Multi-Dimensional Modeling of Ignition, Combustion and Nitric Oxide Formation in Direct Injection Natural Gas Engines

18 Pages·0.367 MB·English

by  Agarwal A., Assanis D.

SAE TECHNICAL
PAPER SERIES 2000-01-1839
Multi-Dimensional Modeling of Ignition,
Combustion and Nitric Oxide Formation
in Direct Injection Natural Gas Engines
Apoorva Agarwal and Dennis Assanis
W. E. Lay Automotive Laboratory, The University of Michigan
Reprinted From: Combustion in Diesel and SI Engines
(SP–1549)
and
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Printed in USA
2000-01-1839
Multi-Dimensional Modeling of Ignition, Combustion and Nitric
Oxide Formation in Direct Injection Natural Gas Engines
Apoorva Agarwal1 and Dennis Assanis
W. E. Lay Automotive Laboratory, The University of Michigan
Copyright © 2000 CEC and SAE International.
ABSTRACT diesel engines, while maintaining the smoke-free
operation of spark-ignited (SI) engines and producing
The heat release and pollutant formation processes in a much lower NOx emissions. Furthermore, it has been
direct injection natural gas engine are studied by coupling seen that fuel metering changes as a result of
detailed chemistry with a multi-dimensional reactive flow compositional changes in natural gas are not a significant
code. A detailed kinetic mechanism consisting of 22 factor in DI applications, as long as condensation does
species and 104 elementary reactions is chosen by not occur due to high injection pressures in the fuel
comparing ignition delay predictions with measurements system [8]. This is a major advantage over carburated,
in a combustion bomb. The ignition model is then premixed SI engines where natural gas composition
coupled with a turbulent combustion model and extended changes can lead to significant equivalence ratio
Zeldovich kinetics to simulate heat release and nitric changes, thus resulting in engine performance changes
oxide production in a direct injection engine. Parametric [9]. Direct injection of natural gas into the cylinder can
studies are conducted to investigate the effect of engine also be used to obtain benefits associated with mixture
operating conditions which include speed, load, injection stratification when the gas is injected very close to the
timing and level of boost. It is shown that use of detailed start of combustion [10].
chemistry is extremely important to predict the correct
The primary challenge in operating DI natural gas
ignition delay period as engine operating conditions
engines is the low ignitability of methane, which is the
change. Use of both time and crank angle as the
primary constituent of natural gas. The optimum ignition
independent variable reveals interesting details of the
delay for DI engines is usually less than 2 ms [11]. A long
heat release process as a function of engine speed.
and variable ignition delay time is undesirable as it leads
Nitric oxide formation is shown to depend on both local
to spiky heat release due to an increase in the premixed
temperature and oxygen concentration and injection
part of combustion. At an engine speed of 1500 RPM, a
timing is found to be the single most important variable
difference in ignition delay of 0.5-0.7 ms represents a 4-6
controlling it. A trade-off between early injection timing
crank angle shift in the start of combustion. So a delay of
for higher IMEP and late timing for lower NO is required
even a few milliseconds may have a significant impact on
to get desirable performance from the engine while
subsequent combustion, peak pressures, and hence
keeping the emission levels under control.
performance and emissions.
INTRODUCTION Experiments in a combustion bomb have shown that
autoignition of natural gas under diesel-like conditions
Growing concerns about the harmful effects of requires temperatures as high as 1100-1200 K for a
conventional fossil fuel emissions have made natural gas reasonable ignition delay [8]. This high temperature
a very attractive alternative fuel to power prime movers requirement mandates that either a high compression
and stationary energy conversion devices. ratio or a high intake air temperature be used, both of
Comprehensive reviews of the advantages of natural gas which have negative effects on engine performance,
as an environmentally friendly, clean burning (very low durability and nitric oxide emissions. Obviously, knowing
propensity to soot), economical and efficient fuel have at what conditions of temperature, pressure, and
been reported in the literature [1-7]. For heavy-duty composition a natural gas-air mixture can auto-ignite, or
engine applications in particular, direct injection (DI) of whether ignition assist in the form of a glow plug [12] or
natural gas promises thermal efficiencies comparable to pilot fuel injection [13] is required is very important.
those attained by high compression ratio, unthrottled
1. Presently at Scientific Research Laboratories, Ford Motor Company, Dearborn, Michigan, USA
1
Since DI natural gas is a new technology, a detailed ignition. The process of selection of the detailed kinetic
understanding of the physical processes involved is mechanism is described in a previous paper [16].
