# Possibilities of Simultaneous  In-Cylinder  Reduction of  Soot andNOx   Emissions for Diesel   Engines with  Direct Injection

Authors: U. Wagner; P. Eckert; U. Spicher
Journal: International Journal of Rotating Machinery
 (2008)

---

## Abstract

Up to now, diesel engines with direct fuel injection are the
propulsion systems with the highest efficiency for mobile
applications. Future targets in reducingCO2 -emissions with regard
to global warming effects can be met with the help of these
engines. A major disadvantage of diesel engines is the high soot
and nitrogen oxide emissions which cannot be reduced completely
with only engine measures today. The present paper describes two
different possibilities for the simultaneous in-cylinder reduction
of soot and nitrogen oxide emissions. One possibility is the
optimization of the injection process with a new injection
strategy the other one is the use of water diesel emulsions with
the conventional injection system. The new injection strategy for
this experimental part of the study overcomes the problem of
increased soot emissions with  pilot injection by separating the
injections spatially and therefore on the one hand reduces the
soot formation during the early stages of the combustion and on
the other hand increases the soot oxidation later during the
combustion. Another method to reduce the emissions is the
introduction of water into the combustion chamber. Emulsions of
water and fuel offer the potential to simultaneously reduce NOx
and soot emissions while maintaining a high-thermal efficiency.
This article presents a theoretical investigation of the use of
fuel-water emulsions in DI-Diesel engines. The numerical
simulations are carried out with the 3D-CFD code KIVA3V. The use
of different water diesel emulsions is investigated and assessed with the numerical model.


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## Body

## 1. Introduction

Up
to now  diesel  engines with  direct fuel injection
are  the propulsion systems
with a  very high efficiency for mobile applications. Ever more
stringent emission  regulations, especially  for particulate andNOx emissions due to environmental and health risks, are ambitious challenges for
engine manufacturers. Conventional diesel engines offer high-thermal efficiency
and, therefore, will play a key role for reducing fuel consumption with regard
to CO2-induced global warming. However, the major drawbacks of these
engines are the relatively high soot and NOx emissions because of
the direct fuel injection and the resulting combustion process. The combustion
in conventional diesel engines leads to locally rich regions of λ≪1
and to local air/fuel ratios of λ=1. This heterogeneous mixture simultaneously causes
soot because of the rich areas and NOx because of high-burning
temperatures zones (Zeldovich mechanism). Utilizing conventional methods, both
emission components cannot be reduced at the same time which is well known as
the soot-NOx-tradeoff [1].Against
this background, with additionally limited resources of fossil fuels, the main
focus should be put on the development and use of advanced engine technologies
with very low-fuel consumption and even further reduced pollutant emissions.
This pressure is even more increased when taking the rapidly growing individual
mobility in countries with a high-population number like China and India
into account. The need to
reduceCO2 emissions and the directly related green house effect,
especially in relation with the particulate matter issue of diesel engines, is
due to the fact that during the last two decades dust and particulate emissions
were reduced by a number of different measures and, therefore, the atmosphere
is getting clearer. This results in more sunlight coming through the atmosphere
which may accelerate the global warming [2]. As an important consequence of
these very complex atmospheric phenomena, the climate changing effects of
fossil fuels should not be lost out of sight. Hence all measures to further
reduce pollutant emissions should be put under the premise of technical
solutions with optimal efficiency to achieve the lowest possible primary energy
effort. Two possible ways to achieve a significant pollutant reduction already
during the combustion in the cylinder are described in the following. On the
one hand, an advanced injection strategy to optimize the combustion process is
proposed. On the other hand, the advantages of an improved fuel by means of a
diesel/water emulsion are pointed out.


## 2. Advanced Combustion Process

### 2.1. Fundamentals

A
significantly contributing source for soot formation during the combustion can
be the interaction of a very rich air/fuel-mixture or even liquid fuel with the
flame. This effect appears especially in modern direct injection diesel engines,
where the injection is often split in a pilot and a main injection due to noise
reasons. After the ignition of the preinjected fuel, a part of the main
injection can interact with the flame still in liquid phase or as a rich
air/fuel mixture as the fuel is injected straight toward the already burning
cylinder areas. This increases the formation of high amounts of soot [3].In
modern diesel engines with direct fuel injection, the injection process itself
is often split in a pilot injection, a main injection, and possibly a
postinjection. Usually, these part injections are done with the same injection
hardware and injector and, therefore, the fuel is injected in the same cylinder
areas. Hence a strong interaction between the separate part injections
regarding the combustion conditions is evident.With
a small amount of fuel injected before the main injection significant
reductions of combustion, noise, andNOx emissions can be achieved
compared to an engine operation without pilot injection. The preinjected fuel
starts to burn before the combustion of the fuel injected during the main
injection and increases the temperature in the combustion chamber. This
increased temperature reduces the ignition delay for the main fuel, and the
ratio of premixed to diffusion-controlled combustion is decreased. Consequently,
the maximum pressure gradient and peak temperature during the combustion decrease
which lowers both the engine noise and the formation of nitrogen oxides.Unfortunately,
the larger fraction of diffusion-controlled combustion leads to higher soot
emissions. This becomes even more apparent with an increasing amount of preinjected
fuel, where the soot emissions rise disproportionately. But, not only the
larger part of diffusion-controlled combustion but also several other
parameters are responsible for the higher soot emissions. The local air/fuel
ratioλ for the main injected fuel is reduced due to
the already burned pilot fuel, which both increases the soot formation and
decreases the soot oxidation during the combustion. Additionally, a direct
contact of injected and liquid fuel  with the potentially still burning pilot fuel
can take place and  lead to an intense soot formation. Figure 1 shows an example
of the cylinder pressure with and without pilot injection. The advantage of the
pilot injection strategy in reducing the maximum pressure gradient due to a
smaller part of premixed  combustion during the combustion can be seen clearly.
Nevertheless, a major disadvantage is the significantly increased soot emission
[4, 5].Figure 1
Cylinder pressure
with and without pilot injection.In
Figure2, the characteristic progression of the soot concentration in the
cylinder during the combustion process is shown. This qualitative soot
formation and oxidation process is found both in simulation models and
different diesel and also in direct injection gasoline engines [8]. The whole
process is divided into a first soot formation phase, where the in-cylinder
soot concentration is rising, and afterwards two soot oxidation phases with
decreasing in-cylinder soot concentration. Additionally, the corresponding main
influencing parameters for the different phases are listed. It can be seen that
for both the soot formation part and the second oxidation phase, the oxygen
concentration is a main parameter to influence the final engine out soot
emission [9].Figure 2
Qualitative
characteristic process of the soot formation and oxidation [6, 7].The
effect of the local air/fuel ratioλ and the temperature on the amount of emitted
soot is shown in Figure 3 [10, 11]. These results were obtained in premixed
flame experiments but under pressure and temperature condition representative
for diesel engines. Only a small increase in the local air/fuel ratio λ results in a significant decrease of the soot
emission. Accordingly, Hansen [12] and Böhm et al. [11] also showed in
extensive investigations that the soot formation could be suppressed completely
when local air/fuel ratio can be kept always above λ=0.6 to 0.7 during the combustion. Hopp [9]
demonstrated in both calculations and engine experiments that the locally
available oxygen content is a key parameter especially for the soot  oxidation process during the combustion and hence for
the engine out soot emissions.Figure 3
Temperature andλ conditions for
soot formation [10, 11].Starting
from these boundary conditions, which are influencing the soot formation, the
combustion process was designed in a way to increase the locally available
amount of oxygen during the combustion to lower the engine out soot emission.


