation the temperature, power input the into power the VCCinput decides into the the VCC mass decides flow ratethe ofmass the refrigerantflow rate in theof the VCC. refrigerant In other in words, the VCC. the In cooling other words, capacity the is cooling decided capacity by the poweris decided input by of the the power VCC. inputFigure of6 thealso VCC. indicates that the EER decreases with the ORC evaporation pressure. According to EquationFigure (32), EER 6 alsois definedindicates as that the the cooling EER capacitydecreases of with evaporation the ORC evaporation in the VCC topressure. the power According generated to by theEquation ORC. As (32), the ORCEER evaporationis defined as pressure the cooling increases, capacity more of power evaporation is supplied in the to theVCC VCC to the subsystem, power and thegenerated cooling by capacity the ORC. of As the the evaporator ORC evaporation also increases pressure because increases, of the more increasing power refrigerantis supplied flow to the in VCC.VCC This subsystem, can be attributed and the to coo theling fact capacity that the of power the evaporator consumed also bythe increases compressor because increased, of the increasing although the coolingrefrigerant capacity flow in also VCC. rises. This The can growth be attributed rate of to the the cooling fact that capacity the power is notconsumed comparable by the to compressor that of the compressorincreased, power, although resulting the cooling in a smaller capacityEER also. rises. The growth rate of the cooling capacity is not comparable to that of the compressor power, resulting in a smaller EER. Appl. Sci. 2019, 9, x FOR PEER REVIEW 10 of 20

Appl. Sci. 2019, 9, x FOR PEER REVIEW 10 of 20 Appl. Sci. 2019, 9, 4242 9 of 18 10000 2.90

100009500 2.90 2.85 95009000 2.85 2.80 90008500 2.80 85008000 2.75

80007500 2.75 2.70 R245ca 75007000

R245fa Energy efficiency ratio Cooling capacity [kW] capacity Cooling 2.70 R245ca R1234zez 2.65 70006500 R1233zd R245fa Energy efficiency ratio Cooling capacity [kW] capacity Cooling R1234zez 2.65 65006000 R1233zd 2.60 1000 1500 2000 2500 3000 3500 6000 ORC evaporation pressure [kPa] 2.60 1000 1500 2000 2500 3000 3500

ORC evaporation pressure [kPa] Figure 6. Effect of ORC evaporation pressure on cooling capacity and energy efficiency ratio for

Figuredifferent 6. Eff workingect of ORC fluids. evaporation pressure on cooling capacity and energy efficiency ratio for Figure 6. Effect of ORC evaporation pressure on cooling capacity and energy efficiency ratio for different working fluids. differentFigure working 7 shows fluids. the variation of exergy efficiency when the ORC is operated under different evaporationFigure7 shows pressures. the variation In this system, of exergy the exergy e fficiency output when refers theto the ORC cooling is operated exergy of underthe evaporator different Figure 7 shows the variation of exergy efficiency when the ORC is operated under different evaporationin the VCC, pressures. and the Inexergy this system, input consists the exergy of the output heat exergy refersto of the thecooling evaporator exergy in the of theORC evaporator and the evaporation pressures. In this system, the exergy output refers to the cooling exergy of the evaporator in thepump VCC, power. and theAt the exergy same inputevaporation consists temperature, of the heat the exergy variation of the trend evaporator of exergy in efficiency the ORC in andFigure the in the VCC, and the exergy input consists of the heat exergy of the evaporator in the ORC and the pump6 is power. similar At to the that same of the evaporation cooling capacity. temperature, This is the because variation the trendexergy of output exergy of effi coolingciency is in linearly Figure6 pump power. At the same evaporation temperature, the variation trend of exergy efficiency in Figure is similarcorrelated to that to the of cooling the cooling capacity capacity. at a given This VCC is because evaporation the exergy temperature. output Furthermore, of cooling is a linearlyhigher 6 is similar to that of the cooling capacity. This is because the exergy output of cooling is linearly correlatedORC evaporation to the cooling pressure capacity leads at to a a given better VCC temp evaporationerature match temperature. between the Furthermore,organic working a higher fluid correlatedand the exhaustto the cooling gas, and capacity the exergy at a degivenstruction VCC perevaporation mass unit temperature. decreases. Furthermore, a higher ORCORC evaporation evaporation pressure pressure leads leads to to a a better better temperature temperature match match between between the the organic organic working working fluid fluid and theand exhaust the exhaust gas, and gas, theand exergy the exergy destruction destruction per massper mass unit unit decreases. decreases.

0.16 R245ca 0.160.15 R245fa R245ca R1234zez R1233zd 0.150.14 R245fa R1234zez R1233zd 0.140.13

0.130.12

0.120.11

0.110.10 Exergy efficiency of combined cycle cycle of combined efficiency Exergy 0.100.09

Exergy efficiency of combined cycle cycle of combined efficiency Exergy 750 1250 1750 2250 2750 3250 3750 0.09 ORC evaporation pressure [kPa] 750 1250 1750 2250 2750 3250 3750

Figure 7. Eﬀect of ORC evaporationORC evaporation pressure pressure on exergy [kPa] eﬃciency of combined cycle. Figure 7. Effect of ORC evaporation pressure on exergy efficiency of combined cycle.

