ORC WHR from marine engine.pdf

Analyzing the optimization of an organic Rankine cycle system for recovering waste heat from a large marine engine containing a cooling water system Min-Hsiung Yang a,⇑ , Rong-Hua Yeh b a Department of Naval Architecture and Ocean Engineering, National Kaohsiung Marine University, Taiwan, Republic of China b Department of Marine Engineering, National Kaohsiung Marine University, Taiwan, Republic of China a r t i c l e i n f o Article history: Received 9 May 2014 Accepted 15 Septemb
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  Analyzing the optimization of an organic Rankine cycle system forrecovering waste heat from a large marine engine containing a coolingwater system Min-Hsiung Yang a, ⇑ , Rong-Hua Yeh b a Department of Naval Architecture and Ocean Engineering, National Kaohsiung Marine University, Taiwan, Republic of China b Department of Marine Engineering, National Kaohsiung Marine University, Taiwan, Republic of China a r t i c l e i n f o  Article history: Received 9 May 2014Accepted 15 September 2014 Keywords: ORCWaste heat recoveryOptimalEvaporationCondensationWorking fluid a b s t r a c t Inthisstudy,sixworkingfluidswithzeroozonedepletionpotentialandlowglobalwarmingpotentialareused in an organic Rankine cycle (ORC) system to recover waste heat from cylinder jacket water of largemarine diesel engines. Thermodynamic analysis and a finite-temperature-difference heat-transfermethodare developedto evaluate the thermal efficiency, total heat-exchanger area, objectiveparameter,and exergy destruction of the ORC system. The optimal evaporation and condensation temperatures forachieving the maximal objective parameter, the ratio of net power output to the total heat-transfer areaof heat exchangers, of an ORC system are investigated.The results show that, among the working fluids, R600a performs the best in the optimal objectiveparameterevaluationfollowedbyR1234ze,R1234yf,R245fa,R245ca,andR1233zdatevaporationtemper-aturesrangingfrom58  Cto68  Candcondensationtemperaturesrangingfrom35  Cto45  C.Theoptimaloperating temperatures and corresponding thermal efficiency and exergy destruction are proposed. Fur-thermore, the influences of inlet temperatures on cylinder jacket water and cooling water in the ORC arepresented for recovering waste heat. The results of this work were verified with theoretical solutionsand experimental results in the literature and it was revealed that they were consistent with them.   2014 Published by Elsevier Ltd. 1. Introduction Because of energy shortages, global warming, and environmen-talpollution, conservingenergyandreducingcarbondioxideemis-sions are becoming increasingly critical for efficient energy use.Waste heat recovery has considerable potential for increasingenergy efficiency and reducing fuel consumption. Although a con-ventional steampower cycle is applied in general industrial powerplants, the performance of the Rankine cycle is insufficient forrecoveringlow-grade waste heat. To enhancethe energyefficiencyand economical use of energy sources, an organic Rankine cycle(ORC) is used to recover low-grade waste heat and transform itintousefulpower[1–3].Inaddition,theapplicationoftheORCsys-tem to the cement, steel, glass, oil, and gas industries cannot onlyrecover the thermal energy but also reduce greenhouse gas [4,5].Because the thermodynamic properties of working fluidssubstantially influence performances of systems, assessing theappropriateness of working fluids for use in the ORC system isessential.SeveralresearchersinvestigatedthesuitabilityoforganicfluidsforheatrecoveryinORCsystems[6–9].Furthermore,XieandYang [10] used the Rankine cycle system to recover waste heatenergy from engines. The results displayed that dry and isentropicfluids were superior to wet fluids because the probability of drop-lets formingas a result of their saturatedvapor characteristics wasreduced. Recently, the studies on converting low-temperature dis-charged heat into electrical energy by using an ORC system forindustrial applications have been reported [11,12].To recover waste heat efficiently, thermodynamic analysis forthe optimized ORC system is crucial. Wei et al. [13] used R245faas the working fluid to optimize the thermodynamic performanceof an ORC system. The result revealedthat when the ambient tem-peraturewas excessivelyhigh, theoutput net powerandefficiencydeteriorated by more than 30% from the nominal state. To recoverthe waste heat, the parametric optimizationof performance analy-sis based on the ORC system were conducted numerically [14,15].Furthermore, an economic factor was considered in the optimiza-tion process of the ORC system. In addition, thermodynamic andthermo-economic optimizations of the ORC system for variouswaste heat source temperatures were performed to obtain the   2014 Published by Elsevier Ltd. ⇑ Corresponding author at: No. 142, Haizhuan Rd., Nanzi Dist., Kaohsiung City81157, Taiwan, Republic of China. Tel.: +886 7 3617141x3404; fax: +886 73656481. E-mail address: (M.