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
Page 1 of 10
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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
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.
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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
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.
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:
𝑄̇
𝑄̇𝑚𝑎𝑥
(1)
Where 𝑄̇𝑚𝑎𝑥 = (𝐶𝑝 𝑚̇)𝑚𝑖𝑛 (𝑇𝑜𝑖𝑙,𝑖𝑛 − 𝑇𝑔𝑎𝑠,𝑜𝑢𝑡 ) 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 𝑄̇ is obtained
from:
𝑄̇ = 𝑚̇𝑜𝑖𝑙 (ℎ𝑜𝑖𝑙,𝑖𝑛 − ℎ𝑜𝑖𝑙,𝑜𝑢𝑡 ) = 𝑚̇𝑔𝑎𝑠 (ℎ𝑔𝑎𝑠,𝑖𝑛 − ℎ𝑔𝑎𝑠,𝑜𝑢𝑡 )
(2)
The number of transfer units (NTU) is a dimensionless number
defined by:
𝑁𝑇𝑈 =
𝑁𝑇𝑈 =
1
𝜀−1
𝑙𝑛 (
)
(1 − 𝐶𝑟 )
𝜀𝐶𝑟 − 1
𝐶𝑟 =
(𝐶𝑝 𝑚̇)𝑚𝑖𝑛
(𝐶𝑝 𝑚̇)𝑚𝑎𝑥
𝐻𝑜 = 𝐻𝑖𝑑𝑒𝑎𝑙 (𝐽𝐶 𝐽𝐿 𝐽𝐵 𝐽𝑅 𝐽𝑆 )
𝐻𝑖𝑑𝑒𝑎𝑙 = 𝑗
In this analysis, the heat exchanger length is increased iteratively
until 𝑈𝐴 determined 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:
𝑚̇𝑜
𝜇 0.14 −2/3
𝐶 ( )
𝑃𝑟
𝑆𝑚 𝑝 𝜇𝑠
(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:
𝐻𝑖 =
𝑁𝑢 𝑘𝑐
𝐷𝑖
(10)
Where the Nusselt number Nu for turbulent flow (𝑅𝑒 > 1 × 104)
inside a tube can be determined from the Gnielinski’s correlation
[20]:
𝑁𝑢 =
Using equations (1-5) it is possible to determine 𝑈𝐴 the overall heat
transfer coefficient across a surface area.
(8)
Where: 𝐽𝐶 , 𝐽𝐿 , 𝐽𝐵 , 𝐽𝑅 , 𝐽𝑆 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:
(4)
(5)
(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 𝐻𝑜 is estimated using the BellDelaware method [19]:
(3)
And for a counter-flow configuration the NTU can be determined
from the effectiveness by:
(6)
and the sum of resistivities can be determined from:
𝑈𝐴
(𝐶𝑝 𝑚̇)𝑚𝑖𝑛
1
∑𝑅𝑡
𝐷
𝑙𝑛 𝑜
1
1
𝐷𝑖
𝑅𝑡 =
+
+
𝐻𝑖 𝐴𝑖 2𝜋𝐿𝑘 𝐻𝑜 𝐴𝑜
Thermal-oil heat exchanger modeling
𝜀=
𝑈𝐴 = 𝑈𝑜 𝐴𝑜 = 𝑈𝑖 𝐴𝑖 =
𝑓
(𝑅𝑒 − 1000)𝑃𝑟
8
1
(11)
2
𝑓 2
1 + 12.7 ( ) (𝑃𝑟 3 − 1)
8
𝑓 = (0.79 ln(𝑅𝑒) − 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:
Δ𝑃𝑜 = Δ𝑃𝑐 + Δ𝑃𝑤 + Δ𝑃𝑒
(13)
The pressure drop in the cross flow central regions of the STHE is
determined from:
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2𝑓𝑝 𝑁𝑐 (
Δ𝑃𝑐 = (𝑁𝐵 − 1)
𝑚̇𝑜 2
)
𝑆𝑚
𝜇 0.14
𝜌( )
𝜇𝑠
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 𝑓𝑝 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:
Δ𝑃𝑤 = (𝑁𝐵 )
(2 + 0.6𝑁𝑤 )(𝑚̇𝑜 )2
𝑅𝐿
2𝜌𝑆𝑚 𝑆𝑤
(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:
2𝑓𝑝 𝑁𝑐 (
Δ𝑃𝑒 = 2
𝑚̇𝑜 2
)
𝑆𝑚
(1 +
𝑁𝑤
)𝑅
𝑁𝑐 𝐵
(16)
𝜇 0.14
𝜌( )
𝜇𝑠
The pressure drop in the tube-side of the heat exchanger is
determined from the Zigrang and Sylvester correlation [21]:
(𝜌𝑢2 )
Δ𝑃𝑖 = (−2 log(𝑎))−2
𝐿
2𝐷𝑖
(17)
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
Where u measures the velocity of the fluid inside the tube and the
constant a is determined from:
2𝑒
5.02
2𝑒
13
𝑎= (
)−(
) log (
)+( )
7.54𝐷𝑖
𝑅𝑒
7.54𝐷𝑖
𝑅𝑒
(18)
Pump: The pump is assumed to be electrically powered; in the
absence of leakages and for adiabatic operation, mass and energy
balance are:
𝑃𝑝𝑢𝑚𝑝 =
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.
