Comparison of the performance of two different Dual-loop organic Rankine cycles (DORC) with nanofluid for engine waste heat recovery

Energy Conversion and Management 126 (2016) 99–109

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Comparison of the performance of two different Dual-loop organic Rankine cycles (DORC) with nanofluid for engine waste heat recovery Haozhong Huang ⇑, Juan Zhu, Bo Yan College of Mechanical Engineering, Guangxi University, Nanning 530004, China

a r t i c l e

i n f o

Article history: Received 29 May 2016 Received in revised form 12 July 2016 Accepted 31 July 2016

Keywords: Dual-loop organic Rankine cycle Engine Waste heat recovery Nanoparticle Nanofluid

a b s t r a c t To recover the heat from engine exhaust, coolant liquid and high-temperature loop, two different Dualloop organic Rankine cycles (DORC) are studied in this paper. The two systems differ for the number of stages of heat recovery from engine exhaust, and both include high temperature loop and low temperature loop in each system. R123, R245fa, ethanol, R141b, and water are the candidate working fluids of HT loop, and R143a is the working fluid of LT loop. Because the coolant water in engines has lower temperature, it is more difficult to recover its heat. Therefore, in this study, graphene nanoparticles and carbon nanotubes are added to coolant water to enhance its heat transfer. Net output power, thermal efficiency, and exergy efficiency are selected as the objective functions. Results show that the single stage system (S1) is a little better than the other. Water-based S1 performs the best and the net output power, the thermal efficiency, and the exergy efficiency are 96.92 kW, 14.13% and 64.04%, respectively. High evaporation pressure and turbine inlet temperature are better for performance optimization. And when the coolant water contains 0.5 wt% carbon nanotubes, system reaches the max net output power increment of 3.84 kW. Ó 2016 Elsevier Ltd. All rights reserved.

1. Introduction With the energy crisis and the increase of fossil fuel global consumption, how to enhance fuel efficiency is a way to contain the problem of high fuel consumption. However, the maximum efficiency of a traditional engine does not exceed 42% [1]. Engine exhaust and coolant water mostly take away the heat of combustion. Therefore, how to recover heat from engine exhaust and coolant to enhance fuel efficiency has become a hot research topic. Methods for recovering waste heat mainly include thermoelectric power generator [2–5], turbocharger [6,7], absorption/adsorption refrigeration [8,9] and Rankine cycle [10–12]. Among them, Rankine cycle is nowadays the most studied topic because its best performance in heat recovers. Noboru et al. [13] studied a simple Rankine cycle with and without condenser; in the case with condenser thermal efficiency improved from 2.9% to 3.7%. Iacopo et al. [14] described traditional single-loop Rankine cycle recover exhaust heat, or exhaust heat, and coolant. The results showed that the thermal efficiency rose, but the recovery of coolant water was very low. Gequn Shu et al. [15,16] added a loop based on traditional Rankine cycle to form the DORC, and analyzed different ⇑ Corresponding author at: College of Mechanical Engineering, Guangxi University, Daxuedong Road 100, Xixiangtang District, Nanning 530004, China. E-mail address: [email protected] (H. Huang). http://dx.doi.org/10.1016/j.enconman.2016.07.081 0196-8904/Ó 2016 Elsevier Ltd. All rights reserved.

working fluids and different thermal states. Then, add regenerator based on DORC, three different regenerative systems were compared with traditional DORC: results demonstrated that the use of regenerator was good for organic working fluids. Working fluids also play an important role in ORC systems of waste heat recovery. They can be classified as dry working fluids, wet working fluids, and isentropic working fluids, or normal working fluids and organic working fluids, or single working fluids and mixed working fluids, respectively. Xinming Xi et al. [17] used zeotropic mixtures to deal with constant phase change temperature of pure fluid. Results showed that the use of zeotropic mixtures could observably increase the work output. Tian and Shu [18,19] described the performances of 20 different working fluids in the Rankine cycle. Results showed that R123, R141b and R245fa had the highest thermal efficiency (gth, within 16.60–13.30%), and their respective work output (Wnet) values ranged from 60 to 49 kJ/kg. And then in order to match the temperature difference of waste heat and working fluid, alkanes were used as working fluids and showed good performance in thermal efficiency, power output performances, and exergy destruction rate, while requiring small turbine sizes. The above mentioned literature highlights that Rankine cycle’s configuration and working fluids, simply or dualloop, with or without regenerator, and traditional working fluid, zeotropic mixtures or alkanes.

