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HEAT REJECTION AVOIDANCE IN COMBUSTION ENGINES
Wagner Matos Santos1, Juliano de Assis Pereira2, Josef Klammer3, José Antonio Perrella
Balestieri4, Alex Mendonça Bimbato4, Marcelino Pereira do Nascimento4
1UNESP/FEG – Universidade Estadual Paulista/Faculdade de Engenharia de Guaratinguetá -
Volkswagen Caminhões e Ônibus;
2MAN Truck and Buses / Volkswagen Caminhões e Ônibus;
3MAN Truck and Buses;
4UNESP/FEG – Universidade Estadual Paulista/Faculdade de Engenharia de Guaratinguetá.
wagnermasa@hotmail.com, juliano.pereira@man.eu, josef.klammer@man.eu,
jose.perrella@unesp.br, alex.bimbato@unesp.br, marcelino.nascimento@unesp.br
ABSTRACT
Heat rejection to a cold reservoir is inherent of thermal machines operation. One of them, the
engine cycle, aims to deliver liquid power and its efficiency increases directly proportional to the
hot reservoir and inversely proportional to the cold reservoir. Most of the studies for increasing
engine efficiency leads to increase the hot reservoir. This research instead, instigates the
discussion of a better use of energy and heat rejection avoidance on internal some combustion
engines systems. It demonstrates the principle that the heat rejection improvement using liquid
instead of air ran against the real objective that should have been using it better. The paper shows
feasible design steps with physical, energetic and economical approach to implement an engine
gas recirculation system, which uses tail pipe gases, through a supersonic Laval nozzle parallel
to intake line after charge air cooler. It also purposes waste heat recovery from engine liners as
heat source for a Rankine cycle. All calculus and simulations are done based on a real engine
data submitted to an European Transient Cycle.
INTRODUCTION
The studied engine is an inline 4 cylinder, 4,5
liter, 4 stroke, turbocharged, water cooled,
reciprocating diesel engine. It uses overhead
valves and a common rail with 1800bar
injection pressure for compression ignition
(Figure 1). The choice of a 4 cylinder engine
for this study was due to high load factor that
this hardware is subjected to. As it leads to a
high heat rejection amount, in terms of
percentage, it is expected to reach more
energy recovery than on a 6 cylinder engine
with higher volume and lower load factor. In
addition, the choice is also strategically
better, since low packaging available will
protect the design for bigger Engines. The
base data for this study was obtained from a
ETC objective measurement.
Figure 1: Engine baseline for study
The most common waste heat recovery studies are related to the exhaust gases. For example,
BORGWARNER, (2014) uses an organic Rankine cycle with exhaust pipes and engine
recirculation systems as main Energy sources. CUMMINS, (2008) presented a complete study in
the 2008 DEER (Diesel Engine Efficiency Emissions Research) demonstrating through test
benches analysis that it would be possible to decrease the fuel consumption in ca. 10% by using
organic Rankine cycle to recover energy from exhaust gas streams. However, with the current
legislation EURO 5-CONAMAP7, resolution 403/2008, and the new legislation EURO 6
(current in EUROPE), the opportunity to use this energy is getting lower because the after
treatment emissions processes are demanding high temperatures to allow the chemical reactions
that decreases emissions level.
Based on a real engine measurement submitted to an ETC cycle, it is possible to notice that a high
amount of energy is lost through the cooling process. As soon as a conventional Rankine Cycle
shows efficiency about 30 % and assuming to recover 10% of the energy lost on the cooling
system, the heat balance for the actual proposal is capable to recover up to 10 kW of liquid power.
A mathematical analysis has been performed using GT Power software enable simulations of
different kind channels profiles and disposition around combustion chamber. For the EGR system
without EGR cooler, a complete ETC cycle was calculated with Excel and VB. It was possible do
demonstrate that the use of an supersonic Laval nozzle would be a feasible solution for mixing
gases from low pressure line to the high pressure line on intake clean air flow.