𝑚̇𝑤𝑓 (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.
𝑚̇𝑜𝑖𝑙 (ℎ𝑜𝑖𝑙,𝑖𝑛 − ℎ𝑜𝑖𝑙,𝑜𝑢𝑡 ) = 𝑚̇𝑤𝑓 (ℎ5 − ℎ1 )
(20)
Equation (20) expresses the total energy balance of the evaporator.
Turbine: The turbine power output of the turbine expander is
calculated by:
𝑃𝑡𝑢𝑟𝑏𝑖𝑛𝑒 = 𝜂𝑡𝑢𝑟𝑏𝑖𝑛𝑒 𝑚̇𝑤𝑓 (ℎ5 − ℎ6,𝑖𝑠 )
(21)
Condenser: As mentioned, the energy balance of each component
allows to control the pinch point temperature. The mass balance is
expressed by equation:
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𝑚̇𝑤𝑓 (ℎ6 − ℎ1 ) = 𝑚̇𝑐𝑤 (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):
𝑚̇𝑤𝑓 (ℎ6 − ℎ6𝑎 ) = 𝑚̇𝑤𝑓 (ℎ2𝑎 − ℎ2 )
(23)
Experimental Setup
Figure 3: Qualitative presentation of the diesel engine exhaust waste heat map
and the selected engine operating point for the ORC design analysis.
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.
Results and Discussion
Effect of the exhaust backpressure on engine
performance
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.
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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
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.
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.
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 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.
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 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 offhighway 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
Page 7 of 10
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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.
Although Figure 11 depicts the relationship between weight and
power of the WHR system, it does not present the optimum weightto-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 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 thermaloil cycle. It is illustrated that as the weight of the WHR system is
increased, the power output is almost linearly increased until 18 kWe,
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.
Page 8 of 10
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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 onhighway 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.
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.
References
[1]
Johnson T V. Review of Vehicular Emissions Trends. SAE
Int J Engines 2015;8:1152–67. doi:10.4271/2015-01-0993.
[2]
Wilson DE. The Design of a Low Specific Fuel
Consumption Turbocompound Engine 1986.
doi:10.4271/860072.
[3]
Tennant DWH, Walsham BE. The Turbocompound Diesel
Engine 1989. doi:10.4271/890647.
[4]
Brands MC, Werner JR, Hoehne JL, Kramer S. Vechicle
Testing of Cummins Turbocompound Diesel Engine 1981.
doi:10.4271/810073.
[5]
Hountalas DT, Katsanos C, Lamaris VT. Recovering energy
from the diesel engine exhaust using mechanical and
electrical turbocompounding. SAE Tech Pap 2007:776–90.
doi:10.4271/2007-01-1563.
[6]
Zhang J, Zhang H, Yang K, Yang F, Wang Z, Zhao G, 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 Convers Manag 2014;84:282–94.
doi:10.1016/j.enconman.2014.04.036.
[7]
Shu G, Yu G, Tian H, Wei H, Liang X. Simulations of a
Bottoming Organic Rankine Cycle (ORC) Driven by Waste
Heat in a Diesel Engine (DE). SAE 2013 World Congr
Exhib 2013. doi:10.4271/2013-01-0851.
[8]
Quoilin S, Broek M Van Den, Declaye S, Dewallef P,
Lemort V. Techno-economic survey of organic rankine
cycle (ORC) systems. Renew Sustain Energy Rev
2013;22:168–86. doi:10.1016/j.rser.2013.01.028.
[9]
Katsanos CO, Hountalas DT, Pariotis EG. Thermodynamic
analysis of a Rankine cycle applied on a diesel truck engine
using steam and organic medium. Energy Convers Manag
2012;60:68–76. doi:10.1016/j.enconman.2011.12.026.
