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Theoretical and experimental evaluation of electric
coolant pump benefits in real driving cycles
To cite this article: Di Bartolomeo Marco
et al
2023
J. Phys.: Conf. Ser.
2648 012079
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ATI-2023
Journal of Physics: Conference Series 2648 (2023) 012079
IOP Publishing
doi:10.1088/1742-6596/2648/1/012079
1
Theoretical and experimental evaluation of electric coolant
pump benefits in real driving cycles
Di Bartolomeo Marco, Di Battista Davide, Cipollone Roberto
University of L’Aquila – Via Giovanni Gronchi 67100, AQ, Italy
marco.dibartolomeo2@univaq.it
Abstract. Engine thermal management is a promising option to reduce fuel consumption and
harmful emissions of Internal Combustion Engines. This is particularly suitable to support the
transition towards a carbon-neutral transportation sector, considering that the role of combustion
engines is expected to persist in the near and medium future. In this study, a prototype pump
electrically actuated was compared to a mechanical pump of a downsized gasoline engine that
propels a real vehicle. In the first phase of the analysis, the cooling circuit was tested from a
hydraulic point of view on all its branches using an engine mounted on a bench equal to that
working on the vehicle. The hydraulic circuit was fully characterized via pressure transducers
and flow meters in all branches for different thermostat lifts, representing different coolant
temperatures. On the same bench, the OEM pump and an electrically actuated one, suitably
redesigned on an operating point more representative of the real operating conditions, were
tested. A vehicle propelled with the same tested engine (having a conventional mechanically
actuated pump) was run on the road following three different driving cycles. The engine
revolution speeds were registered, as well as the temperature of the cooling fluid. The electric
and mechanical pumps were compared using the performance maps previously obtained. The
electric pump speed was set to deliver the same coolant flow rate as the OEM pump, following
the same sequence of thermostat lifts. The results show that a 60 % average reduction of the
pump energy consumption is possible, leading to an average specific CO2 emission reduction of
1 g/km. This result is even more relevant during urban driving, with emission savings hitting 2.5
g/km.
1. Introduction
In the transportation sector, the research is mainly involved in developing innovative technologies and
systems to reduce CO2 emissions and harmful pollutants. Electric and hydrogen-based powertrains
represent a solution to these issues but also present some challenges. Indeed, Battery Electric Vehicles
(BEV) are not really carbon-free since the energy mix primarily uses fossil fuels [1]. Moreover, an
adequate infrastructure must be built to convert entirely the present circulation vehicle fleet [2].
Until then, there are several opportunities to improve the efficiency of current transportation systems,
i.e., internal combustion engines. These can be summarized but are not limited to combustion
improvement [3,4], use of biofuels [5], Waste Heat Recovery (WHR) systems [6–8], and engine
downsizing [9]. This latter is particularly relevant regarding mechanical efficiency improvement, the
reduction of parasitic loss [10] and the decrease of propulsion power. Furthermore, downsizing can be
successfully integrated with Internal combustion Engine hybridization leading to even more significant
fuel economy benefits [11].
In addition to the above-referenced technologies, Engine Thermal Management (ETM) plays a
significant role. In ETM systems, the flow of the different fluids evolving inside an internal combustion
engine is controlled to achieve engine efficiency improvement and CO2 emission reduction while
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guaranteeing, at the same time, a reliable operation of the engine. Indeed, these systems are usually
adopted to regulate the engine temperature by managing the air, oil and coolant flows. In most ETM
applications, the ultimate goal is the reduction of the engine warm-up time. Indeed, this latter is
associated with multiple benefits, including higher mechanical efficiency, fuel economy and
consequently lower CO2 and pollutant emission. In [12], the warm-up time of the oil has been reduced
by splitting the oil sump into two sections and reducing the active oil volume for faster heating. In [13]
and [14], the exhaust gases have been used for the same purpose. When these kinds of systems do not
induce any detrimental effects, the faster heating of the oil determines the reduction of the oil viscosity
and, consequently, an increase in the mechanical and overall efficiency of the engine.
