Thermal and economic analysis of an energy system of an ORC coupled with vehicle air conditioning

Accepted Manuscript Title: Thermal and economic analysis of an energy system of an ORC coupled with vehicle air conditioning Author: Chen Yue, Fengqi You, Ying Huang PII: DOI: Reference:

S0140-7007(16)00011-6 http://dx.doi.org/doi: 10.1016/j.ijrefrig.2016.01.005 JIJR 3234

To appear in:

International Journal of Refrigeration

Received date: Revised date: Accepted date:

25-3-2015 4-1-2016 6-1-2016

Please cite this article as: Chen Yue, Fengqi You, Ying Huang, Thermal and economic analysis of an energy system of an ORC coupled with vehicle air conditioning, International Journal of Refrigeration (2016), http://dx.doi.org/doi: 10.1016/j.ijrefrig.2016.01.005. This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting proof before it is published in its final form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.

Thermal and economic analysis of an energy system of an ORC coupled with vehicle air conditioning Chen Yue1,*, Fengqi You2, Ying Huang1 1. Jiangsu Province Key Laboratory of Aerospace Power Systems, Nanjing University of Aeronautics and Astronautics, Nanjing 210016, China 2. Department of Chemical and Biological Engineering, Northwestern University Evanston, IL 60201, USA

Highlights: -A novel vehicle energy system with exhaust heat recovery is proposed; -Thermal and mass integration of the bottoming ORC and VRC subsystem. - Performance improvement of the VCR subsystem. -Decrease of the investment of the vehicle air conditioning is investigated.

Abstract: We propose a novel vehicle energy system (VES), in which the integration of the organic Rankine cycle (ORC) and the vehicle air conditioning are considered. The thermal and economic performance of the system is investigated through comparing with a conventional VES

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and a VES with standalone ORC exhaust heat recovery. Hydrocarbons, including n-propane and cyclopentane, and hydro-fluorocarbons, including R134a and R245fa, are selected as the working fluids of the energy system. We systematically study influences from integration of (vapor compression refrigeration) VCR and ORC subsystem under different operating conditions. The results indicate that the proposed system holds advantages regarding the overall thermal and economic performance compared with the conventional VES. The proposed system improves the overall thermal efficiency by 9.2-9.8%, and reduces diesel consumption by 1.1-1.3 L.h-1 compared with the conventional VES. The different working fluids considered lead to payback periods in the range of 4,818-5,427 h, R134a is the most efficient working fluid for the integrated VCR-ORC subsystem of the VES under the operating conditions considered in this research. Keywords: exhaust heat recovery, thermal and economic performance, internal combustion engine (ICE), organic Rankine cycle (ORC), vehicle air conditioning. Nomenclature A

Area, m2

Subscripts

Cs

Fuel cost saved, RMB Yuan

0,1,2…

State point

Er

Exchange ratio

a

Air

h

Specific enthalpy, kJ.kg-1

c

Condensation

l

Length, m

CC

VCES

LMTD

Log mean temperature difference, cl

Cooling heat of engine

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K m

Mass flow rate, kg.s-1

disp

Displacement volume

n

Rotational speed, PRM

e

Evaporation

OIC

Overall investment cost, RMB f

Fuel

Yuan p

Pressure, MPa

g

Exhaust gases

P

Power, kW

ICE

Internal

combustion

engine P1

Power output of the conventional Labor

Labor cost

system, kW P2

Power output of the proposed LHV

Lower heating value

system, kW PP

Payback period, h

loss

Energy loss

Pr

Price of fuel, RMB Yuan.L-1

ORC

Organic Rankine cycle

Q1

Overall heating value of fuel, kW

p

Pressure/pump

Q2

Waste exhaust heat, kW

R

Refrigerant

Q3

Refrigeration capacity, kW

REF

Refrigeration cycle

Q4

Discharged heat of the condenser, t

Turbine

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kW Q5

Exhaust waste heat recovered, kW tot

Total

heat

transfer

efficiency T

Temperature, K

v

TCC

Total components cost, RMB vol

Volume, vapor Volume efficiency

Yuan u

Internal energy, kJ.(kg.K)-1

V

Volume flow rate, m3.h-1

Vol

Volume of liquid

Vs

Diesel saved, L.h-1

Symbols β

Expand ratio during isentropic expansion stroke

δ

diesel consumption per kg load

η

Efficiency, %

Abbreviations

ΔFp

diesel penalty

A/C

Air conditioning

Δη

Efficiency difference, %

ICE

Internal

combustion

engine

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△p

Pressure difference, MPa

NBP

Normal boiling point

ρ

Density, kg.m-3

ORC

Organic Rankine cycle

τ

Number of crankshaft

VCR

Vapor

compression

refrigeration φ

percentage of the heat exchanger VES

Vehicle energy system

in the ORC subsystem

1 Introduction The air conditioning (A/C) subsystem of a vehicle, which consumes mechanical power to supply the necessary cooling/heating for the cabin space of a modern vehicle, is an important unit of the vehicle’s energy conversion system [1]. The vehicle A/C works at two modes, including the summer mode and winter mode. At summer mode, the vapor compression refrigerant (VCR) cycle is a popular approach used in the conventional vehicle A/C subsystem. At winter model, the waste heat of the engine coolant or exhaust is used to supply heat for the vehicle cabin with little influence on the overall energy efficiency of the engine power system [2]. However, at summer mode, a large amount of mechanical power is consumed to drive the compressor of the VCR subsystem, and the power consumption of the VCR subsystem increases with the increase of the ambient temperature [3]. It is reported the average thermal efficiency of the vehicle’s engine is only approximately 1/3, and the overall diesel consumption of the engine will increase by 7%-38% when the vehicle’s A/C is operated [4].

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Both increasing liquid fuel consumption and the implementation of stringent environmental emission regulations have become key driving forces in the search for sustainable and economically viable technologies for efficient and clean implementation of vehicle engine energy conversion systems in recent years. To improve the thermal and economic performance of the vehicle A/C subsystem, there are three effective approaches, namely, increasing the coefficient of performance (COP) of the VCR cycle, increasing the energy conversion efficiency of the engine through waste heat recovery, and adapting advanced energy management technologies. Cho et al. [5] assessed the COP of the VCR systems of the same vehicle using the refrigerants of R134a and R123yf. Their simulation results indicated that a system using R123yf exhibited a better cooling capacity and COP compared to those achieved using R134a. Han et al. [6] proposed a mixture of R161/R134a (0.6/0.4 in mass fraction), named M5, to be used as a substitute for R134a used in vehicle VCR systems. Theoretical and experimental studies on the VCR system performances of M5 and R134a were conducted. The experimental results indicated that the COP of M5 was 0.1 higher than that of R134a. Alqdah [7] presented an aqua-ammonia absorption refrigeration system for the A/C system of a vehicle using the exhaust waste heat of the diesel engine. Al-Aqeeli et al. [8] studied the design of a vehicle A/C system using an open cycle absorption system with LiBr-H2O as the working fluid. Shah et al. [9] proposed a model for utilizing the exhaust waste heat to run the A/C system of a vehicle. They considered three fluids (ammonia gas, hydrogen gas and water) for a vapor absorption system of a four-stroke, four-cylinder passenger car. Lorentzen et al. [10] investigated the performance of the A/C system of a vehicle and demonstrated that an efficient and environmentally benign CO2 system performed comparably to a R12 system. Crepinsek et al. [11] conducted a performance comparison of absorption refrigeration cycles for different working fluids and recommended the

