Performance evaluation of an electric vehicle thermal management system with waste heat recovery

Applied Thermal Engineering 169 (2020) 114976

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Performance evaluation of an electric vehicle thermal management system with waste heat recovery

T

Zhen Tiana,b, , Bo Gub, Wenzhong Gaoa, Yuan Zhanga ⁎

a b

Merchant Marine College, Shanghai Maritime University, 201306 Shanghai, China Institute of Refrigeration and Cryogenics, Shanghai Jiaotong University, Shanghai 200240, China

HIGHLIGHTS

simulation model for EVTMS is developed. • Performance EVTMS performance is evaluated in energy, exergy, and thermo-economy. • The the increase in waste heat, COP can be increased by 13.2% at most. • With relative irreversibility rate highly depends on condensation temperature. • The • Compared with PTC heater, the payback period of EVTMS is 4.57 to 6.77 years. ARTICLE INFO

ABSTRACT

Keywords: Driving mileage Electric vehicle Payback period Thermo-economy analysis Waste heat recovery

Widespread adoption of electric vehicles is limited by their mileage, which is severely shortened by the positive temperature coefficient (PTC) heater in cold weather. To alleviate this difficulty, an electric vehicle thermal management system (EVTMS) with waste heat recovery from motor and controller unit is presented in this study. Utilizing a simulation model, the performances of EVTMS are investigated from the aspects of energy, exergy, and thermo-economy. In addition, the effects of waste heat amount and condensation temperature on EVTMS performances are exposed and discussed. It is indicated that increasing waste heat amount and decreasing condensation temperature would contribute to EVTMS performance improvement. Results show that the EVTMS with waste heat recovery leads to, in the best case, the coefficient of performance (COP) increase of 13.2%, the exergy destruction reduction of 20.01%, and the driving mileage improvement of 33.64%. With the waste heat varying in 0–2 kW and condensation temperature varying in 30–45 °C, the COP ranges from 2.05 to 4.71 and the system exergy efficiency varies in 28.63%~40.05%. Compared with the PTC heater, EVTMS with waste heat recovery could save the annual operation cost of 162.31–249.44 €. Base on the operation cost savings, the payback period of EVTMS ranges from 4.57 to 6.77 years.

1. Introduction Transportation is part of the high energy consumption and high greenhouse gas emission sectors [1]. Increasing costs of crude oil and environmental pollution have brought challenges for transportation development. The electric vehicle (EV), offering superior energy efficiency, minimum exhaust emissions as well as quiet operation, has aroused wide attention. Unlike the conventional vehicle, EV without internal combustion engine could not provide waste heat for cabin heating. Besides, the waste heat generated by the motor and controller unit should be removed in case of thermal runaway. In terms of cabin thermal comfort, the positive temperature

⁎

coefficient (PTC) heater is generally used in EV. PTC heater transfers battery electricity to heat directly, which provides the maximum efficiency of 1. The application of PTC heater seriously impacts vehicle driving mileage, acceleration, and battery efficiency [2,3]. Therefore, it is eager to find an alternative method for EV cabin heating. The heat pump (HP) system with the utilization of refrigerant latent heat could achieve higher energy efficiency and reduce electricity consumption [4–6]. Since the popularity of EV has not been long, the research of HP system for EV is still in the embryonic stage. Especially when the HP system works under severe cold conditions, the evaporator frost and compressor performance recession would lead to heating performance decrease [7].

Corresponding author at: Merchant Marine College, Shanghai Maritime University, 201306 Shanghai, China. E-mail address: [email protected] (Z. Tian).

https://doi.org/10.1016/j.applthermaleng.2020.114976 Received 16 August 2019; Received in revised form 14 January 2020; Accepted 19 January 2020 Available online 20 January 2020 1359-4311/ © 2020 Elsevier Ltd. All rights reserved.

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Nomenclature a, b A AC AOC cp C CD D ex E Ex Exdes Fr h j Lpip Lday Ldri m M Mr Nday p PP Pr Q rad Re s S T Tc TCC Te T ΔT u U Vfan Vlt W We x Xtt

cold comp cond cool cr evap hex hot in lab liq lt nb out pip pow r rad spec sp sv sys vap w wst 0

coefficients matrix heat transfer area (m2) circulation area (m2) annual operation cost (€) specific heat capacity (kJkg−1K−1) cost (€) flow coefficient hydraulic diameter (m) specific exergy (kJkg−1) enhancement factor exergy rate (kW) exergy destruction rate (kW) froude number specific enthalpy (kJkg−1) factor pipeline length (m) daily driving mileage (km) driving mileage (km) mass flow rate (kgh−1 or kgs−1) mole mass (gmol−1) refrigerant charge (g) number of driving days pressure (MPa) payback period (year) Prandtl number heat rate (kW) radiator Reynolds number specific entropy (kJkg−1°K−1) suppression factor temperature (°C or K) condensation temperature (°C) total capital cost (€) evaporation temperature (°C) entropy-averaged temperature (°C) temperature difference (°C) velocity (ms−1) overall heat transfer coefficient (kWm−2°C−1) volume flow rate (m3s−1) liquid tank volume (L) power consumption (kW) Weber number vapor fraction lockhart–martinelli parameter

Greek letters α δ ζele η ηir λ μ ρ Φbat

heat transfer coefficient (kWm−2°C−1) thickness (m) electricity price (€kWh−1) efficiency relative irreversibility rate thermal conductive coefficient (kWm−1°C−1) dynamic viscosity (Pas) density (kgm−3) battery capacity (kWh) electricity consumption (kWhkm−1)