imperative to guide the design and optimization CHEMKIN-II was used to read the reaction mechanism
processes. However, in-depth experimental and compute the source terms in the energy and species
investigations, particularly on heavy-duty engines, are equations; the latter were integrated using the stiff ODE
relatively sparse [6,12]. Multi-dimensional CFD studies solver LSODE [20]. Correct predictions of the ignition
have advanced our understanding of the physics of delay period, at different loads and speeds, are essential
gaseous injection and mixing in such engines [10,14,15]. as the delay has a significant impact on the subsequent
However, the scope of those studies has not so far combustion process, and thus engine performance and
considered ignition, combustion and pollutant formation. emissions.
Since the DI ignition process encompasses many
In a reactive turbulent engine flow field, ignition and flame
physical and chemical steps, the numerical solution of
propagation occur at different characteristic time-scales.
detailed chemical kinetics coupled with the flow becomes
During ignition delay, it is appropriate to assume that the
essential for an accurate quantitative prediction.
chemical time-scales are much larger than the flow time-
However, one finds that either the details of chemistry or
scales, so the chemistry is not affected directly by the
the details of the flow are usually neglected in the
fluctuations caused in temperature and species
literature, the primary concern being the enormous
concentrations by turbulence. However, such an
computational costs associated with coupling the two.
assumption is not valid when chemical heat release
With advances in our understanding of the fundamental
becomes significant. In such a situation, the interaction
chemistry of hydrocarbon fuels and with increasing
between chemistry and flow needs to be modeled.
computer memory and processing speeds with every
Various approaches to describe this interaction, like the
passing year, this no longer is a major issue.
mixing-controlled combustion models, the Eddy
Agarwal and Assanis have systematically investigated Dissipation Concept (EDC), and flamelet models were
the suitability of alternative, detailed chemical kinetic explored in detail by Agarwal [20]. In this work, a mixing-
schemes to predict ignition delay under typical, end of controlled combustion model is used to simulate
compression conditions in a DI natural gas engine [16]. combustion after the ignition delay period is over. A one-
Subsequently, one of those kinetic schemes was coupled step chemical reaction is then used to convert fuel to
with the multi-dimensional reactive flow code KIVA-3V products of complete combustion. Such a model was
[17]. Ignition delay predictions were validated against shown to be able to give reasonable predictions of
experimental data from a combustion bomb [17]. Nitric thermodynamic parameters and also concentrations of
oxide formation was also modeled by coupling the major species [20].
extended Zeldovich mechanism with KIVA-3V [18]. The
A kinetic mechanism for nitric oxide formation, the
objective of the present study is to combine all our
extended Zeldovich mechanism, was also coupled with
previously developed models and use them to perform
the ignition/combustion models and used to predict NO
detailed parametric studies of the effects of operating
formation. This mechanism uses the equilibrium values
conditions on DI natural gas engine performance and
of concentrations of radicals and the temperatures and
emissions. Such studies can assess complex tradeoffs
pressures predicted by the combustion models. It does
involved in the engine design process without the need
not participate in the heat release process itself.
for expensive physical prototypes.
However, NO chemistry is linked to the fuel oxidation
This paper is organized as follows: in the next section, process in that the latter provides the species and
previous work published by the authors, on which this temperature needed for NO formation. The details of the
paper is based, is described briefly. After that the details implementation of the nitric oxide chemistry are provided
of engine simulations are provided and the baseline in [18].
engine calculations are described. Thereafter,
parameteric studies are conducted to investigate the DETAILS OF ENGINE SIMULATIONS
effect of engine speed, load, injection timing and level of
boost on engine performance and emissions. The engine geometry chosen for the simulations was
based on a medium-size heavy-duty diesel engine typical
MODELING FRAMEWORK of engines used for urban bus and Class 4 through 7
truck applications. Such a geometry was used earlier in
The modeling framework for our studies is based on a injection and mixing studies [10,14,15], as it was felt that,
version of the Los Alamos code KIVA-3V [19], modified at if the direct injection natural gas engine technology
the University of Michigan to simulate gaseous injection proves to be successful, engines in this category will
[15]. A new methodology was developed to couple prove to be the mainstream application. Details of the
detailed chemistry with KIVA-3V [17]. A detailed simulated engine configuration along with injection data,
chemical kinetic mechanism, which is a reduced version baseline operating conditions and modeling choices for
of GRI-Mech 1.2, consisting of 22 species and 104 this study are shown in Table 1.
elementary reactions was chosen to simulate natural gas
2
coefficient (0.8) times the sonic velocity (508.5 m/s) at
Table 1. Simulated Engine Parameters the specified inlet temperature. The fuel was injected at
an angle of 45° with respect to the horizontal. The
Geometric Specifications
injection duration was 20° and the fuel was injected in a
Bore 12.45 cm
slug profile. If the fuel was injected in the middle of the
Stroke 15.5 cm computational domain, which spanned 45°, the three
TDC Clearance Height 0.37 cm components of velocity were calculated in a manner
shown in Fig. 1.