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suitable conditions of heat source temperature and refrigeration temperature for the different working fluids considered. Koehler et al. [12] studied the performance of an absorption cycle for a vehicle energy system (VES) with exhaust waste heat recovery. Sharafian et al. [13] assessed various adsorption refrigeration technologies and recommended an adsorbed bed style for the A/C system of a vehicle. Olsen et al. [14] proposed a high-performance magnetic refrigeration scheme for a vehicle. Miranda et al. [15] proposed a vehicle cooling/heating system using thermoelectric chip technology based on the exhaust heat recovery. Among these vehicle engine waste heat recovery technologies, the organic Rankine cycle (ORC) technology exhibits great potential and has been the subject of a number of studies [16,17]. Khayyam [18] conducted an analysis of adaptive intelligent control of a vehicle A/C system, and the results indicated that the case of using an intelligent control system was more energy efficient compared with a fuzzycontroller based vehicle A/C system. Moreover, some studies on the integration of the ORC with different subsystems have been performed to increase the energy of different energy systems [19-21]. Although there have been many investigations associated with improving the performance of the vehicle A/C subsystem, there are few existing studies that consider in detail the integration of the conventional A/C subsystem and the engine waste heat recovery of a vehicle. In this paper, We propose a novel vehicle energy system scheme, in which the thermal and mass integration of the bottoming ORC subsystem and the VCR subsystem is considered to improve the thermodynamic performance of the VES, and a common condenser is used to integrate the VCR subsystem and ORC subsystem, and the same type of working fluid is used for the two subsystems. In addition, above feature can decrease the investment capital of VES. Overall, the thermal and economic performance of the proposed VES is preliminarily studied,

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and the influence of the coupled condensation temperature, refrigeration evaporation temperature, and ORC evaporation temperature on the overall thermal and economic performance of the proposed system is analyzed for several organic working fluids and compared with a conventional VES. 2 System description Fig. 1 shows a conventional VES. As shown in Fig. 1(a), the system can be divided into two parts: an internal combustion engine (ICE) subsystem and a VCR subsystem. In the ICE subsystem, ambient air (0) is compressed by the turbocharger compressor (0-1) and then cooled in the intercooler (1-2), after which it flows through the intake to the suction end of the ICE, with a fictitious steady-state diesel thermodynamic cycle assumed for the working processes from point 2 to point 5. The exhaust gases (5) are expanded in the turbine of the turbocharger (5-6), which in turn delivers mechanical power to drive the turbocharger compressor. Next, the exhaust gases (6) exiting the turbine are sent to a catalytic purification unit to lower the emissions, such as NOx. Finally, the exhaust passes through a silencer device that reduces the exhaust noise before being expelled into the atmosphere. Because the temperature and flow rate of the exhaust gases are high, there is a large potential to save some thermal energy from the exhaust gases. Two working modes, including the summer mode and winter mode, are involved for the VES scheme shown in Fig.1. At winter mode, some of the warm air from the radiator is used to supply heating, and only a minimal amount of ventilation power is consumed. At summer mode, the gaseous refrigerant from the evaporator is compressed through a compressor (7-8) and cooled through a condenser (8-9) by air, after which it flows into a liquid tank. The temperature of the liquid refrigerant is lowered through a throttle valve (9-10), and then the lower temperature

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refrigerant is sent to an evaporator to cool the air (10-7). Thus, overall energy consumption of the vehicle’s VRC refrigeration process is high due to the high power consumption of the compressor. The utilization factors of the components of the VCR subsystem are low when the utilization time of the VCR subsystem is short. A novel VES with a waste heat recovery ORC, shown in Fig. 2, includes three subsystems: a topping ICE subsystem, a VCR subsystem and a bottoming ORC subsystem. At summer mode, the working processes of the ICE subsystem and VCR subsystem in this system are the same as the conventional VES shown in Fig. 1. A major difference is that an ORC subsystem is used as the bottoming cycle of the topping ICE to recover the exhaust waste heat. Working processes of the proposed bottoming ORC subsystem are as follows. Working fluid from the liquid tank is pumped (9-11) into an evaporator by a pump in the ORC and is then evaporated in an evaporator (11-12) via recovering the sensible heat from the exhaust gases (6-14). The high-pressure vaporous working fluid from the evaporator expands through a turbine (12-13), which can in turn supply some mechanical power. The cooled working fluid exiting the turbine is condensed to liquid in the condenser (13-9). Fig. 2(a) also shows that a condenser and liquid tank are used as common components of both the bottoming ORC and the conventional vehicle’s VCR subsystem, and the working fluids used in the VCR subsystem are the same as that of the bottoming ORC. Only the pump, evaporator and turbine are added in the ORC subsystem. It is expected that the overall size and cost of the ORC subsystem will be low and that the utilization time of the components of the A/C system could be improved at winter mode. In addition, confined by the maximum heat transfer capacity of the condenser in the VCR subsystem, the ORC subsystem in the proposed system can operate (when the ICE is operating) when no cooling demand is present.

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As shown in Fig.3, an extra ORC subsystem is added to recover the exhaust waste heat from the conventional vehicle ICE exhaust, which can operate (including in the refrigeration and nonrefrigeration seasons) when the ICE is in operating. There exists two working modes, including the summer and winter mode for the two vehicle energy systems shown in Fig.3, while compared with the scheme shown in Fig.2, the two modes for scheme 3 in shown in Fig.3 are completely different. For the scheme shown in Fig.2, two gate valves are used to switch working mode. At summer mode, open the valve 1 and valve 2, and amount of external power is consumed to drive the compressor in the VCR subsystem and pump in the ORC subsystem, the compressed working fluid from the VCR subsystem is mixed with the expanded working fluid from the ORC subsystem, and the flows No.8 is obtained. The condenser and liquid tank are used as common components of the VCR subsystem and the ORC subsystem in summer mode. At winter mode, open valve 1 and close valve 2, no extra external power is consumed to drive compressor of the VCR subsystem, and only the ORC subsystem works, and the key components of the condenser and liquid tank in the VCR subsystem work also, which means the component utilization factors of the condenser and liquid tank in the VCR subsystem are high, compared with scheme 3 shown in Fig.2. Above measures are expected to improve thermodynamic performance of the conventional VES and decrease the investment capital of bottoming ORC subsystem. Two single-direction valves includes the SD valve 1 and SD valve 2 are added in the VCR and ORC subsystems for scheme shown in Fig.2 to avoid the reverse flow happening. For the scheme shown in Fig.3, the ORC subsystem and the VCR subsystem work independently. The bottoming ORC subsystem can work at any mode when topping ICE

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subsystem works. The VCR subsystem only works at summer mode, and utilization factors of the components in the VCR subsystem are not influenced by the winter working time obviously. The selection of the working fluids has been found to be of considerable importance in the ORC subsystem and the VCR subsystem [5,10,22]. The thermal physical properties of the working fluids affect not only the thermodynamic performance but also the overall economic performance [23]. Hundreds of organic fluids are commercially available, but only a few of them fit the proposed energy system. Four working fluids are selected based on their critical temperature, and their physical properties are presented in Table 2 according to the literatures [22-23]. With respect to the critical temperature of the four working fluids in Table 2, a subcritical ORC is considered in this study. Note that the current study intends to focus on the thermal and economic performance of the VES, rather than accurate fluids selection optimization for the system. 3 Mathematical model 3.1 Problem statement This paper focuses on the thermal and economic performance of the proposed vehicle energy system considering the influences of the evaporator size and operational conditions. Three mathematical models are developed to establish several configurations, including the conventional ICE+VCR, ICE+VCR +standalone ORC and the ICE+ integrated VCR-ORC models. The flow diagram calculations of the several VES models are shown in Fig. 4. Fig. 4(a) represents the conventional ICE+VCR model, Fig. 4(b) gives ICE+VCR and the standalone ORC model, and Fig.4(c) shows the proposed ICE+ integrated VCR-ORC system model.

