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Abstract
In the present work, an automotive Diesel engine has been
experimentally tested under a New European Driving Cycle
(NEDC) with the aim of getting experimental plots of time
dependent partitioning of energy injected during the warm-up
process. An additional objective of this work is to assess the
energy recovery capacity installed in the engine, i.e., to assess
how much of the energy that leaves the engine with the
exhaust gasses and the coolant is being employed. With this
target, mean values of some parameters (intake and exhaust
pressures and temperatures, coolant ow and coolant inlet and
outlet temperatures, engine speed and torque) together with
instantaneous variables (crankshaft angle, in-cylinder gas
pressure, intake and exhaust mass ows) were continuously
recorded during the warm-up of the engine. As a result of the
work, the dynamics of the thermal balance of the Diesel engine
under transient road conditions during the warm-up period was
obtained. Gross equivalent and detailed cumulative energy
ows were measured. The driving cycle averaged values of
exhaust gases and coolant energy rates make up 3.75 and
4.31 kW respectively in the engine tested. Thermal losses
account for more than 30 % of the input energy, while the
larger part of the input energy goes to the heating of engine
masses during approximately the rst third part of the NEDC
cycle. For the urban parts of the cycle the mean value of
exhaust gases temperature does not exceed 200°C, and the
corresponding averaged energy rates are 2.84 and 2.18 kW for
the exhaust gases and the coolant. The coolant temperature
takes approximately 720 seconds to reach 80°C. The mean
value of the sum of the energy recovered by the exhaust gas
recirculation (EGR) cooler and the passenger cabin heater
core is of the order 1.5 kW. The data obtained can be used to
establish nominal design parameters for efcient waste energy
recovery systems.
Introduction
Nowadays, internal combustion engines (ICE) has to overcome
many problems and new demands that have arisen over time.
One of the most important challenges facing ICE is the
improvement of fuel economy without compromising the engine
performance. A lot of research has been performed in this area,
from new injection systems and strategies passing through
inlet air temperature control [1], energy recovery systems (as
turbochargers) up to improvement of engine external systems
[2].
Concerning the last of these approaches is the engine thermal
management, one of the areas in which great research efforts
have been concentrated, over the last thirty years, with
technologies aimed at both improving cooling system
performance (including their interaction with other thermal
systems in the vehicles) and the development and application
of waste energy recovery devices in the engine, and in a
broader space, in the whole vehicle.
Referring to the improvement of cooling systems, the
researchers have centred on furthering cooling strategies in
order to ensure the passenger comfort (by means of the heater
core), reduce the engine warm-up time, and optimize the heat
dissipation in each of the engine operating points. Nowadays
these researches have been put into production, thanks to
advances in computer control technology, and the introduction
of electronically controlled electric valves, pumps, fans and
shutters in order to control properly coolant ows.
Some of the theoretical and experimental works related to the
study on cooling system improvement are those of Luptowski
et al. [3], Cortona [4], Hnatczuk et al. [5], Eberth et al. [6],
Chanfreau [7], Chalgren and Traczik [8], Chalgren and Allen
[9], Cho et al. [10], Pang and Brace [11], Norris et al. [12],
Robinson et al. [13], Chastain and Wagner [14], Brace and
Energy Balance During the Warm-Up of a Diesel Engine
2014-01-0676
Published 04/01/2014
Carlos A. Romero
Universidad Tecnologica de Pereira
Antonio Torregrosa, Pablo Olmeda, and Jaime Martin
CMT Universitat Politècnica de València
CITATION: Romero, C., Torregrosa, A., Olmeda, P., and Martin, J., "Energy Balance During the Warm-Up of a Diesel
Engine," SAE Technical Paper 2014-01-0676, 2014, doi:10.4271/2014-01-0676.
Copyright © 2014 SAE International
Downloaded from SAE International by Carlos Romero, Monday, April 14, 2014 11:04:12 AM
Burham [15], Robinson et al. [16], Koch and Haubner [17]. At
present manufacturers are working on oil and engine metal
temperature management.