On the other hand, managing the coolant flow optimizes the cylinder wall temperature, increasing
the combustion efficiency and oil thickness, which sticks to the walls. The potential to have different
average engine block and head temperatures could increase engine performance. A colder engine head
than the block could favor a compression ratio increase which increases, by definition, engine efficiency.
The friction produced by the compression and oil rings decreases with benefits on mechanical efficiency.
This goal can be pursued by considering separate circuits for the engine head and block [15], the
modification of the cooling circuit to speed up lubricating oil warm up [16], new solutions for the
thermostat [17], mechanical coolant pumps coupled with the engine shaft through electromagnetic
clutches [18] and flow control through coolant pumps with additional controllable guide vanes [19].
Indeed, conventional pumps pose severe limitations due to the impossibility of controlling the coolant
flow according to the engine temperature. Electric coolant pumps overcome this issue, providing the
possibility to deliver the desired coolant flow rate [20], increasing the overall engine efficiency and
providing a reliable operation through knock mitigation [21] and after-boiling phenomena reduction
[22,23]. Furthermore, adjusting the coolant flow rate according to the engine’s thermal needs can reduce
the radiator size, [24]
The advantages of electric coolant pumps are not limited to the optimization of the coolant flow. In
mechanical or belt-driven pumps, the design point must be chosen according to the maximum power
conditions of the engine. This feature results in the frequent off-design operation of the pump during
real driving, contributing to poor performance, being the engine almost never operated close to the
maximum power. When a coolant pump is electrically actuated, the design point can be set more freely
because eventual increases in flow rates can be obtained by increasing revolution speeds. This aspect
has been evaluated in [25], where a 15% pump power reduction was achieved, designing the pump
according to the most frequent engine operating conditions faced by the engine during a WLTP type-
approval cycle. Furthermore, making the pump speed independent from the engine allows the use of
volumetric pumps, whose efficiency is less dependent on speed compared to the more conventional
centrifugal pumps.
Volumetric pumps are widely used in the engine lubricating circuit [26], but only a few applications
can be found for engine cooling. In [27], a triple-screw pump for engine cooling applications was tested
and modelled, achieving a specific CO2 emission reduction of up to 2 g/km compared with conventional
centrifugal pumps during a WLTP cycle. Similar results have also been found in [28], where a sliding
vane pump was considered.
This study compared a mechanically and an electrically actuated centrifugal cooling pump, focusing
the attention on the energy absorbed during three real driving cycles. A cooling circuit of a downsized
turbocharged gasoline engine was reconstructed on a test bench, accurately reproducing all the branches
and all the components usually present (thermostat, radiator, oil and heater core). The circuit was
instrumented with a set of pressure sensors which allowed the evaluation of the hydraulic resistances of
all the components and branches. Two flow meters were used on the main branches, one just downstream
of the pump and the other upstream the radiator. On this circuit, the OEM pump and an electrically
actuated pump were operated, evaluating their performances at different thermostat lifts, which have
been varied using suitable spacers. The electric coolant pump was suitably redesigned to match the
cooling needs at a medium engine load, modifying the conventional design criterion, which considers
the engine maximum power.
A vehicle propelled with a similar engine was run on three real driving cycles with urban, rural and
motorway sections in different proportions. During the travel, engine speed was continuously recorded,
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as well as coolant temperature. The sequence of the working points of the pump has been determined,
having previously evaluated the thermostat lift during opening and closing as a function of the coolant
temperature. Once known the pumps characteristic curves and the sequence of operating conditions, the
same coolant flow rate has been fixed by properly actuating the electric pump. In this way, the electric
and mechanical-driven coolant pump has been compared through a model-based procedure in each
driving cycle.