3.2. Sharing a Common Shaft Figure 7. Effect of ORC evaporation pressure on exergy efficiency of combined cycle. In this section, the case where the expander and compressor shared a common shaft is studied, wherein the transmission ratio was 1. The marine engine is operated under 100% MCR. From the above analysis, R245ca and R1233zd show comparable performance. R1233zd is selected as the ORC working ﬂuid because it is more environmentally friendly than R245ca. The ORC condensation temperatures Appl. Sci. 2019, 9, x FOR PEER REVIEW 11 of 20

3.2. Sharing a Common Shaft In this section, the case where the expander and compressor shared a common shaft is studied, wherein the transmission ratio was 1. The marine engine is operated under 100% MCR. From the above analysis, R245ca and R1233zd show comparable performance. R1233zd is selected as the ORC working fluid because it is more environmentally friendly than R245ca. The ORC condensation Appl. Sci. 2019, 9, 4242 10 of 18 temperatures are set as 308 K by using sea water as the heat sink. In this system, the pressure difference across the expander is equal to that across the compressor (P3-P4 = P8-P7). are set as 308Figure K by 8 shows using the sea variation water as of the cooling heat sink. capacity In this under system, different the VCC pressure evaporation difference temperatures. across the It can be seen that the cooling capacity decreases with ORC evaporation pressure and increased with expander is equal to that across the compressor (P3-P4 = P8-P7). Figurethe cold8 shows air temperature. the variation With of the cooling increase capacity of the underORC evaporation di fferent VCC pressure, evaporation the VCC temperatures. condensation It canpressure be seen that is enhanced the cooling for capacity a fixed decreasesPPTD and with a consta ORCnt evaporation heat sink, resulting pressure andin a increasedhigher discharged with the temperature. The enthalpy difference across the compressor outlet and inlet become greater. cold air temperature. With the increase of the ORC evaporation pressure, the VCC condensation pressure Consequently, the cooling capacity of the VCC evaporator is reduced because of the decreasing flow is enhanced for a fixed PPTD and a constant heat sink, resulting in a higher discharged temperature. The rate of the VCC refrigerant. enthalpy difference across the compressor outlet and inlet become greater. Consequently, the cooling capacity of the VCC evaporator is reduced because of the decreasing flow rate of the VCC refrigerant.

8000 T15=273 K 7000 T15=268 K T15=263 K T15=258 K 6000

5000

4000

3000 Cooling capacity [kW] capacity Cooling 2000

1000 1000 1500 2000 2500 3000 3500 ORC evaporation pressure [kPa]

Figure 8. Effect of ORC evaporation pressure on cooling capacity for the case of sharing a common shaft. Figure 8. Effect of ORC evaporation pressure on cooling capacity for the case of sharing a common The calculatedshaft. EER of cooling strongly depends on the operating conditions, especially absolute temperature and the temperature difference between the heat sink and the target refrigeration The calculated EER of cooling strongly depends on the operating conditions, especially absolute temperature. Figure9 illustrates that the EER of cooling decreases with ORC evaporation pressure. temperature and the temperature difference between the heat sink and the target refrigeration For a given heat sink temperature, the VCC condensation pressure and temperature rise when the temperature. Figure 9 illustrates that the EER of cooling decreases with ORC evaporation pressure. ORC evaporation pressure is increased. In that case, the temperature lift of the refrigerant between the For a given heat sink temperature, the VCC condensation pressure and temperature rise when the condenserORC evaporation and the evaporator pressure is is enhanced, increased. resultingIn that ca inse, a the decreased temperatureEER .lift of the refrigerant between Inthe this condenser system, and both the ORC evaporator and VCC is enhanced, used sea waterresulting as thein a heatdecreased sink. EER The. effect of condensation temperature on the performance of both the ORC power cycle and the VCC refrigeration cycle was analysed. The condensation temperatures in both subsystems were assumed to be the same, as they both used the same heat sink. The curves in Figure 10 present an ascending trend as the condensation temperature increases. By sharing a common shaft, the pressure difference between the ORC expander inlet and outlet is equal to that between the compressor outlet and inlet, which can be expressed as P3-P4 = P8-P7. The variation of ORC condensation temperature influences the ORC and VCC. When the ORC is operated at the same evaporation pressure, the pressure difference across the expander is reduced if the condensation temperature is increased. As a result, the VCC evaporation pressure P7 was decreases, as the VCC expander outlet pressure P8 increases with the VCC condensation temperature, explaining why the evaporation temperature increased with the condensation temperature. From Figure 10, it can be concluded that the ORC subsystem should be operated at a higher evaporation pressure to achieve a lower required refrigeration temperature. Figure 11 shows the variation of the cooling capacity with variable condensation temperatures. It can be observed that the cooling capacity increases with rising condensation temperature. For the ORC subsystem, the expander could generate less power output to drive the VCC compressor. In the VCC subsystem, the power consumed by per unit mass refrigerant in the VCC is reduced because the Appl. Sci. 2019, 9, 4242 11 of 18 pressure difference across the compressor decreases with the increase of condensation temperature, as mentioned above. The enthalpy difference of the refrigerant at the compressor outlet and inlet decreases as well, leading to a larger mass flow rate of refrigerant. Furthermore, the evaporation temperature Appl. Sci. 2019, 9, x FOR PEER REVIEW 12 of 20 becomes higher, explaining why the cooling capacity becomes larger for a higher condensation temperature, although the power provided by the ORC expander decreases.