Acronym COP EV EVTMS EXV HP M&C PTC PWHR UDDS

Subscript a ave

cold side compressor condensation coolant critical evaporation heat exchanger hot side inlet labour liquid liquid tank nucleate boiling outlet pipeline power system refrigerant radiator specific single phase saved system vapor wall waste reference state

air side average

With regard to the improvement of HP performance, modifications have been proposed including vapor injection, refrigerant substitutes and mixtures, combined defrost methods, integrated desiccant, and sorption [8–13]. Han et al. [14] studied the performance of a vapor injection air-source HP for an electric rail vehicle. Compared to the noninjection cycle, the heating capacity was found to be maximumly increased by 19.3% with the injection ratio of 0.29. The experimental results of Yu et al. [15] demonstrated that energy consumption of the HP system for an electric vehicle could be reduced up to 41% and 72% with the utilization of R744/R290 and R152a, respectively. In their another study [16], an autocascade HP with CO2-propane mixture as working fluid for EV heating was proposed, which proved that COP

coefficient of performance electric vehicle electric vehicle thermal management system electronic expansion valve heat pump motor and controller positive temperature coefficient percentage of waste heat recovery urban dynamometer driving schedule

increase could be up to 12.3% compared to the CO2 single stage system at −20 °C. Zühlsdorf et al. [17] concluded that the thermodynamic performances of HP could be improved by 35% with a mixture of 30% propylene and 70% R-1234ze(Z). Liu et al. [6] experimentally studied the heating performance of a HP system for EV under various operational conditions. They concluded that propane was the possible refrigerant for vehicle HP system when the ambient temperature was above −10 °C, while the HP system with CO2 showed higher efficiency when the ambient temperature was −20 °C. To ease frost problem, the vehicle HP system with a desiccant-coated heat exchanger was proposed by Zhang et al. [18]. The simulation results revealed that the compressor power consumption was reduced by 38% at an ambient 2

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temperature of −20 °C, which contributed to a cruising mileage of 172 km. Pan et al. [19] evaluated the HP system using an annual energy consumption model. To reduce electricity consumption, the recirculated air was used in their study. The results demonstrated that the heating energy saving in HP system could be up to 57% and the driving mileage could be increased by 11–30%. On careful inspection, the methods mentioned above are aimed at increasing the evaporation temperature and improving the refrigerant mass flow rate. It should be noted that the two parameters highly depend on the quality of the heat source. Therefore, an additional heat source with high quality is required to restrain HP performance decrease. The best solution is the combination of waste heat recovery and HP system, which forms the electric vehicle thermal management system (EVTMS). Lee et al. [20] studied the EVTMS heating performance for an electric bus with waste heat generated by the electric devices. Their results showed that the heating coefficient of performance (COP) was decreased with increasing ambient temperature and compressor frequency. Ahn et al. [21] experimentally investigated an EVTMS with the dual-source HP system, which demonstrated that the EVTMS heating capacity was increased 31.5% and the COP was increased 9.3% with the waste heat recovery. In another study [22], performances of a dehumidifying HP system were experimentally studied. The HP system with waste heat recovery showed 15.8% higher heating capacity and 5.2% higher COP compared to the air source HP system. Tian et al. [23] proposed an EVTMS with battery cooling and motor waste heat recovery. Experiments for EVTMS performance test were carried out in an environmental chamber, which showed that the COP was increased up to 25.55% with waste heat recovery and the driving range was accordingly increased by 31.71%. Studies on the EVTMS exergy analyses have been conducted by researchers. The areas of low exergy efficiency in the EVTMS were examined in Hamut et al.′s study [24]. The results demonstrated that the exergetic COP was in the range of 0.26–0.39 and the ambient temperature was the most significant parameter for EVTMS performance. Later on, Hamut et al. [25] carried out the optimization for maximizing the EVTMS exergy efficiency based on multi-objective evolutionary algorithms, which showed that the exergy efficiency could be increased by 14% compared to the baseline case. Exergy analyses of the EVTMS conducted by Javani et al. [26] showed that the increase in evaporation and condensation temperature would result in exergy efficiency decrease and exergy loss increase. They pointed out that decreasing the mean temperature difference between working fluids is beneficial to exergy efficiency improvement. Tian et al. [27] conducted exergy study on the proposed EVTMS, which revealed that the largest component exergy destruction occurred in the external heat exchanger, accounting for 46.1%. The exergy destructions in scroll compressor and external heat exchanger accounted for more than 80% of the system exergy destruction. Therefore, they suggested these two components to be the most promising components for system performance optimization. Despite the growing interest in EVTMS, further studies are still required to provide more practical results. The innovation of this paper originates from thermo-economy analysis of the EVTMS under various conditions. The effects of condensation temperature and waste heat amount on energy and exergy efficiency of the EVTMS have been investigated. Based on the analysis results, the operation cost, driving mileage, and payback period of the EVTMS have been clarified.

the cabin thermal comfort is realized by the HP subsystem. Refrigerant from the liquid tank is compressed into the condenser (1–2) and then separated into two paths by electronic expansion valve 1 (EXV1) (2–3) and EXV2 (2–14). With low temperature and pressure, the refrigerant in the two paths is evaporated in evaporator (3–4) and radiator (14–4), respectively. Meanwhile, the waste heat generated by M&C (5–6) is dissipated to the coolant, which is pumped to the radiator (7–5). Under cooling mode, the coolant in the radiator is cooled down by air (8–9). Under heating mode, the waste heat in the coolant is taken away by the refrigerant (14–4). The conversion of cooling mode and heating mode is completed by the four-way valve. The circulations, working fluids, and heat transfer direction are illustrated in Fig. 1(b). The waste heat generated by M&C is equivalent to an additional heat source. 2.2. Experimental setup and uncertainties Experimental setup for EVTMS performance test is established in a climate chamber. The air temperature and velocity are controlled by the air handling unit and blower. PTC heaters are utilized to simulate the waste heat generated from the battery unit and motor unit. The specifications of the main components are shown in Table 1. The refrigerant used in the EVTMS is R134a. The coolant is ethylene glycol with a mass fraction of 50%. The detailed experimental setup, test facilities, and test procedures could be referred to [23,27]. The uncertainties of the derived parameters are within 4.13%. 3. Mathematical model 3.1. EVTMS simulation model Scroll compressor: The scroll compressor model is defined as a polynomial function of condensation temperature (Tc) and evaporation temperature (Te), as shown in Eq. (1).