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Nomenclature s h T c P E _ m k R b d Q W

specific entropy (kJ/kg K) specific enthalpy (kJ/kg) temperature (K) specific heat capacity (J/(kg K)) pressure (MPa) exergy (kW) mass flow rate (kg/s) heat transfer coefficient (W/(m2 K)) fouling resistance (m2 K/W) tube wall thickness (m) tube diameter (m) heat injected or rejected (kW) work consumed or output (kW)

Greek letters v mass fraction g efficiency a convective heat transfer coefficient (W/(m2 K)) Da the increment of convective heat transfer coefficient k heat conductivity coefficient (W/(m K))

out evp exh p T f e 0 HT LT EHT ELT,1 ELT,2 ELT,3 THT TLT CHT CLT PHT PLT DORC

output by the system evaporation exhaust pump turbine working fluid exergy ambient state high temperature low temperature evaporator in the HT cycle first evaporator in the LT cycle second evaporator in the LT cycle third evaporator in the LT cycle turbine in the HT cycle turbine in the LT cycle condenser in the LT cycle condenser in the LT cycle pump in the HT cycle pump in the LT cycle dual-loop organic Rankine cycle

Subscripts a all in input the system

However, much waste heat and coolant water have low temperature. Among them, coolant water has the lower temperature, usually below 100 °C, and it is more difficult to recover. Some papers assume the full recovery of coolant water, thus the heat transfer area of coolant water and working fluid is huge, and exergy efficiency declines. In recent years, nanoparticles became an attractive research field. By adding nanoparticles to water, because of attractive van der Waals forces, nanoparticles tend to form aggregates in nanofluids, generating the so-called aggregate-laden nanofluids. The aggregation of them affects physical nanofluid properties such as thermal viscosity and conductivity, and then affects convective heat transfer [20]. According to many studies Al2 O3 , TiO2 and Fe3 O4 nanoparticles were added to water, and the results showed that the nanofluid heat transfer coefficients were better than that of water [21–25]. Graphene nanoparticle is a two-dimensional crystal material and also has very high thermal conductivity, so the nanofluid of it is expected to show a significant convective enhancement. By adding graphene nanoparticles to water, in a circular tube, it gain 14.2% enhancement at most [26] and in a horizontal stainless steel tube, it gain 200% enhancement at most [27]. Carbon nanotube is a kind of one-dimensional quantum materials with special structure. It has very high thermal conductivity and hence great potential for significant convective heat transfer enhancement. By the addition of carbon nanotubes in a double tube heat exchanger, performance enhanced of 78% at most [28], while in a horizontal tube the enhancement was of 350% [29]. From the above literature, it is known that nanoparticles strongly influence heat transfer. Although nanoparticle’s enhancement of heat transfer is very popular, nanoparticles are barely used in Rankine cycle. This study investigates the effect of addition of nanoparticles to coolant water to enhance its heat transfer capability. In this paper, to recover engine waste heat, two different DORC with nanoparticles are proposed. Several working fluids are used in the high temperature loop and results in terms of net output power, thermal efficiency, and exergy efficiency are compared. Moreover, different nanoparticles are added to coolant water to

enhance heat transfer between coolant water and working fluid in low temperature loop, and the influences of the addition of nanoparticles are highlighted by comparing the two systems’ net output powers, thermal efficiencies, and exergy efficiencies. 2. Methodology 2.1. Topping engine A 6-cylinder 4-stroke marine engine was chosen and engine’s main parameters are listed in Table 1. The waste heat of engine running at 700 r/min and 1000 r/min was recovered. Some parameters of exhaust, including specific heat (cp), enthalpy (h) and entropy (s), at given temperature (T) and given pressure (P), could be calculated using methods for exhaust [30]. For example, s could be calculated by:

sexh ðT; PÞ ¼ vCO2 sCO2 ðT; PÞ þ vH2 O sH2 O ðT; PÞ þ vO2 sO2 ðT; PÞ þ vN2 sN2 ðT; PÞ

ð1Þ

2.2. Bottoming DRC system This study discusses two dual-loop Rankine cycles with different exhaust recovery schemes, which are denoted as S1 [31] and S2 [15] and illustrated in Fig. 1a and b, respectively. Either of them includes a high temperature loop and a low temperature loop. S1 and S2 share the same high temperature loop configuration that consists of an evaporator, a turbine, a condenser, and a pump.