[10]
Walraven D, Laenen B, D’Haeseleer W. Comparison of
shell-and-tube with plate heat exchangers for the use in lowtemperature organic Rankine cycles. Energy Convers Manag
2014;87:227–37. doi:10.1016/j.enconman.2014.07.019.
[11]
Mastrullo R, Mauro AW, Revellin R, Viscito L. Modeling
and optimization of a shell and louvered fin mini-tubes heat
exchanger in an ORC powered by an internal combustion
engine. Energy Convers Manag 2015;101:697–712.
doi:10.1016/j.enconman.2015.06.012.
[12]
Di Battista D, Mauriello M, Cipollone R. Effects of an ORC
Based Heat Recovery System on the Performances of a
Diesel Engine. SAE Int 2015;1608:1–11. doi:10.4271/201501-1608.
[13]
Bastani P, Heywood JB, Hope C. A forward-looking
stochastic fleet assessment model for analyzing the impact
of uncertainties on light-duty vehicles fuel use and
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 tradeoff of the ORC WHR system characteristics should be selected based
on the vehicular application.
Page 9 of 10
7/20/2015
emissions. SAE Tech Pap 2012;20:1–33. doi:10.4271/201201-0647.
ECOS 2016, 2016.
[25]
[14]
Cheah L, Bandivadekar A, Bodek K. The trade-off between
automobile acceleration performance, weight, and fuel
consumption. SAE Int 2008;1:771–7. doi:10.4271/2008-011524.
[15]
Clark NN, Khan ABMS, Wayne WS, Gautam M, Thompson
GJ, Mckain DL, et al. Weight Effect on Emissions and Fuel
Consumption from Diesel and Lean-Burn Natural Gas
Transit Buses. Eng Conf 2007:776–90. doi:10.4271/200701-3626.
[16]
Wang L, Kelly K, Walkowicz K, Duran A. Quantitative
Effects of Vehicle Parameters on Fuel Consumption for
Heavy-Duty Vehicle 2015:0–7. doi:10.4271/2015-01-2773.
[17]
Usman M, Imran M, Yang Y, Park BS. Impact of organic
Rankine cycle system installation on light duty vehicle
considering both positive and negative aspects. Energy
Convers Manag 2016;112:382–94.
doi:10.1016/j.enconman.2016.01.044.
[18]
Quoilin S. Past and current research trends in ORC power
systems. ASME ORC 3rd Int. Semin. ORC Power Syst.,
Liege: n.d., p. Keynote speech.
[19]
[20]
Bell K, Muller. Final Report of the Cooperative Research
Program on Shell and Tube Heat Exchangers ( Bulletin No.
5, University of Delaware Engineering Experimental Station
Newark, Delaware, 1963). 1963.
Gnielinski, A. AZ and AS. Banks of plain and finned tubes,
in Heat Exchanger Design Handbook, Hemisphere
Publishing Corp., New York; 1988.
[21]
Zigrang DJ, Sylvester ND. Explicit approximations to the
solution of Colebrook’s friction factor equation. AIChE J
1982;28:514–5. doi:10.1002/aic.690280323.
[22]
Mavrou P, Papadopoulos AI, Stijepovic MZ, Seferlis P,
Linke P, Voutetakis S. Novel and conventional working
fluid mixtures for solar Rankine cycles: Performance
assessment and multi-criteria selection. Appl Therm Eng
2015;75:384–96. doi:10.1016/j.applthermaleng.2014.10.077.
[23]
Mavrou P, Papadopoulos AI, Seferlis P, Linke P, Voutetakis
S. Selection of working fluid mixtures for flexible Organic
Rankine Cycles under operating variability through a
systematic nonlinear sensitivity analysis approach. Appl
Therm Eng 2015;89:1054–67.
doi:10.1016/j.applthermaleng.2015.06.017.
[24]
Franchetti B.; Pesiridis A.; Pesmazoglou I.; Tocci L.;
Sciubba E.; Thermodynamic and technical criteria for the
optimal selection of the working fluid in a mini- ORC.
Page 10 of 10
7/20/2015
Alshammari, F.; Karvountzis-Kontakiotis A. PA. Radial
Turbine Expander Design for Organic Rankine Cycle, Waste
Heat Recovery in High Efficiency, Off-Highway Vehicles.
3rd Bienn. Int. Conf. Powertrain Model. Control, 2016.
Contact Information
Dr. Apostolos Karvountzis-Kontakiotis
[email protected]
Department of Mechanical, Aerospace & Civil Engineering
Brunel University London, Uxbridge, UB8 3PH, United Kingdom
Dr. Benjamin Franchetti
[email protected]
Entropea Labs Ltd
2A 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.
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