20XX-01-XXXX

Effect of an ORC Waste Heat Recovery System on Diesel Engine Fuel Economy for Off-Highway Vehicles

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Abstract

Modern heavy duty diesel engines can well extend the goal of 50% brake thermal efficiency by utilizing waste heat recovery (WHR) technologies. The effect of an ORC WHR system on engine brake specific fuel consumption (bsfc) is a compromise between the fuel penalty due to the higher exhaust backpressure and the additional power from the WHR system that is not attributed to fuel consumption. This work focuses on the fuel efficiency benefits of installing an ORC WHR system on a heavy duty diesel engine. A six cylinder, 7.25ℓ heavy duty diesel engine is employed to experimentally explore the effect of backpressure on fuel consumption. A zero-dimensional, detailed physical ORC model is utilized to predict ORC performance under design and off-design conditions. The ORC model includes a detailed exhaust gas heat exchanger model and a thermodynamic ORC submodel to explore the effect of recovering various amounts of waste heat on ORC thermal efficiency under the same engine load and speed conditions. This study focuses on maximum engine power conditions where the engine exhaust gas and temperature are maximized. The results show that increasing the heat exchanger surface area leads to higher heat recovered at the expense of higher exhaust backpressure and higher WHR system weight, as the ΔT between the fluids approaches zero. At the same time, the weight increase of the heat exchange is illustrated as the main parameter that limits the ORC system design in vehicular applications. Finally, the optimum heat exchanger length is a trade-off between exhaust backpressure, the required net ORC power and weight increase.

Introduction

Automotive and heavy duty vehicles are under high pressure by public regulatory agencies to decrease pollutants and CO2 emissions. State of the art vehicles embody both sophisticated aftertreatment technologies to decrease exhaust pollutants and advanced combustion technologies for low CO2 emissions. However, the goal of over 50% brake thermal efficiency cannot be achieved with the currently existing technology without the utilization of the waste heat recovery technology, as the majority of the fuel energy is wasted [1].

Waste heat recovery (WHR) technology utilizes the exhaust gases as the main heat source to recover heat; however, importing any additional device in the exhaust system increases engine backpressure and has an additional cost on fuel consumption. A widely known WHR technology is turbocompounding (T/C), which can be either

mechanically or electrically connected to the powertrain. Many engine manufacturers such as Caterpillar, U.S. Cummins and Scania tested various T/C configurations and reported an average bsfc reduction between 3% and 6%, depending on the engine load conditions [2–4]. A more recent study validates that a highly efficient T/C system can improve bsfc in the range of 3.3-6.5% [5]. Although the potential efficiency benefit from T/C can be much higher than 6%, this is limited due to the high exhaust backpressure caused by the T/C. The increased pumping losses of the engine result in an additional penalty on fuel consumption that compromises the total engine net efficiency.

ORC is an alternative and more efficient WHR solution that in recent years has been gaining ground in the automotive industry due to the stricter CO2 emission standards. Many theoretical studies regarding integrated ORC systems on vehicle powertrain present thermal efficiency between 6%-15% [5–9]. This variation on the ORC thermal efficiency mainly depends on the heat sources which are used and the engine operating conditions. Exhaust gases are a source of high grade heat to extract energy through a heat exchanger for an ORC system. Fitting a heat exchanger in the exhaust manifold results in an increase on backpressure, but compared to T/C technology this increase is approximately one order of magnitude lower. Both the amount of the extracted heat and the exhaust backpressure depend on the ORC configuration and the heat exchanger type [9–11]. Engine waste heat can be transferred directly through the evaporator to the ORC loop but in some studies an intermediate thermal oil loop between exhaust gases and the ORC is used [7]. Although such a cycle slightly decreases the heat transfer efficiency, it guarantees steady state conditions for the ORC operation while any potential decomposition of the working fluid at high exhaust enthalpy conditions can be avoided [7].