The super heater for the Rankine cycle will assume some the dimensional boundary conditions
from the available engine. As the conduction area will be decreased on the liners
walls (Figure 9), the heat transfer area “lost” for the forced convection cavities will need to be
compensated in order to keep the internal combustion chamber temperature.
The application of this mechanism brings potential to suppress the conventional cooling system
or at least downsize it. The concept of this super heater is developed together with the definition
of a fluid that matches the proposed Rankine cycle.
The energy recovered using this system, converted into decrease of fuel consumption, will need
to overcome the investments to worth it the Customer TCO (Total Cost of Ownership).
The development of this energy recovery system is being supported from a known commercial
vehicle OEM in Brazil. All theory studies were done using a 4.5 Liter Engine.
1. GENERAL INFORMATION
All results presented in this chapter are based in a real ETC cycle of an 4 cylinder 4.5 liter
diesel engine that is currently in production. Circa of 18.000 measurement points were analyzed.
No confidential strategy data is showed in the results and the specific know how is registered on
the patent request number BR102019002884.
1.1 Baseline Engine Energy Consumption
Thermal machines presents low efficiency. Reciprocating combustion engines are not
different and heat is wasted on the conversion of the chemical energy into mechanical. The most
explored recovering alternatives are related to the exhaust gases, but this work intends to
demonstrate the feasibility of recovering energy from the cooling system, that is nowadays it is
barely explored from scientific community.
An interesting way to observe the heat balance is analyzing the complete vehicle. [2]
describes it in his paper and concludes that ca. of 62% of the fuel demand is converted in Themal
Energy (Figure 2), which, nowadays is wasted. This is exactly why the researches about WHR
are so intensive at this moment.
Figure 2: Vehicle Energy flow diagram, Hay, et. al., 2011
For the engine in study, the coolant
system rejects less heat. Figure 3
organizes the energy distribution
showing the corresponding spent at
cruise speed of 1750 rpm with full
load. It shows that 28% from the
Energy generated by the fuel
combustion is rejected through the
radiator.
Figure 3: Energy Distribution on cruise speed of 1750rpm
28% of rejection on the radiator corresponds to 105 kW. This is the result of the calculation based
on the inlet and output temperatures of the Charge Air Cooler (CAC). For the current situation,
the relation Q = m.c.∆T is determinated from:
QRad = Energy rejection on the Radiator (28%
of the fuel energy generation) = 105 kW
Q = QH = 105 kW
m = 270 kg of coolant per minute
c = 3,74 kJ/kg.K (From coolant at working
temperature – data provided from supplier)
∆T = 6,2 K
Figure 4: Cooling cycle [32]
It is important to notice that at this point (CAC), the coolant flows at high rates, which leads to
a high surface coefficient on forced convection. This convection receives the heat from engine
coolant chamber also at high rates. On this region (around liners) the temperature gradient is
extremely high showing a big chance of waste heat recovery.
1.2 Heat Recovery from Linners
As the research assumes the combustion chamber temperature as an important boundary
condition, the heat transfer flow must kept, as well as the material behavior due to thermal fatigue.
With this system, it is possible to generate work using the engine liners as a steam generator for a
Rankine cycle. A side effect will also be noticed: it possible at least to downsize the original
cooling system or even suppress it.
Figure 5 shows that the average
temperature from exhaust gases are
around 405oC reaching peaks of
630oC. The exhaust manifold itself
reaches temperatures above 500oC.
Figure 5: Temperature from exhaust gases before turbocharger
Figure 6: Temperature around the exhaust gases flow
inside engine
Analyzing the cylinder walls, the path of the
exhaust gases and its heat transfer to the coolant
is showed in Figure 6. The Bottom Dead Center
(BDC) wall temperature is considered 1000C
and the highest temperature of 5000C is the
contact between exhaust manifold and engine
head. This path means the complete flow of
exhaust gases inside the engine (crankcase and
cylinder head). Notice that there is not a linear
model to simulate forced convection from
Exhaust gases to cylinder wall and from
cylinder wall to the coolant. The cylinder
temperature gradient needs to be considered on
the simulation.