Energy recovery strategies in the engine and in a broader
space, in the entire vehicle go through regenerator devices,
passing through engine insulation [18] and utilization of the
wasted energy at the tail pipe [19] and in the cooling systems
[20]. Wambsganss [21] reviewed technologies and research
that could be carried out to optimize thermal management
systems. The following waste heat recovery systems are
quoted: thermal storage heating and A/C system that is
charged from the vehicle's engine coolant or air conditioning
system, latent heat storage battery that collects and stores
engine waste, use of exhaust gases as an energy source in an
absorption refrigeration unit to provide cooling for the charge
air cooler, thermochemical heat pump (an absorption system
using a hydroscopic salt, Na
2
S) that can be charged with waste
heat taken from the coolant or from the exhaust gases, and
Rankine cycle to recover the energy available in the cooling
water to produce electric power or to contribute to the motive
power. Heat recovery can also be used for endothermic fuel
reforming reactions, in which exhaust gas thermal energy is
transformed into reformed fuel chemical energy, implying
improvements in overall engine efciency. Heat recovery might
also be used in thermal electric converters to power fuel-
reforming devices such as plasmatron.
Currently, thermoelectric devices are commonly used in a
variety of cooling and power generation applications. The
conguration of the heat-exchanger - thermoelectric energy
conversion system is nowadays under study in order to be
used in power train system. The main problem of this issue is
their cost [22].
Although the potential exists for substantial energy recovery at
full power engine output, passenger vehicles are not subject to
energy usage regulations [23]. Realistic duty cycles must be
examined to evaluate critically the potential energy recovery in
vehicle engines [22].
Even though regulations for medium- and heavy-duty truck fuel
economy already exist, there is no any standard or regulatory
policy related to engine or vehicle energy balance [23, 24, 25].
In order to make a contribution to the ways of measuring
engine energy ows and correlate them with energy savings
potential, in the present work we propose the measurement of
vehicle engine energy balance to be performed on a “driving
cycle” basis.
Usually, energy balances on engines are performed on
constant running conditions in order to guarantee stabilized
conditions and comparable results. Automotive engines,
however, always run on unpredicted driving conditions, which
is quite different from a succession of stabilized points. With
individual tests the estimation of total energy balance over a
city driving cycle is not very accurate and do not cover the
transient heat unbalances.
The normalized driving cycles are representative of the
common utilization of the commercial engines. Accumulated
measurements during the cycle can give a more
comprehensive insight into the energy distribution as compared
with discrete single operating point measurements. It can be
added also that assessing the energy ows during the engine
warm-up over an entire driving cycle, rather than focusing on
individual points of operation, may provide a more complete
and meaningful information for a global vehicle “driving cycle
based” energy saving strategy.
To the authors' knowledge, no experimental data relative to the
time-dependent energy balance of Diesel engines, under
driving cycle conditions, has been reported in the literature.
Many researchers have conducted engine thermal evaluation
under driving conditions to validate temperature prediction
models. Morel [26] exposed a detailed engine simulation model
to calculate engine thermal transients, but as can be inferred
from this work the model was validated using a particular
controlled transient (the coolant temperature was increased
exponentially from 300 K to 380 K, taking care of stabilizing the
points of measurements). Additionally, the aim of Morel's work
was not to obtain an energy balance.
An important contribution in the thermal balance of an engine
was made by Jarrier et al. [27], who enclosed a Diesel engine
in a calorimeter to record the instantaneous energy balance for
different running conditions, identifying the main heat paths.
Though it is suggested in the paper that the experimental is
harnessed to simulate ECE cycles, no experimental results of
the thermal balance under an ECE + MEVG driving cycle were
shown, but rather the mean heat balance during the warm-up
period at constant brake power was presented. Jarrier et al.
assumed that during the warm-up period, the instantaneous
in-cylinder pressures are identical for steady as well as for
unsteady test conditions, not considering the fact that during
the warm-up the mean indicated pressure is affected by the
friction mean effective pressure, which depends on oil viscosity
(i.e. on temperature).
Shayler et al [28] developed a computational model to predict
the warm-up characteristics of a combustion engine, using a
lumped capacity model as well as empirical correlations for
heat transfer (Taylor and Toong correlation) and friction losses
(Patton correlation). The main objectives found in other works
(Taymaz [29], Perez-Blanco [30], Lehner [31]) were different
from obtaining an energy balance; besides, steady-state
working points were used in these works.
One of the objectives of the present work is to contribute with
information relative to temperature gradients and heat uxes in
transient conditions. This information can be used in foreseeing
the technologies that best suits those levels of temperature
gradients and heat uxes, and also, as is the case of the
present work, to give an appraisal of the available energy that
can be recovered just to improve the overall energy efciency.
