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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 Chinab 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 September 2014
Keywords:
ORC
Waste heat recovery
Optimal
Evaporation
Condensation
Working fluid
a b s t r a c t
In this study, six workingfluids with zero ozone depletion potential and lowglobal warmingpotential are
used in an organic Rankine cycle (ORC) system to recover waste heat from cylinder jacket water of large
marine diesel engines. Thermodynamic analysis and a finite-temperature-difference heat-transfer
method are developed to evaluate the thermal efficiency, total heat-exchanger area, objective parameter,
and exergy destruction of the ORC system. The optimal evaporation and condensation temperatures for
achieving the maximal objective parameter, the ratio of net power output to the total heat-transfer area
of heat exchangers, of an ORC system are investigated.
The results show that, among the working fluids, R600a performs the best in the optimal objective
parameter evaluationfollowed by R1234ze, R1234yf,R245fa,R245ca,and R1233zdat evaporation temper-
atures rangingfrom 58 Cto68 C andcondensation temperatures rangingfrom 35 Cto45 C. Theoptimal
operating 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 are
presented for recovering waste heat. The results of this work were verified with theoretical solutions
and 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-
tal pollution, conserving energy and reducing carbon dioxide emis-
sions are becoming increasingly critical for efficient energy use.
Waste heat recovery has considerable potential for increasing
energy efficiency and reducing fuel consumption. Although a con-
ventional steam power cycle is applied in general industrial power
plants, the performance of the Rankine cycle is insufficient for
recovering low-grade waste heat. To enhance the energy efficiency
and economical use of energy sources, an organic Rankine cycle
(ORC) is used to recover low-grade waste heat and transform it
into useful power [13]. In addition, the application of the ORC sys-
tem to the cement, steel, glass, oil, and gas industries cannot only
recover the thermal energy but also reduce greenhouse gas[4,5].
Because the thermodynamic properties of working fluids
substantially influence performances of systems, assessing the
appropriateness of working fluids for use in the ORC system is
essential. Several researchers investigated the suitability of organic
fluids for heat recovery in ORC systems [69]. Furthermore, Xie and
Yang [10] used the Rankine cycle system to recover waste heat
energy from engines. The results displayed that dry and isentropic
fluids were superior to wet fluids because the probability of drop-
lets forming as a result of their saturated vapor characteristics was
reduced. Recently, the studies on converting low-temperature dis-
charged heat into electrical energy by using an ORC system for
industrial applications have been reported[11,12].
To recover waste heat efficiently, thermodynamic analysis for
the optimized ORC system is crucial. Wei et al. [13]used R245fa
as the working fluid to optimize the thermodynamic performance
of an ORC system. The result revealed that when the ambient tem-
perature was excessively high, the output net power and efficiency
deteriorated by more than 30% from the nominal state. To recover
the waste heat, the parametric optimization of 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 and
thermo-economic optimizations of the ORC system for various
waste heat source temperatures were performed to obtain the
http://dx.doi.org/10.1016/j.enconman.2014.09.044
0196-8904/2014 Published by Elsevier Ltd.
Corresponding author at: No. 142, Haizhuan Rd., Nanzi Dist., Kaohsiung City
81157, Taiwan, Republic of China. Tel.: +886 7 3617141x3404; fax: +886 7
3656481.
E-mail address: mhyang@webmail.nkmu.edu.tw(M.-H. Yang).
Energy Conversion and Management 88 (2014) 9991010
Contents lists available at ScienceDirect
Energy Conversion and Management
j o u r n a l h o m e p a g e : w w w . e l s e v i e r . c o m / l o c a t e / e n c o n m a n
http://dx.doi.org/10.1016/j.enconman.2014.09.044mailto:mhyang@webmail.nkmu.edu.twhttp://dx.doi.org/10.1016/j.enconman.2014.09.044http://www.sciencedirect.com/science/journal/01968904http://www.elsevier.com/locate/enconmanhttp://www.elsevier.com/locate/enconmanhttp://www.sciencedirect.com/science/journal/01968904http://dx.doi.org/10.1016/j.enconman.2014.09.044mailto:mhyang@webmail.nkmu.edu.twhttp://dx.doi.org/10.1016/j.enconman.2014.09.044http://-/?-http://-/?-http://-/?-http://-/?-http://-/?-http://-/?-http://crossmark.crossref.org/dialog/?doi=10.1016/j.enconman.2014.09.044&domain=pdfhttp://-/?- -
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maximal net power output and the minimal apparatus cost
[1619].
Employing a geothermal heat source, Shengjun et al. [20]
applied two optimization methods to various working fluids in
an ORC system. They reported that through the thermodynamic
analysis, 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 electricity
production cost of the optimized ORC system and reported that
R141b, R123, and R245fa demonstrated more suitable performance
compared with those of various working fluids used for recovering
the exhaust heat of internal combustion engines. Wang et al. [22]
analyzed an ORC system operated with R134a to achieve system
optimization by maximizing the exergy efficiency and minimizing
the overall apparatus cost under the waste heat source conditions.
