Proceedings of the ASME 2014 International Mechanical Engineering Congress and Exposition IMECE2014 November 14-20, 2014, Montreal, Quebec, Canada

IMECE2014-39976

THERMODYNAMIC ANALYSIS AND OPTIMIZATION OF A COMBINED ORC-VCC SYSTEM USING WASTE FROM A MARINE DIESEL ENGINE

Oumayma Bounefour Ahmed Ouadha Laboratoire d’Energie et Propulsion Navale, Laboratoire d’Energie et Propulsion Navale, Faculté de Génie Mécanique, Université des Faculté de Génie Mécanique, Université des Sciences et de la Technologie Mohamed Sciences et de la Technologie Mohamed BOUDIAF d’Oran, 31000 Oran, Algérie BOUDIAF d’Oran, 31000 Oran, Algérie

ABSTRACT amount of energy produced onboard ships. In some cases, This paper examines through a thermodynamic analysis the onboard and air conditioning systems consume a feasibility of using from marine Diesel engines to similar amount of fuel as propulsion not only for its high drive a vapor compression refrigeration system. Several energy demand but by its continuity in time. Efforts should be working fluids including propane, butane, isobutane and focused on technologies that reduce the energy consumption of propylene are considered. Results showed that isobutane and these systems. Butane yield the highest performance, whereas propane and In the investigation of fuel saving options onboard ships, a propylene yield negligible improvement compared to R134a for great attention is devoted to the study of waste heat recovery operating conditions considered. from Diesel engines. Traditionally used to generate steam that drives turbines dedicated to generate electric power or to INTRODUCTION produce additional mechanical energy to be connected to the Diesel engines are regarded as thermodynamically efficient propulsion shaft in order to reduce fuel consumption, this engines promoted for marine use. Diesel engines whatever their source of energy is nowadays the subject of several designs, four-stroke or two-stroke cycle, have become the most applications. Therefore, many techniques have been explored or common energy production equipment onboard ships and this under exploring to achieve this objective. Especially, practical situation is expected to continue for the foreseeable future. As applications of heat-powered refrigeration cycles where the for road engines, emissions from marine Diesel engines energy required to drive them is mainly in the form of heat and seriously affect the environment and are considered one of the only a very small amount of mechanical or electrical energy is major sources of air pollution. Pollutants from ship emissions needed to circulate the working fluid have been expected may be transported in the atmosphere over several hundred of recently. Among these cycles, absorption refrigeration cycles kilometers, contributing to air quality problems on land. These have been proposed as alternatives to conventional vapor types of engines should face stringent regulations on emissions compression refrigeration systems [1-3]. Although that these control. In particular, efficient use of energy onboard ships is systems can be operated using low grade thermal energy, their one of the prime concerns in design of Diesel engine based performances are very low as compared to conventional vapor modern marine propulsion systems. compression refrigeration systems. Performances of absorption In recent years, there is an increasing need for cooling refrigeration cycles can be increased using multiple stages such onboard ships due to global warming. Cooling and refrigeration as double and triple-effects systems. However, their cost are no less important on-board ships then they are for domestic, increases dramatically and technical problems such as corrosion commercial and industrial applications. Cooling and appear at high temperatures [4-11]. refrigeration systems are necessary for human comfort and the Alternatively the waste heat from the Diesel engine can be preservation of perishables products during voyages. However, used to operate an organic (ORC), which in turn there is no denying that the energy supplied to drive cooling produces the energy necessary to drive the of a and refrigeration systems is increasing continually. Indeed, vapor compression cycle (VCC). The VCC unit can produce conventional refrigeration systems consume an important refrigeration effect at different temperatures. The advantage of