Instantaneous pump power was compared for both pumps. Furthermore, the overall mechanical
energy and the CO2 emission associated with the pumps’ operation were assessed by integration over
the cycle. The analysis provided a reliable evaluation methodology that uses few experimental data to
assess the performances of newly conceived pump prototypes (in terms of types and actuation strategy)
under real driving scenarios.
2. Material and methods
2.1. Experimental test bench development
The first phase of the analysis was dedicated to the characterization of the engine cooling circuit of the
vehicle under consideration. A sample of the same engine was installed on a specific bench. All the
components of the cooling circuits were added to reproduce the layout and hydraulic characteristics
closely. Care has been given to reproduce the exact height of the elements they have in the original
(OEM) engine on the vehicle. Only slight modifications were made in the sections of the connecting
pipes where sensors have been installed. Pressure transducers were mounted upstream and downstream
of every component to evaluate its hydraulic resistance. Furthermore, two flowmeters were used to
measure the volumetric flow rate in each branch of the cooling circuit.
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Figure 1: Experimental test bench layout.
As reported in Figure 1, two separate branches can be observed. When the thermostat is closed, the
whole cooling water delivered by the pump flows towards the engine, the heater core and the water/oil
heat exchanger. When the thermostat opens, part of the flow rate goes toward the radiator, where the
fluid is cooled before mixing with the remaining portion and entering the coolant pump. A tank is placed
upstream of the pump in the same position it has on the vehicle to avoid cavitation phenomena
pressurizing the cooling circuit and, primarily, to increase the boiling temperature of the cooling fluid.
2.2. Experimental characterization
Once the cooling circuit was reproduced, the OEM pump was experimentally characterized. This
latter is mechanically coupled with the engine shaft. In order to run the pump at different speeds, the
mechanical junction connecting the pump and the shaft was disconnected and the mechanical power
was provided externally through a 1.5 kW electric motor. A variable frequency drive ensured the speed
variation. A torque meter was placed between the pump and the electric motor shaft to evaluate torque
and pump speed. Different thermostat opening degrees were set using spacers of different lengths.
The results of the experimental campaign are represented in Figure 2. Efficiency contours have been
reported every two percentage points showing values between 6 % and 16 %. The Best Efficiency Point
(BEP) is located when the thermostat is fully opened, where the coolant flow rate is maximum. This
feature confirms a common approach in designing engine coolant pumps actuated mechanically. Indeed,
the coolant pump must ensure a sufficient flow rate when the engine is running at maximum power
when the shaft drives it. However, this operating condition rarely occurs in real driving situations. As a
result, poor performances are achieved in the most common operating points, as Figure 2 clearly reports.
Figure 2. OEM (a) and prototype pump (b) characteristic curves. Speed curve are represented
every 500 rpm from 1000 rpm to 6000 rpm in (a) and from 1500 rpm to 7000 rpm in (b).
Knowing the hydraulic characteristics of the cooling circuit and the OEM pump performances, an
electrically actuated centrifugal pump has been designed using a new design criterium and prototyped.
The newly designed centrifugal pump, of course, is smaller than its mechanical counterpart, so higher
efficiencies can be achieved at low-load operating conditions. The experimental characterization has
been conducted in the same test bench, slightly modifying the cooling circuit to place the pump in the
same position it would occupy in the engine compartment of the considered vehicle. The input power
was provided by a direct current electric power supplier. The electric power absorption has been
measured. The results of the experimental characterization are reported in Figure 2 b. It can be noted
that the best efficiency point is reached at 60 l/min and 1.6 bar in a hydraulic curve between the
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Journal of Physics: Conference Series 2648 (2023) 012079
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maximum and minimum lift of the thermostat, in contrast with the original pump, showing the highest
efficiency at 110 l/min and 2 bar. Although the overall efficiency of the electric pump also depends on
the efficiencies of the electric components, a much greater performance is reached in the BEP region
(46 % vs. 16 %), which certifies the improvement of the new design. Indeed, being the electric pump
independent from the engine shaft, it can be located far from it, without integrating the volute inside the
engine block (as it frequently happens), avoiding unnecessary pressure losses due to intricate flow
passages.