3.0 T15=273 K T15=268 K 2.5 T15=263 K T15=258 K 2.0

1.5

1.0

Energy efficiency ratio 0.5

0.0 1000 1500 2000 2500 3000 3500 Appl. Sci. 2019, 9, x FOR PEER REVIEW ORC evaporation pressure [kPa] 13 of 20

Figure 9. Eﬀect of ORC evaporation pressure on cooling capacity for the case of sharing a common shaft. Figure 9. Effect of ORC evaporation pressure on coolin g capacity for the case of sharing a common shaft. 300

In this system, both280 ORC and VCC used sea water as the heat sink. The effect of condensation temperature on the performance of both the ORC power cycle and the VCC refrigeration cycle was analysed. The condensation260 temperatures in both subsystems were assumed to be the same, as they both used the same heat sink. The curves in Figure 10 present an ascending trend as the condensation temperature increases.240 By sharing a common shaft, the pressure difference between the ORC

expander inlet and outlet is equal to that between the compressor outlet and inlet, which can be 220 expressed as P3-P4 = P8-P7. The variation of ORC condensation temperature influences the ORC and VCC. When the ORC 200 Peva1=800kPa is operated at the same evaporation pressure, the pressure difference Peva1=900kPa across the expander is reduced if the condensation temperature is increased. As a result, the Peva1=1000kPaVCC evaporation pressure P7 was 180 decreases, as the VCC expander outlet pressure P8 increases with Peva1=1100kPa the VCC condensation temperature,

explaining why the[K] VCC in temperature Evaporation evaporation temperature increased with the condensation temperature. 160 From Figure 10, it 308can be concluded 310 that 312 the ORC 314 subsystem 316 should 318 be operated at a higher evaporation pressure to achieve a lower required refrigeration temperature. Condensation temperature [K]

Figure 10. Effect of condensation temperature in both condensers on evaporation for the case of sharing a commonFigure shaft.10. Effect of condensation temperature in both condensers on evaporation for the case of sharing a common shaft. Figure 12 shows the contribution of each component on the exergy destruction under different condensationFigure temperatures11 shows the variation in the condenser. of the cooling It is capacity evident with that variable under the condensation operating conditionstemperatures. considered,It can be theobserved evaporator that the in thecooling ORC capacity power cycle increa contributesses with rising the most condensation significant temperature. part of the totalFor the exergyORC loss subsystem, in the system. the expander This result could can generate be attributed less power to the significantoutput to drive irreversibility the VCC associatedcompressor. with In the heatVCC transfer subsystem, across thethe large power temperature consumed di byfferences per unit in mass the evaporator. refrigerant in the VCC is reduced because the pressure difference across the compressor decreases with the increase of condensation temperature, as mentioned above. The enthalpy difference of the refrigerant at the compressor outlet and inlet decreases as well, leading to a larger mass flow rate of refrigerant. Furthermore, the evaporation temperature becomes higher, explaining why the cooling capacity becomes larger for a higher condensation temperature, although the power provided by the ORC expander decreases.

12000

10000

8000

6000

4000 Cooling capacity [kW] capacity Cooling Peva1=800kPa Peva1=900kPa 2000 Peva1=1000kPa Peva1=1100kPa 0 309 310 311 312 313 314 Condensation temperature [K]

Figure 11. Effect of ORC condensation temperature on cooling capacity for the case of sharing a common shaft. Appl. Sci. 2019, 9, x FOR PEER REVIEW 13 of 20

300

280

260

240

220

200 Peva1=800kPa Peva1=900kPa Peva1=1000kPa 180 Peva1=1100kPa

Evaporation temperature in VCC [K] VCC in temperature Evaporation 160 308 310 312 314 316 318 Condensation temperature [K]

Figure 10. Effect of condensation temperature in both condensers on evaporation for the case of sharing a common shaft.

Figure 11 shows the variation of the cooling capacity with variable condensation temperatures. It can be observed that the cooling capacity increases with rising condensation temperature. For the ORC subsystem, the expander could generate less power output to drive the VCC compressor. In the VCC subsystem, the power consumed by per unit mass refrigerant in the VCC is reduced because the pressure difference across the compressor decreases with the increase of condensation temperature, as mentioned above. The enthalpy difference of the refrigerant at the compressor outlet and inlet decreases as well, leading to a larger mass flow rate of refrigerant. Furthermore, the evaporation temperature becomes higher, explaining why the cooling capacity becomes larger for a higher condensation temperature, although the power provided by the ORC expander decreases. Appl. Sci. 2019, 9, 4242 12 of 18