Using R12, R123, R134a, and R717 as working fluids superheated at
a constant pressure, Roy et al. [23] numerically studied an ORC sys-
tem and presented parametric optimization. They reported that
R123 exhibited maximal thermal efficiency and minimal irrevers-
ibility at various turbine inlet pressures.
Moreover, the theoretical analysis and exergy evaluation ofsolar thermal energy of an ORC power plant in reverse osmosis
seawater desalination technology were reported [24,25]. Sprouse
et al.[26]reviewed an ORC system for internal combustion engine
exhaust 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.
The application of a cogeneration system, which comprised an ORC
and 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 and
techno-economic analysis of the ORC systems were conducted
numerically[2830].
The thermodynamic and transport properties of working fluids
substantially affect the performance of ORC systems. Moreover, the
heat exchange cost becomes critical when the heat source temper-
ature is low (8090 C). To improve the thermal efficiency of ORC
systems, suitable working fluids and optimal working conditions
for the ORC must be manifest under various conditions. The ther-
modynamic and transport properties of low global warming poten-
tial (GWP) working fluids must be considered when analyzing
optimal operational conditions that yield maximal performance
and minimal heat transfer cost for waste heat recovery in ORC sys-tems. In addition, to improve the energy efficiency design index
Nomenclature
Atot total heat-transfer area of heat exchangers, m2
Acon heat-transfer area of condenser, m2
Aeva heat-transfer area of evaporator, m2
D diameter, mDh hydraulic diameter, m
ED exergy destruction, kWf dimensionless friction factorg acceleration due to gravity, m s2
h heat-transfer coefficient, kW m2 C1
I irreversibility, kWi enthalpy, kJ kg1
k thermal conductivity, kW m1 C1
L length of tube or pipe, mLt thickness of tube wall, mM molecular weight of working fluid, g mole1
m mass flow rate, kg s1
N section number of each part in the heat exchangersp pressure, kPaPr Prandtl numberQ heat transfer rate, kW
q heat flux, kW m2
Re Reynolds numberT temperature, CTcwi cooling water inlet temperature, CThwi cylinder jacket water inlet temperature, CThwo cylinder jacket water outlet temperature, CTri working fluid inlet temperature, CTro working fluid outlet temperature, CDT temperature difference between inlet and outlet of the
heat exchanger, CDTmean logarithmic mean temperature difference, CU overall heat-transfer coefficient of the heat exchanger
kW m2 C1
v specific volume, m3 kg1
W power of the turbine or pump, kW
Greek symbolsc ratio ofWnetto Atote effectiveness
g efficiencyl dynamic viscosity, Pasq density, kg m3
Subscripts0 ambientcon condensation, condensercw cooling watereva evaporation, evaporatorf liquidg vaporhw cylinder jacket wateri inside, inletII second lawj sectionmax maximalnet neto outside, optimizationpre pre-heaterpum pumpr working fluids isentropicsat saturationsup superheatingt tubetot totaltur turbineth thermalver verificationw wall, waterwp water pump
AcronymsEEDI energy efficiency design index
ODP ozone depletion potentialORC organic Rankine cycleGWP global warming potential
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(EEDI) and reduce greenhouse gas emissions from merchant ships,
recovering waste heat from large diesel engines is an essential
method [31]. This study investigates the maximal objective
parameters that represent the maximal ratio of net power output
to heat transfer area for an ORC system for recovering waste heat
from the cooling water system of large marine diesel engine. In
consideration of environmental protection, the criteria used to
select 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 heat
transfer theory of heat exchange are used in this study for calculat-
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 optimal
condensation, evaporation temperatures, and thermal efficiency
are 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 heat
from 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 that
steady-state conditions 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 and
becomes saturated vapor, and then becomes superheated vapor
at the inlet of the turbine. The superheated vapor produces power
as it passes through the turbine and expands. The low-pressure
superheated vapor then enters the pre-heater and heats the liquid
working fluid from the condenser outlet. Subsequently, the cooling
water cools the working fluid in the condenser. After condensation,
the liquid working fluid is pumped back into the pre-heater and
evaporator to complete the cycle. Moreover, to supply cylinderjacket water and cooling water, water pumps are installed in the
ORC system.Fig. 2presents a diagram depicting the temperature
and entropy of the ORC system. Furthermore, the temperature
variations caused by the heat transfer among the cylinder jacket
water, working fluid, and cooling water are also presented.