Table 1 Main engine parameters. Rotation speed (r/min)

Energy power output (kW)

Exhaust temperature (°C)

Exhaust mass flow rate (kg/h)

Engine coolant temperature (°C)

1000 700

760 670

395 430

4594.38 3380.00

73.3 72.4

H. Huang et al. / Energy Conversion and Management 126 (2016) 99–109

101

in the evaporator ELT,3 doing work in turbine TLT; finally, it is liquefied in the condenser CLT and enters the pump PLT for a new round of cycle. The thermodynamic process in low temperature loop of S1 is ⑨-⑤-⑥-⑦-⑧-⑨, and the T-s diagram is shown in Fig. 2d. The process is the same of S2, except for that the working fluid enters the turbine TLT and does work right after being discharged from the evaporator ELT,2. 2.3. Working fluids High temperature loop requires the working fluid has good stability and good capability of withstanding high temperature. The Rankine cycle used R123, R245fa, ethanol, R141b, and water [32,33] as working fluids in high temperature loop and low boiling point organic working fluid R143a in low temperature loop [15]. The nanoparticles added into the engine coolant were graphene nanofluid with mass fractions of 0.025 wt%, 0.05 wt%, 0.075 wt% and 0.1 wt%, respectively, and carbon nanotube with the mass fraction of 0.5 wt% [26,28]. 2.4. Modeling

(a) S1 Before modeling, it should be assumed first that: (1) All the components are stable; (2) the heat and pressure losses from pipes and other parts are ignored; (3) kinetic and potential energies are neglected; (4) isentropic turbine efficiency is 0.75, and isentropic pump efficiency is 0.8; (5) the convective heat transfer coefficient increments of coolant (Da) corresponding to the additions of the graphene nanofluid (with the mass fractions of 0.025%, 0.05%, 0.075% and 0.1%) and the carbon nanotube with the mass fraction of 0.5% are 31%, 63%, 130%, 200% and 350%, respectively [26,28]. Exergy at point i is defined by [34–39]:

_ i  h0 Þ  T 0 ðsi  s0 Þ Ei ¼ m½ðh

(b) S2

ð2Þ

The input useful work of the system consists of the input work from exhaust and engine coolant as well as the consumed work of two pumps, while the output work of the system equals to the output work of two turbines:

Ein ¼ ðEA  EC Þ þ ðED  EE Þ þ W p

ð3Þ

Eout ¼ W T

ð4Þ

Fig. 1. Configuration diagram of DORC system.

Regarding low temperature loop, S1 has two heat exchangers, a turbine, a condenser, and a pump, while S2 has an additional heat exchanger. Thermodynamic process is the same for both high temperature loops in S1 and S2 configurations: ④-①-②-③-④, and the corresponding T-s diagrams are shown in Fig. 2a and c. Working fluid flows through the pump PHT generating high-pressure fluid, which then flows into evaporator EHT. Subsequently, exhausts heat the fluid turning it into vapor. In turbine THT, it expands doing work, then turns into low-pressure gas that enters the condenser ELT,2 (the evaporator in the low temperature loop) and it is cooled into liquid. Finally, it enters the pump PHT again and a new round of cycle starts. The thermodynamic process in low temperature loop of S2 is ⑩-⑤-⑥-⑦-⑧-⑨-⑩, and the corresponding T-s diagram is shown in Fig. 2b. The high-pressure fluid, produced from working fluid in the pump PLT, is pre-heated by the engine coolant in the evaporator ELT,1 and gasified in evaporator ELT,2. Then it absorbs the waste heat of exhaust flowing through high temperature loop