f=

aT ·T bT

(1)

a = [a1 a2 a3 a4 a5 a6 a7 a8 a9 a10 ]T , f = [Qcomp Wcomp where, T b = [b1 b2 b3 b4 b5 b6 b7 b8 b9 ] , T= [Te Te 2 Te 3 Tc Tc 2 Tc 3 Te Tc Te 2Tc TeTc 2]T . The coefficients in a and b are generated from the experimental data, which are summarized in Table 2. EXV: The mass flow rate through the EXV (mEXV ) is predicted by Eq. (2). ]T ,

mEXV = CD AC 2

in (pin

(2)

pout )

where, CD is flow coefficient, CD = 0.02005 in + 0.634 out ; AC is the circulation area, m2; ρ is the density, kgm−3. Heat exchanger: The heat transfer amount (Q ) of a heat exchanger is calculated by: (3)

Q = UA TLMTD −2

−1

where, U is the overall heat transfer coefficient kWm °C ; ΔTLMTD is the logarithmic mean temperature difference between hot and cold fluids, °C. U and ΔTLMTD are expressed as follows:

1 = UAspec TLMTD =

2. System description

1 hot Aspec, hot

Tout ln

+

w w Aspec

+

1 cold Aspec, cold

(4)

Tin Tout Tin

(5)

where, αhot and αcold are local heat transfer coefficients for hot and cold fluid, kWm−2°C−1; Aspec is the specific heat transfer area, m2; λw is the tube wall thermal conductive coefficient, kWm−1°C−1; δw is the tube wall thickness, m; ΔTout and ΔTin are the two fluids temperature difference at the outlet and inlet of heat exchangers. The heat transfer coefficient for single-phase (αsp) turbulent flow is

2.1. EVTMS working principle Fig. 1 demonstrates the layout of the EVTMS and its working principle on the temperature-entropy (T-s) diagram. As shown in Fig. 1(a), the system includes two subsystems: a HP subsystem and a motor and controller (M&C) cooling subsystem. Under heating mode, 3

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calculated according to the Dittus-Boelter correlation [28], as shown in Eq. (6). sp

= 0.023

D

Re0.8Pr n

cond

evap

= jcp, a ua a Pr

2 3

sp

(1

x )0.8 +

3.8x 0.76 (1 x )0.04 (p pcr )0.38

(8)

where, αsp is calculated by the Eq. (6); p is the pressure, MPa; pcr is the critical pressure, MPa; x is the vapor fraction. Evaporator: With respect to the evaporation process, the evaporation heat transfer coefficient (αevap) is determined by the single-phase heat transfer coefficient and nucleate heat transfer coefficient (αnb), as shown in Eqs. (9)–(12).

(6)

where, D is the hydraulic diameter, m; the component n is 0.3 and 0.4 for condensation and evaporation process, respectively. The heat transfer coefficient for the air side (αa) is calculated by Eq. (7). a

=

(7)

=E

sp

+S

(9)

nb

0.32 E = (1 + 9.8Xtt 0.2)We vap

where, the j factor can be referred to [29]; cp,a is the air specific heat, kJkg−1°K−1; ua is the air velocity, ms−1; ρa is the air density, kgm−3; Re is the Reynolds number. Condenser: The condensation heat transfer coefficient (αcond) is calculated by the correlation suggested by Yang and Webb [30].

S= nb

(10)

Bo0.02Fr 0.04exp 0.81x 1

(11)

10 4Reliq E0.79

= 55(p pcr )0.12 ( lg(p pcr ))

0.55M 0.5q 0.67

Fig. 1. (a) The layout of EVTMS and (b) working principle on T-s diagram. 4

(12)

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Table 1 Specifications of the components in the EVTMS. Components

Type

Specifications

Compressor

Scroll

Condenser Four-way valve Evaporator

Parallel flow mini channel – Laminated mini channel

Expansion valve Radiator Motor and controller Liquid tank

Electronic type Plate type PTC heater –

Discharge: 27 cm3 s−1 Nominal power: 2.5 kW Refrigerant side area: 0.59 m2 Air side area: 4.6 m2 Nominal capacity: 10 kW Refrigerant side area: 0.89 m2 Air side area: 8.55 m2 Nominal capacity: 6.2 kW Plate area: 0.008 m2 Heat capacity: 0–5 kW Vlt = 0.3 L

3.2.1.1. Energy analysis. The capacity of the condenser (Qcond ) is regarded as the EVTMS heating capacity (Qsys ), which is expressed in terms of mass flow rate and enthalpy differences between the inlet and the outlet of the condenser. To decrease the calculating error, both air side and refrigerant side are taken into consideration. That is,

Qcond = Qsys =

D

Re0.55Pr1

3

µ µw

(13)

Qevap =

where, Pr is the Prandtl number; μ and μw are the coolant dynamic viscosities at the tube radius and wall, Pas.

Qrad =

3.2.1. Theoretical basis For the sake of analysing EVTMS performances, the following assumptions have been made: (1) The pressure drops in all fluids and components are negligible; (2) The compression process is adiabatic, and the expansion process is isenthalpic; (3) The ambient air is ideal gas; (4) The kinetic, potential, and chemical energy changes are not considered; (5) The reference state is T0 = 0 °C and p0 = 0.101 MPa; (6) The directions of mass and heat transfer to the EVTMS and work transfer from the EVTMS are positive. Each component in the EVTMS can be regarded as a control volume with the mass flow (m ), heat transfer (Q ), and work interactions (W ). The mass and energy balance equations are as follows:

mr ,1 (hr ,1

COP =

mr ,3 (hr ,4

mr ,14 (hr ,4

(21)

hr ,14) + mcool cp, cool (Tcool,7

Tcool,5 )

(22)

Qsys (23)

Qrad Qsys

(24)

3.2.1.2. Exergy analysis. As for the compressor, exergy destruction (Exdes, comp ) is mainly caused by electrical and mechanical losses, which can be expressed as:

Exdes, comp = Wcomp + mr (ex1 = Wcomp + mr [(h1

(16)

h4)

ex 4 )

T0 (s1

(25)

s4 )]

Exergy destructions in heat exchangers are due to the fluid temperature difference. The exergy destructions of the condenser (Exdes, cond ) and evaporator (Exdes, evap ) are calculated by Eqs. (26) and (27), respectively.