-H. Yang).Energy Conversion and Management 88 (2014) 999–1010 Contents lists available at ScienceDirect Energy Conversion and Management journal homepage:  maximal net power output and the minimal apparatus cost[16–19].Employing a geothermal heat source, Shengjun et al. [20]applied two optimization methods to various working fluids inan ORC system. They reported that through the thermodynamicanalysis, R123, R600, R245fa, R245ca, and R600a were suitable.However, through the energy cost evaluation, R152a, R600,R600a, R134a, R143a, R125, and R41 were favorable. In addition,Tian et al. [21] investigated the thermal efficiency and electricityproduction cost of the optimized ORC system and reported thatR141b,R123,andR245fademonstratedmoresuitableperformancecompared with those of various working fluids used for recoveringthe exhaust heat of internal combustion engines. Wang et al. [22]analyzed an ORC system operated with R134a to achieve systemoptimization by maximizing the exergy efficiency and minimizingthe overall apparatus cost under the waste heat source conditions.UsingR12,R123,R134a,andR717asworkingfluidssuperheatedataconstantpressure,Royetal.[23]numericallystudiedanORCsys-tem and presented parametric optimization. They reported thatR123 exhibited maximal thermal efficiency and minimal irrevers-ibility at various turbine inlet pressures.Moreover, the theoretical analysis and exergy evaluation of solar thermal energy of an ORC power plant in reverse osmosisseawater desalination technology were reported [24,25]. Sprouseet al. [26] reviewed an ORC system for internal combustion engineexhaust heat recovery. The results showed a potential improve-ment in fuel economy of approximately 10% through the use of current working fluids and advancements in expander technology.Theapplicationofacogenerationsystem,whichcomprisedanORCand a heat pump, was evaluated numerically [27]. The results of the system performance evaluation revealed that, among theworking fluids used in their study, R236ea and R245ca were supe-rior. Additionally, by using a program code, thermodynamic andtechno-economic analysis of the ORC systems were conductednumerically [28–30].The thermodynamic and transport properties of working fluidssubstantiallyaffecttheperformanceofORCsystems.Moreover,theheat exchange cost becomes critical when the heat source temper-ature is low (80–90  C). To improve the thermal efficiency of ORCsystems, suitable working fluids and optimal working conditionsfor the ORC must be manifest under various conditions. The ther-modynamicandtransportpropertiesoflowglobalwarmingpoten-tial (GWP) working fluids must be considered when analyzingoptimal operational conditions that yield maximal performanceandminimal heat transfer cost for waste heat recoveryin ORCsys-tems. In addition, to improve the energy efficiency design index Nomenclature  A tot   total heat-transfer area of heat exchangers, m 2  A con  heat-transfer area of condenser, m 2  A eva  heat-transfer area of evaporator, m 2 D  diameter, m D h  hydraulic diameter, m ED  exergy destruction, kW  f   dimensionless friction factor  g   acceleration due to gravity, m s  2 h  heat-transfer coefficient, kW m  2  C –1 I   irreversibility, kW i  enthalpy, kJ kg  1 k  thermal conductivity, kW m  1  C –1 L  length of tube or pipe, m L t   thickness of tube wall, m M   molecular weight of working fluid, g mole  1 m  mass flow rate, kg s  1 N   section number of each part in the heat exchangers  p  pressure, kPa P  r   Prandtl number Q   heat transfer rate, kW q  heat flux, kW m  2 Re  Reynolds number T   temperature,   C T  cwi  cooling water inlet temperature,   C T  hwi  cylinder jacket water inlet temperature,   C T  hwo  cylinder jacket water outlet temperature,   C T  ri  working fluid inlet temperature,   C T  ro  working fluid outlet temperature,   C D T   temperature difference between inlet and outlet of theheat exchanger,   C D T  mean  logarithmic mean temperature difference,   C U   overall heat-transfer coefficient of the heat exchangerkW m  2  C  1 v  specific volume, m 3 kg –1 W   power of the turbine or pump, kW Greek symbols c  ratio of   W  net   to  A tot  e  effectiveness g  efficiency l  dynamic viscosity, Pa–s q  density, kg m  3 Subscripts 0 ambient con  condensation, condenser cw  cooling water eva  evaporation, evaporator  f   liquid  g   vapor hw  cylinder jacket water i  inside, inlet II   second law  j  section max  maximal net   net o  outside, optimization  pre  pre-heater  pum  pump r   working fluid s  isentropic sat   saturation sup  superheating t   tube tot   total tur   turbine th  thermal ver   verification w  wall, water wp  water pump  Acronyms EEDI energy efficiency design indexODP ozone depletion potentialORC organic Rankine cycleGWP global warming potential 1000  M.