The paper will also evaluate the installed energy recovery
capacity in the vehicle. The cycle used to study the thermal
balance is the New European Driving Cycle, NEDC, which
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consists of four consecutive Urban Driving Cycles, UDC,
followed by one Extra Urban Driving Cycle, EUDC (according
to directive 98/69/EC).
In order to meet the mentioned objectives, the calculation of
the heat uxes in all the engine uids were performed by
means of both a combustion diagnosis model [32] and the
experimental measurements taken in an automotive Diesel
engine. The engine was mounted on a dynamometer test
bench that is capable to reproduce the normalized driving
cycle. Measurements of intake, exhaust, and coolant ows and
temperatures, fuel consumption, engine speed and torque
were used to calculate the engine brake power, the energy
rejected with exhaust gases and coolant, and also the energy
transferred in the heat exchangers of the cooling system (EGR
cooler and passenger heater). Though not described in this
paper, instantaneous in-cylinder pressure was used to
calculate in-cylinder temperatures and cylinder wall
temperatures to analyze the shape of the metal temperature
trace. A term was dened, called “miscellaneous” losses that
accounts for the part of the input energy remaining after
subtracting from the equivalent input energy the useful energy
and the energy lost with the exhaust gasses and the coolant,
so the energy balance is closed. The major parts of these
miscellaneous losses are the energy transferred to the
surrounding air and the energy stored by the engine masses;
losses associated to defective combustion may also be
included here.
The use of a combustion diagnosis model in transient
conditions faces with an important difculty. This is related to
the delays among the different measurements. As a
consequence, a special post-processing has to be carried out
to synchronize the signals from the different transducers [33],
e.g. in-cylinder pressure with air mass ow and fuel ow
measurements.
The paper is organized as follows: The rst part focuses on
both the description of the basis of engine energy balance and
the particularities of the experimental setup used in this work.
The second part of the paper is related to the experimental
results, together with the energy balances and the comments
about the factors that inuence both the warm-up and the fuel
energy use in general in the engine. Finally, the main
conclusions of this work are outlined at the end of the paper.
1. Theoretical and Experimental Considerations
In this section, an introduction of some particular aspects of the
engine energy balance is presented. Additionally, the
experimental setup used in this work is described.
Energy Balance Fundamentals
In an ICE engine, according to the First Law of
Thermodynamics, part of the introduced energy to the engine
(with the fuel and the intake air) is converted to effective power;
another part is transferred to the coolant, and the remaining
part leave the engine with the exhaust gases, as shown in
Figure 1. The energy conservation equation, for the engine
control volume, is [34]:
(1)
where and are, respectively, the fuel and air mass ows,
h refers to enthalpy, is the energy corresponding to the
input fuel, N
b
is the brake mean effective power, is the heat
transfer rate to the cooling uids, assembles the heat
rejected to the oil (if separately cooled) plus convection and
radiation from the engine external surface, and also the energy
stored by engine masses during transient processes. is
the exhaust gases energy. The friction losses are not present
in equation (1), since part of them is transferred to the cooling
system, while the remainder is transferred to the oil and to the
surroundings. During the engine warm-up, part (if not all) of the
miscellaneous losses ( ) are used for heating-up the
engine metal parts.
Figure 1. Schematic to describe the engine energy balance..
The thermal energy rejected by convection to the coolant is
calculated as:
(2)
where and c
pc
represent the mass ow and specic heat of
the coolant, respectively, while and are the inlet and
outlet coolant temperatures, respectively.
One way to assess the thermal energy transported with the
exhaust gases is multiplying the exhaust mass ow rate by the
difference between the exhaust and intake mass enthalpy-
temperature products:
(3)
During the engine warm-up operation part of the rejected heat
is stored by the metallic parts of the engine, and other part is
progressively transferred to the coolant according to the
expression:
(4)
Here, Q
rej
is the overall rejected heat transfer (W), and T
c
is the
mean coolant temperature.