(17)

Ex = mex

ha,10 )

Wcomp

PWHR =

(15)

s0 )

hr ,3) + ma,10 (ha,11 2

2

(14)

T0 (s

(20)

The percentage of waste heat recovery (PWHR) defined in Eq. (36) is used to characterize the contribution of waste heat recovery to EVTMS heating performance.

where, h is the specific enthalpy, kJ kg ; the subscript in and out refers to the inlet and out of the volume. The specific exergy (ex) and exergy (Ex ) are defined in Eqs. (16) and (17), respectively.

h0)

hr ,4 )

Compared to the compressor power, the fans power is small enough (less than 2%) and could be neglected. The COP of the EVTMS is calculated by:

−1

ex = (h

(19)

The heat amount recovered from the M&C is defined as the capacity of the radiator (Qrad ), which is calculated by Eq. (22).

3.2. Performance analysis model

(mh)in + Qin = (mh)out + Qout + W

hr ,2 )

where, ηcomp is the isentropic efficiency of the compressor, which is assumed as 0.8. The capacity of the evaporator (Qevap ) is the heat amount absorbed from the ambient air, which is expressed in Eq. (21).

0.14

min = mout

ha,12) + mr ,1 (hr ,1 2

comp

where, E is the convective boiling enhancement factor and S is the nucleate boiling suppression factor, which can be referred to [31]. Radiator: The Mc Adams correlation [32] is used for calculating the coolant heat transfer coefficient (αcool).

= 0.36

ma,12 (ha,13

The compressor power consumption (Wcomp ) can be evaluated as:

Wcomp =

cool

(18)

Ex in = Exout + Exdes

where, T is the temperature, K; s is the specific entropy, kJ kg−1K−1; the subscript of 0 refers to the reference state. The main reasons for exergy destruction (Exdes ) in the main components of EVTMS include: the compression and expansion processes are not isentropic and the mismatching temperature between heat exchangers and heat/cold sources. TheExdes is determined by

Exdes, cond = mr (ex1 = mr [(h1

h2)

ex2)

T0 (s1

s2)]

Qcond

Tcond T0 Tcond

Qcond

Tcond T0 Tcond

(26)

Table 2 Coefficients for the scroll compressor model. a0

a1

a2

a3

a4

a5

a6

a7

a8

a9

3.39 × 10−2 b0 444.51

9.94 × 10−4 b1 −13.28

1.88 × 10−5 b2 −0.36

4.05 × 10−7 b3 4.94 × 10−2

−3.81 × 10−4 b4 14.97

8.06 × 10−6 b5 0.33

−6.09 × 10−8 b6 −2.37 × 10−3

−1.06 × 10−6 b7 1.71 × 10−2

−9.54 × 10−8 b8 −1.83 × 10−3

2.06 × 10−8 b9 9.92 × 10−3

5

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Z. Tian, et al.

Exdes, evap = Qevap = Qevap

T0

Tevap Tevap

T0

Tevap Tevap

mr [(h4

mr (ex 4 h3)

proportion (0.2–0.3) of the system capital cost. Therefore, the TCC for EVTMS is calculated by Eq. (36).

ex3)

T0 (s4

s3)]

TCC = Ccomp + Cpump + Chex + Cr + Cpip + Clt + Cfan + CEXV + CPTC + Clab

(27)

(36)

where, Tcond and Tevap are the entropy-averaged temperatures of the heat exchanger inlet air and out air, which are calculated by Eqs. (28) and (29) [33].

Tcond =

h13 s13

h12 T T12 = 13 s12 ln(T13 T12 )

(28)

Tevap =

h10 s10

h11 T T11 = 10 s11 ln(T10 T11)

(29)

It is assumed that the number of heating days per year (Nday) is 150 days. For the vehicle, the useful time during the day is 2 h. The average velocity (uave) is assumed as 31.5 kmh−1 according to the urban dynamometer driving schedule cycle. Therefore, the daily driving distance (Lday) is thus determined as 63 km. The battery capacity (Фbat) is assumed as 100 kWh. The electricity consumed by power system ( pow ) is 0.4 kWhkm−1 according to Falcão et al.′s [34] suggestion. AOC is calculated by Eq. (37). The driving mileage (Ldri) is an important parameter for EV promotion. Ldri is determined by the ratio of battery capacity and vehicle electricity consumption, as shown in Eq. (38).

The exergy destructions in the EXV1 and EXV2 are determined by Eqs. (30) and (31), respectively.

Ex des, EXV 1 = mr ,3 (ex3 Ex des, EXV 2 = mr ,14 (ex14

ex2) = mr ,3 T0 (s2

(30)

s3)

ex2) = mr ,14 T0 (s2

AOC =

(31)

s14)

The exergy destruction of the radiator (Ex rad ) is obtained from Eq. (32). Exdes,rad = mr ,14 (ex14 ex 4 ) + mcool,7 (ex7 ex5) = mr ,14 [(h14 h 4 ) T0 (s14 s4 )] + mcool,7 [(h7 h5) T0 (s7

Ldri =

(32)

(33) The relative irreversibility rate (ηir,i) is defined as the component exergy destruction (Exdes, i ) over the system exergy destruction, which is shown in Eq. (34).