### 2.2. Combustion Process and Test Engine

As
a reference process to evaluate the soot reduction potential of advanced
heterogeneous combustion processes, the new injection strategy was chosen which
is characterized by a spatial separation of the different part injections of
one cycle. The spatial separation of the particular part injections as proposed
should both avoid the negative interaction between pilot and main injection, in
terms of fuel and flame contact, and, even more, increase the local air
fuel/ratioλ for the main injection and, therefore, reduce
the soot emissions as already mentioned above.The
pilot fuel injection is directed toward the central area of the combustion
chamber and later the main fuel is injected using a conventional spray pattern
with a cone angle greater than160°. With this spatial separation the fuel,
injected during the pilot injection phase, burns in another cylinder area than
the main fuel and hence does not reduce the local air/fuel ratio λ for the main combustion as it would be the
case with a conventional pilot and main injection strategy using the same
injector and spray holes.Figure4 shows a simple schematic drawing for the realization of the spatial
separation with two pilot injection sprays and seven main injection sprays. To
separate the two part injections, two different injectors were used and so the
maximum possible spatial distance between the pilot and main injection is
achieved.Figure 4
Schematic injection
strategy.For
the investigations, a single-cylinder heavy duty research diesel engine from
DaimlerChrysler was used which was significantly modified according to the
experimental requirements. This engine with its four valve low-swirl cylinder head
represents the current engine technology but still offers enough space to
integrate the additional equipment necessary to realize the proposed injection
strategy. To achieve the necessary degree of freedom concerning the injection
parameters, the original unit pump injection system was replaced by a common
rail system which enables the free adjustment of the rail pressure and the
injection timing. The high-pressure fuel pump of the engine was also
electrically driven to be able to choose the rail pressure independent from the
engine operation. The second injector for the pilot injection was connected to
the same high-pressure fuel rail than the main injector so the injection
pressure was constant for both part injections but the timing was freely
adjustable. The main injector was equipped with a seven-hole nozzle with a hole
diameter of 0.2 mm and the pilot injector with a two hole nozzle with a hole
diameter of 0.14 mm.The
test bench was equipped with an eddy-current brake and an electric dynamometer
to either brake or crank the engine at a constant speed. Further, an external
electrically driven supercharger was installed together with a water-cooled
intercooler to boost the engine independent from the operating point. The
engine coolant and oil was conditioned to keep the engine temperature constant
at a preset value. In Table1, the main engine specifications and operating
conditions are listed.Table 1
Engine
specifications and operation conditions.
Displacement1827cm3Stroke142 mmBore128 mmCompression
ratio17.7:1EGRexternal: 0/35%Boost
pressure1.5 barEngine
speed1000 rpmInjection
pressure1400 barCoolant/oil temperature80°C


### 2.3. Cylinder Head Modifications

The
cylinder head of the test engine was a conventional four-valve head with a
centrally placed injector. This injector position was maintained for the main
injection, but the production injector was replaced by a commercially available
heavy duty common rail injector. For the pilot injection, a second injector was
additionally integrated in the cylinder head. This injector was a production
light duty injector with a specially manufactured two-hole nozzle. This smaller
injector is sufficient for the pilot injection due to the significantly smaller
amount of fuel for the pilot injection than for the main injection.Figure5 shows a CAD-model of the cylinder head with the two mentioned injectors, the
schematic fuel sprays, and two additional optical accesses to the combustion
chamber. Additionally, a cutout of the piston top is shown. The bowl in the
piston top had a conventional ω-shape which is commonly used in direct injection
diesel engine. The central area of the bowl is not very deep. This reduces the
possible free spray length for the pilot fuel sprays and may lead to a nonoptimal fuel
distribution for the pilot  fuel injection. For the orientation of the two pilot
fuel sprays, this was taken into account by not injecting directly in the
middle of the combustion chamber but a little sideways.Cylinder head
CAD-model andω-bowl in piston.
(a)(b)The
fuel sprays for both part injections are shown simultaneously. This is only
done for a better picturing of the orientation to each other. Yet, during real
engine operation the injections are temporally separated with the pilot injection
toward the central area of the combustion chamber and the main injection toward
the outer parts.


### 2.4. Engine Operation

To
assess the potential in reducing the soot emissions, the spatial separation of
the pilot and main injection was compared to a conventional injection strategy
where the part injections are done with the same injector. Therefore, a
parameter variation concerning injection timing, fuel quantity distribution,
and EGR-rate was performed when operating the engine at the same parameters but
with the two different injection strategies.In
Table2, the values for the parameter variation are listed. The engine was
operated with no EGR and with 35% EGR. The injection timing was varied between
13° CA BTDC to 5° CA ATDC in 2° CA steps. In all cases, the SOE for the pilot
injection was 10° CA earlier than SOE for the main injection. The fuel amounts
for the two part injection were kept constant for the test with 6 mg for
the pilot injection which is about 12.5% of the totally injected fuel mass. For
the pilot fuel quantity variation, the pilot fuel quantity was increased while
the main fuel amount was reduced to keep the total injected fuel mass constant.Table 2
Parameter variation.
ParameterValuesEGR-rate [%]0 35SOE main injection [CA°]13°  BTDC to 5° ATDCSOE pilot injection10°  CA before SOE mainPilot fuel quantity [mg]4 6  11


### 2.5. Soot Emissions

Figures6 and 7 show the indicated specific soot emissions both for a conventional and
a spatially separated pilot injection strategy without EGR and 35% EGR against
the SOE of the main injection. As expected in both EGR cases, the soot
emissions are rising with later injection timing when using a conventional
injection. For the new injection strategy and no EGR, the soot emissions are
slightly decreasing with later injection which was actually not expected. A
possible explanation for this behavior could be worse mixture formation
condition for the pilot fuel spray in terms of
poor combustion chamber geometry with a lower piston position for
earlier injection timings. The two pilot fuel sprays may reach the edge of the
piston bowl and subsequently increase the soot formation. With an EGR-rate of
35%, the soot emissions remain almost on a constant level independent of the
injection timing. Both a longer ignition delay and a better evaporation caused
by the recirculated exhaust gas may improve the mixture formation and avoid an
increase of the soot formation with earlier injection timings.Soot andNOx
emissions with injection timing variation and 0% EGR.
(a)(b)Figure 7
Soot emissions with
injection timing variation and 35% EGR.These
results approve the assumed influence of the local air/fuel ratioλ on the soot
formation and oxidation process. Only a minor increase of the local λ by separating the pilot fuel burning areas
spatially from the main fuel burning areas reduces the soot emissions
significantly as expected according to the soot formation theory [10, 12].