12000

10000

Appl. Sci. 2019, 9, x FOR PEER8000 REVIEW 14 of 20

Figure 12 shows the contribution of each component on the exergy destruction under different condensation temperatures6000 in the condenser. It is evident that under the operating conditions

considered, the evaporator in the ORC power cycle contributes the most significant part of the total exergy loss in the system.4000 This result can be attributed to the significant irreversibility associated with heat transfer across[kW] capacity Cooling the large temperature differences in the evaporator. Peva1=800kPa In the VCC subsystem, the irreversibility across the compressor Peva1=900kPa is the maximum under most 2000 working conditions. However, the exergy destruction in the compressor Peva1=1000kPa is higher than that when the condensation temperature is lower than 309.65 K and Peva = 1100 kPa. Peva1=1100kPa This suggests that in order to improve the performance0 of the proposed system, special attention should be paid to reducing the irreversibility that exists in these309 components 310 during 311 machine 312 design. 313 314 The irreversibility of the compressorCondensation and expander temperature essentially [K] depends on their isentropic efficiency; thus, their proper design can reduce this irreversibility. In particular, in order to reduce Figure 11. Eﬀect of ORC condensation temperature on cooling capacity for the case of sharing a the irreversibilityFigure 11. Effect of the of condenser ORC condensation and evaporator, temperature they onshould cooling be designedcapacity forin suchthe case a way of sharingthat the a common shaft. temperaturecommon difference shaft. between the fluids can be maintained as small as possible. Exergy destruction contribution [kW]

Peva=800kPa

Tcon=313.65K Tcon=313.15K Tcon=312.65K Tcon=312.15K Tcon=311.65K Tcon=311.15K Tcon=310.65K Tcon=310.15K Tcon=309.65K Tcon=309.15K 0 1000 2000 3000 4000 5000 6000 7000 8000 9000

Ieva,ORC Icon,ORC Iexp,ORC Ipum,ORC Icon,VCC Iexp,VCC Icom,VCC Ieva,VCC

(a) ORC evaporation pressure Peva = 800 kPa

Exergy destruction contribution [kW]

Peva = 900kPa

Tcon=313.65K Tcon=313.15K Tcon=312.65K Tcon=312.15K Tcon=311.65K Tcon=311.15K Tcon=310.65K Tcon=310.15K Tcon=309.65K Tcon=309.15K 0 1000 2000 3000 4000 5000 6000 7000 8000 9000

Ieva,ORC Icon,ORC Iexp,ORC Ipum,ORC Icon,VCC Iexp,VCC Icom,VCC Ieva,VCC

(b) ORC evaporation pressure Peva = 900 kPa

Figure 12. Cont. Appl. Sci. 2019, 9, 4242 13 of 18 Appl. Sci. 2019, 9, x FOR PEER REVIEW 15 of 20

Exergy destruction contribution [kW]

Peva = 1000kPa

Tcon=313.65K Tcon=313.15K Tcon=312.65K Tcon=312.15K Tcon=311.65K Tcon=311.15K Tcon=310.65K Tcon=310.15K Tcon=309.65K Tcon=309.15K 0 1000 2000 3000 4000 5000 6000 7000 8000 9000

Ieva,ORC Icon,ORC Iexp,ORC Ipum,ORC Icon,VCC Iexp,VCC Icom,VCC Ieva,VCC

Exergy destruction contribution [kW]

Peva = 1100kPa

Tcon=313.65K Tcon=313.15K Tcon=312.65K Tcon=312.15K Tcon=311.65K Tcon=311.15K Tcon=310.65K Tcon=310.15K Tcon=309.65K Tcon=309.15K 0 1000 2000 3000 4000 5000 6000 7000 8000 9000

Ieva,ORC Icon,ORC Iexp,ORC Ipum,ORC Icon,VCC Iexp,VCC Icom,VCC Ieva,VCC

(d) ORC evaporation pressure Peva = 1100 kPa

Figure 12. Figure 12. ContributionContribution of each of component each component on exergy ondestruction exergy under destruction different undercondensation diﬀerent condensationpressures. pressures.

In theFigure VCC 13 subsystem, shows the variation the irreversibility of exergy effici acrossency the when compressor the condensation is the maximumtemperature under varies. most workingRegarding conditions. the exergy However, input, theit consistes exergy of destruction heating exergy in the of the compressor exhaust gas is and higher pump than power that input. when the condensationThe temperature temperature of exhaust is lower gas T than6 calculated 309.65 by K andusing Peva the PPTD= 1100 method kPa. This is lower suggests than 393 that K, in but order it to improveis set the to be performance a constant of of 393 the K proposedto avoid the system, possible special corrosion attention problems. should Therefore, be paid the heat to reducing input is the irreversibilitythe same, thatalthough exists the in condensation these components temperature during changed machine under design. considered conditions. Moreover, Theless pump irreversibility power is required of the compressorfor a rising condensation and expander temperature. essentially depends on their isentropic Regarding the useful exergy output, the cooling exergy is found to be reduced at its higher VCC efficiency; thus, their proper design can reduce this irreversibility. In particular, in order to reduce evaporation temperature, explaining the reason as to why the exergy efficiency declines as the the irreversibility of the condenser and evaporator, they should be designed in such a way that the condensation temperature increases, except at Peva = 1100 kPa. temperature difference between the fluids can be maintained as small as possible. For Peva = 1100 kPa, the exergy destruction of the expansion valve decreases sharply when the Figurecondensation 13 shows temperature the variation increases of fr exergyom 309.2 effi tociency 309.7 K when (see Figure the condensation 12d), and a valley temperature value at 309.7 varies. RegardingK appeared, the exergy explaining input, why it consistes there is the of maximu heatingm exergy exergy ofefficiency the exhaust under gas this and condition. pump power input. The temperature of exhaust gas T6 calculated by using the PPTD method is lower than 393 K, but it is set to be a constant of 393 K to avoid the possible corrosion problems. Therefore, the heat input is the same, although the condensation temperature changed under considered conditions. Moreover, less pump power is required for a rising condensation temperature. Appl. Sci. 2019, 9, x FOR PEER REVIEW 16 of 20