Figure 7: Combustion Chamber and Heat
rejection flow to the coolant
Figure 8: Combustion Chamber and Heat rejection flow to
the coolant
Notice in Figure 7 that the combustion chamber walls faces a permanent energy flow that comes
from the chamber to the coolant galleries. The combustion chamber walls can reach temperatures
up to 250oC. As the coolant evaporation temperature is around 1100C on the max system
pressure, the external chamber wall cannot be over this temperature (TS). The
estimated temperature distribution through the liners walls in a specific thickness position is
shown in Figure 9. Assuming that the engine combustion chamber temperatures cannot be
impacted from the project, the chamber heat extraction must always be kept independently on the
way it is extracted from the walls.
If QS could eventually assume the
QH value, it would mean
completely suppressing the
conventional coolant system. This
schema is part of the hole project
conclusion and can be seen in
Figure 29. The heat extraction
from engine liners will be limited
by the heat rejection from the
material and channel profile used
around the liners (heat
exchanger). The
construction of the super heater
would also influent on the heat
transfer. The concept of a liner
with internal channel could
Figure 9.
Figure 9: Proposal for combustion Chamber and Heat rejection flow
To evaluate if an alternative heat extraction could eventually substitute the complete cooling
system, an analysis of the coolant galleries was performed. The engine coolant galleries are
represented in Figure 10.
Figure 10: Engine current coolant Galeries
Assuming the engine as a control volume,
the mass flow (in/out) is 270kg per min. It
flows constantly through the galleries with a
complete heat exchange surface of 4,4 m2.
Despite the losses through convection
around the engine and periphery, the heat
transferred to the coolant in the engine in
study is 105 kW which would mean roughly
23,86 kW per m2. It is clear that not all the
surface is in contact with a high heat source,
but considering the turbulence and high
speed flow, this estimation is a confident
value for the objective in discussion.
Figure 11: Fluid phase change inside a heated circular
tube [24]
For the simulation, round section channels
(Figure 11) were considered around the
liners with thermal conductivity equal to the
current crankcase cast iron. The heat transfer
with phase transfer inside a circular tube is
defined by [24]:
Figure 12: Liner temperature profile
Figure 13: Convection correlations for flow in a circular tube
[21]
The liner temperature profile is showed in Figure 12. The OEM design proposal for the super
heater itself cannot be yet shown in this thesis due to rights protection, but the simulations confirm
the feasibility of this proposal, which has already been tried and registered on other patents [23].
The consideration for internal flow into circular tubes are based on Figure 13.
Using the GT Power software with application of the chosen geometry and premises, 7 different
scenarios were simulated trying to find a way to get super-heated fluid on the liners output. All
cases were analyzed considering the cast iron thermal conductivity, ethanol as working fluid, same
distance between the channels rings, 10 bar of pressure input, 800C of
temperature input, 400 discretization steps and same liner diameter. The term channel geometry
is a composition or channel diameter with distance to chamber wall;
Figure 14: Liner evaporator: Case 1 [39]
Figure 14 shows case 1, which considered channel geometry 1 (R=68,5 mm and channel diameter
of 3,5 mm). The mass flow is 1g/s. The simulation showed a laminar flow and low pressure loss,
but low heat transfer (0,2 kW) and no evaporation.
The case 2 uses the same channel geometry from case 1, but higher mass flow (2g/s). The result
was an almost completely evaporated output (1,74 kW). 9.9 bar output.
The case 3 used the same geometry from case 2, but increased mass flow to 5g/s. It resulted in
less evaporation than case 2 (approximately 42%), but had a high heat output absorbed by the
working fluid (2.61 kW). It would be an interesting condition to work in series downstream of
one next evaporator. For example post after treatment system.
The case 4 can be good compared with 2, since the mass flow is also 2g/s and just the channel
diameter was changed to 3 mm (Channel geometry 2). It leaded to a slightly better heat exchange
(1,84 kW), but increased pressure drop (9,7 bar).
The case 5 increases mass flow to 5g/s keeping geometry 2. It improved heat exchange (2.67 kW
at output) but showed higher pressure drop (9.1 bar output).