Fig. 3presents the relationship between the temperature and
entropy of the working fluids used in the ORC system. To prevent
damage to the turbine caused by the working fluid becoming sat-
urated after generating power in the turbine, working fluids that
yield a saturation line with a positive or nearly vertical slope in
theTsdiagram are used in this study. Obviously, the entropy dif-
ference between the saturated liquid and vapor of R600a is the
largest among the working fluids, suggesting that R600a exhibits
the 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 R600a
Molar mass (kg/kmol) 130.5 114.04 114.04 134.05 134.05 58.122
Tcri (C) 165.6 94.7 109.36 174.42 154.01 134.66
Pcri (kPa) 3570.9 3382.2 3634.9 3940 3651 3269
ODP 0 0 0 0 0 0
GWP 7 4 6 1030 693 20
SAFE 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.
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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
Qeva mri2i1a 1
Ieva T0mr s2s1a i2i1aTeva
2
The power output and irreversibility demonstrated by the
working fluid in the turbine can be shown as
Wtur mri3i2=gt 3
Itur T0mrs3s2 4
The effectiveness and irreversibility of the pre-heater is defined
as
e T3T3aT3T1
5
Ipre T0mrs3s3a s1as1 6
The heat flow rate and irreversibility exhibited in the condenser
are expressed as
Qcon mri3ai4 7
Icon T0mr s4s3a i4i3aTcon
8
The power consumption and irreversibility of the working fluid
pump can be calculated as
Wpum mrv4p1p4=gpum 9
Ipum T0mrs1s4 10
The power consumption of the cylinder jacket water and cool-
ing water pumps can be defined as
Wwp mwqwgp
f LwDw
qwV2
w
2
! 11
where fis a dimensionless friction factor, and Lw and Dw are the
length and inner diameter, respectively, of the cylinder jacket water
and cooling water pipes.
The net power output of the ORC system can be determined by
Wnet Wtur WpumWwp;hwWwp;cw 12
The net thermal efficiency of the ORC system is calculated by
gth Wnet=Qeva 13
The exergy destruction of the working fluid in the ORC system
can be obtained by
ED IevaItur IconIpumIpre T0mr i2i1aTeva
i4i3aTcon
14
The second law efficiency is calculated by
gII gth=1 T0=Thw 15
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 divided
into three parts (the superheating, evaporating, and liquid regions)
for the simulation method, as shown in Fig. 2. Similarly, the con-
denser comprises two parts: the superheating and condensing
regions. The logarithmic mean temperature difference (LMTD) is
widely used in calculating heat transfer rate of heat exchangers.
The properties of working fluids and cylinder jacket water and
cooling water vary according to the temperature during heat trans-fer between heat exchangers. In this study, to decrease the influ-
ence of in transport properties caused by the temperature during
heat transfer and to improve the accuracy of the simulation results,
each part of the heat exchangers is subdivided into N equal sec-
tions. The variations of net power output and total heat-exchanger
area of the ORC system for six sets ofNusing R1234yf as working
fluid are evaluated and given in Table 2. It is clearly, that the differ-
ences in net power output are insignificant for various section
numbers,N, but the deviations in total heat-exchanger area, which
are evaluated using the transport properties, are obvious. From
Table 2,thec of the ORC system becomes consistent as the section
number increases. It can be obtained that the relative error of c
between N= 20 and N= 40 is less than 0.1%. Therefore, the number
of sections in each part of the heat exchangers is set as N= 20
throughout this study.
The heat-transfer rate between the working fluid and cylinder
jacket water of one section of each part in the evaporator can be
expressed as[33]
Qj UjAjFDTmean;j 16
wherej represents one of the sections of one part in the evaporator,
Fis a correction factor for the evaporator, and DTmean,j is the LMTD
between the cylinder jacket water and working fluids in the section
and is obtained by[33]
DTmean;j Thwi;jTro;j Thwo;jTri;j
lnThwi;jTro;j=Thwo;jTri;j 17
whereThwi,j and Thwo,j are the inlet and outlet temperatures of the
cylinder jacket water respectively, and Tri,j and Tro,j are the inlet
and outlet temperatures of the working fluid in the section, respec-
tively. The overall heat-transfer coefficient of the section is defined
by[33]
Uj 1
1=ho;j Ao;j=hw;j Ao;j=Ai;j1=hi;j 18
1 1.2 1.4 1.6 1.8 2 2.2 2.4 2.620
60
100
140
180R1233zdR1234yfR1234zeR245caR245faR600a
T(oC
)
s (kJ/kg-oC)
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 DThw= 10 C, DTcw= 8 C,Teva = 65 C, and Tcon= 40 C.
N 1 2 5 10 20 40
wnet(kW) 238.12 238.27 238.39 238.44 238.43 238.44
A(m2) 379.26 376.36 374.45 373.81 373.83 373.83
c (kW/m2) 0.6284 0.6328 0.6357 0.6381 0.6384 0.6385
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whereho,j and hi,j are the heat-transfer coefficients of the working
fluid and cylinder jacket water respectively, and Ao represents the
outside surface area of the tubes in the section.