1 Copyright © 2014 by ASME a combined ORC-VCC system compared to absorption a recovery , the cycle efficiency depends mostly refrigeration systems is that when refrigeration is not needed, on the boiler temperature. Considering the cycle efficiency and all the thermal energy can be converted to power and used for environmental issues, they concluded that R245ca is the most others applications. promising out of the cycles considered in their study. Thermally driven refrigeration cycles that combine organic Dubey et al. [17] have presented an energy analysis of a Rankine cycle and vapor compression refrigeration cycle have coupled power-refrigeration cycle which eliminates the received less attention than the others types of thermally requirement of electrical power for driving the compressor of activated systems. Only few studies have been published the vapor compression refrigeration cycle. The coupled cycle recently. Nazer and Zubair [12] have analyzed a Rankine cycle which uses R245ca as the working fluid in top power loop and air-conditioning system using R114 in the power cycle and R22 bottom refrigeration loop have been assessed with different in the vapor compression cycle. The results obtained showed combinations such as cycle with recuperator, reheater, and that the system was more sensitive to vapor compression cycle economizer. Aphornratana and Sriveerakul [18] have described condenser temperature than other system parameters. They a theoretical analysis of a heat-powered refrigeration cycle, a suggested that system performance could be significantly combined Rankine–vapor–compression refrigeration cycle improved by using two separate condensers for the power and which combines an organic Rankine cycle and a vapor vapor compression cycles. Egrican and Karakas [13] have compression cycle. The cycle can be powered by low grade presented an analysis based on second law of thermodynamics thermal energy as low as 60°C and can produce cooling of Rankine cycle/vapor compression cycle using R22 as the temperature as low as −10°C. In the analysis, two combined working fluid. They calculated the maximum reversible work, Rankine–vapor–compression refrigeration cycles have been lost work and availability for each component of the system. investigated: the system with R22 and the system with R134a. Kaushik et al. [14] have presented a thermodynamic analysis Calculated COP values between 0.1 and 0.6 of both systems and assessment of a Freon fluid Rankine cycle cooling system. have been found. Wang et al. [19] have developed the concept A number of working fluid combinations for the Rankine of using waste heat from stationary and mobile engine cycles to engine cycle and vapor compression cycle subsystems have generate cooling for structures and vehicles. It combines an been chosen on the basis of their thermodynamic properties and organic Rankine cycle with a conventional vapor compression their suitability judged in terms of the performance parameters, cycle. A nominal 5 kW cooling capacity prototype system has namely, the thermal efficiency of the power cycle and the been developed based on this concept and tested under coefficient of performance of the refrigeration cycle. They laboratory conditions. In order to maintain high system found that R114+R22 give the best overall system performance performance while reducing size and weight for portable and the presence of the recovery heat exchanger improves the applications, microchannel based heat transfer components and system COP significantly. Kaushik et al. [15] have presented a scroll based expansion and compression were used. Although thermodynamic modeling and a comparative study of the system has been tested off of its design point, it performed single/dual fluid Rankine cycle cooling systems with well achieving 4.4 kW of cooling at a measured heat activated regenerative heat exchangers in the Rankine engine cycle and COP of 0.48. Both conversion and second law efficiencies have the vapor Compression cycle subsystems. They compared, in been close to the model results, proving it to be an attractive particular, the numerical results for single fluids like R12, R22, technology. The measured isentropic efficiency of the scroll R113 and R114 and dual fluids like R113+R12, R113+R22, expander reached 84%, when the pressure ratio was close to the R114+R12, R114+R22, R114+R113 and R113+R114. They scroll intrinsic expansion ratio. The reduced cooling capacity found that in general dual fluid systems give better overall was attributed to off design operation. Wang et al. [20] have system performance as compared to the single fluid systems. introduced a thermally activated cooling concept that combines Amongst single fluids, R114 and amongst dual fluids R114 + an organic Rankine cycle and a vapor compression cycle. A R113 give the best system performance. The presence of brief comparison with other thermally activated cooling recovery heat exchanger in the vapor compression cycle is technologies has been conducted. A systematic design study has more pronounced as compared to that in the Rankine engine been carried out to investigate effects of various cycle cycle in improving the system performance marginally. Jeong configurations on the overall cycle COP. With both subcooling and Kang [16] have developed a novel cycle with Rankine and and cooling recuperation in the vapor compression cycle, the refrigeration cycles, and discussed the thermodynamic analysis overall cycle COP reaches 0.66 at extreme conditions with an of the cycle. Three different have been evaluated to outdoor temperature of 48.9 °C. A parametric trade-off study find the best candidate for the novel combined cycle: R123, has been conducted afterwards in terms of performance and R134a and R245ca. They found that the R123 cycle gives the weight, in order to find the most critical design parameters for highest cycle efficiency among all cycles considered in the the cycle configuration with both subcooling and cooling present study. The base cycle has a low efficiency because of recuperation. Five most important design parameters have been the high temperature at the turbine outlet. By recovering the selected, including expander isentropic efficiency, condensing heat at the turbine outlet, the overall COP increases by 47% in and evaporating temperatures, /boiling pressure and the case of R245ca cycle. In the base cycle, COP depends recuperator effectiveness. At the end, two additional cycle mostly on the boiler pressure, while in the modified cycle with concepts with either potentially higher COP or practical