### 2.6. Combined Soot andNOx
Reduction Potential

To
identify the full potential of the combustion process with the spatially
separated pilot and  main injection strategy, the emission behavior of the
engine regarding both the soot andNOx emissions in combination has
to be taken into account. Figure 8 shows the soot and NOx reduction
potential for an SOE of the main injection at 1° CA BTDC. This injection timing
still allows an engine operation with a high-thermal efficiency. Starting point
for the emission comparison, defined as 100%, is an engine operation with a
conventional pilot injection strategy and 0% EGR. If the EGR-rate is increased
to 35% for the conventional injection, the common soot-NOx-tradeoff
is still present and the soot emissions rise to 160%, while the  NOx emissions go down to 30% (q. v.
Figure 8 upper diagram). Comparing the
same starting point of the conventional injection with the new injection
strategy and 35% EGR both the soot and NOx emissions are roughly
halved (q. v. Figure 8 lower diagram). The reason for this simultaneous
reduction is the high soot reduction potential of the new injection strategy
which allows an engine operation with higher EGR-rates to achieve significant
NOx reductions without having the drawback of severely increased
soot emissions. Incidentally, even a further reduction of the soot emissions would
be possible if the EGR-rate would not be increased that far as can be seen in Figure
6 (upper diagram) for the no EGR case. Without adding EGR soot, emission
reductions of up to 80%, compared to the conventional split injection, are
possible. A 20 to 40% increase of the NOx emission has to be
accepted then as can be seen in Figure 6 (lower diagram).Potential for
simultaneous soot andNOx reduction.
(a)(b)Generally
speaking, the obtained results have shown that advanced heterogeneous
combustion processes still offer a significant potential to reduce the soot andNOx emissions simultaneously. When applying EGR, this NOx reduction can be achieved without any negative effects on the combustion
efficiency compared to a conventional pilot injection strategy. Concerning the
HC and CO emissions, the experiments have shown that there is virtually no
difference between conventional and new pilot injection strategy and they are,
as expected for heterogeneous diesel combustion, on a low absolute level.


### 2.7. Combustion Behavior for Different Pilot Fuel Quantities

Figures9 and 10 show the cylinder pressure traces for the conventional and the new
pilot injection with 4, 6, and 11 mg fuel for the pilot injection. For the
conventional pilot injection, a significant difference in the cylinder pressure
traces can be seen. For the smallest amount of pilot fuel of 4 mg, no effect on
the combustion could be observed and the combustion process is similar to a
single main injection. With a fuel quantity of 6 mg, only a very small pilot
combustion in the cylinder pressure can be seen, but this is sufficient to
reduce the ignition delay, the premixed combustion part, and hence the maximum
pressure rise significantly. For the 11 mg pilot fuel amount almost no
difference in the main combustion, but a further increased pilot combustion is
evident. The shift in the combustion timing for the main injection and the
ignition delay of the main injection changes reduces the ratio of premixed to
diffusion-controlled combustion and hence, in combination with the already
described λ-effect,
increases the soot emissions.Figure 9
Cylinder pressure with conventional pilot
injection.Figure 10
Cylinder pressure with spatial separation
of pilot and main injection.This
looks completely different for the spatially separated pilot injection as shown
in Figure10. In contrast to the conventional injection also for the smallest
amount of 4 mg fuel a pilot combustion and its effect on the main combustion in
reducing the maximum pressure gradient can be seen in the cylinder pressure. By
increasing the pilot fuel quantity, there is almost no shift in combustion
timing, only the amount of released heat for the pilot combustion is certainly
higher. Because of the constant combustion timing and ignition delay for the
main injection, the ratio of premixed to diffusion-controlled combustion stays
constant resulting in almost constant soot emissions.Additionally,
the spatial separation of the fuel injection has two further advantages.First,
an increasing amount of pilot fuel is not reducing the initial local air/fuel
ratioλ for the main injection. In fact with a higher
amount of pilot fuel the combustion conditions for the main injection
concerning the available oxygen are even improved as less main fuel is then
injected for the same load conditions.Second,
as the pilot fuel is not injected through the same nozzle holes as the main
injection the nozzle layout can be optimized for the much smaller amount of
pilot fuel. For example, only two holes are used for the pilot injection
instead of using the seven holes of the main injector. This also improves the
mixture formation for the spatially separated pilot injection compared to the
conventional pilot injection.Summarizing
the pilot fuel quantity variation results, it can be stated that the spatially
separated pilot injection has a significantly higher pilot quantity tolerance
concerning the soot formation and therefore the complexity of the small fuel
amount control can be reduced.


## 2.1. Fundamentals

A
significantly contributing source for soot formation during the combustion can
be the interaction of a very rich air/fuel-mixture or even liquid fuel with the
flame. This effect appears especially in modern direct injection diesel engines,
where the injection is often split in a pilot and a main injection due to noise
reasons. After the ignition of the preinjected fuel, a part of the main
injection can interact with the flame still in liquid phase or as a rich
air/fuel mixture as the fuel is injected straight toward the already burning
cylinder areas. This increases the formation of high amounts of soot [3].In
modern diesel engines with direct fuel injection, the injection process itself
is often split in a pilot injection, a main injection, and possibly a
postinjection. Usually, these part injections are done with the same injection
hardware and injector and, therefore, the fuel is injected in the same cylinder
areas. Hence a strong interaction between the separate part injections
regarding the combustion conditions is evident.With
a small amount of fuel injected before the main injection significant
reductions of combustion, noise, andNOx emissions can be achieved
compared to an engine operation without pilot injection. The preinjected fuel
starts to burn before the combustion of the fuel injected during the main
injection and increases the temperature in the combustion chamber. This
increased temperature reduces the ignition delay for the main fuel, and the
ratio of premixed to diffusion-controlled combustion is decreased. Consequently,
the maximum pressure gradient and peak temperature during the combustion decrease
which lowers both the engine noise and the formation of nitrogen oxides.Unfortunately,
the larger fraction of diffusion-controlled combustion leads to higher soot
emissions. This becomes even more apparent with an increasing amount of preinjected
fuel, where the soot emissions rise disproportionately. But, not only the
larger part of diffusion-controlled combustion but also several other
parameters are responsible for the higher soot emissions. The local air/fuel
ratioλ for the main injected fuel is reduced due to
the already burned pilot fuel, which both increases the soot formation and
decreases the soot oxidation during the combustion. Additionally, a direct
contact of injected and liquid fuel  with the potentially still burning pilot fuel
can take place and  lead to an intense soot formation. Figure 1 shows an example
of the cylinder pressure with and without pilot injection. The advantage of the
pilot injection strategy in reducing the maximum pressure gradient due to a
smaller part of premixed  combustion during the combustion can be seen clearly.
Nevertheless, a major disadvantage is the significantly increased soot emission
[4, 5].Figure 1
Cylinder pressure
with and without pilot injection.In
Figure2, the characteristic progression of the soot concentration in the
cylinder during the combustion process is shown. This qualitative soot
formation and oxidation process is found both in simulation models and
different diesel and also in direct injection gasoline engines [8]. The whole
process is divided into a first soot formation phase, where the in-cylinder
soot concentration is rising, and afterwards two soot oxidation phases with
decreasing in-cylinder soot concentration. Additionally, the corresponding main
influencing parameters for the different phases are listed. It can be seen that
for both the soot formation part and the second oxidation phase, the oxygen
concentration is a main parameter to influence the final engine out soot
emission [9].Figure 2
Qualitative
characteristic process of the soot formation and oxidation [6, 7].The
effect of the local air/fuel ratioλ and the temperature on the amount of emitted
soot is shown in Figure 3 [10, 11]. These results were obtained in premixed
flame experiments but under pressure and temperature condition representative
for diesel engines. Only a small increase in the local air/fuel ratio λ results in a significant decrease of the soot
emission. Accordingly, Hansen [12] and Böhm et al. [11] also showed in
extensive investigations that the soot formation could be suppressed completely
when local air/fuel ratio can be kept always above λ=0.6 to 0.7 during the combustion. Hopp [9]
demonstrated in both calculations and engine experiments that the locally
available oxygen content is a key parameter especially for the soot  oxidation process during the combustion and hence for
the engine out soot emissions.Figure 3
Temperature andλ conditions for
soot formation [10, 11].Starting
from these boundary conditions, which are influencing the soot formation, the
combustion process was designed in a way to increase the locally available
amount of oxygen during the combustion to lower the engine out soot emission.