-H. Yang, R.-H. Yeh/Energy Conversion and Management 88 (2014) 999–1010  (EEDI) and reduce greenhouse gas emissions from merchant ships,recovering waste heat from large diesel engines is an essentialmethod [31]. This study investigates the maximal objectiveparameters that represent the maximal ratio of net power outputto heat transfer area for an ORC system for recovering waste heatfrom the cooling water system of large marine diesel engine. Inconsideration of environmental protection, the criteria used toselect the working fluids are zero ozone depletion potential valueand low  GWP  . Table 1 lists the properties of the working fluids[32]. The first and second laws of thermodynamics and the heattransfertheoryofheatexchangeareusedinthisstudyforcalculat-ing the turbine power output, thermal efficiency, exergy destruc-tion, and heat-exchanger area of the ORC system. Furthermore,the maximal objective parameters with the corresponding optimalcondensation, evaporation temperatures, and thermal efficiencyare obtained using R1233zd, R1234yf, R1234ze, R245ca, R245fa,and R600a as working fluids. 2. Thermodynamic modeling and analysis In this study, an ORC system used for recovering waste heatfrom a large marine engine is investigated. This ORC systemprimarily consists of a working fluid pump, evaporator, turbine,condenser, and pre-heater, as shown in Fig. 1. It is assumed thatsteady-stateconditions are applied to all components. In the evap-orator, the working fluid absorbs heat transferred from cylinder jacket water released from the engine and approaches the satura-tion temperature. The working fluid continues to be heated andbecomes saturated vapor, and then becomes superheated vaporat the inlet of the turbine. The superheated vapor produces poweras it passes through the turbine and expands. The low-pressuresuperheated vapor then enters the pre-heater and heats the liquidworkingfluidfromthecondenseroutlet. Subsequently, thecoolingwatercoolstheworkingfluidinthecondenser.Aftercondensation,the liquid working fluid is pumped back into the pre-heater andevaporator to complete the cycle. Moreover, to supply cylinder jacket water and cooling water, water pumps are installed in theORC system. Fig. 2 presents a diagram depicting the temperatureand entropy of the ORC system. Furthermore, the temperaturevariations caused by the heat transfer among the cylinder jacketwater, working fluid, and cooling water are also presented.Fig. 3 presents the relationship between the temperature andentropy of the working fluids used in the ORC system. To preventdamage to the turbine caused by the working fluid becoming sat-urated after generating power in the turbine, working fluids thatyield a saturation line with a positive or nearly vertical slope inthe  T  – s  diagram are used in this study. Obviously, the entropy dif-ference between the saturated liquid and vapor of R600a is thelargest among the working fluids, suggesting that R600a exhibitsthe largest amount of enthalpy change during phase changes thatoccur in heat exchangers. In addition, the critical points of   Table 1 The properties of working fluids [32]. Item R1233zd R1234yf R1234ze R245ca R245fa R600aMolar mass (kg/kmol) 130.5 114.04 114.04 134.05 134.05 58.122 T  cri  (  C) 165.6 94.7 109.36 174.42 154.01 134.66 P  cri  (kPa) 3570.9 3382.2 3634.9 3940 3651 3269ODP 0 0 0 0 0 0GWP 7 4 6 1030 693 20SAFE A1 A2 A2 A1 B1 A3 Note:  ODP: Ozone depletion potential, GWP: Global warming potential.1: No flame propagation; 2: Lower flammability; 3: Higher flammability;A: Lower toxicity; B: Higher toxicity. Fig. 1.  Schematic diagram of the ORC system. Fig. 2.  Temperature and entropy diagram of the ORC system. M.-H. Yang, R.