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Because the calculation procedure for the standalone ICE subsystem model is similar in the three VES models, the details of the model calculation for the standalone ICE subsystem are not introduced in Fig. 4(b) and Fig.4(c). First, the model shown in Fig. 4(a) is used to calculate the thermodynamic performance of the conventional ICE+VCR system. The simulation problem assumes the ambient conditions, air flow rate, fuel flow rate, the size of the cylinders, the clearance of the cylinder, intake pressure and exhaust gas temperature are known; The control variables of the model are expanding ratio in isentropic expand stroke, β; An initial value of the expand ratio in the constant-volume combustion stroke, β is assigned. Power output, PICE of the ICE system is calculated using the mathematical model. Then the A/C load, Q3, structure parameters of evaporator and condenser in VCR are given, the compression work consumption, PREF of the VCR subsystem is calculated. And then COP of the VCR is verified. At last, the overall thermal performance of the VES is calculated. The comparison model, shown in Fig. 4(b), is used to calculate the ICE+VCR+ standalone ORC model. Based on the ICE system exhaust results obtained from the calculation procedure shown in Fig.4 (a), sizes and heat-transfer coefficients for evaporator and condenser of ORC subsystem are given, the heat recovered from the exhaust and heat discharged to ambient through condenser are calculated according to the flow patterns and operational states of the exhaust gas and the organic working fluid. Overall thermal and economic performances of the overall VES are investigated through using different pure working fluid. Fig. 4(c) is used to calculate the overall thermal and economic performance of the proposed VES. Calculation procedure of Fig. 4(c) is similar to the Fig.4 (b). Thermal performance

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calculation model of Fig.4(c) is same as Fig.4 (b). The main difference lies in economic performance calculation model, since the condenser and liquid tank of the VCR subsystem is used as the common component of ORC subsystem, thus the component cost investment of the condense and liquid is removed in the economic performance calculation model for the ORC subsystem. 3.2 Calculation model 3.2.1 Assumptions To simplify and clarify the analysis, the following assumptions are made: - Steady-state conditions are assumed for all the components; - The heat and friction losses in the system are omitted; -The exhaust and ambient air are treated as ideal gases; -The kinetic and potential energy changes in the system are negligible; -A subcritical turbine is used in the ORC subsystem. 3.2.2 Vehicle top ICE subsystem The energy balance of the ICE subsystem can be expressed as PICE  PREF  Q 2  Q cooling  m fuel  Q LHV

(1)

Where mfuel is the mass flow rate of diesel, and QLHV is the lower heating value of diesel, 44 MJ/kg [24]. Pref is the power consumption of the VCR subsystem, Qcooling is the engine heat

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taken away by the intercooler, cooling lubricated oil, and cylinder coolant. Q2 is waste heat of the exhaust gases exiting the turbocharger. Power output of the ICE subsystem is calculated by

PICE 

MEP  V disp  n



(2)

The τ term is introduced by Heywood [25] to account for the number of crankshaft revolutions per power stroke, where, for a two-stroke cycle, τ is 1 and for a four-stroke cycle, τ is 2. n represents the rotational speed of the engine in revolutions per unit time. The mean effective pressure (MEP) is a fictitious constant pressure acting on the piston during the power stroke that generates the net work in the cycle (Wnet) as the piston sweeps out the displacement volume in a single expansion (Vdisp). MEP is calculated by

M EP 

W net V disp

(3)

The Wnet term in MEP is split into various work terms corresponding to each stroke. Thus, W n et   W com p ression  W com b u stion  W ex p an sion  abs (W ex h au st  W in take )

(4)

The work consumed in the adiabatic compression stroke (2-3) is W compressio

n

  0  V disp  u 3  u 2 

(5)

For the constant-volume combustion stroke in this study, Wcombustion is zero. The work performed by the cylinder gases on the piston during exhaust is W exhaust  p 5  (V disp  V cl )

(6)

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where Vdisp is the cylinder volume, and Vcl is the clearance volume of the cylinder. The work performed by the gases in the cylinder on the piston during intake is W in take  p 2  (V d isp  V cl )

(7)

The overall pressure loss during intake and exhaust (caused by pumping the gases) in the ICE thermodynamic cycle is calculated by

 p pum p  p1  p 5 

W exhaust  W intake V disp  V cl

(8)

For the combustion process within the cylinders, the average chemical formula for common diesel is taken to be C12H23 [24]. Therefore, the combustion process of the diesel with air in the cylinders can be expressed simply as x C 12 H 23  y ( O 2  3 . 75 N 2 )  12 x CO 2  11 . 5 x H 2 O  3 . 75 y N 2  ( y  17 . 75 x ) O 2

(9)

Where x is the molar flow rate of diesel, and y is the molar flow rate of air. The molar flow rates of diesel and air are calculated by

y 

ma M

x 

m fu e l M

(10)

a

(11)

fu e l

m g  m a  m fuel

(12)

The energy balance during the constant-volume combustion stroke (3-4) can be expressed as [25]

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m g  ( u 4  u 3 )  m fuel  Q LHV

(13)

The next expression represents the constant-volume combustion stroke (3-4)

p4 

M

a

 R a  T4

M

g

 R a  T3

 p3

(14)

The work performed by the combustion gases in the cylinder during the expansion stroke (4-5) is

W expansion 

V disp

0

 (u 4  u 5 )

(15)

The expand ratio of the cylinder gas, β, during the isentropic expansion process is calculated by

 

p4 p5

(16)

In the conventional VES shown in Fig. 1, the waste heat of the exhaust emitted to the environment is Q 2  m g  c p , g  (T6  T0 )

(17)

In the VES with waste heat recovery shown in Fig. 2 and Fig. 3, the waste heat recovered in evaporator 2, Q5, is calculated by Q 5  m g  c p , g  ( T 6  T1 4 )

(18)

In the VES with waste heat recovery shown in Fig. 2 and Fig. 3, the exhaust heat emitted to environment, Q2, is given by

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Q 2  m g  c p , g  ( T14  T 0 )

(19)

3.2.3 ORC subsystem For the waste heat recovery process of the evaporator in the ORC subsystem, the energy balance during the heat transfer process is expressed as Q 5  m R ,O R C  ( h1 2  h1 1 )

(20)

For an evaporator with a given size, the heat transfer process can also be expressed as Q 5  h tot  A  LM TD

(21)

in which the log mean temperature difference (LMTD) in the evaporator is given by

LMTD 

(T 6  T12 )  (T14  T11 ) ln

(T 6  T12 )

(22)

(T14  T11 )

The power output of the ORC subsystem is calculated by PO R C  m R ,O R C  w O R C  Pt,O R C  Pp,O R C

(23)

The power output of the turbine in the ORC subsystem is calculated by Pt,ORC   t  m R, ORC  h12  h13 

(24)

in which ηt is the isentropic efficiency of the turbine. Pp,ORC is the power consumption of the pump, which is calculated by Pp ,O R C  m R ,O R C  ( h1 1  h 9 ) /  p