AOCsv = PP =

(34)

= Exdes, i Exdes, sys

The exergy efficiency of the EVTMS (ηsys) is defined as the ratio between the net output exergy of the system and the input power of the system, which is expressed in Eq. (35).

=

Ex cond, in Excond, out Wcomp

pow

+

bat WEVTMS uave

(38)

where, ζele is electricity price, €kWh ; WEVTMS is the electricity consumed by EVTMS, kW. Under the same driving conditions, thermo-economy analyses of EVTMS and PTC heater are investigated. The saved annual operation cost (AOCsv) is calculated by Eq. (39). The payback period (PP) is determined by Eq. (40).

Exdes,sys = Exdes, comp + Exdes, cond + Exdes, EXV 1 + Ex des, EXV 2 + Exdes, evp + Exdes,rad

sys

=

(37)

−1

The system exergy destruction (Exdes, sys ) of the EVTMS is calculated by Eq. (33).

ir , i

bat EV

s5)]

WEVTMS uave

ele Nday Lday

ele Nday Lday

WPTC

WEVTMS uave

(39)

TCCEVTMS - TCCPTC AOCsv

(40)

3.3. Solving method The calculation flow diagram of EVTMS model is illustrated in Fig. 2. First, input the practical parameters of each component and the operating conditions. The condensation and evaporation temperatures are assumed. After that, each sum-model is calculated consecutively. In order to calculate the refrigerant mass flow rate through EXV1 and EXV2, the condensation temperature is adjusted. The evaporation temperature is adjusted to achieve the convergence of mass flow rate through compressor and expansion valves. Thermodynamic properties of the refrigerant are calculated with REFPROP 9.1 [41]. Afterward, the overall thermal and economic performance criteria would be

(35)

3.2.1.3. Thermo-economy analysis. Thermo-economy analysis of the EVTMS is carried out based on total capital cost (TCC) and annual operation cost (AOC). With respect to TCC, the main components such as compressor, pump, heat exchanger, refrigerant, pipeline, liquid tank, EXV, and PTC heater are considered. The capital cost estimation is summarized in Table 3. The labor cost (Clab) is typically treated as a Table 3 Capital cost calculation for EVTMS components. Components Scroll compressor Pump Heat exchanger Refrigerant cost Pipe cost

Capital cost

Ccomp = 98400 Cpump = 2100

Wcomp 250 Wpump 10

0.46

0.26 1

pump

0.5

Chex = 190 + 310Ahex Cr = 20Mr Cpip = (0.87 + 0.21Dpip ) Lpip

[35,36]

2.5 kW

[35,36]

0.2 kW; 0.8

[35] [37] [38]

7.32 m2 800 g 10–15 m

[39] Manufacture

0.11 kgs−1 6 kW

[39] [40]

Clt = 31.5 + 16Vlt Cfan = 139.5(Vfan + 1.43)

EVTMS cost

CPTC = 15.3Qheat Csys = Ccomp + Cpump + Chex + Cr + Cpip + Clt + CEXV + CPTC

Total cost

TCC = 1.3Csys

Labour cost

Values in this paper

pump

Liquid tank Fan

EXV PTC heater

Reference

CEXV = 101.4mEXV

0.3Csys

6

0.3 L 0.14 m3s−1

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Z. Tian, et al.

Fig. 2. Flow chart of EVTMS model.

calculated.

absolute error for the system power consumption are 6.12% and 13.82%, respectively. The refrigerant mass flow rate is found to have a mean absolute error of 4.84% and a maximum absolute error of 7.99%. The simulation results are in good agreement with the experimental data, and all deviations are within the error band of ± 15%. Therefore, the simulation model could be used for EVTMS performance prediction.

4. Results and discussions In this section, the EVTMS simulation model is validated with the experimental data. Afterward, EVTSM performances are evaluated from these three aspects: energy, exergy, and thermo-economy. The effects of condensation temperature and waste heat amount on EVTMS performances are discussed. The energy performance of EVTMS is evaluated by PWHR and COP. EVTMS exergy performance is evaluated by system exergy destruction, exergy efficiency, and relative irreversibility rate. As for the thermo-economy performance, the annual operating cost, driving mileage and payback period are estimated. To make the results more referential, comparisons between EVTMS and PTC heater are made.

4.2. EVTMS performance 4.2.1. Energy analysis Variations of PWHR and COP with condensation temperature and waste heat amount are shown in Fig. 4. As shown in Fig. 4, PWHR varies in the range of 0–0.69. PWHR increases with the condensation temperature increase, indicating that the waste heat is more easily obtained under unfavourable conditions. Moreover, PWHR increases with the increase in waste heat mount, namely the ratio of refrigerant flowing through the radiator would be increased. COP varies in the range of 2.05–4.71. Increasing the waste heat amount leads to COP increase. The contribution of waste heat to the COP increase is as high as 13%. This could be due to the fact that the waste heat generated by the motor and controller unit is a high-temperature heat source. The temperature of the refrigerant flowing through the radiator would be improved, which profits compressor performance improvement. COP decreases as the condensation temperature increases, but the decrease rate becomes narrowed. This can be ascribed to the larger compressor pressure ratio caused by higher condensation temperature, which would lead to compressor power consumption increase.