Then the detailed energy balance [38] is modeled. The system exergy efficiency is:

ge ¼

Eout Ein

ð5Þ

Pump PHT:

_ f ;HT ðh1  h4 Þ W P;HT ¼ m

ð6Þ

Evaporator EHT:

_ f ;HT ðh2  h1 Þ ¼ m _ g cpg;HT ðT A  T B Þ Q E;HT ¼ m

ð7Þ

Turbine THT:

_ f ;HT ðh2  h3 Þ W T;HT ¼ m

ð8Þ

Condenser CHT or Evaporator ELT,2:

_ f ;HT ðh3  h4 Þ ¼ m _ f ;LT ðh7  h6 Þ Q E;LT;2 ¼ m

ð9Þ

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H. Huang et al. / Energy Conversion and Management 126 (2016) 99–109

800

800 A

WATER

700

ut

s ha

A

700

2

Ex

t

su

a xh

2

E

600

600

T (K)

T (K)

WATER

500 Water

3

B 1

400 4

300

Water

3

B 1

400

500

4

7

300

6

200

7

6

200 0

1

2

3

4

5

6

7

8

9

0

1

2

3

4

5

6

7

s (kJ/kg·K)

s (kJ/kg·K)

(a) the HT loop in S1

(b) the LT loop in S1

480

R143a

400

B

400

C 7

T (K)

7

T (K)

9

480

R143a

D

E

6

320

8

5

8

D

E

320

6

5

9

9

10

240

240

0.50

8

0.75

1.00

1.25

1.50

0.50

1.75

0.75

1.00

1.25

1.50

s (kJ/kg·K)

s (kJ/kg·K)

(c) the HT loop in S2

(d) the LT loop in S2

1.75

Fig. 2. T-s diagram.

Pump PLT:

W P;LT

Evaporator ELT,3:

_ f ;LT ðh5  h10 Þ ¼m

ð10Þ

Evaporator ELT,1: Heat transfer of heat exchanger can be calculated by heat transfer coefficient, heat transfer area and temperature difference.

_ f ;LT ðh6  h5 Þ ¼ m _ c cpc ðT D  EE Þ ¼ K E;LT;1 AE;LT;1 DtE;LT;1 Q E;LT;1 ¼ m K E;LT;1 ¼

1 d0 d0 b d0 1 þ Rsi þ þ Rso þ ai di di k dm a0

ð11Þ ð12Þ

_ f ;LT ðh8  h7 Þ ¼ m _ g cpg;LT ðT B  T C Þ Q E;LT;3 ¼ m

ð13Þ

Turbine TLT:

_ f ;LT ðh8  h9 Þ W T;LT ¼ m

ð14Þ

Condenser CLT:

_ f ;LT ðh9  h10 Þ ¼ m _ w cpw ðT G  T F Þ Q C;LT ¼ m

ð15Þ

The total net output power is:

W net ¼ W T;HT þ W T;LT  W P;HT  W P;LT

ð16Þ

Table 2 Comparison of the calculated results with Ref.[40] Parameter unit

PORC (kW)

gORC

Pcond (kPa)

Pvap (kPa)

Tvap (K)

_ f (kg/s) m

Dh3–4 (kJ/kg)

Benzene Benzene [40] D (%) R11 R11 [40] D (%) R134a R134a [40] D (%)

346.9 349.3 0.69 281.6 290.3 3.00 144. 38 147.5 2.12

19. 97% 19. 86% 0.55 16. 47% 16. 58% 0.66 8. 27% 8. 52% 2.93

21 19.6 7.14 147.9 147.9 0 883.3 883.3 0

2000 2000 0 3835.9 3835.9 0 3723.4 3723.4 0

494.6 494.5 0.02 466 461 1.08 378 369.9 2.19

2.73 2.737 0.26 7.487 7.487 0 8.9667 8.9667 0

129.8 130.5 0.54 40.6 41.9 3.10 19.3 19.4 0.52

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The system thermal efficiency is gth:

ðW T;HT  W P;HT Þ þ ðW T;LT  W P;LT Þ Q E;HT þ Q E;LT;1 þ Q E;LT;2 þ Q E;LT;3

ð17Þ

3. Validation The present study employs classical thermodynamics formulae and validates them by the simple Rankine cycle and the same parameters as described in Ref. [40]. The results are shown in Table 2. The differences are primarily caused by the difference in the simulation software sets. REFPROP Software [41] was used in Ref. [40], while in the present study working fluid properties were calculated by EES (Equation Evaluation Solution) Software [42].