(25)

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3.2.4 Vehicle’s VCR subsystem The compression work consumption of the VCR subsystem is calculated by PR E F  m R ,R E F  w R E F

(26)

In which, wREF is the specific work consumption and is calculated by w REF  ( h8  h7 )

(27)

The refrigeration capacity of the evaporation process is calculated by Q 3  m R ,R E F  ( h 7  h10 )

(28)

The coefficient of performance (COP) of the refrigeration cycle is calculated by

COP 

Q3

(29)

PR E F

3.2.5 Overall thermodynamic performance indices The formula for the ICE effective thermal efficiency is given as

 ICE 

P1



Q1

PICE  PREF

100 %

Q1

(30)

The thermal efficiency of the conventional VES is given as  CC1   ICE

(31)

The thermal efficiency of the ORC subsystem is

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 ORC 

PORC

100 %

Q5

(32)

The overall thermal efficiency of the proposed VES is given as

 CC2 

P2



P1  PORC

Q1

100 %

Q1

(33)

Thus, the thermal efficiency improvement of the proposed VES is

  CC 

 CC2   CC1  CC1

100 %

(34)

The diesel saved per hour for the proposed VES is expressed by

Vs 

3 . 6   P2  P1  /  CC2 Q LHV   fuel

(35)

in which, ρfuel is the liquid diesel density in kg/L. 3.2.6 Economic performance model The purpose of this section is to propose a primary economic performance model. The comparative analysis approach is used to investigate the economic performance of the two schemes shown in Fig.2 and Fig.3, and the scheme shown in Fig.1 is used as the comparison base. The ICE subsystem in three schemes shown in Fig.1 to Fig.3 possesses the same working condition at any working modes (summer/winter), and the main difference lies the VCR and ORC subsystem at different working modes for scheme 2 and scheme 3. Comparing the working modes for the two schemes shown in Fig.2 and Fig.3, it is known all the components in the VCR and ORC subsystem work at summer mode for the two schemes

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shown in Fig.2 and Fig.3, the main difference lies the utilization factors of the condenser and liquid tank, which increases for the proposed VES scheme in winter mode, and only the topping ICE and bottoming ORC subsystem work. Thus, a simplified economic performance analysis is conducted, and the economic performance calculations of the ORC subsystem are taken into account. To arrive at an economic performance analysis, it is necessary to define several indices in this section. The overall investment cost of the ORC subsystem is calculated by [28] OIC  C labor  TCC

(36)

Where Clabor is the labor cost, equaling to 0.3TCC in this research [28]. TCC is the total component cost of the ORC subsystem. In this research, considering the localization of the ORC subsystem, the overall investment cost of the bottoming ORC subsystem in China is calculated by OIC  f  Er  C labor  TCC



(37)

In which f is the local factor. f refers the components cost factor at some certain district, decided by the production level and material cost in that district. f reflects the cost difference for the same component at different production district. Er is the exchange ratio of the Euro Dollar to the RMB Yuan, set at 8.41 [30]. The basic economic data of this calculation model refers to a Euro-based study [29]. Note that impact of the exchange ratio and the price fluctuation of diesel is not considered in this study. Under the same mechanical driving power requirement for different schemes, economic performances of the scheme 2 and scheme 3 are investigated. Since the use of bottoming exhaust

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waste heat recovery ORC, thermal efficiency of the VES is expected to be improved. The save energy cost of the ORC subsystem is converted to the cost of the saved diesel fuel in this research. However, when introducing the bottoming ORC subsystem compared to the scheme shown in Fig.1, the overall weight of vehicle increases, resulting in the increase of diesel fuel. Thus, overall diesel cost saved each hour is calculated by Cs  (Vs   F p )  Pr fuel

(38)

in which, Vs is the diesel saved, L.h-1. ΔFp is the diesel penalty caused by weight increase due to the use of the vehicle’s ORC subsystem. Since the heat-exchanger accounts for a big weight fraction in the ORC subsystem, and diesel penalty, ΔFp caused by the ORC subsystem is calculated by  F p  m Ex  A /   

(39)

in which mEx is the weight per unit heat transfer area, kg.m-2; A is the heat exchanger area, m2; φ is the percentage of the heat exchanger in the ORC subsystem; δ is the diesel consumption per kg load, L.kg-1. Prfuel is the price of diesel, which is taken as 7.74 RMB Yuan.L-1 in this study [29]. The payback period is calculated by PP  OIC / Cs

(40)

The main system components considered in this research are presented in Table 1. The corresponding costs and calculation expressions are also presented according to the study of Quoilin et al. [28]. 4 Results and discussion

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4.1 Solving method In this paper, three vehicle energy system models are built. Through comparing with two VES schemes shown in Fig.1 and Fig.3 under the same working conditions, the thermal and economic performance analysis is conducted for the proposed vehicle ORC-VRC system shown in Fig.2. Thermal and economic performance of the proposed system scheme is demonstrated through comparing with the scheme shown in Fig.1. The economic performance improvement of the proposed system is studied through comparing with the scheme shown in Fig.3. The Aspen plus V7.1 simulation platform via the PENG-ROM physical properties method, which is similar to the analysis tool used in [31], is used to simulate performances of the several VES schemes. Based on the plate evaporator and condenser models build in Aspen Exchanger Design & Rating V7.1 and an excel user self-defined ICE work processes model, thermal and economic performance of the proposed VES is discussed. Based on the four working fluids shown in Table 2, thermal and economic performances of the proposed system are discussed in this section through analyzing the impaction from the three operation variables: the evaporation temperature of the VCR subsystem, Te,REF, the condensing temperature of the VCR subsystem, Tc, and the evaporation temperature of ORC subsystem, Te_ORC. The input parameters and boundary conditions are presented in Table 3. The minimal evaporation temperature of the ORC subsystem was set to 360 K, which corresponds to the heat source temperature. The temperature of the ICE exhaust gases is at approximately 850 K. According to equation (7), much of the exhaust waste heat will be recovered under the low evaporator temperature and fixed heat transfer area. Yet, the temperature of the exhaust gases

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exiting the evaporator is reduced to the dew point of some of the acidic components in the gases, possibly causing corrosion problems, which is not discussed in detail in this study. The corrosion problem is the focus of future work. 4.2 Overall thermal and economic performances Table 4 presents the overall thermal-economic performance indices of different VES types. All the results are obtained based on the ambient temperature of T0 at 303 K. The operation time of the two VES schemes only refers the overall working time in winter mode in this research. Table 4 shows that thermal efficiency of the bottom ORC subsystem is of 9.7% for the scheme 2 and scheme 3, and overall thermal efficiency of the ICE is improved by 9.8% when using R245fa as working fluid and at evaporating temperature of 360 K. The research of Tian et al. [32] shows the thermal efficiency is in range of 8.0%-10.0%, when using the same working fluid and at the same evaporating temperature. Thermal efficiency of ORC subsystem in range of 9.8%-10.7% is found when using the same working fluid in literature [26]. Literature [26] also shows the overall thermal efficiency improvement of the ICE reach 10.54% at rated condition. Compared the above literature results with our investigation, it is known the thermal performance calculation results of the bottoming ORC subsystem in this research are very close to above studies, which indirectly verified the precision of the thermal performance calculation model in this study. Base on thermal performance improvement from the ORC subsystem, it can be seen from column 2 and 3 in Table 4, the thermal efficiency of the VES is increased by 9.8% through using