Figure 3. Thermostat lift as a function of temperature in opening and closing conditions.
In order to take into account the opening of the thermostat, which is significant from a hydraulic
point of view, the relationship between the thermostat lift and engine coolant temperature has been
investigated. A correlation has been identified on the bench, heating the thermostat through an electric
power supplier simulating in this way the presence of the hot fluid inside. The stem temperature through
a surface thermocouple has been measured. For each temperature, the lift was measured during both
heating and cooling in order to catch an expected typical hysteresis cycle. The results are reported in
Figure 3. The thermostat starts opening at 80 °C, reaching the maximum lift at 120 °C showing that the
real engine operates at high coolant temperatures. Indeed, for a given temperature, the thermostat lift is
higher when the thermostat is closing, determining an operating region that increases the inertia of the
thermostat, making it less sensitive to sudden coolant temperature variations and stabilizing the pump
operation.
2.3. Evaluation procedure
An evaluation procedure has been developed after the characterization of the engine cooling circuit and
of the operating curves of the OEM pump and the prototype electrical one. The aim of the analysis is
the comparison of the two pump options over real driving cycles.
The comparison has been organized in this way according to Figure 4:
1. The vehicle equipping the tested engine runs a real driving cycle on the road. The sequence of
engine speeds (Neng) is registered over time, as well as the most relevant data (coolant
temperature, fuel consumption) from ECU and OBD. The knowledge of the transmission ratio
(τoem) gives the sequence of the pump speeds (Npump) during time;
2. The thermostat lift is calculated knowing the coolant temperature and its behavior (temperature
increasing/decreasing) from Figure 3;
3. Once pump speed and thermostat lift are known, the data previously obtained from the engine
tested in the bench give the volumetric flow rate (Q) delivered by the OEM pump and the
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pressure head (ΔP) during the real driving. The mechanical power (Pmech) of the OEM pump can
be derived.
4. Once Q and ΔP are known for each operating condition, the working points of the electric pump
can be fixed. Since the prototype pump is smaller than the OEM one, the pump speed will be
higher than that of the OEM. A unique value of the speed of the electrical cooling pump is easily
identified, knowing its characteristic curves. The power absorbed by the electric coolant pump
can be derived from the data of the testing activity previously done and, definitively, the energy
absorbed over the overall driving cycle;
5. Finally, the knowledge of the different absorbed power makes it possible to compare the
performances and evaluate the potential energy savings.
A flowchart of the different phases of the procedure is reported in Figure 4.
Figure 4. Evaluation procedure flowchart.
It is worth noting that when the electric pump is operated via speed of rotation, a fine control of the
flow rate circulation delivered by the pump is obtained. So, the setup of the same sequence of flow rates
provided by the OEM cooling pump during real driving following the three driving cycles was easily
reached. The same pressure at the pump outlet at the same flow rates on all branches were guaranteed
since the cooling circuit on the bench was the same and experimental data were obtained at different
thermostat lifts. The potentiality of the electric coolant pump to precisely deliver the desired flow rate
by varying its revolution speed gives a reasonable expectation of when the flow rate control will be used
for engine thermal management.
3. Results
3.1. Definition of driving cycles
The developed methodology has been applied using experimental data obtained in three different driving
cycles, with different characteristics in terms of driving profiles and similar initial conditions. The
reference vehicle is equipped with a downsized gasoline engine, which is particularly interesting for its
hybridization potential. Its main characteristics are listed in Table 1.
OBD data have been acquired in all cases to evaluate the differences between the different paths.