S1 R245fa S2 R245fa S1 R123 S2 R123

0.14

thermal efficiency ηth

gth ¼

0.16

0.12

0.10

S1 R141b S2 R141b S1 Ethanol S2 Ethanol S1 Water S2 Water

0.08

0.06

4. Results and discussion

1

Since many factors can greatly affect the whole system in HT loop of Rankine cycle, following Sections 4.1 and 4.2 will only discuss the effects of the parameters of HT loop. The engine was run at 700 r/min, and the variation of Wnet, gth, and ge of the two systems were compared. The maximum evaporation pressure of the HT loop (Pevp,HT) is 7 MPa, which is limited by the pressure limitation of EHT. An appropriate Pevp,LT should be selected to achieve the greatest Wnet for the system. The condensation temperature of HT loop is limited by the temperature difference between the working fluid in HT cycle and that in LT loop in ELT,2 (CHT). The condensation temperatures of R123, R245fa, ethanol, R141b, and water in HT loop are 353 K, 347 K, 349 K, 348 K, and 369 K, respectively. 4.1. Evaporation pressure of high temperature loop Fig. 3 shows the variation of the net output power with Pevp,HT. Wnet rises with Pevp,HT for both S1 and S2 because as Pevp,HT rises, _ f ;HT conthe enthalpy drops at the turbine increase rapidly, while m tinuously decreases, thereby leading to the first dramatic and then steady increase of WT,HT. Meanwhile, WP,HT rises, and therefore, as Pevp,HT rises, Wnet first increases rapidly and then grows steadily. Fig. 4 shows thermal efficiency variation with Pevp,HT. As Pevp,HT rises, the values of gth of both S1 and S2 rise. As Pevp,HT rises, Wnet _ f ;HT declines rises rapidly at first and then slows down, while m steadily and QC,HT decreases. Therefore, the overall thermal load

2

exergy efficiency ηe

net output power Wnet (kW)

0.60

S1 R245fa S2 R245fa S1 R123 S2 R123

50

7

0.55 0.50 0.45

S1 R141b S2 R141b S1 Ethanol S2 Ethanol S1 Water S2 Water

0.40 0.35 0.30

40 1

2

3

4

5

6

Pevp,HT (MPa) Fig. 3. Variation of net output power with Pevp,HT.

7

8

8

S1 R245fa S2 R245fa S1 R123 S2 R123

0.65

S1 R141b S2 R141b S1 Ethanol S2 Ethanol S1 Water S2 Water

6

decreases. As Wnet rises and the overall thermal load decreases, the system gth grows rapidly at first and then steadily rises. Fig. 5 shows the variation of the exergy efficiency with Pevp,HT. ge rises with Pevp,HT for both S1 and S2, because as Pevp,HT rises, the irreversible loss of EHT decreases significantly, and simultaneously, the irreversible loss of THT increases. Both two factors affect ge. In the variation range of Pevp,HT, the amount of reduction in irreversible loss of EHT is greater than that of increase in irreversible loss of THT, which thus can explain the increasing trend of ge with Pevp,HT. From Figs. 3–5, Wnet, gth, and ge increase with Pevp,HT, and the system performance is optimized. The value of S1 is always a little bit better than that of S2. In fact, the irreversible loss of EHT in S1 is greater than that in S2; however, compared with S1, S2 has an irreversible loss of ELT,3, and the irreversible loss of TLT of S2 is greater than that in S1. Therefore, the total irreversible loss in S1 is less than that in S2. Recovering the exhaust in HT loop, its temperature is high and the quality is better, so Wnet in S1 is slightly higher than that in S2. This is because greater temperature difference leads to greater heat exchange between exhaust and working fluid. Besides, _ f ;HT in S1 is greater than that of S2. Both variations of WT and WP m result in the difference of Wnet between the two systems. The dif_ f ;HT in S1 ference in gth is smaller than that in Wnet and ge. In fact, m

90

60

5

Fig. 4. Variation of thermal efficiency with Pevp,HT.