The dimensional empirical expression for nucleate boiling in
the evaporator is used to calculate the nucleate-boiling heat-
transfer coefficient of the working-fluid side[34]:
ho 55Pr0:120:4343lnRPr 0:4343 lnPrr
0:55M0:5q
0:67 19
whereMis the molecular weight of the working fluid, q is the heat
flux of the tube, andRpis set to 1.0 lm for the surface roughness of
the tube. The heat-transfer coefficient of the water side can be cal-
culated using the DittusBoelter correlation for 6000 < Re < 107 and
0.5 < Pr < 120[33]:
hi 0:023Re0:8w Pr
nw
Dhkw
20
where n =0.4 is used for the condenser and n =0.3 is used for the
evaporator. The corresponding heat-transfer coefficient of the tube
wall is calculated as [33]
hw 2pktLtlnDo=Di
21
Furthermore, the correlation proposed by Zukauskas [35] is
applied to calculate the heat-transfer coefficient on the working
fluid side for superheating vapor or subcooling liquid, which gives
ho krDo
0:71Re0:5r Pr
0:36r
PrrPrw
n22
wheren =0 is used for the superheating vapor and n =0.25 is used
for the liquid. In addition, Prw is evaluated at the wall temperature
of the tubes.
Similarly, Eqs. 17, 18, 20 and 21 can be applied to calculate
heat-transfer in the condenser and pre-heater. In the condenser,
for the working-fluid side around the horizontal tubes, the correla-
tion of the average heat-transfer coefficient for the film condensa-
tion is applied[36]:
ho 0:729gqfqf qgk
3
ri0fg
lfTsatTwDo
!1=423
where qfand qgare the liquid and vapor densities of the working
fluid, respectively;Tsatrepresents the condensation temperature in
thecondenser, andi0fgis themodified latent heat of theworking fluid.
In the pre-heater, the liquid working fluid is released from the
pump outlet by high-pressure flows in the tubes, and the vapor
working fluid is released from the turbine outlet by low-pressure
flows on the shell side. The heat-transfer calculation of pre-heater
can be obtained by applying the process as mentioned previously.
Therefore, the total heat-exchanger area in the ORC system can be
obtained by
Atot Aeva;1Aeva;2Aeva;3Acon;1Acon;2Apre 24
The total cost of heat exchangers contributes largely to the total
ORC system cost in low-temperature heat source power plant and
is assumed to be representative of complete system cost [2,19,37].
Finally, the objective parameter that represents the ratio of the net
power output Wnetto total heat-transfer area Atotin the ORC sys-
tem is defined as[38]
c Wnet=Atot 25
In this study, the ORC simulation is performed using a calcula-
tion program written in FORTRAN. The thermodynamic and trans-
port properties of the working fluids are obtained from the
National Institute of Standards and Technology (NIST) database
REFPROP 9.0[39]. The simulation procedure used by the program
is presented inFig. 4.
4. Results and discussion
4.1. Verification
To evaluate the accuracy of the thermodynamic simulation
results for the ORC system, the numerical solution of the evapora-
tion and condensation pressures, power output of the turbine,
thermal efficiency, and exergy destruction are verified using
R245fa at Teva= 106.85 C, Teva= 31.3 C, and a net power output
Wnet fixed at 10 kW [9], as shown in Table 3. The comparison
results for the ORC thermal efficiency, gth,verare evaluated exclud-ing the power consumption of the water pumps in the ORC system.
In addition, the calculated data on the R600a used in this study are
compared with the previously published results of an ORC system
that was evaluated [14] at Teva= 87.15 C and Teva= 25 C with
mr = 3.61 kg/s, as shown in Table 4. In this comparison, the thermo-
dynamic parameters of the working fluid are analyzed in the ORCFig. 4. Flow chart of the calculation procedures for the ORC system.
Table 3
Comparison of present calculated results with those of Ref. [9].
Parameter unit Teva (C) Tcon (C) Wnet(kW) Peva (kPa) Pcon (kPa) mr(kg/s) gth,ver(%) Wtur(kW) ED (kW)
R245fa[9] 106.85 31.3 10 1492.3 187.4 0.4988 8.4 10.615 45.08
R245fa 106.85 31.3 10 1482.1 186.23 0.5 8.36 10.561 44.9
D(%) 0.69 0.6 0.24 0.47 0.51 0.4
Note: D represents absolute error.
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system excluding the pre-heater. Furthermore, numerical calcu-
lated solutions of this study are validated with the experimental
results of Declaye et al. [40] for a ORC system with R245fa at
Teva= 85.3 C and Tsup= 10 C, as shown in Fig. 5. The corresponding
evaporation pressure, Peva, is maintained at 900 kPa and the con-
densation pressure,Pcon, varies with various condensation temper-
atures which results in the pressure-ratio variation from 2.6 to 5.8.