Furthermore, it can also be observed that a higher exergy efficiency could be obtained when the ORC power systemis operated at a higher evaporation pressure. This is because the evaporation temperature in the VCC is much lower, as shown in Figure 9. Appl. Sci. 2019, 9, 4242 14 of 18

6 Peva1=800kPa 5 Peva1=900kPa Peva1=1000kPa 4 Peva1=1100kPa

Exergy efficiency of the combined cycle [%] cycle combined the of efficiency Exergy 309 310 311 312 313 314 Condensation temperature [K]

Figure 13. Effect of condensation temperature on exergy efficiency in the case of sharing a common shaft. Figure 13. Effect of condensation temperature on exergy efficiency in the case of sharing a common Regardingshaft. the useful exergy output, the cooling exergy is found to be reduced at its higher VCC evaporation temperature, explaining the reason as to why the exergy efficiency declines as the Table 1 summarises the optimal working conditions when the required cold air temperature is condensation temperature increases, except at P = 1100 kPa. not higher than 263 K for a given heat sink temperatureeva of 298 K. It can be understood that the belt FordrivePeva transmission= 1100 kPa, unit the could exergy result destruction in the maximu of them expansion cooling output valve of decreases 9823 kW. sharply Unfortunately, when the its condensationORC evaporation temperature is 3500 increases kPa, which from 309.2 is much to 309.7 higher K (see than Figure that of12 d),the andtransmission a valley value unit atsharing 309.7 Ka appeared,common explaining shaft. It why is thereobvious is thethat maximum a thermo-economic exergy efficiency evaluation under is this needed condition. to decide which Furthermore,transmission unit it can is the also best be by observed making a that trad ae-off higher between exergy the ethermodynamicfficiency could performance be obtained and when the the ORCcompactness/reliability power systemis operated of the ORC-VCC at a higher combined evaporation system. pressure. This is because the evaporation temperature in the VCC is much lower, as shown in Figure9. Table1 summarisesTable the1. Optimal optimal operational working conditions conditions at when263 K VCC the requiredevaporation cold temperature. air temperature is not higher than 263 K for a given heatParameters sink temperature of 298Belt K. Drive It can beCommon understood Shaft that the belt drive transmission unit could resultORC in theworking maximum fluid cooling output R245ca of 9823 R1233zd kW. Unfortunately, its ORC evaporation is 3500 kPa, whichVCC is much refrigerant higher than that of R134a the transmission R134a unit sharing a common shaft. It is obvious that aCooling thermo-economic capacity at evaluation263K [kW] is needed 9823 to decide which 6902 transmission unit is the best by making a trade-oORCff between evaporation the thermodynamicpressure [kPa] performance 3500 and the 900 compactness/reliability of the ORC-VCC combinedEnergy system. efficient ratio (EER) 2.74 3.06 Power output of expander [kW] 3584 2180 Table 1. OptimalTotal exergy operational destruction conditions [kW] at 263 K VCC 7482 evaporation 8132 temperature.

ORCParameters thermal efficiency [%] Belt Drive 13.79 Common 9.10 Shaft Exergy efficiency [%] 13.10 9.50 ORC working fluid R245ca R1233zd VCC refrigerant R134a R134a 4. Discussion Cooling capacity at 263K [kW] 9823 6902 ORC evaporation pressure [kPa] 3500 900 Because exhaustEnergy gas effi cienttemperature ratio (EER) varies signific 2.74antly when the 3.06 engine loads change, it is essential to studyPower the outputpotential of expander of the [kW]WHR system 3584under different engine 2180 loads. The layout sharing Total exergy destruction [kW] 7482 8132 a common shaft isORC considered thermal effi ciencyfor its [%] simple struct 13.79ure, especially for 9.10 transportation. R1233zd and R134a are used in theExergy ORC eandfficiency VCC, [%] respectively. By 13.10 considering the influence 9.50 of the engine load on the operation of the ORC-VCC WHR system, an evaluation of the performance was conducted in this 4. Discussion Because exhaust gas temperature varies significantly when the engine loads change, it is essential to study the potential of the WHR system under different engine loads. The layout sharing a common shaft is considered for its simple structure, especially for transportation. R1233zd and R134a are used in the ORC and VCC, respectively. By considering the influence of the engine load on the Appl. Sci. 2019, 9, 4242 15 of 18 Appl. Sci. 2019, 9, x FOR PEER REVIEW 17 of 20 paperoperation with of respect the ORC-VCC to the cooling WHR output system, at the an evaluationsame temperature. of the performance The following was analysis conducted is based in this on thepaper flowing with respectoperating to theconditions. cooling outputThe ambient at the temper same temperature.ature and the The sea followingwater (heat analysis sink) are is basedboth 298 on the flowing operating conditions. The ambient temperature and the sea water (heat sink) are both K, the refrigeration temperature T15 is 263 K, which is low enough for both air conditioning and food preservation.298 K, the refrigeration The condensation temperature rangesT15 fromis 263 309 K, to which 313 K, is and low the enough low temperature for both air conditioningis limited by andthe heatfood sink preservation. temperature. The condensationThe ORC evaporation ranges from pressure 309 to is 313 the K, optimal and the lowpressure temperature of 900 kPa is limited listed byin Tablethe heat 1. sink temperature. The ORC evaporation pressure is the optimal pressure of 900 kPa listed in TableFigure1. 14 shows the cooling capacity of ORC-VCC when the engine is operated under different loads.Figure It is noted 14 shows that thea higher cooling cooling capacity output of ORC-VCC could be whenachieved the enginewhen increasing is operated the under engine. different The wasteloads. heat It is recovery noted that is closely a higher related cooling to the output thermal could properties be achieved of the when heat increasingsource. As shown the engine. in Figure The 2,waste the heat heat energy recovery available is closely for related recovery to the increases thermal with properties the engine of the load. heat source.The higher As shownthe temperature, in Figure2, thethe better heat energy the quality available of the for heat recovery energy, increases and ther withe is thea peak engine temperature load. The at higher 40% MCR. the temperature, These two factorsthe better resulted the quality in an upward of the heat trend energy, for the and ORC there power is a output, peak temperature and a sharp atchange 40% MCR. appeares These at 40% two MCR.factors This resulted variation in an trend upward is then trend transmitted for the ORC to the power cooling output, capacity and ain sharp VCC, change as it is appearespowered by at 40%the ORCMCR. expander. This variation trend is then transmitted to the cooling capacity in VCC, as it is powered by the ORCFigure expander. 14 also shows that the cooling capacity increased with the condensation temperature. However,Figure the 14 condensationalso shows that temperature the cooling is capacitylimited by increased the parameters with the of condensation sea water, including temperature. its temperatureHowever, the and condensation flow rate. The temperature operational is condition limited by of thethe parameterscombined cycle of sea could water, be controlled including by its adjustingtemperature the andORC flow operation rate. The pressure operational and the condition flow rate of of the the combined heat sink cycleand heat could sources be controlled according by toadjusting the cooling the ORCdemand operation on ship. pressure and the flow rate of the heat sink and heat sources according to the cooling demand on ship.