0.70

70

4

Pevp,HT(MPa)

100

80

3

1

2

3

4

5

6

Pevp,HT (MPa) Fig. 5. Variation of exergy efficiency with Pevp,HT.

7

8

104

H. Huang et al. / Energy Conversion and Management 126 (2016) 99–109

Wnet(kW)

S1 S2

75.17 73.96

68.47 67.65

93.92 92.25

81.77 80.34

96.92 95.01

is greater than that in S2, leading to greater heat exchange in ELT,2 and thus greater total heat exchange in S1. Meanwhile, Wnet in S1 is greater than that in S2. Owing to the variations of two parameters, there is a slight difference in gth between S1 and S2. Table 3 shows the optimal performance of S1 and S2.

gth

S1 S2

10.58% 10.54%

9.54% 9.52%

13.63% 13.55%

11.63% 11.57%

14.13% 14.02%

4.2. Turbine inlet temperature of high temperature loop (T2)

ge

S1 S2

52.32% 50.78%

48.49% 47.24%

62.53% 60.59%

55.95% 54.24%

64.04% 61.91%

Table 3 Optimal system performance. R123

R245fa

Ethanol

R141b

Water

By comparing the resulting T2 of the two systems, the influence of T2 on system thermodynamic performance can be studied. At

69.3

net output power Wnet (kW)

net output power Wnet (kW)

75.3 75.0 74.7 74.4

S1 S2

74.1 73.8 73.5 560

580

600

620

640

660

680

68.4

67.5

66.6

S1 S2

65.7

480

700

500

520

540

560

T2(K)

T2(K)

(a) R123

(b) R245fa

580

600

620

96

net output power W net (kW)

82.0 94

92

90

S1 S2

88

81.5

560

580

600

620

640

660

680

S1 S2

81.0 80.5 80.0 79.5

86

560

700

580

600

620

640

T2(K)

T2(K)

(d) R141b

(c) Ethanol 97

net output power Wnet(kW)

net output power Wnet (kW)

Pevp,HT = 7 MPa

96 95 94 93

S1 S2

92 91 560

580

600

620

640

660

680

T2(K)

(e) Water Fig. 6. Variation of net output power with T2.

700

660

680

700

105

H. Huang et al. / Energy Conversion and Management 126 (2016) 99–109

fixed temperature at exhaust outlet and condensation temperature of HT loop, T2 changes below the system’s safe operation temper_ f ;HT . ature (702 K) by adjusting m Fig. 6 shows the variation of the net output power with T2. In case of wet working fluids (i.e. ethanol and water), as T2 rises, _ f ;HT decreases continuously, the enthalpy drop of THT increases, m _ f ;HT and T2 together affect the variation of and WP,HT declines. m Wnet. The dry working fluid (R245fa) and the isentropic working fluids (R123 and R141b) show the same trend as wet working fluids at first. After overheating, the enthalpy drop of overheated working fluid in THT keeps increasing; meanwhile the reduction _ f ;HT prevails, and WT decreases. Under the combined action in m _ f ;HT , Wnet increase first and then decreases as the workof T2 and m ing fluid is overheated and the system performance deteriorates. Fig. 7 shows the variation of the thermal efficiency with T2. _ f ;HT decreases, QC, Regarding the wet working fluids as T2 rises, m HT decreases, whereas Wnet keeps rising, thereby leading to the

continuous rising of gth. As for the dry working fluid and the isentropic working fluids, gth rises at first as T2 rises; this is similar to the condition of wet fluids. After overheating the working fluid, the temperature at the turbine exit rises. Due to the combined effects _ f ;HT and the temperature at turbine exit, QC,HT increases, the of m system overall thermal load rises, and simultaneously Wnet declines and therefore gth reduces. Fig. 8 shows the variation of the exergy efficiency with T2. As T2 _ f ;HT , WP,HT and Ein decline. In case of wet working fluids, WT, rises, m HT keeps increasing, so the system efficiency ge rises with T2. As for dry and isentropic working fluids, WT,HT first increases and then drops, so ge varies with WP,HT and WT, rising and then declining. As shown in Figs. 6–8, the performance of S1 is better than that of S2 as long as T2 varies in operation safe range. At the same T2, in high temperature loop, the recovered heat energy of S1 is higher _ f ;HT in S1 is greater than that in S2, and the increthan that of S2, m _ f ;HT is very a small. Therefore, the net ment of WP,HT induced by m 0.0963