It can be observed that there is a slight deviation occurred between
the numerical results and experimental data at lower pressure
ratios, Peva/Pcon. This may be resulted from the constant pump
efficiency assumed in the simulation. As a whole, the numerical
solutions obtained in this study are consistent with those reported
in Wang et al.[9]and Dai et al.[14], and Declaye et al. [40], as canbe clearly seen inTables 3 and 4andFig. 5.
4.2. Problem description
The heat source of the ORC system used in this study is the
waste heat of the cylinder jacket water released from the cooling
water system installed in a large marine internal combustion
Table 4
Comparison of present calculated results with those of Ref. [14].
Parameter unit Teva (C) Tcon (C) mr(kg/s) Peva (kPa) Pcon (kPa) Qev a (kW) gth,ver(%) Wtur(kW) ED(kW)
R600a[14] 87.15 25 3.61 1550 350 1456.85 11.52 180.91 224.13
R600a 87.15 25 3.61 1552.5 350.7 1449.8 11.63 180.27 225.41
D(%) 0.16 0.2 0.48 0.95 0.35 0.57
Note: D represents absolute error.
(a) (b)
(c) (d)
58 60 62 64 66 683.6
4
4.4
4.8
5.2
R1233zdR1234yfR1234zeR245caR245faR600a
Tcon
= 40o
CT
sup= 5
oC
Teva
(oC)
(o/o)
th
58 60 62 64 66 68280
320
360
400
440
480
R1233zdR1234yfR1234zeR245caR245faR600a
Atot
(m2
)
Tcon
= 40o
CT
sup= 5
oC
Teva
(oC)
58 60 62 64 66 680.5
0.55
0.6
0.65
0.7
0.75
R1233zdR1234yfR1234zeR245caR245faR600a
Tcon
= 40oC
Tsup
= 5o
C
Teva
(oC)
(kW/m2)
Fig. 6. The effect ofTeva on (a) g th, (b)Atot, (c) c , and (d)ED and g IIin the ORC system.
2 3 4 5 6 70
2
4
6
8
10
Peva
= 900 kPa
Teva
= 85.3o
C
Tsup
= 10o
C
Declaye et al. [40]
This study
Peva
/ Pcon
th
(o/
o)
R245fa
Fig. 5. Validation of the proposed thermal efficiencies with those from experimental
work[40]of the ORC system with R245fa.
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R245fa achieves excellent thermal efficiency at Teva= 6168C. As
shown in Fig. 6(b), the total heat-transfer area, Atot, gradually
increases initially and then rises steeply as the evaporation tem-
perature increases, because the temperature difference between
the cylinder jacket water and working fluid decreases in the evap-orator. In addition,Fig. 6(b) shows that the Atotcurve of R1234zd
exhibits the highest value but the smallest value for R600a among
those of the working fluids.
Fig. 6(c) shows the effects ofTeva on the objective parameter c,
which is the ratio of net power output to the total heat-transfer
area of the ORC system. For all of the working fluids, the valuesofc increase initially, attain a maximal value, and finally decrease
(a) (b)
(c) (d)
(e) (f)
0.4
0.45
0.5
0.55
0.6
58
60
62
64
66
68Teva
(oC)
35
37
39
41
43
45
Tcon( oC)
0.640
0.629
0.618
0.608
0.597
0.586
0.575
0.565
0.554
0.553
0.552
0.543
0.532
0.5220.511
0.500
(kW/m2)
R1233zd
0.4
0.45
0.5
0.55
0.6
58
60
62
64
66
68Teva
(oC)
35
37
39
41
43
45
Tcon( oC)
0.644
0.641
0.630
0.620
0.610
0.600
0.590
0.580
0.570
0.560
0.550
0.540
0.530
0.5200.510
0.500
(kW/m2)
R1234yf
0.4
0.45
0.5
0.55
58
60
62
64
66
68
Teva(oC)
35
37
39
41
43
45
Tcon( oC)
0.646
0.640
0.629
0.618
0.608
0.597
0.586
0.575
0.565
0.554
0.543
0.532
0.522
0.511
0.500
(kW/m2)
R1234ze
0.4
0.45
0.5
0.55
0.6
58
60
62
64
66
68
Teva(oC)
35
37
39
41
43
45
Tcon( oC)
0.640
0.629
0.618
0.608
0.597
0.586
0.582
0.580
0.575
0.565
0.554
0.543
0.532
0.522
0.511
0.500
(kW/m2)
R245ca
0.4
0.45
0.5
0.55
0.6
58
60
62
64
66
68Teva
(oC)
35
37
39
41
43
45
Tcon( oC)
0.640
0.629
0.618
0.608
0.601
0.597
0.586
0.575
0.565
0.5540.543
0.532
0.522
0.511
0.500
(kW/m2)
R245fa
0.4
0.45
0.5
0.55
0.6
58
60
62
64
66
68Teva
(oC)
3537
39
41
43
45
Tcon( oC)
0.677
0.670
0.656
0.640
0.629
0.618
0.608
0.597
0.586
0.575
0.565
0.554
0.543
0.532
0.522
0.511
0.500 (kW/m2)
R600a
Fig. 8. Contours ofc for (a) R1233zd, (b) R1234yf, (c) R1234ze, (d) R245ca, (e) R245fa, and (f) R600a.