There are many different types of heat exchangers which can be used to extract heat from exhaust gases. A comparison study between a shell-and-tube heat exchanger and a plate heat exchanger showed that more heat is gained by the latter [10]. In another study a plate counter flow heat exchanger model was used to simulate the extracted heat from the exhaust gases [7]. A plate heat exchanger has been also adopted in another experimental work with the aim to investigate the effect of backpressure on the engine performance, although it is mentioned by the authors that this type of heat exchanger is not suitable in terms of exhaust pressure drop [12]. Houndalas et al. used a shell-and-tube evaporator to explore the different amounts of extracted heat under various engine load conditions [9]. In most of the above referred simulation studies, the negative effect of

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additional exhaust backpressure on fuel consumption is not mentioned as they primarily focus on the amount of extracted heat and the ORC thermal efficiency. Furthermore only a few experimental works consider the exhaust backpressure effect of the heat exchanger, but they do not give insights in the heat exchanger design and optimization.

Another drawback of the integration of an ORC system on a vehicle platform is the weight increase. The effect of weight increase on fuel consumption has been investigated in the past, mainly by using vehicle simulation tools [13–16]. In the case of a regional delivery truck it was found that a 20% reduction of its weight can lead to a 12% reduction in fuel consumption [16]. High vehicle weight is a negative performance factor as due to the higher vehicle inertial mass both engine load and fuel consumption during acceleration increase while at the same time frictional forces are enhanced under all operating conditions. To this end, the installation of an ORC system is expected to constrict the vehicle performance, especially under partial engine load conditions. A recent study showed that the weight of an ORC system that utilizes 200kWth exhaust waste heat of a tractor and presents a theoretical 10% thermal efficiency can reach approximately 300 kg [17]. The latter means that an additional 20kWe

is accompanied by a significant increase on vehicle weight mass, although in the case of an off-highway vehicle this drawback is less important as most of the operating time the vehicle may work as a stationary machine (such as excavators, cranes etc).

Research in ORC is increasingly gaining momentum, in 2010 less than 20 papers were published in this field while as of 2014 there were more than 200 papers published [18]. However, most of these studies are thermodynamic analysis of the ORC system, evaluating the effect of various working fluids, heat exchangers and expanders on thermal efficiency and power output, while only few of them are experimental works. In fact, there is a huge gap between the calculated ORC efficiency in theoretical and experimental works, as the drawbacks of the weight increase and the exhaust backpressure are not considered in modeling. It is expected that heat exchanger area is related to the ORC system weight, thermal efficiency and exhaust backpressure. Therefore, it becomes of significant importance to investigate the trade-off between the extracted heat, the exhaust backpressure and the additional weight loading on powertrain performance and fuel consumption.

The aim of this study is to explore the effect of implementing an ORC WHR system on an off-highway vehicle in terms of ORC efficiency, engine efficiency and powertrain performance. The analysis considers the effect of exhaust backpressure on engine fuel consumption and the potential changes of power to weight ratio due to the ORC additional weight. The study includes the experimental data of a heavy-duty diesel engine, a detailed shell-and-tube exhaust gas-thermal oil heat exchanger model and a simplified thermodynamic ORC model that operates with various working fluids. The heavy duty diesel engine fuel consumption was measured at maximum engine power for various exhaust backpressure values. By using the thermal oil heat exchanger model, it was possible to match different volumes of the heat exchanger for different backpressure measured values. The thermal oil heat exchanger model is able to calculate the additional weight of more intrusive heat exchangers. The thermal efficiency of the ORC system is calculated for various amounts of extracted heat and for two different working fluids. Finally, the new integrated powertrain unit fuel consumption and the vehicle engine-ORC power/weight ratio are calculated as a function of exhaust backpressure and ORC weight.

Modeling Approach of the Waste Heat Recovery System

A schematic representation of the WHR system that is adopted in this simulation study is given in Figure 1. The designed ORC system exchanges heat with the thermal-oil loop that guarantees the smooth power output of the ORC system and avoids any potential decomposition of the working fluid. The only heat source of the ORC system is the thermal oil flow, while the condenser is assumed to exchange heat with an ultimate heat sink (cooling air). Assuming a worst case scenario where the OHV is stationary, the cumulative air mass flow rate is supplied by electrically driven fans. Using data charts for air fan-cooling the equivalent power required to drive the fans to produce the mass flow rate has been calculated. This power is 10% of the total ORC power.