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Experimental Setup
The engine used in this study is a high speed direct injection
(HSDI) turbocharged Diesel engine; its main characteristics are
shown in Table 1. The engine is mounted on a test bench, with
the required equipment and instrumentation to control the
engine operation and to measure the desired variables. The
dynamometer is an AC motor, operating at variable frequency,
with the ability to control braking torque with a very rapid
response. Specic software for the control of the dynamometer
allows load transients and simulations of road load. The
instantaneous variables (in-cylinder pressure, boost pressure,
exhaust pressure) were measured by means of a high
frequency acquisition system, up to 100 kHz per channel. The
chosen acquisition frequency was every 0.5 crank angle
degrees, i.e. 1440 data points every engine thermodynamic
cycle. So, the sampling frequency ranged from 15 kHz to 27
kHz, depending on the engine speed.
Table 1. Basic characteristics of the engine tested.
In order to synchronize precisely the acquired data of
instantaneous variables with the actual crankshaft position, the
output of an optical angle encoder with 0.1° crank angle
resolution was used as a sampling signal. Pressure variation in
the intake and exhaust ows was measured using water-
cooled piezoresistive transducers. In-cylinder pressure was
measured by means of a non cooled piezoelectric transducer.
The reference level for the pressure is xed by assuming that
the pressure at bottom dead centre (BDC) after the intake
stroke is equal to the mean intake pressure.
The fuel ow was measured or estimated by means of two
independent procedures. On the one hand, a gravimetric
balance that provides mass ow values at a frequency of 10
Hz was installed in the fuel line feeding the injection pump. On
the other hand, engine's electronic control unit (ECU) provides
an estimation of the volumetric fuel ow at a frequency of 200
Hz. In order to obtain an accurate measurement of the
instantaneous fuel mass ow delivered to the engine, the
combination of the two signals was required. The signals from
the load cell in the dynamometer and the inductive pick-up
speed sensor were used to calculate the effective power of the
engine.
The intake air mass ow was evaluated by a hot wire sensor
ow meter. This device yields directly values of mass ow rate
with a time response which is fast enough for the tests
requirements.
The coolant ow rate was measured with magnetic type ow
meters. Gas temperatures were measured with ungrounded
K-type thermocouples. Coolant and oil temperatures were
measured with metal sheath thermocouples. Exhaust gas
temperatures were measured with radiation shielded
thermocouples. Two thermocouples were placed in the cylinder
head of the engine to measure the wall temperature of external
surfaces in contact with the ambient air. The measured
external temperatures are used as reference for the engine
warm-up studies. The uncertainties of the relevant quantities
measured are given in Table 2.
Table 2. Uncertainties of the relevant quantities measured
Because of the importance of the cooling system design in the
scope of the present paper, a schematic of its conguration is
sketched in Figure 2.
Figure 2. Schematic of the engine cooling system. 1- engine structure;
2- radiator; 3- coolant pump; 4- pump driving mechanism; 5- electrofan;
6- thermostat; 7- outlet coolant box; 8- expansion tank; 9- thermal core
(passenger cabin heater); 10- EGR cooler; V1- bypass valve. T and F
- temperature and flow sensors.
The cooling system of the engine tested consists of the engine
structure, engine driven pump, expansion tank, engine oil
cooler, EGR cooler, radiator module and heater core, which
serve to dissipate the heat required to maintain the coolant
temperature and passenger cabin at the desired level, and
thermostat, designed to permit the coolant to ow to the
radiator once the engine reaches the operating temperature.
The coolant mass ows through the cooling circuit change in
proportion to engine speed, thanks to the mechanical gearing
of the coolant pump.
2. Experimental Energy Balance
Energy Input to the Engine
Figure 3 shows the measured fuel ow rate, engine speed and
the NEDC cycle prole. The fuel ow rate follows the pattern of
the driving cycle consisting of various partial load working
conditions, which are mild conditions for the engine. Assuming
a complete burning process, i.e. all the supplied to the engine
cylinders fuel is entirely converted to thermal energy, the plot of
the latter results in a scaled version of the fuel ow rate. In this
work the complete thermal energy released has been called
the equivalent energy input with the fuel ( ), and is
computed as the fuel ow rate times the fuel heating value
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H
u
(42500 kJ/kg). The energy input with the air is not taken
here into account since it will be accounted for in the
computation of energy losses with exhaust gases.
Figure 3. Measured fuel flow rate as a function of engine speed and
NEDC driving cycle profile.