Each test starts in the same location with the engine at ambient temperature, evolving differently as a
function of the separate destination. Engine torque and speed are used to evaluate the instantaneous
mechanical power of the vehicle and cumulative energy absorbed over the driving cycle. Furthermore,
vehicle speed has been used to differentiate the urban, rural and motorway sections. The different
sections have been distinguished according to the European Real Driving Emission measurement
legislation. Vehicle speed ranges are 0 km/h to 45 km/h for the urban part, 45 km/h to 80 km/h for the
rural part and 80 km/h to 130 km/h for the motorway section.
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Table 1. Engine and vehicle characteristics.
Displacement
1300 cc
Number of cylinders
4 in-line
Maximum power
110 kW
Charging
Waste-Gated
Turbocharger
Fuel
Gasoline
Mixture formation
Direct Injection
Kerb weight
1430 kg
Number of gears
6
Transmission
Automatic
The driving cycles analyzed in this study show different characteristics to evaluate the potential
benefits of the electric coolant pumps in terms of energy saving in a broad set of operating conditions.
Figure 5a reports the speed profile as a function of time. It can be observed that all of them show similar
vehicle speeds in the first part, because of the common starting point and the shared urban road. After
approximately 600 seconds, average speed profiles differ. Driving cycle “A” maintains an average speed
below 80 km/h for almost the whole test. In driving cycle “B”, despite the similar duration, a motorway
section is provided, with instantaneous vehicle speeds between 80 km/h and 120 km/h in the last part of
the cycle. Finally, driving cycle “C” has a similar trend to driving cycle “B”, with an extended duration
of the rural and motorway section. Synthesis results are reported in Figure 5b. All three driving cycles
have a similar urban section with a total length of approximately 4 km. No motorway section is present
in A, with a rural part lasting 11 km. On the other hand, the motorway section lasts 5 km in driving cycle
B at the expense of a shorter rural part compared to the previous cycle. Finally, both the motorway and
the rural section are extended in driving cycle C, with a duration of 13 km and 11 km, respectively.
Figure 5. Vehicle speed as a function of time (left) and urban, rural and motorway distances for each
driving cycle (right).
As already observed the cumulative energy for each driving cycle can be evaluated from engine
speed and torque measurements. The mechanical power is usually higher during the motorway section
because of higher engine speed and torque. This trend can be observed in Figure 6a, where the
cumulative energy is represented for each driving cycle. After 600 seconds, the cumulative energy in
driving cycles B and C shows a sharper increase in the motorway section due to the higher average load
needed in these parts to reach higher vehicle speeds. The overall energy is higher in cycle C,
characterized by the longest distance and the highest average load because of the extended rural and
motorway section.
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In contrast, cycles A and B show similar cumulative energy of around 4 kWh despite being
characterized by significant differences in the driving profiles. This aspect is also outlined in Figure 6b,
where the energy share associated with each section is represented. As previously noted, driving cycle
A does not have a motorway section, and the rural part dominates the absorbed overall energy with a
share of 80 %. On the other hand, in driving cycles B and C, the most significant part is related to the
rural and motorway sections, especially for the latter, where their contribution approaches 90%.
Figure 6. Cumulative energy as a function of time (left) and energy share of urban, rural and
motorway sections for each driving cycle (right).
3.2. Performances evaluation
Knowing the engine speed and the coolant temperature from ECU data and evaluating the thermostat
lift at each instant thanks to the previous testing activity (Figure 3), the performances of the OEM pump
in terms of flow rate, pressure head and mechanical power can be assessed. The power absorbed by the
redesigned electric coolant pump can be determined by imposing the same flow rate acting only on its
revolution speed. The same hydraulic resistance guarantees the same pressure head (which is the one
requested by the circuit). The opening of the thermostat caused by the increased temperature of the
cooling fluid during driving condition is evaluated when the correlation between temperature and
thermostat lift is known (Figure 3).
The results of the analysis in terms of instantaneous power and cumulative energy required during
the driving cycle are reported in Figure 7. For driving cycle A, the average power absorbed over the
cycle is under 300 W for the OEM pump. A peak of power can be seen when the engine load increases,
determining higher engine speed and consequently increasing the pump speed proportionally. This
feature is more evident in cycles B and C, where driving sections with higher average loads are present.