-H. Yeh/Energy Conversion and Management 88 (2014) 999–1010  1001  R245ca and R1234yf are the highest and lowest, respectively,among the six working fluids.The heat flow rate and irreversibility exhibited in the evapora-tor are calculated as Q  e v  a  ¼  m r  ð i 2    i 1 a Þ ð 1 Þ I  e v  a  ¼  T  0 m r   ð s 2    s 1 a Þ   i 2    i 1 a T  e v  a   ð 2 Þ The power output and irreversibility demonstrated by theworking fluid in the turbine can be shown as W  tur   ¼  m r  ð i 3    i 2 Þ = g t   ð 3 Þ I  tur   ¼  T  0 m r  ð s 3    s 2 Þ ð 4 Þ The effectiveness and irreversibility of the pre-heater is definedas e  ¼  T  3    T  3 a T  3    T  1 ð 5 Þ I   pre  ¼  T  0 m r  ½ð s 3    s 3 a Þ  ð s 1 a    s 1 Þ ð 6 Þ Theheatflowrateandirreversibilityexhibitedinthecondenserare expressed as Q  con  ¼  m r  ð i 3 a    i 4 Þ ð 7 Þ I  con  ¼  T  0 m r   ð s 4    s 3 a Þ   i 4    i 3 a T  con   ð 8 Þ The power consumption and irreversibility of the working fluidpump can be calculated as W   pum  ¼  m r  v  4 ð  p 1    p 4 Þ = g  pum  ð 9 Þ I   pum  ¼  T  0 m r  ð s 1    s 4 Þ ð 10 Þ The power consumption of the cylinder jacket water and cool-ing water pumps can be defined as W  wp  ¼  m w q w g  p  f  L w D w q w V  2 w 2 !  ð 11 Þ where  f   is a dimensionless friction factor, and  L w  and  D w  are thelengthandinnerdiameter, respectively, ofthecylinderjacketwaterand cooling water pipes. The net power output of the ORC system can be determined by W  net   ¼  W  tur     W   pum    W  wp ; hw    W  wp ; cw  ð 12 Þ The net thermal efficiency of the ORC system is calculated by g th  ¼  W  net  = Q  e v  a  ð 13 Þ The exergy destruction of the working fluid in the ORC systemcan be obtained by ED  ¼  I  e v  a  þ  I  tur   þ  I  con  þ  I   pum  þ  I   pre  ¼  T  0 m r    i 2    i 1 a T  e v  a   i 4    i 3 a T  con  ð 14 Þ The second law efficiency is calculated by g II   ¼  g th = ð 1   T  0 = T  hw Þ ð 15 Þ 3. Heat transfer analysis A shell-and-tube heat exchanger is designed for the evaporator,condenser, and pre-heater. To calculate the heat transfer coeffi-cient for each phase of the working fluid, the evaporator is dividedintothree parts (the superheating, evaporating, and liquidregions)for the simulation method, as shown in Fig. 2. Similarly, the con-denser comprises two parts: the superheating and condensingregions. The logarithmic mean temperature difference (LMTD) iswidely used in calculating heat transfer rate of heat exchangers.The properties of working fluids and cylinder jacket water andcoolingwatervaryaccordingtothetemperatureduringheattrans-fer between heat exchangers. In this study, to decrease the influ-ence of in transport properties caused by the temperature duringheattransferandtoimprovetheaccuracyofthesimulationresults,each part of the heat exchangers is subdivided into  N   equal sec-tions. The variations of net power output and total heat-exchangerarea of the ORC system for six sets of   N   using R1234yf as workingfluidareevaluatedandgiveninTable2. Itisclearly,thatthediffer-ences in net power output are insignificant for various sectionnumbers,  N  , but the deviations in total heat-exchanger area, whichare evaluated using the transport properties, are obvious. FromTable 2, the  c  of the ORC systembecomes consistent as the sectionnumber increases. It can be obtained that the relative error of   c between N   =20and N   =40 is less than0.1%. Therefore, the numberof sections in each part of the heat exchangers is set as  N   =20throughout this study.The heat-transfer rate between the working fluid and cylinder jacket water of one section of each part in the evaporator can beexpressed as [33] Q   j  ¼  U   j  A  j F  D T  mean ;  j  ð 16 Þ where  j  representsoneof thesectionsof one part intheevaporator, F   is a correction factor for the evaporator, and D T  mean ,  j  is the LMTDbetweenthe cylinder jacket water and workingfluids inthesectionand is obtained by [33] D T  mean ;  j  ¼ ð T  hwi ;  j    T  ro ;  j Þ  ð T  hwo ;  j    T  ri ;  j Þ ln ½ð T  hwi ;  j    T  ro ;  j Þ = ð T  hwo ;  j    T  ri ;  j Þ ð 17 Þ where  T  hwi ,  j  and  T  hwo ,  j  are the inlet and outlet temperatures of thecylinder jacket water respectively, and  T  ri ,  j  and  T  ro ,  j  are the inletand outlet temperatures of the working fluid in the section, respec-tively. The overall heat-transfer coefficient of the section is definedby [33] U   j  ¼  1 ð 1 = h o ;  j Þ þ ð  A o ;  j = h w ;  j Þ þ ð  A o ;  j =  A i ;  j Þð 1 = h i ;  j Þ ð 18 Þ 1 1.2 1.4 1.6 1.8 2 2.2 2.4 2.62060100140180 R1233zdR1234yf R1234zeR245caR245faR600a    T   (   o  C   ) s (kJ/kg- o C) Fig. 3.  The temperature and entropy plots of working fluids.  Table 2 The effect of sections in each part of heat exchangers in calculated results for R1234yf at  D T  hw  = 10   C,  D T  cw  = 8   C,  T  eva  = 65   C, and  T  con  = 40   C. N   1 2 5 10 20 40 w net   (kW) 238.12 238.27 238.39 238.44 238.43 238.44  A  (m 2 ) 379.26 376.36 374.45 373.81 373.83 373.83 c  (kW/m 2 ) 0.6284 0.6328 0.6357 0.6381 0.6384 0.63851002  M.-H. Yang, R.-H. Yeh/Energy Conversion and Management 88 (2014) 999–1010
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