2 Copyright © 2014 by ASME advantages have been proposed. Demierre et al. [21] have best methodology would be to begin with a simple and reliable studied the concept of a low power ORC–ORC thermodynamic model which could quickly evaluates and system. The system provides about 20 kW heat at the condenser compares thermodynamically a large list of refrigerants. More and uses R134a as the working fluid for both cycles. A first efficient refrigerants could then be investigated in-depth using experimental setup has been built to test the pump and the experimental and full system modeling approaches. The supercritical evaporator. A comparison between the results development of the previously mentioned thermodynamic obtained with an in-house supercritical evaporator simulation model is, in fact, the aim of this paper. program and measurements is presented. The design steps of Nowadays, analysis is became a powerful tool in the compressor-turbine are briefly presented. The compressor- the design and performance judgment of any energy system. It turbine unit has been balanced and tested, with air, at speeds up mainly consists to identify the components of a system which to 140,000 rpm. Li et al. [22] have analyzed and evaluated are responsible for losses in order to minimize them. Although hydrocarbons including propane, butane, isobutane and exergy analysis has been used for several years in the analysis propylene as working fluids used in organic Rankine cycle of all types of energy systems, its application to combined powered vapor compression refrigeration system. With the ORC-VCC systems has to some extent been limited in the overall COP and working fluid mass flow rate per kW cooling literature. In this study, energy and exergy analyses of a capacity as key performance indicators, the results have combined organic Rankine cycle (ORC) and vapor compression indicated that butane is the best refrigerant for the combined refrigeration cycle (VCC) using waste heat from a marine system as the boiler exit temperature is between 60 and 90°C, Diesel engine are performed. A number of natural fluids, the condensation temperature varies from 30 to 55°C and the including propane (R290), butane (R600), isobutane (R600a) evaporation temperature ranges from −15 to 15°C. When the and propylene (R1270) have been analyzed in order to identify boiler exit temperature reaches 90°C and the other input the most suitable fluids for different operating conditions. parameters are in typical values, the overall COP of the butane Exergy losses in each component of the system and the exergy case reaches 0.470. Bu et al. [23] have developed a efficiency of the overall system are determined in order to thermodynamic model for an organic Rankine cycle/vapor identify the causes and locations of thermodynamic compression cycle system that uses from hot irreversibilities. springs. Six working fluids (R123, R134a, R245fa, R600a, R600 and R290) have been selected and compared in order to MARINE DIESEL ENGINE ENERGY BALANCE identify suitable working fluids which may yield high system Although the great technological development of modern efficiencies. Results showed that because of high system Diesel engines, only a part of the energy contained in the fuel is pressure for R290 and R134a, R600a is the more suitable converted to power output. The maximum efficiency remains working fluid for ORC in terms of expander size parameter, lower than 45%. The main losses are dissipated as heat in the system efficiency and system pressure. In addition, R600a is exhaust , coolants, and transferred to the environment. also the most appropriate working fluid for VCC in terms of An energy balance of a marine Diesel engine indicates how pressure ratio and coefficient of performance. R600 and R600a the energy contained in the fuel is used or lost. Indeed, the are more suitable working fluids for ORC-VCC in terms of injection of a mass of fuel in the hot air in the overall coefficient of performance, refrigerating capacity per chamber produces a large amount of heat. However, the unit mass flow rate and chilled water yield from per ton hot mechanical work requires only a fraction of the heat produced. water. The residual heat is, in fact, discharged at various places during Onboard ships, whatever their types, HCFC22 has been the its stay in the cylinder. For a marine Diesel engine, a first-law most commonly used refrigerant in air conditioning and analysis yields refrigeration systems. However, accordingly to the Montreal QWQQQ    (1) protocol which prohibited the HCFCs in equipment produced in c ex r after 31 December 1997, alternative refrigerants, mainly where, Qin is the heat supplied to the Diesel engine, W the hydrofluorocarbons (HFCs) such as R134a, R404A, R410A and mechanical energy output, Q the heat transferred to the cooling R407C have been used. Unfortunately, alternative refrigerants c are now considered as potential global warming gases. This systems, Qex the heat rejected in exhaust gases, Qr the heat situation combined to energy efficiency of refrigeration systems transferred to the environment by radiation. has encouraged manufacturers to reconsider some natural Figure 1 shows an example of an energy balance of a refrigerants such as , CO2 and some hydrocarbons. modern marine Diesel engine. The heat transferred to coolants These refrigerants are naturally available and environment includes charge air cooling (17.8%), jacket water cooling friendly and it has been widely reported in literature that natural (4.8%) and lubricating oil cooling (3.2%). In addition, 25.1% of refrigerants performances are higher as compared to the others the total energy is lost, released into the atmosphere during the synthetic refrigerants. exhaust outlet. Finally, a small part of the energy (0.6%) is The analysis of systems performance of a large number of released by radiation. At first glance, the analysis of heat flows possible refrigerants using detailed experimentation or full shows that there are three engine waste heat streams, at system modeling is time consuming and cost prohibitive. The different temperature levels, that have potential to be recovered:

3 Copyright © 2014 by ASME exhaust (300-600°C); charge air (200°C); jacket water (80- acceptable by our environment. Their contribution to global 100°C). warming problems is not negligible.

Type Wartsila 4L20 Configuration 4 in-line cylinders Engine speed, RPM 1000 Engine output, kW 720 Cylinder bore, mm 200 Stroke, mm 280 Swept volume, dm3 35.2 Compression ratio 15 Exhaust gas flow (100 % load), kg/s 1.46 Exhaust gas temperature (100 % load), °C 350 Fuel consumption (100% load), g/kW h 196 Tab. 1: Engine characteristics

Fig. 1: Typical heat balance of a marine Diesel engine In addition to thermophysical and chemical properties, essential criteria, it has now become crucial to consider An example of main Diesel engine characteristics at 1000 rpm environmental and economic aspects in the choice of a working is summarized in Table 1 [24]. fluid. Some hydrocarbons such as propane (R290), butane (R600), isobutane (R600a) and propylene (R1270) can be WORKING FLUIDS SELECTION considered as objective choices because they are abundant, Organic fluids such as R123 and R245fa are commonly used as inexpensive and they have excellent thermophysical properties. the working fluids for the ORC-VCC system. Although these The main thermophysical properties, security properties based fluids present several advantages including excellent on ASHRAE 34 and environmental properties of the selected thermodynamic properties and safety, they are not completely fluids are shown in Table 2. Thermophysical data Environmental properties Security Substance Type M T T p b c c group ODP GWP (g/mol) (°C) (°C) (bar) R134a i 102.0 -26.1 101.1 40.6 A1 0 1430 R290 w 44.1 -42.1 96.7 42.5 A3 0 -20 R600 d 58.1 -0.5 152.0 38.0 A3 0 -20 R600a d 58.1 -11.7 134.7 36.3 A3 0 -20 R1270 d 42.08 -47.62 91.75 46.0 A3 0 -20 Tab. 2: Thermophysical, security and environmental properties of the selected working fluids