Energy used for Effective Power
The readings of the dynamometer torque and crankshaft
rotational speed are used to calculate the engine brake power:
N
e
= M · πn/30, where M is the engine torque, and n is the
crankshaft rotational speed. The averaged driving power for
the whole cycle is 4.75 kW, while the maximum value of this
power is about 40 kW (at the extra urban cycle). In addition,
the maximum power demanded at the urban cycles is 20 kW.
Energy Rejected with the Coolant
Coolant temperatures and ows were measured during engine
testing at different points of the engine, the passenger heater,
and the EGR cooler, as Figure 2 shows. Figure 4 shows the
evolution of mean block, inlet and outlet coolant temperatures
through the engine block. The knowledge of coolant
temperatures and the coolant ow allowed to calculate, by
using expression (2), the heat carried off with the coolant
mass, as it passes through the engine water jacket under the
warm-up operating conditions. The plot of these heat losses is
presented in gure 4, where for comparison purposes the fuel
equivalent energy rate has been also reproduced. The trends
in both plots are very similar, suggesting some kind of linear
relationship between them. This fact could be worthy for engine
control algorithms.
The warm-up period of the engine is almost 720 s, time at
which coolant temperature is 80°C that corresponds to the
thermostat opening temperatures. For the four ECE parts of
the NEDC cycle the averaged value of energy rates to the
coolant is 2.18 kW.
Figure 4. Heat rejected to coolant calculated after measuring coolant
flow through the engine and inlet and outlet coolant temperatures. The
trace of the fuel equivalent energy and mean block temperature has
been also plotted.
The time traces of the coolant and block temperatures give an
idea of the thermal inertia of the engine system roughly
described by the equation (4). In fact, a complex identication
process could lead to the identication of the thermal time
constants of the system, and their relationships with engine
and coolant masses. This work is beyond the scope of the
present work, though some qualitative comments on the
inuence of the engine masses will be added later.
Energy Evacuated with the Exhaust Gases
Measured air and fuel ow rates, and also intake and exhaust
gas temperatures were used in the equation (3) to calculate
the energy evacuated with exhaust gases. The evolution of
intake air and exhaust gas temperatures together with the
calculated exhaust gas energy are presented in gure 5. The
mean value of exhaust gas temperature is lower than 200°C
during the four urban cycles (780 s), after that time this
temperature raises up to 500°C.
Miscellaneous Term of the Energy Balance
The residual of the difference between the equivalent energy
rate input and the sum of the calculated brake effective power,
coolant and exhaust gas losses corresponds to the
miscellaneous term in equation (1), i.e. the term that closes the
energy balance. Figure 6 shows the behavior of the
components of this difference during the warm-up of the engine
under NEDC. In gure 7 the time evolution of the
miscellaneous term is plotted, the major part of which goes to
heating the engine.
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Figure 5. Heat evacuated with exhaust gases.
Figure 6. Equivalent energy input with the fuel compared to the
cumulative of the energy converted to useful work and energy
evacuated with coolant and exhaust gases.
The averaged values for exhaust and coolant energies at the
whole cycle are (3.75 ± 0.10) kW and (4.31 ± 0.15) kW,
respectively, considering the uncertainties associated with the
measurements. Thermal losses account for more than 30 % of
the input energy, while the larger part of the input energy goes
to the heating of engine masses.
Figure 7. Experimental miscellaneous energy rate obtained after
subtracting the cumulative of effective power and the rates of energy
evacuated with exhaust gases and coolant from the equivalent energy
rate input with the fuel.
Figure 8 shows the average percentage, with respect to the
equivalent input energy (energy available at fuel), of the four
aforementioned energies, i.e., effective power, exhaust gases,
coolant, and miscellaneous losses. Unfortunately, there was
not enough information to calculate in detail the miscellaneous
losses (heat exchanged with lubricant oil, thermal energy
storage and ambient losses), but gross results are in
agreement with Jarrier's results [27], who quoted that during
the rst minute of the engine warm-up about 65 % of the
energy of combustion is used for engine mass heating, being
negligible the heat transfer from the engine to the ambient. As
the time passes the share of the energy transferred to the
ambience increases, reaching an important value. An
estimation of the energy transferred to the surrounding air at
the end of the last ECE part of the driving cycle entails 1.8 kW,
calculated with the experimental external block temperature
(Figure 4), an average ambient temperature of 24 °C, and an
average convection coefcient between the engine block and
the air of 35 W/(m
2
K). This heat corresponds to a 12% of the
input energy, giving an idea of its importance.