## 2.2. Combustion Process and Test Engine

As
a reference process to evaluate the soot reduction potential of advanced
heterogeneous combustion processes, the new injection strategy was chosen which
is characterized by a spatial separation of the different part injections of
one cycle. The spatial separation of the particular part injections as proposed
should both avoid the negative interaction between pilot and main injection, in
terms of fuel and flame contact, and, even more, increase the local air
fuel/ratioλ for the main injection and, therefore, reduce
the soot emissions as already mentioned above.The
pilot fuel injection is directed toward the central area of the combustion
chamber and later the main fuel is injected using a conventional spray pattern
with a cone angle greater than160°. With this spatial separation the fuel,
injected during the pilot injection phase, burns in another cylinder area than
the main fuel and hence does not reduce the local air/fuel ratio λ for the main combustion as it would be the
case with a conventional pilot and main injection strategy using the same
injector and spray holes.Figure4 shows a simple schematic drawing for the realization of the spatial
separation with two pilot injection sprays and seven main injection sprays. To
separate the two part injections, two different injectors were used and so the
maximum possible spatial distance between the pilot and main injection is
achieved.Figure 4
Schematic injection
strategy.For
the investigations, a single-cylinder heavy duty research diesel engine from
DaimlerChrysler was used which was significantly modified according to the
experimental requirements. This engine with its four valve low-swirl cylinder head
represents the current engine technology but still offers enough space to
integrate the additional equipment necessary to realize the proposed injection
strategy. To achieve the necessary degree of freedom concerning the injection
parameters, the original unit pump injection system was replaced by a common
rail system which enables the free adjustment of the rail pressure and the
injection timing. The high-pressure fuel pump of the engine was also
electrically driven to be able to choose the rail pressure independent from the
engine operation. The second injector for the pilot injection was connected to
the same high-pressure fuel rail than the main injector so the injection
pressure was constant for both part injections but the timing was freely
adjustable. The main injector was equipped with a seven-hole nozzle with a hole
diameter of 0.2 mm and the pilot injector with a two hole nozzle with a hole
diameter of 0.14 mm.The
test bench was equipped with an eddy-current brake and an electric dynamometer
to either brake or crank the engine at a constant speed. Further, an external
electrically driven supercharger was installed together with a water-cooled
intercooler to boost the engine independent from the operating point. The
engine coolant and oil was conditioned to keep the engine temperature constant
at a preset value. In Table1, the main engine specifications and operating
conditions are listed.Table 1
Engine
specifications and operation conditions.
Displacement1827cm3Stroke142 mmBore128 mmCompression
ratio17.7:1EGRexternal: 0/35%Boost
pressure1.5 barEngine
speed1000 rpmInjection
pressure1400 barCoolant/oil temperature80°C


## 2.3. Cylinder Head Modifications

The
cylinder head of the test engine was a conventional four-valve head with a
centrally placed injector. This injector position was maintained for the main
injection, but the production injector was replaced by a commercially available
heavy duty common rail injector. For the pilot injection, a second injector was
additionally integrated in the cylinder head. This injector was a production
light duty injector with a specially manufactured two-hole nozzle. This smaller
injector is sufficient for the pilot injection due to the significantly smaller
amount of fuel for the pilot injection than for the main injection.Figure5 shows a CAD-model of the cylinder head with the two mentioned injectors, the
schematic fuel sprays, and two additional optical accesses to the combustion
chamber. Additionally, a cutout of the piston top is shown. The bowl in the
piston top had a conventional ω-shape which is commonly used in direct injection
diesel engine. The central area of the bowl is not very deep. This reduces the
possible free spray length for the pilot fuel sprays and may lead to a nonoptimal fuel
distribution for the pilot  fuel injection. For the orientation of the two pilot
fuel sprays, this was taken into account by not injecting directly in the
middle of the combustion chamber but a little sideways.Cylinder head
CAD-model andω-bowl in piston.
(a)(b)The
fuel sprays for both part injections are shown simultaneously. This is only
done for a better picturing of the orientation to each other. Yet, during real
engine operation the injections are temporally separated with the pilot injection
toward the central area of the combustion chamber and the main injection toward
the outer parts.


## 2.4. Engine Operation

To
assess the potential in reducing the soot emissions, the spatial separation of
the pilot and main injection was compared to a conventional injection strategy
where the part injections are done with the same injector. Therefore, a
parameter variation concerning injection timing, fuel quantity distribution,
and EGR-rate was performed when operating the engine at the same parameters but
with the two different injection strategies.In
Table2, the values for the parameter variation are listed. The engine was
operated with no EGR and with 35% EGR. The injection timing was varied between
13° CA BTDC to 5° CA ATDC in 2° CA steps. In all cases, the SOE for the pilot
injection was 10° CA earlier than SOE for the main injection. The fuel amounts
for the two part injection were kept constant for the test with 6 mg for
the pilot injection which is about 12.5% of the totally injected fuel mass. For
the pilot fuel quantity variation, the pilot fuel quantity was increased while
the main fuel amount was reduced to keep the total injected fuel mass constant.Table 2
Parameter variation.
ParameterValuesEGR-rate [%]0 35SOE main injection [CA°]13°  BTDC to 5° ATDCSOE pilot injection10°  CA before SOE mainPilot fuel quantity [mg]4 6  11


## 2.5. Soot Emissions

Figures6 and 7 show the indicated specific soot emissions both for a conventional and
a spatially separated pilot injection strategy without EGR and 35% EGR against
the SOE of the main injection. As expected in both EGR cases, the soot
emissions are rising with later injection timing when using a conventional
injection. For the new injection strategy and no EGR, the soot emissions are
slightly decreasing with later injection which was actually not expected. A
possible explanation for this behavior could be worse mixture formation
condition for the pilot fuel spray in terms of
poor combustion chamber geometry with a lower piston position for
earlier injection timings. The two pilot fuel sprays may reach the edge of the
piston bowl and subsequently increase the soot formation. With an EGR-rate of
35%, the soot emissions remain almost on a constant level independent of the
injection timing. Both a longer ignition delay and a better evaporation caused
by the recirculated exhaust gas may improve the mixture formation and avoid an
increase of the soot formation with earlier injection timings.Soot andNOx
emissions with injection timing variation and 0% EGR.
(a)(b)Figure 7
Soot emissions with
injection timing variation and 35% EGR.These
results approve the assumed influence of the local air/fuel ratioλ on the soot
formation and oxidation process. Only a minor increase of the local λ by separating the pilot fuel burning areas
spatially from the main fuel burning areas reduces the soot emissions
significantly as expected according to the soot formation theory [10, 12].