Figure 14. Cooling capacity under different engine loads at 263 K required cooling temperature. 5. ConclusionsFigure 14. Cooling capacity under different engine loads at 263 K required cooling temperature. A refrigeration system that essentially integrates an ORC power cycle and a vapour compression 5. Conclusions refrigeration combined cycle by waste heat recovery of the marine engine was modelled and analysed in thisA refrigeration paper. The thermal system energythat essentially of the exhaust integrates gas an of anORC engine power was cycle recovered and a vapour by the compression ORC power refrigerationcycle to drive combined a compressor cycle inby the waste VCC. heat On recovery the basis of of the comprehensive marine engine energywas modelled and exergy and analysed analyses, inthe this influence paper. ofThe transmission thermal energy ratio of between the exhaust the expander gas of an and engine the compressor was recovered on system by the performanceORC power cyclewas clarified. to drive Ita compressor was shown thatin the both VCC. ORC On and the VCC basis could of comprehensive be operated under energy their and optimal exergy operational analyses, theconditions influence simultaneously of transmission with ratio the between maximum the 9823expander kW cooling and the capacity compressor at 263 on K system cold temperature performance in wasterms clarified. of the belt It drivewas shown transmission that both unit. ORC Although and VCC the transmission could be operated unit sharing under a commontheir optimal shaft operationalwas more compact conditions and simultaneously reliable than the with belt the drive maximum unit, the 9823 ORC kW and cooling VCC could capacity not beat operated263 K cold at temperaturetheir optimal in conditions terms of the at the belt same drive time, transmission leading to unit. a lower Although cooling the capacity. transmission The ORC unit evaporator sharing a common shaft was more compact and reliable than the belt drive unit, the ORC and VCC could not be operated at their optimal conditions at the same time, leading to a lower cooling capacity. The Appl. Sci. 2019, 9, 4242 16 of 18 contributed the most significant part of the exergy loss in the ORC-VCC combined system because of its significant irreversibility associated with heat transfer across large temperature differences. The cooling capacity showed an upward trend with increasing engine load. A sharp change appeared at 40% MCR because of the high temperature in the exhaust gas. The thermo-economic evaluation of the system and corresponding experimental validations will be conducted in the future work.

Author Contributions: Y.L. built the model of the combined ORC-VCC cycle, analysed the results and write the whole papers. Z.Y. provided the supervision of the paper writing and revision. W.L. participated in the paper revision. All the authors read and approve the manuscript. Funding: Engineering and Physical Sciences Research Council: EP/N020472/1, Engineering and Physical Sciences Research Council: EP/N005228/1, Engineering and Physical Sciences Research Council: EP/R003122/1, Engineering and Physical Sciences Research Council: EP/P028829/1 Acknowledgments: This research was funded by Engineering and Physical Sciences Research Council (EP/N020472/1, EP/N005228/1, EP/R003122/1, and EP/P028829/1) in the United Kingdom. We gratefully acknowledge the assistance of Hongyang Li in proofreading. Conﬂicts of Interest: The authors declare no conﬂict of interest.