Figure 1: Schematic presentation of the engine-ORC system

Direct heat transfer from the exhaust gases to the organic fluid is often preferred in transport applications as it increases the heat transfer efficiency and reduces the weight of the WHR system, while the thermal oil loop requires an extra heat exchanger and pump. However, the thermal oil eases the control of the thermodynamic conditions in the ORC circuit and for the purpose of this study facilitates the comparison between different working fluid performances. The thermal oil loop consists of a Shell and Tube Heat Exchanger (STHE) directly exposed to the engine exhaust gases. The thermal oil is pumped through the STHE extracting thermal energy from the exhaust gases and flows through a Brazed Plate Heat Exchanger (BPHE). In the BPHE the thermal oil exchanges thermal energy with the organic fluid and gets cooled down to its initial temperature before flowing back to its reservoir and being pumped back to the STHE thereby resetting the cycle.

In the ORC circuit an organic liquid fluid is pumped into the regenerator where it is preheated as it exchanges heat with the organic fluid at vapor phase. Then it is pumped into the BPHE where it extracts heat from the thermal oil and increases its energy content to obtain superheated vapor. The superheated working fluid then

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drives an expander coupled to an electric generator. A power conversion unit converts the turbine’s mechanical work into conditioned electricity. The organic fluid leaves the turbine in vapor phase and is primarily cooled down in the regenerator before condensed back to liquid form in a condenser and being pumped back into the Rankine cycle system.

Thermal-oil heat exchanger modeling

The thermal-oil heat exchanger model describes a STHE where the exhaust gas exchanges heat with a thermal oil. The model calculates the heat transfer coefficients and the pressure drop at both sides of the heat exchanger. Both sub-models are described in this subsection.

In this study, the simulation of the heat exchanger is based on the ε−NTU method, where ε measures the effectiveness of the heat exchanger:

ε= QQmax

(1)

Where Qmax=(C pm )min ¿ is the heat transfer that is achieved in principle by a counterflow heat exchanger with infinite length. The total extracted heat from the exhaust gases Q is obtained from:

Q=moil¿ (2)

The number of transfer units (NTU) is a dimensionless number defined by:

NTU= UA(Cp m )min

(3)

And for a counter-flow configuration the NTU can be determined from the effectiveness by:

NTU=1

(1−C r )ln( ε−1

ε C r−1 ) (4)

C r=(C pm )min

(C pm )max

(5)

Using equations (1-5) it is possible to determine UA the overall heat transfer coefficient across a surface area.

In this analysis, the heat exchanger length is increased iteratively until UAdetermined by the ε−NTU method using equations (1-5) matches the values obtained from the heat transfer coefficients as described in the following paragraphs.

Heat Transfer Coefficient: the overall heat transfer coefficient across a surface area is determined by:

UA=U o Ao=U i Ai=1

∑ Rt(6)

and the sum of resistivities can be determined from:

Rt=1

H i Ai+

lnDo

Di

2πLk+ 1

H o Ao

(7)

Where L denotes the length of the tube, which is updated iteratively by the simulation and k is the thermal conductivity of steel. The shell side heat transfer co-efficient H o is estimated using the Bell-Delaware method [19]:

H o=H ideal(JC J L J B J R J S) (8)

Where: JC , J L , JB , J R , J S are the correction factors for: baffle window flow, baffle leakage effects, bundle bypass effect, laminar flow correction factor and correction factor for un-equal baffle spacing [19]. The heat transfer co-efficient for an ideal tube bank hideal is:

H ideal= jmo

SmCp( μ

μs )0.14

Pr−2/3 (9)

Where Sm is the cross-flow area of the central region of the heat exchanger and j factor is a dimensionless number obtained from correlations based on the Reynolds umber (Re) and the tube diameter

(Do) and tube pitch (PT ) [19].