## 2.6. Combined Soot andNOx
Reduction Potential

To
identify the full potential of the combustion process with the spatially
separated pilot and  main injection strategy, the emission behavior of the
engine regarding both the soot andNOx emissions in combination has
to be taken into account. Figure 8 shows the soot and NOx reduction
potential for an SOE of the main injection at 1° CA BTDC. This injection timing
still allows an engine operation with a high-thermal efficiency. Starting point
for the emission comparison, defined as 100%, is an engine operation with a
conventional pilot injection strategy and 0% EGR. If the EGR-rate is increased
to 35% for the conventional injection, the common soot-NOx-tradeoff
is still present and the soot emissions rise to 160%, while the  NOx emissions go down to 30% (q. v.
Figure 8 upper diagram). Comparing the
same starting point of the conventional injection with the new injection
strategy and 35% EGR both the soot and NOx emissions are roughly
halved (q. v. Figure 8 lower diagram). The reason for this simultaneous
reduction is the high soot reduction potential of the new injection strategy
which allows an engine operation with higher EGR-rates to achieve significant
NOx reductions without having the drawback of severely increased
soot emissions. Incidentally, even a further reduction of the soot emissions would
be possible if the EGR-rate would not be increased that far as can be seen in Figure
6 (upper diagram) for the no EGR case. Without adding EGR soot, emission
reductions of up to 80%, compared to the conventional split injection, are
possible. A 20 to 40% increase of the NOx emission has to be
accepted then as can be seen in Figure 6 (lower diagram).Potential for
simultaneous soot andNOx reduction.
(a)(b)Generally
speaking, the obtained results have shown that advanced heterogeneous
combustion processes still offer a significant potential to reduce the soot andNOx emissions simultaneously. When applying EGR, this NOx reduction can be achieved without any negative effects on the combustion
efficiency compared to a conventional pilot injection strategy. Concerning the
HC and CO emissions, the experiments have shown that there is virtually no
difference between conventional and new pilot injection strategy and they are,
as expected for heterogeneous diesel combustion, on a low absolute level.


## 2.7. Combustion Behavior for Different Pilot Fuel Quantities

Figures9 and 10 show the cylinder pressure traces for the conventional and the new
pilot injection with 4, 6, and 11 mg fuel for the pilot injection. For the
conventional pilot injection, a significant difference in the cylinder pressure
traces can be seen. For the smallest amount of pilot fuel of 4 mg, no effect on
the combustion could be observed and the combustion process is similar to a
single main injection. With a fuel quantity of 6 mg, only a very small pilot
combustion in the cylinder pressure can be seen, but this is sufficient to
reduce the ignition delay, the premixed combustion part, and hence the maximum
pressure rise significantly. For the 11 mg pilot fuel amount almost no
difference in the main combustion, but a further increased pilot combustion is
evident. The shift in the combustion timing for the main injection and the
ignition delay of the main injection changes reduces the ratio of premixed to
diffusion-controlled combustion and hence, in combination with the already
described λ-effect,
increases the soot emissions.Figure 9
Cylinder pressure with conventional pilot
injection.Figure 10
Cylinder pressure with spatial separation
of pilot and main injection.This
looks completely different for the spatially separated pilot injection as shown
in Figure10. In contrast to the conventional injection also for the smallest
amount of 4 mg fuel a pilot combustion and its effect on the main combustion in
reducing the maximum pressure gradient can be seen in the cylinder pressure. By
increasing the pilot fuel quantity, there is almost no shift in combustion
timing, only the amount of released heat for the pilot combustion is certainly
higher. Because of the constant combustion timing and ignition delay for the
main injection, the ratio of premixed to diffusion-controlled combustion stays
constant resulting in almost constant soot emissions.Additionally,
the spatial separation of the fuel injection has two further advantages.First,
an increasing amount of pilot fuel is not reducing the initial local air/fuel
ratioλ for the main injection. In fact with a higher
amount of pilot fuel the combustion conditions for the main injection
concerning the available oxygen are even improved as less main fuel is then
injected for the same load conditions.Second,
as the pilot fuel is not injected through the same nozzle holes as the main
injection the nozzle layout can be optimized for the much smaller amount of
pilot fuel. For example, only two holes are used for the pilot injection
instead of using the seven holes of the main injector. This also improves the
mixture formation for the spatially separated pilot injection compared to the
conventional pilot injection.Summarizing
the pilot fuel quantity variation results, it can be stated that the spatially
separated pilot injection has a significantly higher pilot quantity tolerance
concerning the soot formation and therefore the complexity of the small fuel
amount control can be reduced.


## 3. Fuel/Water Emulsion

### 3.1. Fundamentals

Another
promising method to reduce emissions of nitrogen oxides and particulates in
direct injection diesel engines is water introduction into the combustion
chamber. Various introduction strategies with their particular advantages and
disadvantages with respect to emissions and applicability in different engine applications
are possible. For obtaining maximum improvements, water has to be added at the
right spot at the right time [14]. In conventional, heterogeneous diesel
combustion, nitrogen oxides are mainly formed by the highly temperature-dependent
Zeldovich mechanism. All water introduction strategies aim at reducing the
temperatures in the combustion chamber. The effect of water introduction is
two-fold: water reduces the temperature by its large enthalpy of vaporization
and the larger heat capacity compared to dry air.Injection
of emulsions places the water at the right spot in the spray region. As a
result,NOx emissions can be reduced significantly. Furthermore,
emulsions offer the potential to also reduce particulate emissions, for example
[14–17]. However, some authors also report a neutral or even negative
effect of emulsions on particulates, at least in some points of operation (e.g.,
[17, 18]). For example, Matheaus et al. [17] found drastically increased values
of particulate matter, unburned hydrocarbons, and CO at idle operation in a
heavy duty diesel engine. Musculus et al. [19] attributed this to a possible
cylinder wall wetting at this point of operation due to the increased liquid
penetration of diesel fuel-water emulsion compared to pure diesel fuel.There
are different chemical and physical mechanisms that possibly explain a
reduction of soot emissions. One possible mechanism is the occurrence of the
so-called microexplosions of emulsified fuel droplets that lead to a better atomization
and thus air-fuel mixing [20]. This violent rupture of droplets has been
observed in numerous single droplet evaporation experiments, for example [21].
While some investigations on sprays have been conducted, there does not seem to
be clear evidence that microexplosions occur in modern DI diesel engine
combustion process.In
principle, the use of diesel fuel-water emulsions is possible without any major
modification of the injection equipment. Of course, since the emulsion has a
lower heating value than pure diesel fuel, the injection system has to be
modified in order to keep the rated engine power constant. A drawback of
emulsified fuels is the longer ignition delay and problems associated with
cold-start and idling operation.The
various water introduction strategies offer a high potential in improving diesel
engine emissions. Thus it is necessary to gain a better understanding of the
different processes involved during combustion with these techniques and
develop numerical models able to capture these effects. This section of the paper
presents a first step toward developing these capabilities in multidimensional
combustion modeling and focuses on diesel fuel-water emulsions. The
computational model has been already validated on single droplet evaporation
experiments [22]. Here, the model was used to investigate diesel fuel-water
emulsion combustion in a heavy duty diesel engine.The
3D-CFD code KIVA-3V [23] was used for the calculations in this study. The code
solves the mass, momentum, and energy conservation equations coupled with the
RNG k-ε turbulence model in three dimensions as a function of time. Various
differences to the original code in the physical and chemical submodels
describing the interactions between spray droplets and the gas phase, ignition
and combustion were employed in this study. For example, the primary breakup of
droplets was modeled by the Blob-model and the secondary breakup by the
Rayleigh-Taylor and Kelvin-Helmholtz hybrid model, the droplet evaporation by
using a semicontinuous evaporation model, ignition by the Shell ignition model
and the emission formation for nitrogen oxides by the extended Zeldovich
mechanism, soot formation by the Hiroyasu model, and its oxidation by the Nagle
and Strickland-Constable model. For further details of the different models used
for the calculation please refer to [22].