1006 M.-H. Yang, R.-H. Yeh / Energy Conversion and Management 88 (2014) 9991010
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as the evaporation temperature increases. An optimal c, which rep-
resents the maximal net power output per unit area of the heat
exchangers, can be obtained for each working fluid based on the
results. In addition, cmax=0.73 kW/m2 is obtained at the corre-
sponding optimal evaporation temperature Teva,o=64.2 C for
R600a. The c values of R1234yf are higher than those of R1234ze
atTeva= 5860.5 C; however, R1234ze yields higherc values than
R1234yf atTeva=
60.568 C. Thecmax
of R1234yf and R1234ze are
0.7 and 0.69 atTeva,o= 63.5 C and 63 C, respectively. As shown in
Fig. 6(d), the exergy destruction decreases with evaporation tem-
perature whereas the second law efficiency increases as the evap-
oration temperature increases when Tcon= 40 C and Tsup = 5 C.
Among these working fluids, R1234yf exhibits the lowest exergy
destruction and highest second law efficiency at Teva= 5867 C.
Generally, working fluids having high thermal efficiency exhibit
low exergy destruction in the ORC system.
By contrast, high condensation temperatures decrease the ther-
mal efficiency because the power output of the turbine decreases,
as shown in Fig. 7(a). The figure also indicates that R245ca and
R245fa exhibit high thermal efficiency at Tcon=3543 C and
Teva=65 C. Similarly, at Tcon= 3543 C, the ORC system using
R1234yf obtains maximal thermal efficiency. As shown in
Fig. 7(b), theAtotcurves tend to declineas the condensation temper-
ature increases in the ORC system. This is because as the condensa-
tion temperature increases, the temperature difference between
the cooling water and working fluid in the condenser increases,
causing a decrease in the total heat-transfer area. Based on
Figs. 6(b) and7(b), R600a, R1234ze, and R1234yf exhibit superior
transport properties in the heat exchange process in the ORC sys-
tem. Fig. 7(c) shows the influence ofTconon cat Teva=65 C for each
working fluid. As expected, the values ofc increase initially, then
approach the peak points, and finally decrease as the condensation
temperature increases. The maximal value, cmax=0.73, occurs at
Tcon,o= 41 C for R600a. In the objective parameter evaluation,
R600a evidently performs more satisfactorily compared with the
other working fluids tested. Although R245ca and R245fa demon-
strate excellent performance in thermal efficiency at high evapora-
tion and low condensation temperatures, low transport properties
yield inferior values in the objective parameter estimation. Increas-
ing the condensation temperature causes the exergy destruction of
the system to increase and the second efficiency to reduce, respec-
tively, as can be observed in Fig. 7(d). This figure also indicates that
R1234ze and R1233zd demonstrate unfavorable performance in
exergy destruction and the second efficiency at most condensation
Table 5
The cmax and its corresponding DThw,o, DTcw,o, Teva,o,Tcon,o, g th,o, EDo, and g II,o for the
ORC system atTcwi = 25 C and Thwi= 85 C.
Item R1233zd R1234yf R1234ze R245ca R245fa R600a
cmax (kW/m2) 0.58 0.66 0.67 0.62 0.64 0.71
DThw,o (C) 7.4 8.9 8.4 7 7.5 7.4
DTcw,o (C) 5.1 6 5.5 4.6 4.9 5
Teva,o (C) 64.2 63.2 63.7 63.9 63.9 64.6
Tcon,o (C) 38.1 40.2 39.3 37.6 38.1 39.4
gth,o (%) 4.46 4.08 4.2 4.56 4.48 4.38
EDo (kW) 413.42 519.97 484.55 387.3 418.44 418.67
gII,o (%) 30.71 30.85 30.56 31.05 30.79 30.97
(a) (b)
(c) (d)
85 87 89 91 93 9563
65
67
69
71
R1233zdR1234yfR1234zeR245caR245faR600a
Teva,o(
oC
)
Tcwi
= 25oC
Thwi
(oC)
85 87 89 91 93 9536
38
40
42
44
R1233zdR1234yfR1234zeR245caR245faR600a
Tcon,o
(oC
)
Tcwi
= 25oC
Thwi
(oC)
85 87 89 91 93 950.5
0.6
0.7
0.8
0.9
1
R1233zdR1234yfR1234zeR245caR245faR600a
Tcwi= 25o
C
Thwi
(oC)
max
(kW/m2)
85 87 89 91 93 953.6
4
4.4
4.8
5.2
R1233zdR1234yfR1234zeR245caR245faR600a
Tcwi
= 25oC
Thwi
(oC)
th,o
(o/o
)
Fig. 9. The influence ofThwi on (a)Teva,o (b)Tcon,o (c) cmax, and (d) gth,o at Tcwi= 25 C.