The tube side heat transfer co-efficient is determined from:

H i=N u kc

Di(10)

Where the Nusselt number Nu for turbulent flow (ℜ>1 ×104) inside a tube can be determined from the Gnielinski’s correlation [20]:

Nu=

f8

( ℜ−1000 ) Pr

1+12.7( f8 )

12 ( Pr

23 −1)

(11)

f =(0.79 ln ( ℜ)−1.64 )−2(12)

Pressure drop: The pressure drop in the shell-side of the heat exchanger is also determined from the Bell-Delaware method [19]. The pressure drop is determined from the sum of the pressure drop in the cross flow central region, window section and entry and exit baffle sections such that:

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Δ Po=Δ Pc+Δ Pw+ Δ Pe (13)

The pressure drop in the cross flow central regions of the STHE is determined from:

Δ P c=(N ¿¿ B−1)

2 f p N c ( mo

Sm)

2

ρ( μμs )

0.14 RL RB ¿(14)

Where NB is the number of baffles in the STHE, NC is the number

of tubes along the central width of the STHE and f p factor is a dimensionless number obtained from correlations based on the Reynolds umber (Re), tube diameter (Do) and tube pitch (PT ). The

values of RL and RB are the correction factors for baffle leakage effects and bundle bypass effect. Similarly, the pressure drop in the baffle window section of the STHE is determined from:

Δ Pw=(N ¿¿B)(2+0.6 N w) ( mo )2

2ρ Sm SwRL¿ (15)

Where NB is the number of tubes along the width of the baffle

section of the heat exchanger and Sw is the area of the baffle region of the heat exchanger. Finally, the pressure drop in the entrance and exit region of the STHE exchanger is determined from:

Δ Pe=22 f p N c( mo

Sm)

2

ρ( μμs )

0.14 (1+Nw

N c )RB

(16)

The pressure drop in the tube-side of the heat exchanger is determined from the Zigrang and Sylvester correlation [21]:

Δ Pi=(−2 log (a ) )−2 (ρ u2)2 Di

L (17)

Where u measures the velocity of the fluid inside the tube and the constant a is determined from:

a=( 2 e7.54 Di )−( 5.02

ℜ ) log( 2e7.54 Di )+( 13

ℜ )(18)

ORC system modeling

Figure 2 shows the thermodynamic cycle of an organic Rankine cycle, in which a thermal oil is used as an intermediate heat transfer fluid between the exhaust gasses and the working fluid. The transformations that the working fluid undergoes in the cycle are briefly illustrated in Table 1.

The thermodynamic model of the ORC has been developed using an in-house MATLAB code. The cycle simulator defines the cycle properties at each of the points (numbered 1 through 7 in Figure 2) of

the ORC. To this end, the mass and energy balance equations are used for each point.

Figure 2: Schematic representation of the engine-ORC system

Table 1: ORC thermodynamic cycle description.

Transformation Component Description

1 – 2 Pump Fluid pressurization

2 – 2a Regenerator, cold side Fluid pre-heating in the regenerator

2a – 5 Evaporator Fluid vaporization and superheating

5 – 6 Turbine expander Fluid expansion

6 – 6a Regenerator, hot side Heat rejection to the pressurized fluid

6a – 1 Condenser Fluid condensation

Pump: The pump is assumed to be electrically powered; in the absence of leakages and for adiabatic operation, mass and energy balance are:

Ppump=mwf (h2 , is−h1 )

ηpump ηmotor

(19)

Evaporator: Thermal energy is provided to the working fluid by the heated thermal oil. With the objective to control the pinch point temperature difference, the evaporator energy balance is split in 3 transformations; namely: pre-heating, vaporization and super-heating.

moil¿ (20)

Equation (20) expresses the total energy balance of the evaporator.

Turbine: The turbine power output of the turbine expander is calculated by:

Pturbine=ηturbinemwf (h5−h6 , is) (21)

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Condenser: As mentioned, the energy balance of each component allows to control the pinch point temperature. The mass balance is expressed by equation:

mwf ( h6−h1)=mcw(hcf , 4−hcf ,1) (22)

Equation (22) expresses the energy balance for the condenser.