The case 6 is better comparable with case 4 because it changes just wall thickness to chamber
assuming the channel geometry 3 and mass flow. It was noted a sensible improvement in heat
transfer (1,94 kW) and complete phase change with superheated outlet. This condition is indicated
for installation of expander in series upstream evaporator.
The case 7 kept channel geometry 3 and, in relation to case 6, the mass flow is increased to 3 g/s.
A sensible improvement of heat transfer was observed.
Heat transfer ability increases progressively on the seven simulated steps, showing the influence
of the channel geometry as well as the mass flow ratio.
Based on the data from case 7, the spring profile around the liners presents a 13,7 m long tube
with 3 mm diameter and total shell area of 0,13 m2. Considering the total heat transfer ability of
2892 W, the heat extraction is 22,2 kW/m2.
Smaller channels diameter improves heat transfer at low and medium mass flows due to turbulent
flow. Low wall thickness improves heat absorption of the working fluid, limitation by material
strength. Liner evaporator is suitable for generation of steam or mixed steam for subsequent
superheating in other evaporator in series. Considerable heat absorption into the working fluid.
Smaller channels diameter increases pressure loss, especially in the superheated area. Limitation
by the permissible wall temperature prevents overheating at higher pressures.
The mathematical model was simplified for a round section in order to use existing empirical data,
but improvements on the profile and use of better conductivity materials can increase the potential
transfer ability.
The evaporator to run a Rankine cycle must fulfill the dimensional boundary conditions from the
available engine liners. As the conduction area will be decreased on the liners walls, like showed
in the model from Figure 9, the heat transfer area “lost” for the forced convection cavity needs to
be dimensioned in a way to compensate the Heat extraction from combustion chamber. This
situation can potentially mean QH = QS (Figure 9) and is necessary to keep the internal
combustion chamber temperature.
The working fluid for the purposed Rankine cycle shall be ideally the same from the cooling
system, which shall be easily available on the market. Nevertheless, the fluid that matches this
requirement (ethanol), brings a lot of concerns regarding safety of the system.
The use of ethanol in a Waste Heat Recovery (WHR) system from the OEM shows corrosion and
performance effectiveness to support the decision of using it in a OCR cycle. A system, at this
moment not integrated to the cooling, is now running tests, but unfortunately the specifications
and safety features are non-disclosure at the moment.
With the definition of the cycle fluid, one positive displacement pump needs to be specified in
order to control the working fluid flow and an expander can be defined based on the reached steam
generation obtained from the mathematical model.
The application of a system able to recovery energy from the cooling system has potential to
substitute the conventional cooling system or at least downsize it. The concept of this super heater
allows the application of an WHR without need of extra installation package that is one of the
biggest challenges on the existing WHR vehicle proposals.
1.3 Cold Engine Gas Recirculation System
Using the exhaust gases parameters combined with the measured temperatures T_AGR_VK and
T_AGR_NK (Figure 15), the first penalty on the energy balance is a consumption of 19,5 kW to
cool down the exhaust gases before it comes back into the intake line. The second energy penalty
it that the Exhaust gases that could be used for performance on the turbocharger are drained to
supply the Engine Gas Recirculation (EGR) demand. It causes an intangible loss of performance
when the gases that could drive the turbocharger are used to come back into intake line. This side
effect is not calculated in this thesis.
Figure 15: Measurement points from the ETC cycle [17]
EGR systems are used to decrease combustion temperature at specific engine loads and
consequently the NOx emissions. The Figure 16 shows the EGR demand in comparison with the
engine intake fresh air. The EGR system from the base line engine (Figure 17) uses the inert
gases from the exhaust manifold before the turbocharger. This point is directly connected to the
intake manifold and on its way there is a control valve and a heat exchanger that decreases the
exhaust gases temperature before mixing it with the fresh air.
In a conventional 4 cylinder Diesel Engine, decreasing exhaust gases temperatures from 6000C
to 1000C can mean a sensitive loss of energy. Some concepts has been developed to decrease the
inlet temperature, but most of them stops on the presence of particulate on the turbocharger
compressor wheels.