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the bottoming ORC exhaust heat recovery subsystem, and the fuel saved is of 1.3 L.h-1, compared with the conventional VES without exhaust heat recovery. The specific investment cost (SIC) of the ORC subsystem in Fig.3 is verified in this section. According to Fig.5 in [28], for a 5 kW scale ORC, the SIC is about 2136-4260 €/kW when using different working fluids, while according to the literature [34], SIC for a 10 kW scale ORC is of 7500 €/kW, which is much higher than the data given in [28]. Thus, it is difficult to verify the economic performance results in this research due to the kW scale ORC systems have not been put into commercial application and no a standard market price can be used as reference. Table 4 shows the payback period of the referenced VES is of 7,520 h, which is increased by 56% compared with the proposed VES due to the condenser accounts for a big fraction in overall capital investment. Table 5 lists the overall performance indices of the proposed VES with different working fluids; these results are obtained based on a power output requirement of 168.5 kW and a refrigeration capacity requirement of 30 kW. The overall thermal efficiency of the proposed VES is found to be 37.8-38.1%, which represents an improvement of the overall thermal efficiency by 9.2-9.8%, a fuel saving of 1.1-1.3 L.h-1, and a payback period in the range of 4,818-5,420 h. The use of R134a results in the best thermodynamic and economic performance, and Pentane is the next-best-performing fluid. The use of R245fa results in the worst overall thermal and economic performance. 4.3 Key operation parameter impact analysis 4.3.1 Evaporation temperature of the VCR subsystem (Te_REF)

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Fig. 5 and Fig. 6 show the influence of the evaporation temperature, Te_REF, on COP of the VCR subsystem and the overall thermal efficiency of the proposed VES. The VES using cyclopentane exhibits an advantage in refrigeration performance (a higher COP), but the VES using R134a exhibits an advantage on the overall thermal performance. Fig. 5 shows that COP of the vehicle’s VCR subsystem increases as Te_REF is increased. The subsystem using cyclopentane exhibits prominent thermal performance advantages due to its lowest compression ratio through the refrigeration compressor of the four working fluids considered. It can also be seen from Fig. 5 that COP of the vehicle’s VCR subsystems using R134a is very near to that of R245fa and much lower than that of the other two working fluids. This result is due to the similarity of the compression ratios of the two working fluids (R134a and R245fa), which are much higher than those of the other two fluids (pentane and cyclopentane) under the operation conditions explored in this research. Fig. 6 illustrates the influence of Te_REF on the overall thermal efficiency of the VES. Fig. 5 indicates that the refrigeration performance of the VCR subsystem using R134a is a slightly worse than those using the other three working fluids, while the power output of the ORC subsystem (Table 5) using R134a is the highest value. Therefore, as shown in Fig.6, the overall thermal efficiency of the VES using R134a is the highest in this study. Fig. 6 also shows that the thermodynamic performance of the refrigeration subsystem using R245fa is close to that of using R134a, but the power output and thermal efficiency of the ORC subsystem using R245fa is much lower due to a lower expansion ratio in the turbine. Therefore, the scheme using R134a as the working fluid illustrates prominent advantages on the overall thermodynamic performance. It also is known from above results, working fluids shows good refrigeration performance are not

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necessary for the optimal selection of ORC system, the detailed study on working fluids selection of the VCR-ORC system should be conducted in the future. 4.2.2 Condensation temperature (Tc) Fig. 7 to Fig. 13 reveal the influences of Tc, on the refrigeration power consumption, COP of VCR, overall thermal efficiency, diesel saved and payback period of the proposed system. Fig. 8 and Fig. 10 in particular indicate that the thermodynamic performance of both the VCR subsystem and the ORC subsystem become worse as the condensation temperature rises. Fig. 8 shows the VCR subsystem using cyclopentane demonstrates a prominent advantage in the refrigeration performance, which is similar as results shown in Fig.5. This advantage exists because the compression ratio of the VCR subsystem is low in the cyclopentane system. Fig. 9 and Fig. 11 indicate that the system using R134a exhibits prominent thermodynamic performance advantages in both the ORC subsystem and the VES system due to the high COP of VCR subsystem. Fig. 9 and Fig. 10 show the effects of the condensation temperature on the ORC subsystem. Both the power output and the thermal efficiency of the ORC using R134a are the highest in the VES among the four working fluids under the same temperature difference between the evaporation temperature and the condensation temperature (Te_ORC-Tc). The pressure difference (pe_ORC-pc) of R134 is the highest among the four working fluids at the same saturated temperature difference between the evaporation temperature and the condensation temperature (Te_ORC-Tc), which explains these results. Fig.9 also shows PORC using R134a decreases steeply as Tc is increased, compared with the other three working fluids, due to the remarkable increase of compression ratio of compressor at high condensing temperature. PORC using R134a is lower

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than that using Cyclopentane with Tc over 314 K, and almost no advantage on thermal performance of the ORC system is shown with Tc over 318 K. Fig. 11 shows the relationship between the overall thermal efficiency and the condensation temperature. The overall thermal efficiency is found to decrease with increasing condensation temperature. The system using R134a as a working fluid holds a thermodynamic performance advantage when the condensation temperature is lower than 312 K. Fig. 12 demonstrates relationship between the diesel saved and the condensation temperature of Tc. The diesel saved in all the systems decreases as increase of condensation temperature is increased, and it is evident for the scheme using R134a as working fluids. The scheme using R134a as working fluid can save much more diesel compared to a system using the other three working fluids considered in this study, due to the aforementioned prominent thermodynamic performance advantages of the ORC subsystem. Fig. 13 displays influence of Tc on the payback period. The lowest payback period is found in the case of R134a, followed by pentane, R245fa, and cyclopentane at low Tc. The resulting payback periods of these systems differ greatly, due to high thermal efficiency improvement of the scheme using R134a as working fluids at low Tc and large vapor volume flow rate (Vt) differences through the organic turbine under the same saturated evaporation temperature, as shown in Table 5. 4.2.3 Evaporation temperature of the ORC subsystem (Te_ORC) Fig. 14 to Fig. 17 show the influence of the ORC evaporation temperature on exhaust waste heat recovered, thermal efficiency of ORC, overall thermal efficiency, amount of diesel saved, and payback period, respectively. These Figures illustrate that the ORC subsystem and the

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VES exhibit improved thermal performances when evaporation temperature increases. When the evaporation temperature rises, the evaporation pressure of the ORC subsystem and the expansion ratio of the turbine are increased, and overall thermal performance of the proposed VES is improved. Fig.14 to Fig.16 show that impacts from the fixed evaporating temperature ranges on performances of different working fluids have been studied due to the subcritical ORC is assumed in this research, and the operation temperature for different working fluids confined by their critical temperature. Fig. 14 shows the influence of the ORC evaporation temperature on waste heat recovered, which illustrates that the waste heat recovered decreases with an increase of Te_ORC. The LMTD in the evaporator of the ORC subsystem decreases with increasing Te_ORC, and waste heat recovered,Q5, decreases as Te_ORC is increased due to LMTD is decreased at high Te_ORC according to Eq.(21), under the given overall heat transfer coefficient. Thus, it is expected that the waste heat recovered should decrease with an increasing ORC evaporation temperature. Fig. 15 displays influence of the ORC evaporation temperature on the thermal efficiency of the ORC subsystem. The thermal efficiency of the ORC subsystem is observed to increase with increasing Te_ORC because the expansion ratio through the organic turbine increases with increasing Te_ORC and the enthalpy drop through the organic turbine is also increased. Fig. 14 and Fig. 15 also show power output of the ORC system, PORC, and thermal efficiency of the ORC subsystem, ηORC, is mainly affected by the Te_ORC, and impact from working fluids selection can be ignored. Fig. 16 shows the relationship between the thermal efficiency of the VES and the ORC evaporation temperature. The overall thermal efficiency is observed to increase as the