Figure 8. Driving cycle averaged energy balance of the engine.
From Figure 8, it is clear that exhaust gases and coolant heat
losses increase from the cold start until the thermal
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stabilization is reached. This is due to the heat demand of the
cold metallic parts, which behave like a thermal storage
reservoir. In addition, during the warming up, the miscellaneous
energy decreases, but the energy transferred to the
surrounding air increases as the engine temperature increases.
The relative importance of the miscellaneous component of the
energy balance as compared to measured losses during the
warm-up can be followed in the Figure 9.
Figure 9. Relative importance of the miscellaneous losses as
compared to the Cumulative energy supplied with the fuel and the total
of useful energy and energy wasted with exhaust gases and coolant.
As is derived from Figure 8, for the rst third part of the driving
cycle the non effective energy decomposes equally between
coolant and exhaust gas energy losses, and miscellaneous
losses. After the rst 720 seconds the miscellaneous losses
tend to stabilize to a value conditioned by the losses to the
surrounding air (close to the value estimated before), but just
after this period the extra urban part of the cycle demands
more energy and the miscellaneous losses again increase.
Assuming that the miscellaneous term is entirely used for
increasing the engine block temperature, and knowing the
average engine temperature (the engine external surface
temperature measured), the thermal capacitance of the engine
can be estimated, i.e. the equivalent engine thermal mass,
after approximating the engine thermal response to a second
order dynamic process. Conversely, knowing, for instance, a
desired reduction in the miscellaneous losses, and giving the
target stabilized engine temperature during the warm-up, the
reduction of the engine mass required to meet that target can
be found. In this way, the signicance of reducing the engine
mass in the reduction of engine warm-up time is illustrated.
Figure 8 allows concluding about the reduction of engine
masses as a measure to increase the available heat, and also
to reach the nominal operating temperature of the engine in a
shorter time. The increase of the available heat could be used
also to improve the passenger cabin comfort during the
warm-up. In engine vehicles there is a compromise between
reaching the shortest possible engine warm-up and also
reaching sufcient passenger cabin heat performance, which is
linked with the heat received by the coolant.
3. Energy Recovered in the Vehicle Engine
under Test
As it was mentioned in the introduction, the second objective of
this paper aims at the calculation of the energy recovered in
the vehicle engine under study. In the previous section the
thermal energy evacuated with the exhaust gases and the
coolant at the outlet of the engine was measured and plotted.
As it has been represented in Figure 2, the engine cooling
system comprises three heat exchangers: engine radiator to
dissipate the non used engine coolant heat to the air, EGR
cooler to cool down the exhaust gases that are returned to the
engine cylinders, and heater core installed to warm the air of
the vehicle passenger cabin.
Actually the exchange of heat in the radiator takes place only
after the temperature imposed by the thermostat valve
characteristics is reached, which is achieved in the engine
tested after approximately 720 seconds of running operation,
as follows from Figure 4. After this time the thermostat valve
starts to regulate the coolant ow diverting part of the coolant
through engine radiator. The coolant dissipates part of its
internal energy in the radiator without any utility in return.
Therefore all the energy dissipated in the radiator is potentially
subject to recover, once the thermostat valve has opened.
As for the EGR cooler, equation (2) was used to calculate the
coolant energy gains in this device, given the values of coolant
ow and coolant temperatures at the inlet and outlet ports of
this device. The same equation was applied to nd out the heat
transferred from the coolant to the air of the passenger cabin.
So, this are two devices that in fact recover energy separately
from exhaust gases (the EGR cooler), and from coolant (the
passenger cabin heater). The timed energy rates recovered by
these devices during a driving cycle are plotted in the graphics
of 10.
It follows from Figure 10 that energy exchanges in EGR cooler
and passenger heater are of comparable magnitude, and not
being restricted to engine emissions control the EGR cooler
might provide adequately the energy required for the
passenger comfort. In the engine tested, the EGR cooler suits
only the purposes of emissions control, cooling the exhaust
gases and transferring the resulted thermal energy to the
coolant. The heat evacuated by the engine that is truly being
recovered is the energy used for passenger cabin heater.
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Figure 10. Heat recovered from the exhaust gases in the EGR cooler
and heat positively used to heat the passenger cabin.