Regenerator: The regenerator extracts energy from the exhaust gas of the turbine to pre-heat the pressurized working fluid. Notice that, since none of the fluid changes phase in the process, the lowest ΔT appears either at the component inlet or outlet; for this reason, there is no need for a pinch point check. The energy balance of this heat exchanger is expressed by equation (23):

mwf ( h6−h6 a )=mwf (h2 a−h2 ) (23)

Experimental Setup

The design of the ORC system is based on the exhaust heat flow of a 7.25ℓ Yuchai engine. This heavy duty diesel engine is turbocharged, direct injection and fulfills the EURO III regulatory constraints. The engine that was tested in Brunel University does not present any EGR or VGT system while any potential aftertreatment system has been removed. The detailed characteristics of the heavy duty diesel engine are presented in Table 2.

Table 2: Yuchai YC6A280-30 diesel engine characteristics.

Displaced volume 7255 cc

Stroke 132 mm

Bore 108 mm

Compression ratio 17.5:1

Number of Cylinders 6

Number of Valves 4

Maximum Torque 1100Nm @ 1400-1600rpm

Maximum Power 206kW @ 2300rpm

Optimum bsfc point ≤205 g/kWh

The whole engine operating map was measured on the engine dyno and a schematic presentation of the exhaust heat is presented in Figure 3. It is illustrated that the maximum power point at 2300rpm presents the highest enthalpy flow, which is 284.3 kWth; therefore this engine operating point was selected for the design of the WHR system. Then, a throttle valve was installed in the exhaust pipe, enough distance downstream to the turbocharger not to affect its operation. The throttle experimentally simulates the effect of exhaust backpressure on fuel consumption for the ORC selected engine operating point. The tests were conducted by using two pressure sensors upstream and downstream of the throttle valve. The Gems 3100 Series pressure sensors have a 0 to 7 bar operating range, while an AVL Dynamic Fuel Balance 7131 was used for the measurements of the fuel consumption.

Figure 3: Qualitative presentation of the diesel engine exhaust waste heat map and the selected engine operating point for the ORC design analysis.

Results and Discussion

Effect of the exhaust backpressure on engine performance

Figure 4 presents the relationship between the exhaust backpressure and engine fuel consumption, when the engine operates at its maximum power. It was found that after the installation of the exhaust throttle valve, the backpressure at wide opened throttle conditions was 60mbar higher compared to the original engine measurements. The original bsfc is 209 g/kWh and is depicted in Figure 4 as the baseline curve.

Figure 4: Effect of the exhaust backpressure on engine bsfc at the maximum engine load and speed conditions.

The investigation was performed for a backpressure range between 60 mbar and 280 mbar (Figure 4). Over this backpressure value of 280kPa the engine is not able to keep the same power output, as mixture becomes too rich. It can be seen that 60 mbar increase of the exhaust back pressure leads to 0.25% increase on bsfc while at higher exhaust backpressure values the fuel penalty is proportionally

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increased. The error in pressure measurement is less than ±1% full scale reading (FSO). It is expected that as exhaust backpressure is increased, engine has to consume more fuel in order to maintain the same engine power output that leads to higher exhaust temperature and slightly lower exhaust massflow. The latter means that the available exhaust enthalpy is higher which can potentially increase the ORC system efficiency. On the other hand, higher backpressure corresponds to a higher length and heavier heat exchanger which can significantly affect the power to weight ratio of the vehicle. This study tries to compromise all these parameters in the design of an exhaust gas-thermal oil heat exchanger.

Other recent studies present that 100mbar backpressure leads to 1% increase in fuel consumption [12,17]. However in this study it was found that the effect of such a backpressure results 0.5% increase in bsfc, which is the half compared to literature. Further experimental work to other engine operating points was necessary to deeply understand the reason of this observed deviation. It was found that exhaust backpressure at low engine load and speed operating conditions is more noticeable compared to the selected engine operating point of this study. In fact, a backpressure of 100mbar can increase up to 3.5% fuel consumption at low speed and torque engine operating conditions, but at the maximum engine power conditions the effect of backpressure is minimized.