4.1. Model validation To validate the EVTMS simulation model, the predicted data is compared with the experimental dada under the same working conditions. In the experiments, the environmental temperature varies in the range of −5 ~ 5 °C, while the cabin temperature is kept at 20 °C. The operating range of scroll compressor speed is from 2000 to 5000 rpm. The inlet and outlet coolant temperatures are set at 20 °C and 60 °C, respectively. The air velocities for condenser and evaporator are 3 ms−1 and 4 ms−1, respectively. Waste heat amount varies in the range of 0–2 kW. Fig. 3 demonstrates comparisons of experimental data and predicted data for heating capacity, power consumption, and refrigerant mass flow rate. The heating capacity shows a mean absolute error of 6.37% and a maximum absolute error of 14.89%. The mean and the maximum

4.2.2. Exergy analysis Fig. 5 reveals the effects of waste heat amount and condensation temperature on the system exergy destruction and exergy efficiency. 7

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Fig. 3. Comparisons of experimental data and predicted data for (a) heating capacity, (b) power consumption, and (c) refrigerant mass flow rate.

amount. The results in Figs. 5 and 6 are consistent, that is, the reduction of system exergy destruction indicates the improvement of system exergy efficiency. The system exergy efficiency ranges in 28.63–40.05%. In order to locate the most likely component to improve system exergy performances, the relative irreversibility rate is calculated. The relative irreversibility rate changes with waste heat amount and condensation temperature are demonstrated in Fig. 7. Compared Fig. 7(a)–(d), the relative irreversibility rate for all components varies greatly for different waste heat amount. The ηir,evap remains almost the same for different condensation temperatures, but the ηir,rad is slightly

Fig. 4. Variations of PWHR and COP with Tcond and Qwst .

The system exergy destruction is in the range of 1.28–2.63 kW. At a given waste heat, the system exergy destruction increases with the increase of condensation temperature, which is caused by the temperature difference between refrigerant and air. The waste heat could reduce the system exergy destruction. The reason is that the increasing evaporation temperature profits the reduction of fluids temperature difference in the heat exchangers. The system exergy destruction is maximumly decreased by 20.01% with 2 kW waste heat. Accordingly, the results of the system exergy efficiency are shown in Fig. 6. The system exergy efficiency is found to decrease with increasing condensation temperature, while increase with increasing waste heat

Fig. 5. Variation of Ex des, sys with Tcond and Qwst . 8

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the range of 27.31%~32.94% and 0–14.78%. When the condensation temperature is 30 °C, the ηir,EXV1 and ηir,EXV2 keep almost steady. As shown in Fig. 7(b), the ηir,comp and the ηir,EXV1 are increased compared to that in Fig. 7(a), which is caused by the pressure ratio increase. Because of the increase of PWHR, the ηir,evap decreases from 28.48% to 15.97% when the waste heat amount increases from 0 to 2 kW. When the condensation temperature is 35 °C, the ηir,EXV2 increases from 0 to 2.19%. Fig. 7(c) and (d) show the component relative irreversibility rate when the condensation temperature is 40 °C and 45 °C, respectively. As can be seen from Fig. 7(c), the ηir,evap and the ηir,EXV1 respectively decreases from 20.98% to 10.36% and 14.91% to 7.36% with waste heat amount increasing. The reason for this phenomenon is the decrease in mass flow rate, which could be proved by the PWHR (as shown in Fig. 4). The ηir,EXV2 increases to 4.82% when the waste heat increases to 2 kW. Fig. 7(d) shows the maximum ηir,comp increases to 39.05% while the maximum ηrad decreases to 9.01%. From these figures, it can be concluded that the sensitive parameters for the exergy destruction and the relative irreversibility rate are refrigerant mass flow rate, temperature difference, and pressure ratio, which are affected by waste heat amount and condensation temperature. Potential locations for exergy performance improvement are scroll compressor, evaporator, and condenser. Utilization of ejectors and heat exchangers with hydrophobic surfaces in EVTMS might be the possible ways to further improve the exergy efficiency.

Fig. 6. Variation of ηsys with Tcond and Qwst .

decreased as the condensation temperature is increased. The relative irreversibility rate for other components shows different trends with condensation temperature variation. Specific discussions are listed as following: As shown in Fig. 7(a), the ηir,comp decreases from 31.18% to 26.64% with the waste heat amount increasing from 0 to 2 kW. The increase in waste heat amount reduces the pressure ratio, which would decrease the irreversibility in compressor. The ηir,evap decreases from 37.91% to 23.08%, which can be attributed to the increase in evaporation temperature leading to the temperature difference reduction between refrigerant and environment temperature. The variation of refrigerant mass flow rate contributes to ηir,cond and the ηir,rad respectively varies in

4.2.3. Thermo-economy analysis Comparisons of annual operation cost for PTC heater and EVTMS are shown in Fig. 8. The annual operation cost of PTC heater is estimated as 316.8 €, while the annual operation cost of EVTMS varies greatly with the variation of waste heat amount and condensation temperature. The annual operation cost of EVTMS decreases with the

Fig. 7. Variations of ηir with Tcond and Qwst (a) Tcond = 30 °C, (b) Tcond = 35 °C, (c) Tcond = 40 °C, (d) Tcond = 45 °C. 9

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consumption and thus extend the vehicle driving mileage. However, the condensation temperature has a negative effect on the vehicle driving mileage, which could be supported by the COP reduction (as shown in Fig. 4). When the waste heat amount is 2 kW, the vehicle driving mileage decreases from 239.94 km to 217.18 km with the condensation temperature increasing from 30 °C to 45 °C. The vehicle driving mileage is 179.53 km under PTC heater working mode. Compared to the PTC heater, the vehicle driving mileage with waste heat recovery could be maximumly increased by 33.64%, 29.62%, 25.35%, and 20.97% when the condensation temperature is 30 °C, 35 °C, 40 °C, and 45 °C, respectively. It is worth noting that for different waste heat amount, the driving mileage could be accordingly increased by 1.28%~1.52%. The proportion is not remarkable owing to the electricity consumed by the EVTMS is not comparable to the electricity consumed by the power system. Fig. 10 demonstrates the results of annual operation cost saving and payback period. The annual operation cost saving ranges from 162.31 € to 249.44 €. The increase in waste heat amount results in the increase of annual operation cost saving. For a specific waste heat amount, the operation cost saving decreases with the increased condensation temperature. The variation of payback period is in the opposite direction of the operation cost saving. The shortest payback period is 4.57 years and the longest payback period is 6.77 years, which could supply useful information for the electric vehicle manufactures.