### 3.2. Numerical Results

The
engine considered here was a heavy duty DI diesel engine with a bore of 128 mm,
a stroke of 142 mm, and a compression ratio of 17.7:1. The engine was retrofitted
with a common-rail fuel injection system and utilizes a low swirl-combustion
process. A seven-hole nozzle in a centrally located injector was used.When
using the same injection equipment with emulsified fuels, it might not be
possible anymore to obtain maximum power because the injection durations become
too long. Thus in addition to the base nozzle, a modified nozzle was used in
some of the calculations. The modified nozzle has been adjusted in order to
obtain the same injection durations for an emulsion with 25% water content
compared to the original nozzle with pure diesel fuel when injecting the same
fuel energy. For this modification, it was assumed that both nozzles have the
same discharge coefficient. The original nozzle (nozzle A) had a hole diameter
of 200μm, the modified
nozzle (nozzle B) a diameter of 225 μm.For
the simulations, only 1/7 of the combustion chamber was modeled, taking
advantage of the axissymmetric nature of the problem. The calculations were
only performed between IVC and EVO. Figure11 shows the computational grid that
was used for the heavy duty engine. The grid consisted of approximately 110500
cells at BDC and 33800 cells at TDC. Figure 12 shows the location of cutplanes
used in subsequent figures.Figure 11
Computational grid
of the heavy duty DI diesel engine.Figure 12
Computational grid
of the heavy duty DI diesel engine.Figure13 exemplarily shows a comparison between experimental and numerical results
for a lower part-load point of operation. Here, the heat release rate was
computed for both the experiment and the simulation by a pressure analysis
software. For all other cases, the heat release rate was taken directly from
the simulation. There is a good agreement for both pressure and heat release
rate.Figure 13
Computational grid
of the heavy duty DI diesel engine.All
of the following calculations were performed at a single point of operation at
upper part load and an engine speed of1750min−1. Except where
otherwise stated, an emulsion with a water content of 25% by mass was used. For
all calculations, the injected mass was adjusted in order to maintain the same
indicated mean effective pressure of the calculated part of the high-pressure
cycle. For larger fuel masses, the injection pressure was kept constant and the
injection duration was prolonged. Table 3 shows the injection parameters for
the base case.Table 3
Injection
parameters for the base case.
Diesel(y0,Diesel=1.0)Emulsion(y0,Diesel=0.75)NozzleAApinj [MPa]140140minj [mg]120160SOI [ATDC°]−4−4DOI [°]17.5422.40


### 3.3. Nonreacting Conditions

To
compare the behavior  of pure diesel and
the   emulsion, a first set of
calculations was performed to investigate the spray behavior under nonreacting
conditions.The
liquid spray length for both fuels is shown in
Figure14. Until the penetration
of diesel reaches a plateau, the penetration of both fuels is nearly identical.
For the emulsion, the steady liquid penetration length is reached later than
for diesel fuel. This is in correspondence with the slower evaporation of
emulsified fuel droplets. While the longer liquid penetration is uncritical
under the given conditions, the emulsified fuel will have a higher risk of fuel
impingement on the piston walls at low load or idle operation.Figure 14
Liquid penetration under nonreacting conditions.Figure15 visualizes the equivalence ratio in cutplane A for the two fuels at 2° ATDC.
Here, the equivalence ratio was defined as the ratio of oxygen atoms necessary
to oxidize all available carbon and hydrogen atoms and the actually available
amount of oxygen atoms. Since water displaces a certain amount of diesel fuel,
the mixture in the emulsion spray is significantly leaner. Especially, the very
high value of Φ≈5.0 near the nozzle when using pure diesel is
reduced to Φ≈3.0 for the emulsion. The reduced equivalence
ratio will have an important impact on soot formation chemistry.Equivalence ratio
in the emulsion spray (upper) and diesel spray (lower), cutplane A.
(a)(b)An
analysis was carried out to estimate the kinetic limit of superheat of fuel
water emulsion droplets. Some of the droplets in the spray exceed this limit.
However, all these droplets are extremely small (below 1μm) and it is thus not likely that these droplets
would microexplode or that these microexplosions would have a dramatic effect
on the spray. However, the current analysis is not adequate to fully assess the
occurrence and influence of microexplosion.


### 3.4. Reacting Conditions

Figure16 compares the pressure history and heat release rate for fuels with different
water contents. The start of injection was held constant at −4° ATDC. The
ignition delay and thus also the premixed burn fraction continuously increases
with increasing water content of the fuel. The end of injection of the
different cases is indicated by arrows in Figure 16. Because of the longer
injection duration, the point where the heat release rate drops is shifted to
later times for higher water fractions. As a result, the MFB50 point is shifted
only slightly to later times and differs by approx. 1° for pure Diesel and the
35% Emulsion. Figure 17 exemplarily shows the temperature distribution in
cutplane A for the 25% emulsion and pure Diesel fuel at 2° ATDC. As can be
seen, the lift-off distance of the diffusion flame is increased for the
emulsion. Flame lift-off is an important characteristic of a Diesel engine
combustion process [24]. An increased lift-off length reduces the equivalence
ratio in the diffusion flame, because the amount of entrained air increases
with distance to the nozzle. This reduction in the equivalence ratio is in
addition to the reduction already shown in the preceding section.Figure 16
Heat release rate
and pressure history for fuels with different water contents.(a) Temperature distribution of the 25% emulsion; and (b) pure diesel fuel flame,
cutplane A.
(a)(b)The
improved equivalence ratio in terms of soot formation is shown in
Figure18.
The idea of Akihama et al. [13] is followed, plotting the equivalence ratio
over temperature for all individual computational cells. Also shown are the
regions of soot and NOx formation, which were also adopted from [13].
Note that the location of these regions depend also on other thermodynamic
parameters and the type of fuel and are given only as a guideline. At TDC there
is significant scatter for both fuels because the diffusion flame is not fully
established. At later times, the typical shape of a diffusion flame is obtained
with the highest temperatures in the slightly rich region. It can be seen that
the equivalence ratio in the cells of the emulsion flame are shifted away from
the soot formation region compared to Diesel fuel. While not as significantly,
the thermodynamic states, where high NOx formation occurs are also
reduced. Since NOx formation is extremely dependent on peak
temperatures, already a small reduction in peak temperature will lead to a
large reduction in nitrogen oxide emissions.Equivalence
ratio-temperature plots for pure diesel fuel and a 25% emulsion; Soot andNOx formation maps according to [13].
(a)(b)(c)(d)(e)Figure19 shows the predicted nitrogen oxide and soot history. As expected, the
formation of NOx is drastically reduced with increasing water
content. While the peak soot mass is continuously reduced with higher water
concentrations, the soot mass at EVO has a minimum for the emulsion with 10%
water fractions and increases again for higher water fractions. While a very
simple soot model was used here, the occurrence of such a minimum seems
reasonable, since there will be a tradeoff between the reduction in soot
formation due to the lower equivalence ratios and a reduction of soot oxidation
due to the lower combustion temperatures. However, the oxidation of soot by
OH-radicals was neglected in the current soot model and the increased
OH-radical concentration with increasing water content might have an important
influence on soot oxidation.Figure 19
Soot andNOx history for different fuels.Finally,
Figure20 shows the soot-NOx and the ISFC-NOx tradeoff
for a variation of the start of injection under constant load. Since the
specific fuel consumption ISFC has been calculated with the IMEP which was in
turn calculated from IVC to EVO, the fuel consumption values can only be seen
in qualitative terms. In addition to the trends of pure diesel fuel and the
emulsion with 25% water content using nozzle A, results are also shown for the
25% emulsion with nozzle B. There is a clear advantage in the tradeoff for the
emulsified fuel with both nozzles. Comparing the optimal point in terms of
emissions, the reduction in nitrogen oxides with the 25% emulsion is approximately
21%. This is in accordance with a “rule of thumb,” predicting a reduction of
10% in nitrogen oxides for 10% of water addition [14, 18]. Specific fuel
consumption is also improved when using the emulsified fuel. This numerical
result is in contrast to some results from the literature [14], while in
accordance with other [16]. The influence of the emulsification on ISFC will
depend strongly on the kind of engine operation and the specific injection
system used in the investigations.Soot-NOx and ISFC-NOx tradeoff.
(a)(b)In
future studies, a validation with engine test data will extend the verification
of the computational model performed in this study. Furthermore, a
phenomenological soot model that better captures physical effects will be used
in the future. This includes the effort to capture the important effect of an
increased OH-radical concentration due to the addition of water on the soot
oxidation.