Page 28 of 61

evaporation temperature of Te_ORC increases. In addition, the overall thermal efficiency, ηcc of R134a is found to be the highest of the four working fluids when the ORC evaporation temperature is below 370 K. The overall thermal efficiency of the VES using cyclopentane is the highest when Te_ORC is over 370 K. Fig. 17 shows the effect of increasing Te_ORC on the amount of diesel saved. The amount of diesel saved increases with increasing ORC evaporation temperature. Using R134a as the working fluid could save more diesel when the evaporation temperature is lower than 370 K, and the system using cyclopentane exhibits advantages in diesel savings when evaporation temperature is over 410 K. Fig. 18 shows the relationship between the payback period and Te_ORC. The payback periods of the four systems decrease with increasing Te_ORC. The system using R134a exhibits prominent economic performance advantages when the evaporation temperature is below 370 K, and the system using cyclopentane as a working fluid exhibits prominent economic performance advantages when the ORC evaporation temperature is in the range of 370 K to 420 K. The payback periods of the proposed scheme using the four working fluids decrease slightly when the ORC evaporation temperature is over 420 K, because the thermal efficiency of the ORC subsystem increases as Te_ORC is increased, while actual exhaust heat recovered, Q5, and PICE are decreased at certain Te_ORC according to Eq.(21) at given heat transfer areas and given heat transfer coefficient. Since the volumetric flow rate and the size of the organic turbine in the R134a case are significantly smaller compared to the other three working fluids cases at the same power output, the working fluid R134a exhibits prominent advantages and is recommended as an ideal working fluid for the VES.

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5 Conclusions A vehicle energy system combining a bottoming ORC with a vehicle air conditioning system was developed and studied in this paper. The common organic working fluids pentane, R245fa, R134a, and cyclopentane were selected for use in the VCR subsystem and the ORC subsystem of the vehicle. After building a thermal and economic performance model, the effects of the key operation parameters were analyzed. The results indicated that the power output of the bottoming ORC is higher than maximum refrigeration capacity requirements and that the overall thermodynamic performance of the conventional VES is significantly improved. The overall thermal efficiency of the proposed VCE is in range of 37.8-38.1%, and the proposed system improves the thermal efficiency by 9.2-9.8%, while saving fuel by as much as 1.1-1.3 L diesel/h. The different working fluids considered lead to payback periods in the range of 4,818-5,427 h. The overall thermal and economic performance largely depends on the type of the working fluid, Te_REF, Tc, and Te_ORC. Specifically, the overall thermodynamic performance is improved with an increase of Te_REF, and it becomes worse with increases of Tc and become worse with decrease of Te_ORC. The overall thermal and economic performance is affected more by Tc than by either Te_REF or Te_ORC. Because the proposed VES using R134a as the working fluid offers the greatest overall advantages regarding thermal and economic performance under the operation conditions considered in this research, it is recommended for use.

Acknowledgements

Page 30 of 61

The work described in this paper is supported by a Project funded by the Program of National Youth Nature Science Foundation of China (51506090) and Youth Natural Science Foundation of Jiangsu Province of China (BK20130799). The authors would also like to gratefully acknowledge fruitful discussions with Daniel Garcia from the Northwestern University at Evanston.

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[7] K.S. Alqdah. Performance and evaluation of aquea ammonia auto air conditioner system using exhaust waste energy. Energy Procedia 6(2011) 467-476. [8] N.A. Aqeeli, P. Gandhidasan. The use of an open cycle absorption system in automobile as an alternative to CFC. The 6th Saudi Engineering Conference, KFUPM, Dhahran, 2002. [9] A.A. Shah. Proposed model for utilizing exhaust heat to run automobile air-conditioner, International Conference on Sustainable Energy and Environment, Bangkok, Thailand, 2006. [10] G. Lorentzen, J. Pettersen. A new efficient and environmentally benign system for car airconditioning, Internal Journal of Refrigeration 16 (1993) 4-12. [11] Z. Crepinsek, D. Goricanec, J. Krope. Comparison of the performances of absorption refrigeration cycles, Journal of WSEAS Transaction on Heat and Mass Transfer 3(2009) 6576. [12] J. Koehler, W.J. Tegethoff, D. Westphalen, M. Sonnekalb. Absorption refrigeration system for mobile applications utilizing exhaust gases, Journal of Heat and Mass Transfer 32 (1997) 333–340. [13] N. Yamada, M.N.A. Mohamad. Efficiency of hydrogen internal combustion engine combined with open steam Rankine cycle recovering water and waste heat. International Journal of Hydrogen Energy 35(2010) 1430-1442. [14] K.S. Kalyan, J.M. Pedro, R.K. Sundar. Analysis of exhaust waste heat recovery from a dual fuel low temperature combustion engine using an Organic Rankine Cycle, Energy 35(2010) 2187-2399.

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[15] A.G. Miranda, T.S. Chen, C.W. Hong. Feasibility study of a green energy powered thermoelectric chip based air conditioner for electric vehicles, Energy 59(2013) 633-641. [16] G.P. Shu, J. Zhao, H. Tian, X.Y. Liang, H.Q. Wei. Parametric and exergetc analysis of waste heat recovery system based on thermoelectric generator and organic rankine cycle utilizing R123, Energy 45(2012) 806-816. [17] I. Vaja, A. Gambarotta. Internal combustion engine (ICE) bottoming with organic Rankine cycles (ORCs), Energy 35 (2010) 1084-1093. [18] H. Khayyam. Adaptive intelligent control of vehicle air conditioning system. Applied Thermal Engineering 51(2013) 1154-1161. [19] B.J.Hipolito-valencia, E.Rubio-Castro, J. M. Ponce-Ortega, M.Serna-Gonzalez, F.NapolesRivera, M.M.El-Halwagi. Optimal design of inter-plant waste energy integration. Applied Thermal Engineering 62(2014)633-652. [20] B.J.Hipolito-Valencia, E.Rubio-Castro, J.M.Ponce-Ortega, M.Serna-Gonzalez, F. NapolesRivera, M.M.El-Halwagi. Optimal integration of organic Rankine cycles with industrial processes. Energy Conversion and Management 73(2013)285-302. [21] S. Gonazalez, M. M. El-Halwagi. Multiobjective design of interplant trigeneration systems. AICHE Journal 60(2014)213-236. [22] G.Q. Qiu. Selection of Working fluids for micro-CHP systems with ORC, Renewable energy 48(2012) 565-570.

