ORC WHR from marine engine.pdf

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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: [email protected](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
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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

1000 M.-H. Yang, R.-H. Yeh / Energy Conversion and Management 88 (2014) 9991010
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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.

M.-H. Yang, R.-H. Yeh / Energy Conversion and Management 88 (2014) 9991010 1001
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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


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


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.



9/12

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


Teva,o(

oC

)

Tcwi

= 25oC

Thwi

(oC)

85 87 89 91 93 9536

38

40

42

44


Tcon,o

(oC

)

Tcwi

= 25oC

Thwi

(oC)

85 87 89 91 93 950.5

0.6

0.7

0.8

0.9

1


Tcwi= 25o

C

Thwi

(oC)

max

(kW/m2)

85 87 89 91 93 953.6

4

4.4

4.8

5.2


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.

http://-/?-http://-/?-


10/12

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


Teva,o

(oC)

Thwi

= 85oC

Tcwi

(o C)

20 22 24 26 28 3032

34

36

38

40

42

44


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


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.



11/12

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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