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temperatures. According to Figs. 6(d) and 7(d), R1234yf
demonstrates satisfactory performance in exergy destruction at
low evaporation and high condensation temperatures.
4.4. Optimization
To show the optimal operating temperatures of the ORC system,
the distributionsofc for various evaporationand condensationtem-
perature ranges at DThw= 8 C and DTcw= 5 C are plotted in
Fig.8(af) for each working fluid. As expected, the optimal evapora-
tion temperatures, Teva,o, and optimal condensation temperatures,
Tcon,o, canbe observed for the maximal ratioofWnettoAtot. Moreover,
these contour plots ofc show thevariations in theoptimal operating
temperatures among the working fluids. The cmax of R600a is thehighest among the six working fluids with a corresponding Teva,oof 64.3 C, and Tcon,o of 41.1 C, followed by R1234ze, R1234yf,
R245fa, and R245ca. Clearly, R1233zd exhibits the lowest maximal
objective parameter, cmax = 0.53 atTeva,o= 63.9 C andTcon,o= 40 C,
among the working fluids.
Furthermore, the maximal objective parameter, cmax, and its
corresponding optimal operating conditions, DThw,o, DTcw,o, Teva,o,
Tcon,o, gth,o, EDo, and gII for the ORC system at Tcwi = 25 C,
Thwi= 85 C, and mhw= 128 kg/s are obtained numerically and are
shown inTable 5. Under the condition in which the cylinder jacket
water and cooling water inlet temperatures are maintained con-
stant, the optimal temperature differences between the inlet and
outlet of the cylinder jacket water and cooling water, DThw,o and
DTcw,o, are obtained according to the maximal objective parameterfor various working fluids. The DThw,o andDTcw,o of R1234yf are
higher than those of the other working fluids. A high temperature
difference between the inlet and outlet of the cylinder jacket water
indicates a large amount of heat energy added in the ORC system.
Similarly, a high temperature difference between the inlet and out-
let of the cooling water suggests that an additional cooling load is
required in the condenser of the system. Conversely, the lowest
values of DThw,o, and DTcw,o are obtained for R245ca. Moreover,
the corresponding optimal evaporation, condensation, thermal
efficiency, and exergy destruction are determined to compare the
working fluids. The thermodynamic properties and transport prop-
erties affect the results for the maximal objective parameter and
optimal operating temperatures. Among the working fluids,
R600a exhibits the highest objective parameter value of 0.71 kW/
m2. The sequence ofcmax
for each working fluid is listed, as men-
tioned previously. According to Table 5, the corresponding Teva,oandTcon,o of cmax vary among the working fluids. Also, note that
R600a and R245ca exhibit the highest Teva,o and lowest Tcon,o,
respectively. The values ofgth,o andEDo, which respectively repre-
sent the thermal efficiency and exergy destruction, are calculated
according to the conditions of the cmax in the ORC system. Under
the optimal conditions, R1234yf exhibits the lowest Teva,oand high-
estTcon,o, resulting in inferior performance in thermodynamic effi-
ciency and exergy destruction. In addition, R245ca exhibits the
lowest exergy destruction under the optimal conditions.
4.5. Effects of cylinder jacket water and cooling water temperatures
Increasing the heat source temperature enhances the perfor-mance of the ORC system. In this study, the cylinder jacket water
(a) (b)
(c) (d)
20 22 24 26 28 3060
62
64
66
68
R1233zdR1234yfR1234zeR245caR245faR600a
Teva,o
(oC)
Thwi
= 85oC
Tcwi
(o C)
20 22 24 26 28 3032
34
36
38
40
42
44
R1233zdR1234yfR1234zeR245caR245faR600a
Tcon,o
(oC
)
Thwi
= 85oC
Tcwi
(oC)
20 22 24 26 28 300.5
0.6
0.7
0.8
0.9
1
R1233zdR1234yfR1234zeR245ca
R245faR600a
Thwi
= 85oC
Tcwi
(oC)
max
(kW/m2)
20 22 24 26 28 303.6
4
4.4
4.8
5.2
R1233zdR1234yfR1234zeR245caR245faR600a
Thwi
= 85oC
Tcwi
(oC)
(o/o
)
th,o
Fig. 10. The influence ofTcwi on (a)Tcon,o (b)Teva,o (c) cmax, and (d) g th,o at Thwi = 85 C.