The present work purposes a high flow variable valve to mix exhaust gases from vehicle tail pipe
(lower possible temperature) with the pressurized fresh air on the intake line (after turbocharger)
either before or after Charge Air Cooler (CAC). In order to study the possibility to the mix the
low pressure line (Exhaust gases) with the high pressure line (fresh air), one ETC cycle
measurement from the base line engine was done.
Figure 16: EGR demand in a complete ETC cycle Figure 17: EGR mechanism from baseline Engine
Notice that the EGR demand reaches picks of 30% of complete intake air, but sensitively varies
on the engine working cycle. The EGR rate average, in comparison with fresh air for the ETC
cycle from Figure 16, is 19,27%.
Based on the data from Figure 18 it is possible to calculate that the average temperature before
the turbocharger is 403oC reaching picks of more than 6000C. As it is necessary to spend energy
to cool down those gases, an obvious solution for that would be capturing the exhaust gases from
a lower temperature point. In a combustion engine vehicle, this point is known as tail pipe, which,
in this case, reaches temperatures lower then 300oC maintaining an average of 253oC as showed
in Figure 18.
Figure 18: Exhaust gases Temperature on low and High pressure lines
Taking inert gases from this
point would mean directly a
40% decrease of the EGR
system heat rejection:
(4030C x 2530C)
Even the fuel consumption
reduction estimation are not
simple, if 6 kW heat rejection
avoidance could be directly
converted in useful power, it
would mean 3,5% of the 164
kW power from the engine in
study.
Unfortunately taking the inert gases from the tail pipe is not an easy task. Alternatives like Low
Pressure EGR are possible solution, but the contact from the particulate with the compressor
wheel from turbo charger are not desirable. The technical requirement for durable systems are
very expensive and doesn´t worth it taking the EGR cooler out from the system. Therefore, the
proposal from this work is taking the inert gases from the tail pipe and mixing it with fresh air
after the turbocharger compressor.
In Figure 19, it is possible to
understand how different
are the pressures from the
tail pipe line and the fresh
air line after turbocharger
compression. Notice that
more than 60% from 18.000
delta pressure points
measured in this ETC cycle
are under 1 bar. The
remaining measured points
works on 2bar or 3bar
difference, 20% each.
Figure 19: Delta pressure between low and high pressure lines.
At the same time that the delta pressure is not constant, the EGR demand are also not. In Figure
20 the EGR demand 100% means 30% of the engine complete intake (fresh air + Recirculated
gases).
Figure 20: EGR demand distribution
The combination of delta pressure and EGR demand is shown in Figure 21. A mechanism to
mix those pressure lines must be developed to mix 100% of EGR demand
Figure 21: EGR rate and Delta pressure from tail pipe and compressed air after turbocharger
As the Bernoulli principle works for all situations were uncompressible air flow increases, a
system was designed to allow pressure drop by increasing air speed just when exhaust gases are
necessary. This design enables the system to decrease the heat rejection in about 9kW, what
means ca of 40% heat rejection avoidance.
The Figure 22 shows the
temperature that would result from
the mixture of the fresh air directly
with the tail pipe exhaust gases.
Note that this configuration does
not consider recuperation of any
energy post after treatment
system. If such system is
considered in parallel, the exhaust
gases would have even lower
temperatures.
Figure 22: Temperature after mixing lines
A conventional VENTURI system uses the Bernoulli principle increasing the flow speed in order
to decrease the flow pressure and consequently mix the low pressure (Exhaust Tail pipe) with high
pressure (Fresh Air) lines. The Figure 23 shows the involver pressures.
Figure 23: Delta pressure between Fresh air line and Exhaust “Tail pipe”
To reach 3 bar reduction on the fresh air intake line, it would be necessary to reduce the cross
section in more than 50%. As can be observed in Figure 21, the full EGR demand is not necessarily
coincident with high delta pressure demands, which leads to the conclusion that a conventional
VENTURI installed in the intake line would cause unnecessary pressure drop throughout the
engine map.