Fig. 8. Comparisons of AOC between EVTMS and PTC heater.

5. Conclusions In this paper, the electric vehicle thermal management system (EVTMS) performances under various conditions have been studied from the perspective of thermodynamic and thermo-economic. Additionally, the effects of the waste heat amount and condensation temperature on the performances of EVTMS were investigated. (1) The heating COP of the EVTMS varies in the range of 2.05–4.71. With the waste heat recovery from the motor and controller unit, the COP maximumly increases up to 13%. However, COP decreases up to 50.72% when the condensation temperature increases from 30 °C to 45 °C. (2) Exergy analysis results demonstrate that the waste heat recovery contributes to 20.01% reduction in system exergy destruction. The maximum system exergy efficiency is 28.63% and 40.05% for EVTMS without and with waste heat recovery, respectively. The irreversibility mainly occurs in compressor, condenser, and evaporator, where EVTMS performance is most likely to be improved. (3) The annual operation cost saving of the EVTMS ranges from 162.31 € to 249.44 € and the payback period is in the range of 4.47–6.77 years. With the EVTMS, the driving mileage could be extended by 33.64% compared to that with PTC heater.

Fig. 9. Comparisons of Ldri between EVTMS and PTC heater.

In conclusion, EVTMS with waste heat recovery is advantageous in energy and operation cost saving compared to PTC heaters. At the same time, electricity saving could reduce battery charge–discharge frequency, which would bring about profits for the life of electric vehicles. Fig. 10. Variations of AOCsv and PP with Tcond and Qwst .

Declaration of Competing Interest

increase of waste heat amount while increases with the increase of condensation temperature. In the most adverse case, the EVTMS has an annual operation cost of 144.8 €, which is decreased by 54.8% compared to that of PTC heater. The annual operation cost of EVTMS could be maximumly reduced by 78.74%, 73.87%, 66.89%, and 59.21% when the condensation temperature is 30 °C, 35 °C, 40 °C, and 45 °C, respectively. Fig. 9 presents the results of vehicle driving mileage with the variation of waste heat amount and condensation temperatures. As for the EVTMS, increasing the waste heat amount would reduce electricity

All authors have declared that there is no conflict of interest for this paper. Acknowledgement The authors would like to acknowledge the support by the National Natural Science Foundation of China (Grant No. 51706129), Postdoctoral Science Foundation of China (Grant No. 2019M650084), “Chenguang Program” (Grant No. 17CG50), and Shanghai Science and Technology Commission (Grant No.18040501800). 10

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Appendix A. Supplementary material [20]

Supplementary data to this article can be found online at https:// doi.org/10.1016/j.applthermaleng.2020.114976.

[21]

References [22]

[1] Z. Li, A. Khajepour, J. Song, A comprehensive review of the key technologies for pure electric vehicles, Energy 182 (2019) 824–839. [2] N. Javani, I. Dincer, G.F. Naterer, Thermodynamic analysis of waste heat recovery for cooling systems in hybrid and electric vehicles, Energy 46 (2012) 109–116. [3] C.-W. Cho, H.-S. Lee, J.-P. Won, M.-Y. Lee, Measurement and evaluation of heating performance of heat pump systems using wasted heat from electric devices for an electric bus, Energies 5 (2012) 658–669. [4] S. Bellocchi, G. Leo Guizzi, M. Manno, M. Salvatori, A. Zaccagnini, Reversible heat pump HVAC system with regenerative heat exchanger for electric vehicles: analysis of its impact on driving range, Appl. Therm. Eng. 129 (2018) 290–305. [5] Y. Wang, W. Li, Z. Zhang, J. Shi, J. Chen, Performance evaluation and prediction for electric vehicle heat pump using machine learning method, Appl. Therm. Eng. 159 (2019) 113901. [6] C. Liu, Y. Zhang, T. Gao, J. Shi, J. Chen, T. Wang, L. Pan, Performance evaluation of propane heat pump system for electric vehicle in cold climate, Int. J. Refrig. 95 (2018) 51–60. [7] H.-S. Lee, M.-Y. Lee, Steady state and start-up performance characteristics of air source heat pump for cabin heating in an electric passenger vehicle, Int. J. Refrig. 69 (2016) 232–242. [8] G. Zhou, H. Li, E. Liu, B. Li, Y. Yan, T. Chen, X. Chen, Experimental study on combined defrosting performance of heat pump air conditioning system for pure electric vehicle in low temperature, Appl. Therm. Eng. 116 (2017) 677–684. [9] G. Zhang, H. Zou, F. Qin, Q. Xue, C. Tian, Investigation on an improved heat pump AC system with the view of return air utilization and anti-fogging for electric vehicles, Appl. Therm. Eng. 115 (2017) 726–735. [10] L. Jiang, R.Z. Wang, J.B. Li, L.W. Wang, A.P. Roskilly, Performance analysis on a novel sorption air conditioner for electric vehicles, Energy Convers. Manage. 156 (2018) 515–524. [11] C. Kwon, M.S. Kim, Y. Choi, M.S. Kim, Performance evaluation of a vapor injection heat pump system for electric vehicles, Int. J. Refrig. 74 (2017) 138–150. [12] J.H. Ahn, H. Kang, H.S. Lee, Y. Kim, Performance characteristics of a dual-evaporator heat pump system for effective dehumidifying and heating of a cabin in electric vehicles, Appl. Energ. 146 (2015) 29–37. [13] C. Fan, G. Yan, J. Yu, Thermodynamic analysis of a modified solar assisted ejectorcompression heat pump cycle with zeotropic mixture R290/R600a, Appl. Therm. Eng. 150 (2019) 42–49. [14] X. Han, H. Zou, W. Kang, H. Xu, Y. Chen, C. Tian, Experimental investigation on vapor injection ASHP for electric rail vehicles in cold region, Appl. Therm. Eng. 153 (2019) 473–482. [15] B. Yu, J. Yang, D. Wang, J. Shi, J. Chen, Energy consumption and increased EV range evaluation through heat pump scenarios and low GWP refrigerants in the new test procedure WLTP, Int. J. Refrig. 100 (2019) 284–294. [16] B. Yu, J. Yang, D. Wang, J. Shi, J. Chen, Modeling and theoretical analysis of a CO2propane autocascade heat pump for electrical vehicle heating, Int. J. Refrig. 95 (2018) 146–155. [17] B. Zühlsdorf, J.K. Jensen, B. Elmegaard, Heat pump working fluid selection—economic and thermodynamic comparison of criteria and boundary conditions, Int. J. Refrig. 98 (2019) 500–513. [18] L. Zhang, K. Hashimoto, H. Hasegawa, M. Saikawa, Performance analysis of a heat pump system with integrated desiccant for electric vehicles, Int. J. Refrig. 86 (2018) 154–162. [19] L. Pan, C. Liu, Z. Zhang, T. Wang, J. Shi, J. Chen, Energy-saving effect of utilizing