Abbreviations

Nomenclature IC Internal combustion ECCS Electricity-cooling combined system MCR Contract maximum continuous rating COP Coefficient of performance EER Energy efficiency ratio HFO Heavy fuel oil PPTD Pinch point temperature difference ORC Organic Rankine cycle SRC Steam Rankine cycle VCC Vapour compression cycle WHR Waste heat recovery Symbols E Exergy (kW) h Specific enthalpy (kJ/kg) I Exergy destruction m Mass flow rate (kg/s) Q Heat (kW) T Temperature (K) W Power (kW) η Efficiency Subscript cond Condenser exp Expander exh Exhaust gas com Compressor eva Evaporator ex Exergy in Input is Isentropic process f1 Working fluid in ORC f2 Working fluid in VCC pump Fluid pump out Output w Water 0 Reference environment Appl. Sci. 2019, 9, 4242 17 of 18

References

1. UNEP. Green Economy in a Blue World; Synthesis Report; The WordFish Center Working Papers: Geneva, Switzerland, 2012. 2. Tsitsilonis, K.M.; Theotokatos, G. A novel systematic methodology for ship propulsion engines energy management. J. Clean. Prod. 2018, 204, 212–236. [CrossRef] 3. Shu, G.; Liang, Y.; Wei, H.; Tian, H.; Zhao, J.; Liu, L. A review of waste heat recovery on two-stroke IC engine aboard ships. Renew. Sustain. Energy Rev. 2013, 19, 385–401. [CrossRef] 4. Bidini, G.; di Maria, F.; Generosi, M. Micro-cogeneration system for a small passenger vessel operating in a nature reserve. Appl. Therm. Eng. 2005, 25, 851–865. [CrossRef] 5. Tien, W.K.; Yeh, R.H.; Hong, J.M. Theoretical analysis of cogeneration system for ships. Energy Convers. Manag. 2007, 48, 1965–1974. [CrossRef] 6. Rigby, G.R.; Hallegraeff, G.M.; Sutton, C. Novel ballast water heating technique offers cost-effective treatment to reduce the risk of global transport of harmful marine organisms. Mar. Ecol. Prog. Ser. 1999, 191, 289–293. [CrossRef] 7. Theotokatos, G.; Livanos, G. Exhaust gas waste heat recovery in marine propulsion plants. In Sustainable Maritime Transportation and Exploitation of Sea Resources; CRC Press: Boca Raton, FL, USA, 2001; pp. 663–671. 8. Theotokatos, G.; Livanos, G. Techno-Economical analysis of single pressure exhaust gas waste heat recovery systems in marine propulsion plants. Proc. Inst. Mech. Eng. Part M J. Eng. Marit. Environ. 2013, 227, 83–97. [CrossRef] 9. Altosole, M.; Laviola, M.; Trucco, A.; Sabattini, A. Waste heat recovery systems from marine diesel engines: Comparison between new design and retrofitting solutions. In Maritime Technology and Engineering; CRC Press: Boca Raton, FL, USA, 2014; pp. 735–742. 10. Benvenuto, G.; Trucco, A.; Campora, U. Optimization of waste heat recovery from the exhaust gas of marine diesel engines. Proc. Inst. Mech. Eng. Part M J. Eng. Marit. Environ. 2014, 230, 83–94. [CrossRef] 11. Hung, T.C. Waste heat recovery of organic Rankine cycle using dry fluids. Energy Convers. Manag. 2001, 42, 539–553. [CrossRef] 12. Singh, D.V.; Pedersen, E. A review of waste heat recovery technologies for maritime applications. Energy Convers. Manag. 2016, 111, 315–328. [CrossRef] 13. Baldi, F.; Gabrielii, C. A feasibility analysis of waste heat recovery systems for marine applications. Energy 2015, 80, 654–665. [CrossRef] 14. Andreasen, J.G.; Meroni, A.; Haglind, F. A comparison of organic and steam Rankine cycle power systems for waste heat recovery on large ships. Energies 2017, 10, 547. [CrossRef] 15. Liang, Y.; Shu, G.; Tian, H.; Liang, X.; Wei, H.; Liu, L. Analysis of an electricity–cooling cogeneration system based on RC–ARS combined cycle aboard ship. Energy Convers. Manag. 2013, 76, 1053–1060. [CrossRef] 16. Liang, Y.; Shu, G.; Tian, H.; Wei, H.; Liang, X.; Liu, L.; Wang, X. Theoretical analysis of a novel electricity–cooling cogeneration system (ECCS) based on cascade use of waste heat of marine engine. Energy Convers. Manag. 2014, 85, 888–894. [CrossRef] 17. Liang, Y.; Shu, G.; Tian, H.; Sun, Z. Investigation of a cascade waste heat recovery system based on coupling of steam Rankine cycle and NH3-H2O absorption refrigeration cycle. Energy Convers. Manag. 2018, 166, 697–703. [CrossRef] 18. Zhang, N.; Lior, N. Methodology for thermal design of novel combined refrigeration/power binary fluid systems. Int. J. Refrig. 2007, 30, 1072–1085. [CrossRef] 19. Pouraghaie, M.; Atashkari, K.; Besarati, S.M.; Nariman-Zadeh, N. Thermodynamic performance optimization of a combined power/cooling cycle. Energy Convers. Manag. 2010, 51, 204–211. [CrossRef] 20. Demirkaya, G.; Vasquez, P.R.; Goswami, D.Y.; Stefanakos, E.; Rahman, M.M. Analysis of a combined power and cooling cycle for low-grade heat sources. Int. J. Energy Res. 2011, 35, 1145–1157. [CrossRef] 21. Prigmore, D.; Barber, R. Cooling with the Sun’s heat design consideration and test data for a Rankine cycle prototype. Sol. Energy 1975, 17, 185–192. [CrossRef] 22. Wang, H.; Peterson, R.; Harada, K.; Miller, E.; Ingram-Goble, R.; Fisher, L.; Yih, J.; Ward, C. Performance of a combined organic Rankine cycle and vapor compression cycle for heat activated cooling. Energy 2011, 36, 447–458. [CrossRef] Appl. Sci. 2019, 9, 4242 18 of 18