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[23] H.S. Li, X.B. Bu, L.B. Wang, Z. Long, Y.W. Lian. Hydrocarbon working fluids for a Rankine cycle powered vapor compression refrigeration system using low-grade thermal energy, Energy and Building 65(2013) 167-172. [24] M.S. Graboski, R.L. Mccormick. Combustion of fat and vegetable oil derived fuels in diesel engines, Progress in Energy and Combustion Science 24(1998) 125-164. [25] J.B. Heywood. Internal combustion engine fundamentals. New York, McGraw-Hill, 1998. [26] J. Zhang, H.G Zhang, K. Yang, F.B Yang, Z. Wang, et al.. Performance analysis of regenerative organic Rankine cycle (RORC) using the pure working fluid and the zeotropic mixture over the whole operating range of a diesel engine. Energy Conversion and Management 84(2014) 282-294. [27] J.Q Fu, J.P Liu, Y.P Yang, C.Q Ren, G.H Zhu. A new approach for exhaust energy recovery of internal combustion engine: Steam turbocharging, Applied Thermal Engineering 52 (2013) 150-159. [28] S. Quoilin, S. Declaye, B.F. Tchanche, V. Lemort. Thermo-economic optimization of waste heat recovery organic Rankine cycles. Applied Thermal Engineering 31(2011) 2885-2893. [29] http://www.x-rates.com/table/ 02-25-2014. [30]http://www.sdpc.gov.cn/zcfb/zcfbtz/2014tz/W020140124560699807886.pdf 12-24-2014. [31] C. Yue. D. Han, W.H. Pu. Analysis of the integrated characteristics of the CPS (combined power system) of a bottoming organic Rankine cycle and a diesel engine. Energy 72(2014) 739-751.

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Figures

ICE power cycle subsystem

Air condition subsystem Air

Radiator

9

Pump

wICE Q1

8

Liquid tank Throttle valve

4

Air flow Flue gas flow Coolant flow Organic working fluid Mechanical power

Condenser

5

Compressor

T

7

10

6

3

Evaporator

ICE

Exhaust end

Suction end

To cooling To heating Flue gas

Q2 Silencer device

1

2 5

2 Inter cooler

6

Catalytic purificaiton

1

T0 Te_REF

Air

9

Tc

Q4

8 wREF

0 10

Q3

7

Mechanical power Power transfer device

0

S Mechanical power

(a)

(b)

Fig. 1. The conventional vehicle energy system. (a) Flow diagram; (b)Temperature- entropy diagram.

Page 36 of 61

ICE power cycle subsystem

Air condition subsystem

ORC subsystem

Air

Radiator

4

Condenser Liquid tank

T

9 8 Valve 1

wICE

Valve 2

Q1 5

13

SD valve1 A

Compressor

Pump

Throttle valve

Turbine

Pump

6

3

10

7

ICE

Exhaust end

Suction end

To cooling To heating

Flue gas

Silencer device

14 5

2 Inter cooler

6

12

Mechanical power

Evaporator

14 Q2 11

Catalytic purificaiton

Tc T0 Evaporator2

0

1

2

1 Air

Q5

Te_ORC

Te-REF SD valve2

Mechanical power

0

11

12

Q4 8

9

wORC 13

wREF 10

Q3

7

Power transfer device

Mechanical power

(a)

S

A

(b)

Fig.2. VES with waste heat recovery ORC technology. (a) Flow diagram;(b) Temperatureentropy diagram.

Page 37 of 61

ICE power cycle subsystem

Air condition subsystem

ORC subsystem

Air

Radiator

4

Condenser2

Condenser

Liquid tank2

Liquid tank

T wICE

9

8

Q1

Compressor

5

13

A Throttle valve

Pump

Turbine

Pump

6

3

10

7

ICE

Exhaust end

Suction end

To cooling To heating

Flue gas

Silencer device

14 5

2 Inter cooler

6

12

Mechanical power

Evaporator

14 Q2 11

Tc T0

Evaporator2

Air

0

1

2

Catalytic purificaiton

1

Q5

Te_ORC

Te-REF

Mechanical power

0

11

12 wORC 13

Q4 8

9

wREF 10

Q3

7

Power transfer device

Mechanical power

(a)

S

A

(b)

Fig. 3. The referenced VES with waste heat recovery ORC technology. (a) Flow diagram;(b) Temperature- entropy diagram.

Page 38 of 61

Given T0, p0, ma, mfuel, Q1, p6, T6, Ti, p1, Vdisp,Vcl, Adjust β

Solve T6c

No

Given T0, p0, PICE, mflue, p6, T6,

Given T0, p0, PICE, mflue, p6, T6,

Given evaporator and condenser structure parameter Ae_ORC,ACon

Given evaporator structure parameter Ae_ORC

Solve ORC loop: hc,hh,htot,Q5,PORC,Q4,ηORC etc.

Solve ORC loop: hc,hh,htot,Q5,PORC,Q4,ηORC etc.

Solve thermal performance: ηCC,ΔηCC,vs, etc.

Solve thermal performance: ηCC,ΔηCC,vs, etc.

Solve economic performance: Cs,PP, etc.

Solve economic performance: Cs,PP, etc.

β=β+Δβ

|T6c-T6|<ε1 Yes Solve PICE , mfuel, T6, etc.

Given Q3 and Δte,Δtc

Solve PREF , ηICE, etc.

(a)

(b)

(c)

Fig.4. The calculation flow diagram. (a) The conventional ICE+VCR model; (b) the ICE+VCR+standalone ORC model; (c) the ICE+integrated VCR-ORC model.

Page 39 of 61

8

n-pentane R245fa R134a cyclopentane

7

COP

6

5

4

3

272

276

280

284

288

Te_REF,K

Fig. 5 COP of the air conditioning subsystem as a function of the evaporation temperature Te_REF.

Page 40 of 61

38.6

n-pentane R245fa R134a cyclopentane

38.4

ηCC,%

38.2

38.0

37.8

37.6

37.4 272

276

280

284

288

Te_REF,K

Fig. 6. Overall thermal efficiency as a function of the evaporation temperature Te_REF.

Page 41 of 61

10

n-pentane R245fa R134a cyclopentane

PREF,kW

8

6

4

300

302

304

306

308

310

312

314

316

318

Tc,K

Fig. 7. Compression power consumption of the air conditioning subsystem as a function of the condensation temperature Tc.

Page 42 of 61

8

n-pentane R245fa R134a cyclopentane

7

COP

6

5

4

3 300

302

304

306

308

310

312

314

316

318

Tc,K

Fig. 8. COP of the air conditioning subsystem as a function of the condensation temperature Tc.

Page 43 of 61

24

n-pentane R245fa R134a cyclopentane

PORC,kW

21

18

15

12

300

302

304

306

308

310

312

314

316

318

Tc,K

Fig. 9. Power output of the ORC subsystem as a function of the condensation temperature Tc.

Page 44 of 61

14 13

n-pentane R245fa R134a cyclopentane

12

ηORC,%

11 10 9 8 7 6 300

302

304

306

308

310

312

314

316

318

Tc,K

Fig. 10. Thermal efficiency of the ORC subsystem as a function of the condensation temperature Tc.

Page 45 of 61

n-pentane R245fa R134a cyclopentane

39

ηCC,%

38

37

36

300

304

308

312

316

320

324

328

Tc,K

Fig. 11. Overall thermal efficiency as a function of the condensation temperature Tc.

Page 46 of 61

n-pentane R245fa R134a cyclopentane

Vs, L.h

-1

1.5

1.2

0.9

0.6 300

304

308

312

316

320

324

328

Tc,K

Fig. 12. Diesel saved as a function of the condensation temperature Tc.

Page 47 of 61

10000

n-pentane R245fa R134a cyclopentane

PP, h

8000

6000

4000

300

304

308

312

316

320

324

328

Tc, K

Fig. 13. Payback period as a function of the condensation temperature Tc.

Page 48 of 61

176

n-pentane R245fa R134a cyclopentane

Q5,kW

168

160

152

144

340

360

380

400

420

440

460

480

500

Te_ORC,K

Fig. 14. Waste heat recovered as a function of the evaporation temperature Te_ORC.