To avoid this condition, a “variable venture” concept was created. In Figure 24 there is a control
valve on the fresh air track which controls the back pressure in order to derivate the flow to the
line where the Venture is located. With this action, when EGR is not necessary, the control
valve opens and the fresh air flows with no restriction.
Figure 24: Variable Venturi
For stationary conditions the pressure drop can be calculated as follows:
[1]
Although [22] shows that the venture
equations are representative just for
uncompressible flow. The identification of this
transition is determinate by the flow speed. In
Figure 25 both situations are showed.
Figure 25: Subsonic and supersonic flow behavior [22]
Mach (M) is the number defined by the division from the current speed by the sound speed. This
parameter leads to calculations for compressible flow in some points of the engine map in study.
Notice in Figure 26 that the pressure drop between fresh air line and tail pipe cannot reach values
over 1,5 bar. Over this limit, the speed inside the venture overcomes sound speed and its behavior
cannot anymore be modelled with the conventional venture theory. There is also a risk of inverting
the behavior. It means that instead of lowering the pressure, it could be even higher, like showed
in Figure 25.
Figure 26: Pressure drop limit using uncompressible fluid equations
To overcome the technical restriction imposed
from the uncompressible flow, the purposed
system had the conventional venture concept
substituted by a supersonic Laval nozzle that
can be observed in Figure 27. It consists in a
system which drops pressure and increases flow
velocity. This system is commonly used in
aircraft jet engines to increase speed. When the
fluid starts a supersonic flow on the throat, the
pressure drop will appear right after the throat
and not directly on it.
Figure 27: Laval nozzle temperature, pressure and
velocity behavior [44]
The calculation point by point on the ETC cycle shows that a Laval nozzle completely in series
with the fresh air flow would need through diameters that varies between 13 mm and 21 mm. As
the with 13 mm, the pressure drop is already enough to fulfill the hole map from Figure 21.
Therefore, the project was dimensioned to reach M=1 on the throat at lowest air available intake
flow. In Figure 28 it is possible to see the purposed supersonic Laval nozzle cross section and
channels for the exhaust gases admission.
Figure 28: Variable supersonic laval nozzle cross section
1.4 Consolidated system
With a cold EGR and an evaporator around liners able to substitute the conventional cooling
system, a complete new system shows feasibility to be built and can be observed in Figure 29.
Figure 29: Heat rejection avoidance system (complete)
The ETANOL would substitute the old coolant as working fluid. The CONDENSER would work
in place of the old coolant radiator. The RECUPERATOR would be a heat exchanger liquid x air
that would substitute the old charge air cooler air x air. The EVAPORATOR around the engine
heat sources would substitute the complete engine coolant galleries.
CONCLUSION
EGR heat rejection was 19,5 kW with a delta temperature of 323 K (403-80). As this delta drops
to 173k (253-80), the energy spent would be 10,4 kW, what means a heat rejection avoidance of
9,1 kW. It can be converted in a smaller radiator for the old system or to a smaller condenser on
the purposed new one.
The conventional coolant system rejects 23,86 kW/m2 with coolant and the mathematical model
of a evaporator with simple round cross section is able to extract 21,5 kW/m2. It was demonstrates
that the coolant system can be completely substituted by evaporators around the “hot spots” in the
engine. It would mean transferring 105 kW to the ORC working fluid.
By changing the dry charge air cooler to a wet one using the same ethanol line on it would mean
extra 23,4 kW (measured data) transferred to the ORC for WHR.
WHR potential = (105 kW – 19,5kW) + (23,4 +10,4kW) = 119,3 kW
Conventional Rankine cycles doesn´t use to have more than 30% of efficiency. If this system can
be able to recover between 10% and 30% of it, it would mean between 12 kW and 36 kW back to
the power line. I would easy represent extra 10% of extra power or 10% less diesel consumption.
Note that the system has potential to be cheaper than the old system and will include one extra
function: the heat recovery. Today the expander cannot still be clearly priced because there wasn’t
until now large scale production of it. As constructively, it doesn´t differ a lot from an air
conditioning compressor, it should not cost more than 50% more than the known part.
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