Thermal-oil heat exchanger design

The measured exhaust gas thermodynamic properties (specifically mass flow rate and temperature) vary as the exhaust backpressure is increased. These exhaust data were used as input for the exhaust gas heat exchanger simulations. The heat exchanger surface area was increased discretely until the calculated backpressure in the simulation matched the measured experimental exhaust backpressure values.

The selection of the working fluid of an ORC system is of major importance, as it significantly affects the ORC efficiency. Compared to pure working fluids, mixture can improve the system efficiency due to their lower irreversibility and higher cycle exergy efficiency [22,23]. However, the aim of this study is not to present another thermodynamic analysis of the ORC system but to a novel methodology on how to consider the negative aspects of an ORC system such as exhaust back pressure and additional weight on the preliminary design. Therefore, the calculations which are presented consider two fluids: Fluid A is R1233zde which is commonly used in many studies in published literature and fluid B, whose properties are confidential. In the following figures, they are referred as Fluid A and Fluid B.

Exhaust backpressure and residence time are highly linked with each other. Gas hour space velocity (GHSV) is defined as the ratio between the volumetric flow and the volume of the device. A smaller heat exchanger presents lower volume, higher GHSV values and consequently low residence time of the exhaust gas in the heat exchanger and lower exhaust backpressure values. On the other hand, a bigger heat exchanger increases the residence time of the exhaust gas and this leads to higher backpressure values. The relationship between GHSV and exhaust backpressure is schematically presented in Figure 5.

Figure 5: Relationship between gas hour space velocity (GHSV) and exhaust pressure drop for both tested fluids.

Figure 6 and Figure 7 show the temperature – energy diagrams for the organic fluids studied. The red continuous line, dot and dash blue line and the dashed black line show the temperature profiles of the exhaust gases, thermal oil and organic fluid respectively. The temperature profiles of the two fluids are very distinct; however the overall energy extraction is similar. This allows direct comparison of the ORC effect for two fluids that have different thermodynamic behavior on the engine performance. It should be mentioned that the exhaust gas temperature at the outlet of the heat exchanger has been limited to 130oC to avoid corrosion problems related to the dew point.

Figure 6: Temperature profile of Fluid A.

Figure 7: Temperature profile of Fluid B.

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Figure 8 compares the change in exhaust gas backpressure (continuous line) and extracted thermal energy from the exhaust gases (dashed line) for varying heat exchanger surface area, for the two fluids considered. As expected the heat exchanger length is proportional to the exhaust gas backpressure. However, the extracted heat resulting from increasing the heat exchanger length are increasingly diminished at the expense of backpressure as the ΔT between the exhaust gases and thermal oil approaches zero. This behavior is observed for both Fluid A and Fluid B, however the ORC employing Fluid A is able to extract more energy before reaching its plateau. The reason for this can be understood considering Figure 6 and Figure 7. The temperature profile of the thermal oil for Fluid A configuration follows more closely the exhaust gases (larger ΔT between inlet and outlet and smaller mass flow rate). Thermodynamically the ORC cycle employing Fluid A is more efficient, meaning that it generates more power for the same heat input. Consequently, for the same ORC power output Fluid A would require a smaller heat exchanger or alternatively, for the same heat exchanger length its temperature profile allows the extraction of more heat.

Figure 8: Simulation results of the effect of thermal oil heat exchanger length on exhaust pressure drop and absorbed heat for both tested fluids.

Powertrain performance

The implementation of an ORC system in the powertrain of an off-highway vehicle can affect powertrain power, fuel consumption and the total weight of the vehicle. The aim of this paper is to explore the positive and negative aspects of an ORC-equipped powertrain.