The sum of the heat exchange rates taking place in the
mentioned EGR cooler and passenger heater can be
reasonably seen as the capacity of the cooling system of the
engine under test to recover thermal energy from exhaust
gases and coolant during the warm-up period. In Figure 11 the
percent of energy “recovered” during the engine warm-up is
presented. The average value of the energy “recovered” during
the operation of the engine under city loads is of approximately
20 percent (approximately 1 kW) of the compounded heat
losses carried off with the coolant and the exhaust gases. This
would be the energy actually recovered, provided that the heat
from the recirculated exhaust gases was transferred to another
independent owing medium different from the coolant that
circulates through the passenger cabin heater.
Figure 11. Percent of the energy recovered in the tested vehicle engine
as related to the sum of energy wasted with exhaust gases and
coolant. Part of the energy otherwise disposed with exhaust gases and
coolant is recovered in the EGR cooler and passenger heater; though
strictly speaking these recovered energies are coupled.
The energy recovered by the turbocharged has not been
accounted here because it makes part of the enclosed
controlled volume for which the energy balance was done.
Analysis of the data from the tested engine has yielded design
energy balance information. Their full potential will not be
realised, however, unless an integrated approach is taken to
the entire engine. By controlling the heat/energy ow in the
engine and reducing component drive power requirements, the
gains are potentially very large. Decoupling the cooling system
components (fan, pump) from the engine crankshaft by driving
them with electric motors, and replacing mechanically actuated
thermostats by electronically controlled valves, it is possible to
improve the system warm-up performance and reduce the
unnecessary waste of power.
A signicant amount of the heat carried away by the exhaust
gases and coolant could be recovered.
Summary/Conclusions
Two were the motivations of the paper herein presented. The
rst one was to present a “diving-cycle-based” approach for the
evaluation of energy balance of a Diesel engine. The second
motivation consisted on the assessment of the energy
recovered in the engine cooling system under warm-up
operating conditions.
Even if the ideas presented may not seem to be new, this study
is, to the authors' knowledge, the rst one that attempts to
explore the instantaneous energy balance during real driving
conditions, and specically under a NEDC driving cycle, with
the idea of revealing the dynamics of energy partitioning, and
also to give an estimate of the energy losses averaged along
the cycle.
The driving cycle averaged values of exhaust and coolant
energy rates are 3.75 and 4.31 kW, respectively. Of this energy
approximately 20 % is being recovered by EGR cooler and
passenger heater. The available average energy rate is of the
order of 4 kW. The cumulative energy losses corresponding to
the sum of energy evacuated with exhaust gases and that
dissipated by the coolant at the end of the driving cycle, make
up 9500 kJ approximately. The cumulative miscellaneous
energy losses are around 8000 kJ. These gures show the
potentials that exist for substantial energy recovery, once
appropriate means to manage the wide variations in exhaust
and coolant mass ows illustrated in this paper are brought into
production. It is the quality of the energy that is being lost what
delays the development of energy recovery systems for
automotive engines.
It is seen from the energy balance that the miscellaneous term
is the major constituent of the energy losses during the
warm-up period. Of this miscellaneous term the metal thermal
storage capacity plays the largest role in the control of the
warm-up time, until the engine temperatures have stabilized.
Hence a reduction in the engine mass entails a reduction in the
warm-up time and a reduction of friction losses, which traduces
in a reduction of the pollutant emissions and fuel consumption.
Finally, the energy recovered in a commercial passenger
vehicle mounting the tested engine was examined.
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Definitions/Abbreviations
C
pbl
- block material specic heat, J/kg·K
C
pc
- coolant specic heat, J/kg·K
- coolant heat capacity rate
h
exh
- specic enthalpy of exhaust gases, J/kg
h
int
- specic enthalpy of intake mass, J/kg
- intake air mass ow rate, kg/s
m
c
- coolant mass, kg
m
bl
- external block mass, kg
- coolant mass ow rate, kg/s
N
b
- effective engine brake power, W
- fuel mass ow rate, kg/s
- heat exchanged between the coolant and the cylinder wall,
W
- cycle-averaged heat ow from the cylinder enclosure
through the combustion chamber walls, W
t - time, s
T
c
- mean coolant temperature in the water jacket, K
- coolant temperature at engine inlet, K
- coolant temperature at engine outlet, K
T
exh
- exhaust gas temperature, K
T
int
- intake gas temperature, K
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