Working fluids for organic Rankine cycles can be classified according to the slope of the saturated vapor line in a T–s 160 R134a diagram into wet, isentropic or dry fluids. Figure 2 indicates 120 R290 that wet fluids have high critical temperatures and large latent R600 heat compared to dry fluids. A large involves smaller 80 R600a mass flow rates. However, dry fluids which are characterized R1270 by a positively sloped saturation curve in the T–s diagram have 40 better thermal performance because the fluid does not condense 0 after its expansion in the expander unlike wet fluids which partially condense after their expansion. For wet working (°C) Temperature -40 fluids, it is important to make sure that the vapor at the -80 expander exit is dry to prevent the expander from liquid-hit. 0.0 0.5 1.0 1.5 2.0 2.5 3.0 Control of the vaporization pressure of the ORC cycle can (kJ/kg°C) avoid the above issue.

Fig 2: T-s saturated curves comparison of the working fluids Thermodynamic parameters including enthalpy and entropy at different states have been obtained using a set of

4 Copyright © 2014 by ASME equations of state. The computational model adopted in this refrigeration cycle. The system uses the same working fluid for study is based on four local equations of state presented below: both power and refrigeration cycles. The power cycle identified - an equation of state for the gas state, as 1-2-3-4-1 produces the mechanical energy required to drive ZZT ,   (2) the compressor of the refrigeration cycle identified as 5-6-7-3- 5. Both subsystems share a common condenser. The combined - a correlation for the saturated vapor pressure, ORC-VCC system under study consists of recovery heat p p T (3) ss  exchanger, an expander, a condenser and a pump for the ORC - a correlation for the saturated liquid density, system and a compressor, a condenser, a throttling valve and an evaporator for the VCC system as illustrated in Fig. 3. The LL T  (4) recovery heat exchanger uses waste heat from the Diesel engine - an equation of the specific heat capacity at constant to heat and vaporize the working fluid. The vapor produced pressure in the ideal gas state. (state 6) is expanded in an expander producing the mechanical c00 c T (5) pp  energy necessary to drive the compressor of the VCC system. Using the above equations and the differential equations of The fluid exiting the expander (state 7) is condensed (7-3) and thermodynamics [25], it is possible to calculate the other pumped (3-5) back to the recovery heat exchanger to undergo a essential thermodynamic functions necessary for the new cycle. For the VCC system, the compressor uses the thermodynamic analysis, namely, the enthalpy, the entropy and mechanical energy produced by the ORC system to increase the exergy. pressure and temperature of the working fluid to attain the The model has been successfully used for a two-stage condenser (state 2) where it is condensed to saturated liquid refrigeration cycle [26], a cascade refrigeration cycle [27], a (state 3) before to be throttled in the throttling valve to reach cycle [28] and a water-to-water heat pump [29]. the evaporator pressure (state 4). In the evaporator, the working fluid vaporizes by extracting heat from the cooling space (4-1). The vapor produced is aspirated by the compressor to undergo a SYSTEM DESCRIPTION new cycle. Figure 3 depicts the schematic and T-s diagrams of the combined organic Rankine cycle and vapor compression

Pressure T , p 5 boil boil 6

T , p 3 cond cond 7 2 t , p 4 evap evap 1

Enthalpy

Fig. 3: Schematic and p-h diagram of the combined organic Rankine Cycle and vapor compression cycle defined in terms of temperature, pressure and chemical THERMODYNAMIC ANALYSIS composition. Habitually, thermodynamic cycles are analyzed using the After the determination of the working fluid properties at energy analysis method which is based on the first law of each point of the cycle represented in Fig. 3, the calculation of thermodynamics, i.e. the energy conservation concept. the system performance can be carried out. The thermodynamic Unfortunately, this method cannot locate the degradation of the model is based on the following assumptions: quality of energy. Instead, exergy analysis which is based on - all processes are marked by steady state and steady flow, both the first and second laws of thermodynamics can - kinetic and potential energy have negligible effects, overcome easily the limitations of the energy analysis. It - heat and friction losses are neglected, permits to quantify the magnitude and the location of exergy - the working fluid leaving the condenser and the losses within the system. Furthermore, the total exergy losses evaporator is saturated, can be considered as an optimization criteria which, by - the temperature and pressure of the environment are minimization, provide optimum processes configuration. The 25°C and 1 atm, respectively. concept and the methodology of exergy analysis are well- Mass, energy and exergy balances for any control volume documented in the literature [30-32]. Bosnjakovic [33] has at steady state with negligible kinetic, potential and chemical defined exergy as the theoretically gainful amount of work energy changes can be expressed, respectively, by obtained by bringing materials into equilibrium with the mmin  out (6) surroundings in a reversible process. The surroundings can be