## 3.1. Fundamentals

Another
promising method to reduce emissions of nitrogen oxides and particulates in
direct injection diesel engines is water introduction into the combustion
chamber. Various introduction strategies with their particular advantages and
disadvantages with respect to emissions and applicability in different engine applications
are possible. For obtaining maximum improvements, water has to be added at the
right spot at the right time [14]. In conventional, heterogeneous diesel
combustion, nitrogen oxides are mainly formed by the highly temperature-dependent
Zeldovich mechanism. All water introduction strategies aim at reducing the
temperatures in the combustion chamber. The effect of water introduction is
two-fold: water reduces the temperature by its large enthalpy of vaporization
and the larger heat capacity compared to dry air.Injection
of emulsions places the water at the right spot in the spray region. As a
result,NOx emissions can be reduced significantly. Furthermore,
emulsions offer the potential to also reduce particulate emissions, for example
[14–17]. However, some authors also report a neutral or even negative
effect of emulsions on particulates, at least in some points of operation (e.g.,
[17, 18]). For example, Matheaus et al. [17] found drastically increased values
of particulate matter, unburned hydrocarbons, and CO at idle operation in a
heavy duty diesel engine. Musculus et al. [19] attributed this to a possible
cylinder wall wetting at this point of operation due to the increased liquid
penetration of diesel fuel-water emulsion compared to pure diesel fuel.There
are different chemical and physical mechanisms that possibly explain a
reduction of soot emissions. One possible mechanism is the occurrence of the
so-called microexplosions of emulsified fuel droplets that lead to a better atomization
and thus air-fuel mixing [20]. This violent rupture of droplets has been
observed in numerous single droplet evaporation experiments, for example [21].
While some investigations on sprays have been conducted, there does not seem to
be clear evidence that microexplosions occur in modern DI diesel engine
combustion process.In
principle, the use of diesel fuel-water emulsions is possible without any major
modification of the injection equipment. Of course, since the emulsion has a
lower heating value than pure diesel fuel, the injection system has to be
modified in order to keep the rated engine power constant. A drawback of
emulsified fuels is the longer ignition delay and problems associated with
cold-start and idling operation.The
various water introduction strategies offer a high potential in improving diesel
engine emissions. Thus it is necessary to gain a better understanding of the
different processes involved during combustion with these techniques and
develop numerical models able to capture these effects. This section of the paper
presents a first step toward developing these capabilities in multidimensional
combustion modeling and focuses on diesel fuel-water emulsions. The
computational model has been already validated on single droplet evaporation
experiments [22]. Here, the model was used to investigate diesel fuel-water
emulsion combustion in a heavy duty diesel engine.The
3D-CFD code KIVA-3V [23] was used for the calculations in this study. The code
solves the mass, momentum, and energy conservation equations coupled with the
RNG k-ε turbulence model in three dimensions as a function of time. Various
differences to the original code in the physical and chemical submodels
describing the interactions between spray droplets and the gas phase, ignition
and combustion were employed in this study. For example, the primary breakup of
droplets was modeled by the Blob-model and the secondary breakup by the
Rayleigh-Taylor and Kelvin-Helmholtz hybrid model, the droplet evaporation by
using a semicontinuous evaporation model, ignition by the Shell ignition model
and the emission formation for nitrogen oxides by the extended Zeldovich
mechanism, soot formation by the Hiroyasu model, and its oxidation by the Nagle
and Strickland-Constable model. For further details of the different models used
for the calculation please refer to [22].


## 3.2. Numerical Results

The
engine considered here was a heavy duty DI diesel engine with a bore of 128 mm,
a stroke of 142 mm, and a compression ratio of 17.7:1. The engine was retrofitted
with a common-rail fuel injection system and utilizes a low swirl-combustion
process. A seven-hole nozzle in a centrally located injector was used.When
using the same injection equipment with emulsified fuels, it might not be
possible anymore to obtain maximum power because the injection durations become
too long. Thus in addition to the base nozzle, a modified nozzle was used in
some of the calculations. The modified nozzle has been adjusted in order to
obtain the same injection durations for an emulsion with 25% water content
compared to the original nozzle with pure diesel fuel when injecting the same
fuel energy. For this modification, it was assumed that both nozzles have the
same discharge coefficient. The original nozzle (nozzle A) had a hole diameter
of 200μm, the modified
nozzle (nozzle B) a diameter of 225 μm.For
the simulations, only 1/7 of the combustion chamber was modeled, taking
advantage of the axissymmetric nature of the problem. The calculations were
only performed between IVC and EVO. Figure11 shows the computational grid that
was used for the heavy duty engine. The grid consisted of approximately 110500
cells at BDC and 33800 cells at TDC. Figure 12 shows the location of cutplanes
used in subsequent figures.Figure 11
Computational grid
of the heavy duty DI diesel engine.Figure 12
Computational grid
of the heavy duty DI diesel engine.Figure13 exemplarily shows a comparison between experimental and numerical results
for a lower part-load point of operation. Here, the heat release rate was
computed for both the experiment and the simulation by a pressure analysis
software. For all other cases, the heat release rate was taken directly from
the simulation. There is a good agreement for both pressure and heat release
rate.Figure 13
Computational grid
of the heavy duty DI diesel engine.All
of the following calculations were performed at a single point of operation at
upper part load and an engine speed of1750min−1. Except where
otherwise stated, an emulsion with a water content of 25% by mass was used. For
all calculations, the injected mass was adjusted in order to maintain the same
indicated mean effective pressure of the calculated part of the high-pressure
cycle. For larger fuel masses, the injection pressure was kept constant and the
injection duration was prolonged. Table 3 shows the injection parameters for
the base case.Table 3
Injection
parameters for the base case.
Diesel(y0,Diesel=1.0)Emulsion(y0,Diesel=0.75)NozzleAApinj [MPa]140140minj [mg]120160SOI [ATDC°]−4−4DOI [°]17.5422.40