In driving cycle B, the first peak identifies the beginning of the motorway section, where a higher engine
load is usually needed to enter the highway. After the first peak, the average power remains higher
compared with the first part of the cycle with values higher than 200 W. Similar considerations can also
be applied to case C. However, multiple peaks occur, mainly caused by the more significant vehicle
speed variations in the highest-load section as reported in Figure 5.
Energy absorption over the cycles reflects the trend observed for the engine. This parameter is the
highest in driving cycle C with a total energy equal to 180 kJ, followed by driving cycle A with no
motorway section, which stands at 140 kW and driving cycle B at 115 kJ. It is worth noting that this
similarity is not straightforward. Indeed, the power required by the vehicle is a function of engine torque
and speed, which increases when a given vehicle speed must be reached in a limited time. In contrast,
the power the pump absorbs is a function of rotational speed and thermostat lift. When these parameters
increase, higher flow rates are delivered, thus increasing the power requests of the pump.
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Figure 7. Cumulative energy and absorbed power by the prototype and OEM pump for driving cycles
A, B and C as a function of distance.
Concerning the redesigned electrical cooling pump, the electric power required is lower all over the
cycle. This aspect is particularly evident during power peaks and the highest-load regions but also
applies to the initial part of the driving cycles, mainly consisting of alternating urban and rural sections.
Power peaks reach a maximum of 250 W, with average values between 30 W and 70 W, depending on
the different vehicle loads. It is interesting to observe that in driving cycle A, despite being characterized
by a 50 % lower distance than C, the energy absorbed by the pump is only 13 % lower.
The energy savings are represented in Figure 8a. The maximum energy savings are reached in driving
cycle A, where the prototype absorbs 61% less energy than the mechanical pump. Similar results are
also achieved in the other driving cycles, with 58% and 56% energy savings in driving cycles B and C,
respectively. Compared to the energy absorbed by the vehicle due to propulsion, the OEM pump
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cumulative energy represents a portion ranging from 0.7% to 1.1 %. However, evaluating potential
specific CO2 emission savings gives birth to interesting considerations.
Fuel rate from ECU data has been integrated over the cycle and for the urban, rural and motorway
sections. Considering an average gasoline density of 0.76 kg/l, a lower heating value of 43.128 MJ/kg
and a CO2 emission factor of 3.16 kgCO2/kgfuel in agreement with IPPC data [29], the mass of CO2
emitted in each section can be evaluated. Knowing the distance covered in each section and comparing
the performances of the OEM and prototype pumps, CO2 emission savings are evaluated. The results
are reported in Figure 8b. The avoided specific CO2 emissions range from 0.8 g/km and 1.5 g/km, with
the most significant potential associated with the urban section. This feature is valid for all the driving
cycles, but it is particularly evident in driving cycle A, where the specific CO2 emission reduction stands
at 2.65 g/km. The main reason for this behavior must be found in the engine’s operating conditions at
the beginning of the driving cycle. The engine operates at lower efficiency when cold-started,
determining higher fuel consumption. As a result, specific fuel consumption and CO2 emission increase.
Consequently, energy savings are far more relevant if the engine operates in off-design conditions,
leading to higher specific CO2 reduction. It is worth noting that these results have been obtained
considering the electric power absorbed by the batteries thus taking into account the efficiency of the
electric motor and the electronics. More accurate evaluations are needed to consider the battery and
alternator efficiencies. However, very high efficiencies can be reached with 48-volt batteries. On the
other hand, the alternator efficiency strongly varies with the operating conditions and its performance
should be compared to the drive belt system used in conventional applications. In any case, the results
obtained in this analysis are even more critical if evaluated in the EU legislative framework. Considering
that starting from 2021, car manufacturers must pay a fine of 95 Euro for every gram of CO2 emitted
per kilometer (on a fleet-wide average basis), the energy-saving potential of coolant pumps electrically
actuated has a non-negligible economic value. Furthermore, it is worth observing that more significant
energy savings can be obtained if the electrical cooling pump is operated according to an optimized
control strategy which is oriented to reduce engine warm up time.