0.1059

0.1056

thermal efficiency ηth

thermal efficiency ηth

S1 S2

0.1053

0.1050

0.1047

0.0954 0.0945 0.0936 S1 S2

0.0927 0.0918

0.1044 560

580

600

620

640

660

680

480

700

500

520

560

580

600

620

(b) R245fa

(a) R123 0.1165

0.1360

0.1160

thermal efficiency ηth

0.1377

0.1343 0.1326 0.1309 S1 S2

0.1292 0.1275

0.1155 0.1150 0.1145

S1 S2

0.1140 560

580

600

620

640

660

680

700

560

580

600

620

640

T2(K)

T2(K)

(c) Ethanol

(d) R141b

0.142

thermal efficiency ηth

thermal efficiency ηth

540

T2(K)

T2(K)

0.140

0.138

0.136

S1 S2

0.134 560 580 600 620 640 660 680 700

T2(K)

(e) Water Fig. 7. Variation of thermal efficiency with T2.

660

680

700

106

H. Huang et al. / Energy Conversion and Management 126 (2016) 99–109

0.525

0.490

0.520

exergy efficiency ηe

exergy efficiency ηe

S1 S2

0.515 0.510 0.505

0.480 S1 S2

0.475 0.470 0.465

0.500 560

580

600

620

640

660

680

480

700

500

520

540

560

T2(K)

T2(K)

(a) R123

(b) R245fa

580

600

620

0.560

exergy efficiency ηe

0.630

exergy efficiency ηe

0.485

0.615

0.600

S1 S2

0.585

0.555

0.550

S1 S2

0.545

0.540 580

600

620

640

660

680

700

560

580

600

620

640

T2(K)

T2(K)

(c) Ethanol

(d) R141b

660

680

700

0.639

S1 S2

exergy efficiency ηe

0.630

0.621

0.612

0.603

560

580

600

620

640

660

680

700

T2(K)

(e) Water Fig. 8. Variation of exergy efficiency with T2.

output power of high temperature loop in S1 is greater than in S2. Moreover, since part of the exhaust heat flows into low temperature loop in S2, the net output power of low temperature loop in S2 is slightly greater than that in S1. Both these two factors result in a greater Wnet in S1 than that in S2. Thermal loads of S1 and S2 are quite similar. Together with the fact that Wnet in S1 is greater _ f ;HT in S1 than that in S2, gth in S1 is slightly higher than in S2. m is higher than in S2, and thus the irreversible losses of EHT and THT are greater in S1 than those in S2; whereas S2 has an

irreversible loss in ELT,3 compared with S1, and the irreversible loss in TLT is greater than that in S1. Accordingly, the values of ge in S1 and S2 are different.

4.3. Effect of rotational speed Using water as high temperature working fluid and high temperature evaporation pressure of 7 MPa, the recovery capabilities

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0.145 1000r/min

115

700r/min

1000r/min

thermal efficiency η th

net output power W net (kW)

120

110 105 100 700r/min 700r/min

1000r/min

700r/min

0.140 1000r/min

0.135

95 90 S1

S1

S2

0.130

S2

S1

(a) Variation of net output powers

S1

S2

S2

(b) Variation of thermal efficiencies

0.8

exergy efficiency ηe

700r/min

1000r/min

700r/min

1000r/min

S2

S2

0.6

0.4

0.2

0.0 S1

S1

(c) Variation of exergy efficiency Fig. 9. Performance at different rotational speeds in different systems.