5 Copyright © 2014 by ASME The rate of exergy, , is given by Q W  mout h out  min h in (7) Ex m ex (10) Exheat  W  Exout  Exin   Ex (8) where, ex is the specific exergy defined as where subscriptions in and out denote inlet and outlet states, Q ex h  h0  T 0 s  s 0  (11) and W the net heat and work inputs, m the mass flow rate, h The subscript 0 is related to the reference state. the enthalpy, Ex the rate of exergy and Ex is the rate of From a second law point of view, it is important to quantify exergy losses. the exergy losses in each component in order to assess the The rate of exergy transfer by heat, Ex , at a temperature T, is overall performance of the system. Mass, energy and exergy heat balance equations are applied to each component of the system. given by The mathematical model for both ORC and VCC subsystems is T0 given by the equations summarized in Table 3. Exheat 1 Q (9) T

Component Energy balance Exergy balance

Pump Wpump m ORC  h53 h  Expump  Ex35  Ex  Wpump

T0 Boiler Qboil m ORC  h65 h  Exboil  Ex56  Ex  Qboil 1  * Tboil

Expander Wexp  mORC  h6 h 7  Extur  Ex6  Ex 7  W exp

Condenser Qcond m ORC  h23 h  Excond  Ex7  Ex 2  Ex 3 T Evaporator Q m h h Ex  Ex  Ex  Q 1  0 evap VCC 14  evap 41evap * Tevap

Compressor Wcomp m VCC  h21 h  Ecomp  Ex12  Ex  Wcomp

Throttling valve hh34 Ethrot  Ex34  Ex

WWQtur pump evap COP   Extot Cycle W ex 1 Qboil comp Exboil ORC COPVCC Tab. 3: Energy and exergy balances equations

(R600a) and propylene (R1270) have been selected as working The exergy supplied to the system, Exboil , and the total exergy fluids. Performance of the system including the power output, losses, Ex , can be expressed as tot thermal efficiency, individual exergy losses, total exergy losses T and exergy efficiency can be determined by varying some Ex Q 1 0 (12) boil boil * operating parameters of the system shown in Fig. 3. The main Tboil parameters used in the analysis are shown in Table 4. Ex   Ex   Ex   Ex   Ex  tot pump boil tur cond (13) Simulation results have been obtained based on 1 kg/s working fluid mass flow rate in the ORC subsystem. Waste heat Exevap   Ex comp   Ex throt from Diesel engines can be provided at temperatures ranging * * Tevap and Tboil refer to low-temperature reservoir and high- from 60 to 300°C. The boiler exit temperatures range from 60 temperature reservoir, respectively. They are calculated using to 90, which corresponds to heat source temperatures ranging * from 70 to 100°C. For a heat source temperature about 100°C, TTTboil boil   (14) the maximum boiler exit temperature is 90°C. The natural  * TTTevap evap   fluids critical temperatures which are larger than 90°C allow the use of subcritical cycles for both power and refrigeration SIMULATION RESULTS AND DISCUSSIONS subsystems. The condensation temperature can be varied The above equations are used to develop a FORTRAN between 30 and 50°C. To cover a wide range of applications, program for the calculation of the system performance in order the evaporation temperature is varied from -10 to 10°C. to enable comparisons among different working fluids for Generally, the isentropic efficiency of a turbine ranges from 80 several operating conditions. Four pure hydrocarbon to 90% whereas the isentropic efficiency of compressor ranges refrigerants, namely propane (R290), butane (R600), isobutane from 70 to 90%.