1008 M.-H. Yang, R.-H. Yeh / Energy Conversion and Management 88 (2014) 9991010
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temperature is considered at Thwi=8595 C. To show the influ-
ence of cylinder jacket water temperature on the operating tem-
peratures and cmax, the variations in the optimal evaporation and
condensation temperature at Tcwi=25 C are presented in
Fig. 9(a) and (b), respectively. Both the optimal evaporation and
condensation temperatures increase as Thwi increases in the ORC
system. Furthermore, the increment ofTeva,ois more apparent than
that ofTcon,o
for each working fluid, suggesting that the influence of
Thwi onTeva,ois more substantial than that on Tcon,oin the ORC sys-
tem. As previously mentioned, R1234yf exhibits the lowest Teva,oand highestTcon,oat Thwi=8595 C. Notably, the optimal evapora-
tion temperature of R600a under all of the conditions is the highest
than those of the other working fluids. Therefore, R600a is suitable
for use at a high evaporation temperature in the ORC system. In
addition, it is observed that the Teva,o and Tcon,o of R245ca and
R245fa are similar.Fig. 9(c) shows the influence ofThwi on thecmaxin the ORC system. Overall, the maximal objective parameters
increase as the cylinder jacket water temperature increases
because of an increase in the net power output. Remarkably, the
increase in the cmax of R1234yf is mitigated as Thwi increases.
Therefore, R1234yf is suitable for use with a low-temperature heat
source. The variations in optimal thermal efficiency, which are
obtained under the conditions corresponding to cmax, at various
Thwi are shown inFig. 9(d). The gth,o values from highest to lowest
are R245ca, R245fa, R1233zd, R600a, R1234ze, and R1234yf.
The variation of the optimal evaporation and condensation tem-
peratures in relation to the cooling water inlet temperature Tcwi at
Thwi=85 C are shown in Fig. 10(a) and (b). Similarly, the Teva,oandTcon,oincrease as the cooling water inlet temperature increases
for each working fluid, and the influence ofTcwi on Tcon,ois stronger
than that on Teva,o in the ORC system. According toFig. 10(a), the
Teva,o values of R1233zd, R245ca, and R245fa tend to be similar at
Tcwi= 2630 C. In addition, Fig. 10(b) indicates that the Teva,o values
of R600a and R1234ze and those of R1233zdand R245fa are similar
at various Tcwi. The optimal objective parameters decrease as Tcwidecreases for each working fluid, as shown inFig. 10(c). Based on
Fig. 9(c) and (d) and Fig. 10(c) and (d), although R600a exhibitsthe most favorable results in the objective parameter analysis,
R245ca exhibits the highest net thermal efficiency under optimal
conditions among the working fluids. The results given in Fig. 9(c)
andFig. 10(c) show the sequence ofcmax, as mentioned previously.
Based on these results, R1233zd performs unfavorably in the object
parameter analysis, and R1234yf exhibits the lowest net thermal
efficiency under optimal conditions for the ORC system.
5. Conclusions
In this study, the thermodynamic and transport properties of
the ORC working fluids used to recover waste heat from a large
marine diesel engine are optimally simulated to increase the EEDI
and reduce greenhouse gas emissions from merchant ships. Anobjective parameter c, which represents the ratio of net power out-
put to total heat-exchanger area, is determined to analyze the per-
formance of the ORC system in recovering waste heat. The optimal
operating temperatures of the ORC system, Tcon,o and Teva,o, are
obtained numerically to achieve the maximal objective parameter
cmax atTcwi=2030 C and Thwi=8595 C for R1233zd, R1234yf,
R1234ze, R245ca, R245fa, and R600a. The results, which are
obtained numerically, support the following conclusions:
1. In the evaluation of the maximal objective parameter for recov-
ering waste heat from the diesel engine containing a cooling
water system, R600a performs the most satisfactorily, followed
by R1234ze, R1234yf, R245fa, and R245ca, and R1233zd
performs the least satisfactorily at Teva=5868 C andTcon= 3545 C.
2. The working fluid demonstrating superior thermodynamic
properties does not necessarily demonstrate excellent perfor-
mance in the heat-transfer process. Although R245ca, R245fa,
and R1234yf exhibit higher thermal efficiency among the work-
ing fluids according to thermodynamic analysis, outstanding
performance in the evaluation of objective parameters for the
ORC system is not guaranteed.
3. In the ORC system, the cylinder jacket water temperature
affects the optimal evaporation temperature more strongly than
it affects the optimal condensation temperature. By contrast,
the cooling water temperature affects the optimal condensation
temperature more substantially than it affects the optimal
evaporation temperature.
Acknowledgements
The financial support for this research from the Engineering
Division of National Science Council, Republic of China, through
contract NSC 101-2221-E-022-004, is greatly appreciated.
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