[23] [24] [25] [26] [27] [28] [29] [30] [31] [32] [33] [34] [35] [36] [37] [38] [39] [40]

[41]

11

recirculated air in electric vehicle air conditioning system, Int. J. Refrig. 102 (2019) 122–129. D.-Y. Lee, C.-W. Cho, J.-P. Won, Y.C. Park, M.-Y. Lee, Performance characteristics of mobile heat pump for a large passenger electric vehicle, Appl. Therm. Eng. 50 (2013) 660–669. J.H. Ahn, H. Kang, H.S. Lee, H.W. Jung, C. Baek, Y. Kim, Heating performance characteristics of a dual source heat pump using air and waste heat in electric vehicles, Appl. Energ. 119 (2014) 1–9. J.H. Ahn, J.S. Lee, C. Baek, Y. Kim, Performance improvement of a dehumidifying heat pump using an additional waste heat source in electric vehicles with low occupancy, Energy 115 (2016) 67–75. Z. Tian, W. Gan, X. Zhang, B. Gu, L. Yang, Investigation on an integrated thermal management system with battery cooling and motor waste heat recovery for electric vehicle, Appl. Therm. Eng. 136 (2018) 16–27. H.S. Hamut, I. Dincer, G.F. Naterer, Exergy analysis of a TMS thermal management system) for range-extended EVs (electric vehicles), Energy 46 (2012) 117–125. H.S. Hamut, I. Dincer, G.F. Naterer, Analysis and optimization of hybrid electric vehicle thermal management systems, J. Power Sources 247 (2014) 643–654. N. Javani, I. Dincer, G.F. Naterer, B.S. Yilbas, Exergy analysis and optimization of a thermal management system with phase change material for hybrid electric vehicles, Appl. Therm. Eng. 64 (2014) 471–482. Z. Tian, B. Gu, Analyses of an integrated thermal management system for electric vehicles, Int. J. Energ. Res. 43 (2019) 5788–5802. A.S. Lavine, F. Incropera, D.P. DeWitt, Fundamentals of Heat and Mass Transfer, seventh ed., John Wiley & Sons, Inc., 2011. Z. Tian, B. Gu, L. Yang, F. Liu, Performance prediction for a parallel flow condenser based on artificial neural network, Appl. Therm. Eng. 63 (2014) 459–467. C.Y. Yang, R.L. Webb, Condensation of R-12 in small hydraulic diameter extruded aluminum tubes with and without micro-fins, Int. J. Heat Mass Transf. 39 (1996) 791–800. Z. Tian, L. Ma, B. Gu, L. Yang, F. Liu, Numerical model of a parallel flow minichannel evaporator with new flow boiling heat transfer correlation, Int. J. Refrig. 63 (2016) 1–13. S. Kakaç, H. Liu, Heat Exchangers : Selection, Rating, and Thermal Design, CRC Press, 2012. A.A. Hasan, D.Y. Goswami, S. Vijayaraghavan, First and second law analysis of a new power and refrigeration thermodynamic cycle using a solar heat source, Sol. Energy 73 (2002) 385–393. E.A.M. Falcão, A.C.R. Teixeira, J.R. Sodré, Analysis of CO2 emissions and technoeconomic feasibility of an electric commercial vehicle, Appl. Energ. 193 (2017) 297–307. R.J.J.W. Smith, Chemical Process Design and Integration, 2005. L. Garousi Farshi, S.M.S. Mahmoudi, M.A. Rosen, M. Yari, M. Amidpour, Exergoeconomic analysis of double effect absorption refrigeration systems, Energy Convers. Manage. 65 (2013) 13–25. S. Quoilin, S. Declaye, B.F. Tchanche, V. Lemort, Thermo-economic optimization of waste heat recovery organic rankine cycles, Appl. Therm. Eng. 31 (2011) 2885–2893. C. Yue, F. You, Y. Huang, Thermal and economic analysis of an energy system of an ORC coupled with vehicle air conditioning, Int. J. Refrig. 64 (2016) 152–167. A. Farsi, S.M.H. Mohammadi, M. Ameri, An efficient combination of transcritical CO2 refrigeration and multi-effect desalination: energy and economic analysis, Energy Convers. Manage. 127 (2016) 561–575. A.H. Mosaffa, L.G. Farshi, C.A. Infante Ferreira, M.A. Rosen, Exergoeconomic and environmental analyses of CO2/NH3 cascade refrigeration systems equipped with different types of flash tank intercoolers, Energy Convers. Manage. 117 (2016) 442–453. E.W. Lemmon, M.L. Huber, M.O. Mclinden, NIST Standard Reference Database 23: Reference Fluid Thermodynamic and Transport Properties-REFPROP, Version 9.1, Gaithersburg: National Institute of Standards and Technology, 2010.