23. Wang, H.; Peterson, R.; Herron, T. Design study of configurations on system COP for a combined ORC (organic Rankine cycle) and VCC (vapor compression cycle). Energy 2011, 36, 4809–4820. [CrossRef] 24. Wali, E. Working fluids for solar, Rankine-cycle cooling systems. Energy 1980, 5, 631–639. [CrossRef] 25. Wali, E. Optimum working fluids for solar powered Rankine cycle cooling of buildings. Sol. Energy 1980, 25, 235–241. [CrossRef] 26. Aphornratana, S.; Sriveerakul, T. Analysis of a combined Rankine-vapor-compression refrigeration cycle. Energy Convers. Manag. 2010, 51, 2557–2564. [CrossRef] 27. Bu, X.; Wang, L.; Li, H. Performance analysis and working fluid selection for geothermal energy-powered organic Rankine-vapor compression air conditioning. Geotherm. Energy 2013, 1, 2. [CrossRef] 28. Bu, X.B.; Li, H.S.; Wang, L.B. Performance analysis and working fluids selection of solar powered organic Rankine-vapor compression ice maker. Sol. Energy 2013, 95, 271–278. [CrossRef] 29. Bu, X.; Wang, L.; Li, H. Working fluids selection for fishing boats waste heat powered organic Rankine-vapor compression ice maker. Heat Mass Transf. 2014, 50, 1479–1485. [CrossRef] 30. Biancardi, F.R.; Sitler, J.W.; Melikian, G. Development and test of solar Rankine cycle heating and cooling systems. Int. J. Refrig. 1982, 5, 351–360. [CrossRef] 31. Demierre, J.; Henchoz, S.; Favrat, D. Prototype of a thermally driven heat pump based on integrated Organic Rankine Cycles (ORC). Energy 2012, 41, 10–17. [CrossRef] 32. Demierre, J.; Favrat, D.; Schiffmann, J.; Wegele, J. Experimental investigation of a Thermally Driven Heat Pump based on a double Organic Rankine Cycle and an oil-free Compressor-Turbine Unit. Int. J. Refrig. 2014, 44, 91–100. [CrossRef] 33. Wan, X.; Cai, L.; Yan, J.; Ma, X.; Chen, T.; Zhang, X. Power management strategy for a parallel hybrid-power gas engine heat pump system. Appl. Therm. Eng. 2017, 110, 234–243. [CrossRef] 34. Shang, S.; Li, X.; Wu, W.; Wang, B.; Shi, W. Energy-saving analysis of a hybrid power-driven heat pump system. Appl. Therm. Eng. 2017, 123, 1050–1059. [CrossRef] 35. Liang, Y.; Al-Tameemi, M.; Yu, Z. Investigation of a gas-fuelled water heater based on combined power and heat pump cycles. Appl. Energy 2018, 212, 1476–1488. [CrossRef] 36. Liang, Y.; Bian, X.; Qian, W.; Pan, M.; Ban, Z.; Yu, Z. Theoretical analysis of a regenerative supercritical carbon dioxide Brayton cycle/organic Rankine cycle dual loop for waste heat recovery of a diesel/natural gas dual-fuel engine. Energy Convers. Manag. 2019, 197, 111845. [CrossRef] 37. Prevention of Air Pollution from Ships-Updated 2000 Study on Greenhouse Gas Emissions from Ships Phase 1 Report; MEPC 58/INF. 6; International Maritime Organization (IMO): London, UK, 2008. 38. Technical File for Hyundai-Man B&W 12K98MC-C Mk6; Hyundai Heavy Industries Co., Ltd.: Ulsan, Korea, 2006. 39. Poling, B.E.; Prausnitz, J.M.; O’Connell, J.P. The Properties of Gases and Liquids; McGraw-Hill: New York, NY, USA, 2001; Volume 5. 40. Shu, G.; Liu, L.; Tian, H.; Wei, H.; Liang, Y. Analysis of regenerative dual-loop organic Rankine cycles (DORCs) used in engine waste heat recovery. Energy Convers. Manag. 2013, 76, 234–243. [CrossRef] 41. Li, X.; Song, J.; Yu, G.; Liang, Y.; Tian, H.; Shu, G.; Markides, C.N. Organic Rankine cycle systems for engine waste-heat recovery: Heat exchanger design in space-constrained applications. Energy Convers. Manag. 2019, 199, 111968. [CrossRef]

© 2019 by the authors. Licensee MDPI, Basel, Switzerland. This article is an open access article distributed under the terms and conditions of the Creative Commons Attribution (CC BY) license (http://creativecommons.org/licenses/by/4.0/).