of exhaust at different engine speeds (700 r/min and 1000 r/min) were compared. Fig. 9 shows the system performances at different engine rotational speeds; the performance of S1 is better than that of S2. With the same system at different rotational speeds, Wnet at 1000 r/min Wnet is significantly higher than that at 700 r/min, while gth is lower. In fact, exhaust temperature at 1000 r/min is slightly lower than that at 700 r/min. In the same system, a higher exhaust temperature leads to better heat source quality and higher thermal efficiency; therefore, gth at 1000 r/min is lower. However, although the exhaust temperature at 1000 r/min is lower, the flow rate is much higher, and thus the overall recovered heat energy is greater, leading to the distinct superiority of S1 in Fig. 9(a). There is a very slight difference in ge between 700 r/min and 1000 r/min from Fig. 9(c). 4.4. Effect of the addition of nanoparticles in engine coolant Fig. 10 and Table 4 show the influence of the addition of different types of nanoparticles to engine coolant on system performance. Increasing convective heat transfer coefficient of engine coolant, Wnet of the system grows. In fact, when AE,LT,1 is fixed, KE,LT,1 and heat exchange per unit time rise as the convective heat transfer coefficient of engine coolant increases, and the working fluid at the exit gains additional heat energy. When wet working fluid R143a is the low temperature working fluid, low temperature

loop gains more heat energy and the enthalpy drops in the turbine increases, thereby leading to the increases of WT,LT, increase Wnet further. gth changes little with the increase of convective heat transfer coefficient of engine coolant because the system thermal load increases while Wnet increases. As the convective heat transfer coefficient of engine coolant becomes greater, the irreversible losses of ELT,1 reduce while that of TLT rises. All these three factors work together and influence the trend of system exergy efficiency. In other words, when heat exchange area is fixed, enthalpy drop of the coolant flowing through heat exchanger ELT,1 become greater as the heat transfer coefficient increases, so Ein increase. When the increment of Wnet is greater than that of Ein, the exergy efficiency rises; otherwise, ge reduces. When the carbon nanotube of 0.5% wt is added to coolant (i.e. Da is 350%), the system Wnet increment reaches its maximum of 3.84 kW. Meanwhile, Fig. 10(a) shows that the enhancement of convection by the addition of nanoparticles is almost the same to S1 and S2. 5. Conclusions In the present study, two different Dual-loop organic Rankine cycles (DORC) are proposed to recover heat from engine exhaust, coolant, and high temperature loop. The one-stage exhaust heat recovery system is denoted as S1, and the two-stage system as S2. Nanoparticles in different concentrations are added to coolant to enhance heat transfer. The following conclusions can be drawn.

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0.20

S1 S2 3

thermal efficiency η th

the increment of net output power(kW)

4

2

1

S1 S2

0.16

0.12

0.08

0.04

0 31%

63%

130%

200%

0.00

350%

0%

31%

63%

130%

200%

Δα

Δα

(a) net output power

(b) thermal efficiency

350%

0.64

exergy efficiency ηe

S1 S2 0.63

0.62

0.61

0.60 0%

31%

63%

130%

200%

350%

Δα

(c) exergy efficiency Fig. 10. Influence of Da on system performance.

Table 4 Performance of adding nanoparticles.

Da

0%

31%

63%

130%

200%

350%

Wnet (kW)

S1 S2

95.50 93.67

96.51 94.68

97.22 95.40

98.13 96.32

98.70 96.88

99.34 97.53

gth

S1 S2

13.94% 13.79%

13.92% 13.77%

13.90% 13.76%

13.88% 13.74%

13.87% 13.72%

13.85% 13.71%

ge

S1 S2

62.89% 62.21%

63.24% 62.59%

62.65% 62.02%

62.77% 62.16%

63.20% 62.61%

63.92% 63.37%

(1) Appropriate increases of high temperature evaporation pressure and the temperature at the high temperature turbine inlet can improve the performances of Rankine cycles in both recovery systems. (2) S1 is better than S2 in engine exhaust heat recovery scheme. (3) Water-based S1 performs the best. (4) When the heat exchange area is fixed, the addition of nanoparticles to engine coolant can enhance heat transfer and improve system efficiency, and the maximum increment of net output power is 3.84 kW.

Acknowledgement The research is sponsored by projects of Natural Science Foundation of Guangxi (project Outstanding Young Scholarship Award, Grant No. 2014GXNSFGA118005), Natural Science Foundation of

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