In this study, the optimum ORC thermal efficiency was designed for approximately 20kWe for both fluids, which is presented in a previous study [24]. For different amounts of extracted heat calculated from the thermal-oil heat exchanger, the mass flow of the working fluid is equally changed without further optimizing the ORC system. Both the isentropic efficiency of the pump and the expander are assumed constant under all off-design conditions. This assumption expects to slightly affect the results, as the authors proved that the isentropic efficiency of a radial expander can be maintained high by controlling the generator rotational speed [25].

Figure 9 presents the effect of ORC net power as a function of exhaust backpressure. It is observed that at low exhaust backpressure values, a 10mbar increase in exhaust backpressure results in a significant increase in ORC power output of the order of 5kWe. However, at higher exhaust backpressure values a 10 mbar increase

in exhaust backpressure results in a 1kWe increase in the ORC net power output. In fact, the ORC power output presents the same plateau, which was observed at Figure 8, as it is correlated with the extracted heat of the thermal-oil heat exchanger. Furthermore, the selection of the working fluid seems to have a small impact on ORC net power. These results contrast with the observations made in the context of Figure 6 and Figure 7 Although Fluid A extracts more energy from the exhaust gases the overall ORC net power output between the two fluids is very similar. The reason for this is that Fluid B relies heavily on regeneration, equation (23) and consequently requires less heat for the same power output. A detailed explanation of this is outside the scope of this study, the interested reader should refer to [24]. It should be noted however that being heavily dependent on regeneration may have negative consequences as it requires a larger, heavier regenerator heat exchanger and a higher working fluid mass flow rate, increasing the overall weight of the system. This effect has not been included in this study. Finally, Figure 9 depicts the powertrain power output as a function of exhaust backpressure. Similar to the ORC net power case, powertrain power is increased at low backpressure conditions, while at higher backpressure values reach a plateau and at even higher backpressure values is expected to drop, as under extreme backpressure values the engine power drops.

Figure 9: Effect of thermal oil heat exchanger exhaust pressure drop on ORC net power and cumulative powertrain power.

An ORC system is not only an additional power assisted device but can also assist the powertrain fuel consumption. The implementation of an ORC system on powertrain can have both positive and negative aspects in terms of fuel consumption. Figure 10 illustrates both the negative effect of exhaust backpressure, due to the thermal-oil heat exchanger length, and the positive impact of the ORC system on engine bsfc. As exhaust backpressure is increased, the rate of the extracted heat from the thermal-oil heat exchanger is decreased which is depicted on the rate of fuel consumption of the integrated system. Furthermore, it is shown that different fluids have a small but not negligible effect on fuel consumption, especially under low exhaust backpressure conditions. Last but not least, further increase of the thermal-oil heat exchanger size is not expected to be beneficial for the fuel consumption of the integrated system, as the effect of the exhaust backpressure on fuel economy is more important to the small amount of the additional extracted heat. In fact, such a case seems to be non-realistic heat exchanger design and therefore was not included in the results of this study.

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Figure 10: Relationship between thermal oil heat exchanger exhaust pressure drop and powertrain fuel consumption.

Weight is another important parameter on the selection of the suitable WHR system. A high power WHR system is expected to weight more compared to a less powerful one. However, most the studies deal only with the thermodynamic characteristics of the WHR system and do not pay attention on the additional weight of the WHR system. In this study, it is possible to calculate the weight of the WHR system which is assumed to consist of the thermal-oil heat exchanger and the ORC system weight. Regarding the thermal-oil STHE, it was possible to calculate its weight for varying surface area. The ORC system weight was determined by adopting the empirical equation that was found in literature [17]:

YORC=14.641 XORC+40.087 (24)

Where XORC is the ORC system power in kWe and YORC is the ORC system weight in kg. Figure 11 illustrates the relationship between the weight and the power of a WHR system, considering the thermal-oil cycle. It is illustrated that as the weight of the WHR system is increased, the power output is almost linearly increased until 18 kW e, where the weight is increased more compared to the additional power benefit. The latter was expected as the heat exchanger needs to be much longer and heavier as the temperature difference between the hot and cold flow is decreased. In the same figure is also presented that the selection of the working fluids has a trivial impact on the total WHR system.