Figure 8. Energy saving for driving cycles A, B and C and specific emission reduction in each section
of the driving cycles.
4. Conclusions
This study assessed the potential benefits of a centrifugal pump electrically actuated in real driving
conditions. The analysis required identifying a reference vehicle and developing a methodology that
used an experimental and theoretical approach to determine reliable results. The OEM pump of the
reference vehicle was characterized in an experimental test bench, suitably built to reproduce the engine
cooling circuit precisely. The performance map and all the flow rates on the different cooling circuit
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branches were obtained. Having taken care that the cooling circuit on the vehicle and that on the bench
are equal from a hydraulic point of view, the testing activity gave the possibility to identify all the
parameters of the cooling circuit on the vehicle only knowing the revolution speed of the pump (which
was mechanically linked to the engine shaft). The effect of the cooling temperature (which was not
reproduced on the bench but registered during the real driving of the vehicle) varies the hydraulic
performance of the cooling circuit due to the thermostat opening and closing. This effect has been
reproduced on the bench by artificially opening the thermostat with suitable spacers.
An electrically actuated cooling pump was redesigned to match its best efficiency point at a medium-
load engine condition (instead of considering the maximum engine power as it is conventionally done).
This new pump was operated on the same test bench and a comprehensive data set was obtained. The
same delivered flow rate and pressure of the OEM pump can be easily reproduced by acting only on the
speed of rotation of this redesigned pump.
The energy savings potential has been evaluated considering three different driving cycles consisting
of urban, rural and motorway sections in different proportions.
The redesigned electrical cooling pump determined an average of 60 % lower energy absorption,
especially in driving cycles consisting mainly of urban and rural sections. Compared to the energy
absorbed by the vehicle due to propulsion, the OEM pump cumulative energy represents a portion close
to 1 %, which is a negligible result in terms of energy. However, according to the driving considered,
the CO2 emission reduction ranges from 0.8 g/km to 1.5 g/km. This result is even more encouraging if
the urban section is considered. In this case, specific CO2 emission reduction increases up to 2.5 g/km,
determining an even higher economic value if European standards are considered.
The obtained results contribute to the energy reduction only due to the pump redesign and the
electrical actuation. A more significant effect is expected when the flow rate control of the cooling fluid
is used to reduce the engine warm-up time and optimize energy flows on board.
5. Acknowledgements
This study has been conducted in the framework of the funded project PEMSO: Pompe Elettriche
per una Mobilità SOstenibile (Electric Pumps for a Sustainable Mobility). The authors are grateful for
the support of the company Metelli S.p.A. in the person of Dr. Sergio Metelli President of the Company.
6. Nomenclature
BEP
Best Efficiency Point
OBD
On-board Diagnostic
cum
cumulative
OEM
Original Equipment Manufacturer
E
energy
P
power
ECU
Electronic Control Unit
prot
prototype
el
electric
Q
volumetric flow rate
eng
engine
T
temperature
ETM
Engine Thermal Management
WHR
Waste Heat Recovery
mech
mechanical
ΔP
pressure head
N
rotational speed [rpm]
7. References
[1] IEA, World Energy Outlook 2022, Paris, 2022. https://www.iea.org/reports/world-energy-
outlook-2022 (accessed June 1, 2023).
[2] IEA, The Future of Hydrogen - Seizing today’s opportunities, 2019.
https://www.iea.org/reports/the-future-of-hydrogen (accessed June 1, 2023).
[3] A. De Vita, L. Allocca, A. Montanaro, G. Meccariello, F. Duronio, S. Ranieri, Under-
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