## 3.3. Nonreacting Conditions

To
compare the behavior  of pure diesel and
the   emulsion, a first set of
calculations was performed to investigate the spray behavior under nonreacting
conditions.The
liquid spray length for both fuels is shown in
Figure14. Until the penetration
of diesel reaches a plateau, the penetration of both fuels is nearly identical.
For the emulsion, the steady liquid penetration length is reached later than
for diesel fuel. This is in correspondence with the slower evaporation of
emulsified fuel droplets. While the longer liquid penetration is uncritical
under the given conditions, the emulsified fuel will have a higher risk of fuel
impingement on the piston walls at low load or idle operation.Figure 14
Liquid penetration under nonreacting conditions.Figure15 visualizes the equivalence ratio in cutplane A for the two fuels at 2° ATDC.
Here, the equivalence ratio was defined as the ratio of oxygen atoms necessary
to oxidize all available carbon and hydrogen atoms and the actually available
amount of oxygen atoms. Since water displaces a certain amount of diesel fuel,
the mixture in the emulsion spray is significantly leaner. Especially, the very
high value of Φ≈5.0 near the nozzle when using pure diesel is
reduced to Φ≈3.0 for the emulsion. The reduced equivalence
ratio will have an important impact on soot formation chemistry.Equivalence ratio
in the emulsion spray (upper) and diesel spray (lower), cutplane A.
(a)(b)An
analysis was carried out to estimate the kinetic limit of superheat of fuel
water emulsion droplets. Some of the droplets in the spray exceed this limit.
However, all these droplets are extremely small (below 1μm) and it is thus not likely that these droplets
would microexplode or that these microexplosions would have a dramatic effect
on the spray. However, the current analysis is not adequate to fully assess the
occurrence and influence of microexplosion.


## 3.4. Reacting Conditions

Figure16 compares the pressure history and heat release rate for fuels with different
water contents. The start of injection was held constant at −4° ATDC. The
ignition delay and thus also the premixed burn fraction continuously increases
with increasing water content of the fuel. The end of injection of the
different cases is indicated by arrows in Figure 16. Because of the longer
injection duration, the point where the heat release rate drops is shifted to
later times for higher water fractions. As a result, the MFB50 point is shifted
only slightly to later times and differs by approx. 1° for pure Diesel and the
35% Emulsion. Figure 17 exemplarily shows the temperature distribution in
cutplane A for the 25% emulsion and pure Diesel fuel at 2° ATDC. As can be
seen, the lift-off distance of the diffusion flame is increased for the
emulsion. Flame lift-off is an important characteristic of a Diesel engine
combustion process [24]. An increased lift-off length reduces the equivalence
ratio in the diffusion flame, because the amount of entrained air increases
with distance to the nozzle. This reduction in the equivalence ratio is in
addition to the reduction already shown in the preceding section.Figure 16
Heat release rate
and pressure history for fuels with different water contents.(a) Temperature distribution of the 25% emulsion; and (b) pure diesel fuel flame,
cutplane A.
(a)(b)The
improved equivalence ratio in terms of soot formation is shown in
Figure18.
The idea of Akihama et al. [13] is followed, plotting the equivalence ratio
over temperature for all individual computational cells. Also shown are the
regions of soot and NOx formation, which were also adopted from [13].
Note that the location of these regions depend also on other thermodynamic
parameters and the type of fuel and are given only as a guideline. At TDC there
is significant scatter for both fuels because the diffusion flame is not fully
established. At later times, the typical shape of a diffusion flame is obtained
with the highest temperatures in the slightly rich region. It can be seen that
the equivalence ratio in the cells of the emulsion flame are shifted away from
the soot formation region compared to Diesel fuel. While not as significantly,
the thermodynamic states, where high NOx formation occurs are also
reduced. Since NOx formation is extremely dependent on peak
temperatures, already a small reduction in peak temperature will lead to a
large reduction in nitrogen oxide emissions.Equivalence
ratio-temperature plots for pure diesel fuel and a 25% emulsion; Soot andNOx formation maps according to [13].
(a)(b)(c)(d)(e)Figure19 shows the predicted nitrogen oxide and soot history. As expected, the
formation of NOx is drastically reduced with increasing water
content. While the peak soot mass is continuously reduced with higher water
concentrations, the soot mass at EVO has a minimum for the emulsion with 10%
water fractions and increases again for higher water fractions. While a very
simple soot model was used here, the occurrence of such a minimum seems
reasonable, since there will be a tradeoff between the reduction in soot
formation due to the lower equivalence ratios and a reduction of soot oxidation
due to the lower combustion temperatures. However, the oxidation of soot by
OH-radicals was neglected in the current soot model and the increased
OH-radical concentration with increasing water content might have an important
influence on soot oxidation.Figure 19
Soot andNOx history for different fuels.Finally,
Figure20 shows the soot-NOx and the ISFC-NOx tradeoff
for a variation of the start of injection under constant load. Since the
specific fuel consumption ISFC has been calculated with the IMEP which was in
turn calculated from IVC to EVO, the fuel consumption values can only be seen
in qualitative terms. In addition to the trends of pure diesel fuel and the
emulsion with 25% water content using nozzle A, results are also shown for the
25% emulsion with nozzle B. There is a clear advantage in the tradeoff for the
emulsified fuel with both nozzles. Comparing the optimal point in terms of
emissions, the reduction in nitrogen oxides with the 25% emulsion is approximately
21%. This is in accordance with a “rule of thumb,” predicting a reduction of
10% in nitrogen oxides for 10% of water addition [14, 18]. Specific fuel
consumption is also improved when using the emulsified fuel. This numerical
result is in contrast to some results from the literature [14], while in
accordance with other [16]. The influence of the emulsification on ISFC will
depend strongly on the kind of engine operation and the specific injection
system used in the investigations.Soot-NOx and ISFC-NOx tradeoff.
(a)(b)In
future studies, a validation with engine test data will extend the verification
of the computational model performed in this study. Furthermore, a
phenomenological soot model that better captures physical effects will be used
in the future. This includes the effort to capture the important effect of an
increased OH-radical concentration due to the addition of water on the soot
oxidation.


## 4. Summary

Two
different measures to improve the soot-NOx and the ISFC-NOx tradeoffs. The first concept uses an advanced heterogeneous CI combustion
processes which still offer substantial possibilities to reduce the critical
soot and NOx emissions while maintaining a high-thermal efficiency.
Due to comfort and NOx reduction reasons modern diesel engines with
common-rail injection systems usually use a small pilot injection before the
main injection which effect is mainly a shorter ignition delay for the main
combustion. But this can lead to increased soot emissions. The combustion
process discussed in this paper with a spatial separation of the pilot from the
main injection represents an advanced CIDI process which avoids the negative
effect of the pilot injection on the soot formation while still maintaining its
positive effects on NOx and noise. The results have shown that
besides reduced soot emissions also substantially reduced NOx emission can be achieved due to the improved EGR-tolerance of the combustion
process without increasing soot.Furthermore,
the potential of using diesel fuel-water emulsions in the diesel combustion
process has been investigated numerically. Here, also improved soot-NOx and ISFC-NOx tradeoffs were predicted with the CFD-calculations. First,
experimental results confirm these predictions and allow a further optimization
of the relevant calculation models.



---

*Source: 175956-2008-10-28.xml*