Figure 11: Relationship between WHR system weight and power output.

Although Figure 11 depicts the relationship between weight and power of the WHR system, it does not present the optimum weight-to-power ratio for the proposed WHR system. This is performed by Figure 12, which presents the power-to-weight ratio as a function of thermal-oil heat exchanger length. It is observed that the maximum power-to-weight ratio is given for a quite short thermal-oil heat exchanger and further increase of the heat exchanger length results in a drop on the power-to-weight ratio. However, the optimum heat exchanger length in terms of power-to-weight ratio can result only in an average benefit of 4.6% on fuel consumption and 5% on additional power output for both fluids, while for a slightly heavier WHR system that reach a ratio of 2 kWe/kg, the averaged benefit on fuel consumption for both fluids is 8% and the powertrain power is increased by 9%.

Figure 12: Effect of thermal oil heat exchanger length on power-to-weight ratio of the WHR system.

The design of the ORC system can be a trade-off between the effect of additional power, the fuel consumption benefit and the device weight, as schematically presented in Figure 13. The optimum ORC characteristics also depend on the application, as in the case of on-highway vehicles the weight increase may be a more important parameter under real driving engine conditions on the design of the WHR system compared to off-highway or stationary applications. In this study, it was selected to explore the relationship between the ORC design characteristics only for the ORC design engine operating conditions.

Figure 13: The WHR system design is a trade-off between ORC characteristics.

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The relationship between bsfc, considering both the exhaust backpressure and the fuel economy due to the ORC system, the extra ORC power and the additional weight is presented in Figure 14, as a result of varying heat exchanger surface area. It was found that up to 18 kWe of additional net power output, the WHR system weight increases linearly with the bsfc reduction and the additional power output. However, over this value, the temperature difference between the exhaust gas and the thermal oil is decreased enough so as to require a bigger and heavier heat exchanger. Differences between the two working fluids are not included in the results, as there is only a small difference at low thermal power points.

Figure 14: Relationship between fuel consumption, additional weight and extra power for the design of a WHR system.

Summary/Conclusions

The installation of an ORC WHR system in the powertrain of a vehicular application has both positive and negative aspects on performance and fuel economy. This study considers both aspects to explore the limits on the design of an ORC heat exchanger. For this reason, a heavy-duty diesel engine was employed to experimentally explore the effect of implementing a shell-and-tube heat exchanger on engine fuel consumption due to the additional exhaust backpressure. Then, a detailed heat exchanger model was utilized to calculate the trade-off relationship between the extracted exhaust heat and the exhaust backpressure by considering various heat exchanger volumes, while an estimation of the WHR system weight as a function of heat exchanger volume was also included in this study.

The selection of the working fluid is very important not only for the improved ORC efficiency but also for the compact design of the heat exchanger which can decrease its weight and its effect on exhaust pressure drop. This study showed that the maximum power to weight ratio of a WHR system is not always the optimum in terms of fuel consumption improvement and additional power, as 10% worst power-to-weight ratio can give up to 42.5% additional benefit on fuel consumption and power increase of the ORC system. The final trade-off of the ORC WHR system characteristics should be selected based on the vehicular application.

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Contact Information

Dr. Apostolos Karvountzis-Kontakiotisa.karvountzis@brunel.ac.ukDepartment of Mechanical, Aerospace & Civil EngineeringBrunel University London, Uxbridge, UB8 3PH, United Kingdom

Dr. Benjamin Franchettibenjamin@entropea.comEntropea Labs Ltd2A Greenwood Rd, London, E8 1AB, United Kingdom

Acknowledgments

The authors would like to acknowledge the financial support provided by Innovate UK through grant TS/M012220/1 in support of this project.

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Definitions/Abbreviations

WHR Waste Heat Recovery

T/C Turbocompounding

bsfc Brake specific fuel consumption

ORC Organic Rankine Cycle

STHE Shell and Tube Heat Exchanger

BPHE Brazed Plate Heat Exchanger

WHRS Waste Heat Recovery System

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