6 Copyright © 2014 by ASME Working fluid mass flow rate in ORC, kg/s 1.0 300 Evaporation temperature, °C 5 R134a Boiler exit temperature, °C 60–90 250 R290 Condensation temperature, °C 40 R600 Expander isentropic efficiency, % 75 200 R600a Compressor isentropic efficiency, % 80 R1270 Pump isentropic efficiency, % 75 150 Tab. 4: Design and operating parameters of the system 100 In absence of relevant suitable model to validate such configuration of the whole combined system, the validation has 50 been conducted for both the two subsystems separately. Thus, Refrigeration effect (kW) the accuracy of simulation results have been checked by 0 60 65 70 75 80 85 90 comparing calculated values using the developed code to values obtained using the free code SOLKANE. Computations have Boiler exit temperature (°C) been carried out according to the parameters listed in Table 5. Compression and expansion processes are assumed isentropic. Figure 5 depicts the effect of the boiler exit temperature on the net power produced by the ORC subsystem. For R134a, Fluid R134a R600 and R600a, the ORC net power output increases linearly Mass flow rate in ORC, kg/s 1 with increasing turbine inlet temperature. However, for R290 Boiler temperature, °C 90 and R1270, the ORC net power output first increases than Condenser temperature, °C 30 decreases with increasing the boiler exit temperature. The Evaporation temperature, °C -10 maximum power output is produced by a system using R600, Refrigeration effect, kW 1 followed, in order, by R600a, R290, R1270 and R134a. Tab. 5: Validation case operating parameters 0.6 Table 6 presents comparisons between the two codes. It is clear that the model proposed is reliable in predicting the performance of both systems. The relative error for all 0.5 parameters considered remains below 2.7%. In addition, the program developed has the advantage to be used more flexibly in order to determine the cycle performance for different 0.4 operating conditions. COP

0.3 Parameter Model Solkane  (%) Pump power, kW 1.4 1.4 0

Turbine power, kW 12.7 12.7 0 0.2 60 65 70 75 80 85 90 ORC Mass flow rate, g/s 505.9 512.0 -1.2 Thermal efficiency, % 11.3 11.0 2.7 Boiler exit temperature (°C) Compressor power, kW 0.19 0.19 0

Fig. 4: Influence of the boiler exit temperature on the Pressure ratio 3.83 3.84 -0.26 refrigerating effect and the coefficient of performance

VCC Mass flow rate, g/s 6.623 6.633 -0.15 COP 5.40 5.40 0 Figure 6 illustrates the percentage of the exergy destroyed Tab. 6: Validation case results in each component with respect to the total system exergy loss Results based on energy analysis, in terms of refrigerating for each refrigerant. For all refrigerants, the highest exergy effect and COP, the standard criterions of performance analysis losses occur in the evaporator (36-39%) followed by the of a refrigeration cycles, are plotted versus the boiler exit condenser (25-27%), the boiler (10-13%) and the expander (10- temperature in Fig. 4. Whatever the refrigerant used, the 11%). Exergy losses in the compressor, the throttling valve and increase of the boiler exit temperature enhances the the pump are negligible comparatively to the above exergy performance of the system. As expected, all hydrocarbons losses. refrigerants give better refrigerating effect than R134a. However, in terms of COP, propane and propylene give comparable results to R134a. Isobutane is the best refrigerant considering the COP.

7 Copyright © 2014 by ASME 50 R134a REFERENCES [1] Fernandez-Seara, J., Vales, A., Vazquez, M., 1998, “Heat R290 40 recovery system to power an onboard NH3-H2O absorption R600 refrigeration plant in trawler chiller fishing vessels”, Appl. R600a 30 Thermal Eng., 12, pp. 1189–1205. R1270 [2] Norihiro, I., Yuta T., Masao, G., 2006, “Application of Low Temperature Waste Heat from Ship Engine to Air- 20 conditioning for reduction of GHG Emissions”, Marine Eng., 41, pp. 444-450.

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