Page 49 of 61

20

ηORC,%

16

12

n-pentane R245fa R134a cyclopentane

8

4 340

360

380

400

420

440

460

480

500

Te_ORC,K

Fig. 15. Thermal efficiency of the ORC subsystem as a function of the evaporation temperature Te_ORC.

Page 50 of 61

41

40

ηCC,%

39

n-pentane R245fa R134a cyclopentane

38

37

340

360

380

400

420

440

460

480

500

Te_ORC,K

Fig. 16. Overall thermal efficiency as a function of the evaporation temperature Te_ORC.

Page 51 of 61

2.5

Vs, L.h

-1

2.0

1.5

n-pentane R245fa R134a cyclopentane

1.0

0.5 340

360

380

400

420

440

460

480

500

Te_ORC,K

Fig. 17. Diesel saved as a function of the evaporation temperature Te_ORC.

Page 52 of 61

PP, h

9000

n-pentane R245fa R134a cyclopentane

6000

3000

340

360

380

400

420

440

460

480

500

Te_ORC,K

Fig. 18. Payback period as a function of the evaporation temperature Te_ORC.

Page 53 of 61

Table 1. Component costs and calculation expressions. Cost(€)

Component

Dependent variable

Typical value of the variable

Turbine

volume flow rate Vt, m3.s-1

-

Heat exchanger

heat exchange effective area A, m2

Pump

electrical power Pp,ORC, kW

-

Liquid tank

volume Vol, L

20

Piping

pipe diameter and length, dpipe=20mm,lpi (0.897+0.21·dpipe) · lpipe mm pe=30m

Working fluid

volume, L

45

-

Working fluid price

20 €.kg-1

130

-

Miscellaneous hardware

-

300

-

Control system

-

500

-

Clabor

-

-

Local factor

f

0.6

0.5-0.7

mEx

kg.m-2

0.5

-

φ

-

0.6

-

δ

L.kg-1

0.005

-

1.5·(225+170·Vt) 190+310·A 900·(Pp,ORC/0.3)0.25 31.5+16·Vol

0.3· TCC

Page 54 of 61

Table 2. Properties of the four working fluids. Refrigerant

Formula

M, kg.kmol-1

NBP,K

Tcrit, K

Pcrit, MPa

n-Pentane

C5H12

72.15

309.20

469.90

3.360

R245fa

CHF2CH2CF3

134.05

288.28

427.16

3.651

R134a

CF3CH2F

102.03

247.08

374.21

4.059

Cyclopentane

C5H10

70.13

322.15

511.75

4.440

Page 55 of 61

Table 3. Input parameter and boundary conditions. Parameter

Typical value

Ranges

Ambient temperature at refrigeration season, T0, K

303

-

Ambient pressure, p0, MPa

0.101

-

Type of engine

TDV8 3.6

-

Liquid diesel density, ρfuel, kg.L-1

0.832

-

Number of cylinder

8

-

Bore×stroke, mm×mm

84.0×98.5

-

Volume efficiency, ηvol, %

87

-

Rated power of engine, kW

176

-

Rotation speed, n, RPM

3,500

-

Diesel fuel consumption, L.100km-1

13.4

10.3-19.1

Compression ratio, p3/p2

16.1

-

Turbocharger exhaust temperature, T6, K

850

[26,27]

Vehicle speed, km/h

100

-

Evaporating temperature, Te_REF, K

278

258-288

Condensing temperature, Tc, K

311

300-328

Subcooling degree of condenser, K

3

0-5

Compressor isentropic efficiency, %

75

60-90%

Refrigeration capacity, Q3, kW

30

5-25% of PICE

Isentropic efficiency of pump in ORC,%

75

-

Isentropic efficiency of turbine in ORC, %

80

60-90%

Evaporation temperature of ORC, Te_ORC, K

360

340-510

Effective heat-transfer area of evaporator 2, A, m2

10

-

Page 56 of 61

Overall heat transfer coefficient of evaporator 2, htot, 0.13 kW.(m2.K)-1

0.06-0.20

Page 57 of 61

Table 4. Overall performance results for the different system schemes. Conventional Proposed VES VES R134a

Referenced VES

ICE subsystem Overall fuel consumption, L.(100km.h)-1

13.4

12.1

12.1

ICE power output, PICE, kW

176

158.8

158.8

Waste exhaust heat, Q2, kW

268

193

193

Thermal efficiency of ICE subsystem, ηICE,%

35.9

35.9

35.9

Refrigeration power consumption, PREF, kW

7.5

7.5

7.5

COP of refrigeration cycle

4.0

4.0

4.0

Exhaust waste heat recovered,Q5, kW

-

169.0

169.0

Fuel temperature out of evaporator,T14,K

-

359

359

Turbine inlet volume flow rate, Vt, m3/s

-

0.0074

0.0074

Power output of turbine, Pt,ORC, kW

-

18.1

18.1

Power consumption of pump, Pp,ORC, kW

-

1.6

1.6

Power output of ORC subsystem, PORC, kW

-

16.5

16.5

Thermal efficiency of ORC subsystem, ηORC,%

-

9.7

9.7

34.4

38.1

38.1

-

9.8

9.61

Air conditioning subsystem

ORC subsystem

Thermo-economic performance Overall thermal efficiency, ηcc,% Thermal efficiency improvement, ΔηCC, %

Page 58 of 61

Fuel saved per hour, Vs, L gasoline.h-1

-

1.3

1.3

△FP, L gasoline.h-1

-

0.04

0.08

Fuel cost saved, Cs, RMB Yuan.h-1

-

9.74

9.41

OIC, RMB Yuan

-

46,926

70,813

Payback period, h

-

4,818

7,520

Page 59 of 61

Table 5. Overall performance indices of the proposed VES with different working fluids. Case of npentane

Case of R245fa

Case of R134a

Case of cyclopentane

ICE power output, PICE, kW

160.7

161.1

158.8

158.2

Overall fuel L.(100km.h)-1

12.2

12.3

12.1

12.2

Refrigeration power consumption, PREF, kW

6.0

6.4

7.5

5.7

COP of the refrigeration cycle

5.0

4.6

4.0

5.2

Waste heat recovered, Q5, kW

171

171

169

170

Fuel temperature out evaporator, T14, K

359

360

359

361

Turbine inlet volume flow rate, Vt, m3/s

0.0333

0.0156

0.0074

0.051

Power output of turbine, Pt,ORC, kW

15.0

14.7

18.1

16.4

Power consumption of pump, Pp,ORC, kW

0.23

0.40

1.6

0.13

Power output of the subsystem, PORC, kW

ORC

14.8

14.3

16.5

16.2

Thermal efficiency of the ORC subsystem, ηORC,%

8.6

8.4

9.7

8.7

Overall thermal efficiency, %

38.0

37.8

38.1

38.0

Thermal efficiency improvement,

9.5

9.2

9.8

9.5

ICE subsystem

consumption,

Air condition subsystem

ORC subsystem

of the

Thermo-economic performance

Page 60 of 61

ΔηCC ,% Diesel saved per hour, Vs, L diesel.h-1

1.2

1.1

1.3

1.2

△FP, L gasoline.h-1

0.04

0.04

0.04

0.04

8.99

8.20

9.74

8.98

OIC, RMB Yuan

43,666

44,457

46,926

43,729

Payback period, PP, h

4,870

5,427

4,818

4,877

Fuel cost saved, Cs, RMB Yuan.h1

Page 61 of 61