Connecting Rod Length 31.0 cm
Piston Bowl Shape Mexican-Hat Injection Angle θ with the horizontal
Compression Ratio 16.01:1
5
Displaced Volume/cylinder 1.86 liter 2.
2
n
Injection Data si V cos θ
θ
Fuel Temperature 400 K os TOP VIEW
Vc ° Y
Injection Velocity 508.5 m/s = 22.5
v
Injection Timing 25° BTDC / 10° BTDC u=Vcosθcos22.5 X
Injection Duration 20° CA
V cos θ
Injection Profile Slug
Injection Angle 45° with respect to θ
θ
horizontal n
si
Baseline Operating Conditions V SIDE VIEW
-
Speed 1500 RPM =
w
Overall Equivalence Ratio 0.6
Z
Injection Timing 25° BTDC
V
Intake Pressure 3 bar X
Modeling Choices Figure 1. Determination of injection velocity
Duration Of Simulations 180° BTDC - 180° ATDC components
Intake Air Temperature 675 K
For the engine simulations, it is important to make
Intake Pressure 3 bar / 4 bar reasonable choices for different engine parameters.
Initial Turbulent Intensity 1.7 x Mean Piston Speed Since it is known that autoignition of methane requires
much higher temperatures than diesel fuel, either very
Initial Swirl Ratio 1.0
high intake air temperatures or very high compression
EGR 10%
ratios are needed to achieve these temperatures at the
Turbulence Model RNG Based k-ε time of fuel injection. A compromise between both of
Cylinder Wall Heat Flux Adiabatic these parameters was obtained by using a compression
ratio of 16:1 (a reasonable value for medium-size heavy-
duty diesel engines) and an intake air temperature of 675
A 45° sector mesh was used for the calculations, K (402°C). This choice of the intake air temperature
assuming symmetry for an eight hole injector. This ensured that the fuel was injected at air temperatures
greatly reduced the computational costs, as only 1/8th of close to 1200 K. Although the value of 402°C is quite high
the cylinder had to be simulated. Periodic boundary for intake air temperature, it can be achieved through a
conditions were assumed for the front and back faces of combination of low heat rejection engine operation, high
the sector mesh, which consisted of 42×3×25 cells in the boost (3-4 bar manifold absolute pressure), no
three dimensions. Although grid-independence tests intercooling and EGR. For our purposes, adiabatic
were not done, the grid used in these calculations was engine operation was assumed with 10% EGR,
finer than that used in an independent work [15] for a consisting of products of complete combustion (carbon
similar engine geometry. Since a “Mexican-Hat” shape dioxide and water). The mass fractions of CO and H O
2 2
for the piston bowl was assumed, the height of the were calculated by assuming stoichiometric combustion
cylinder head was adjusted till the desired compression of methane and pure air.
ratio of 16:1 was obtained. The values shown for engine
Only the closed part of the engine cycle was simulated
geometry in Table 1 were obtained after all the desired
since a sector mesh was used. However, it is important
adjustments in cylinder head height were done.
to prescribe the correct value of turbulence intensity at
The temperature of the fuel (pure methane) was the start of the calculation. Experiments in engines have
assumed to be 400 K. Assuming the flow is choked, the shown that the intake generated flow velocity is
injection velocity was specified to be equal to a discharge proportional to the engine speed [11]. Based on the
3
normalized values of turbulent intensity with mean piston
200
speed, it can be assumed that at 180° BTDC, the Motoring (a)
following relationship exists between turbulence intensity Firing
′
(u ) and mean piston speed (S ): 150
p 1500 RPM
ar)
u′=1.7Sp (1) sure (b 100 SOI
s
e
where the mean piston speed is defined as S = 2LN, L Pr
p
being the stroke and N the engine crankshaft speed. 50
The constant 1.7 was also confirmed by running a
thermodynamic diesel engine simulation with a
0
turbulence model [21] at different engine speeds. An
-150 -100 -50 0 50 100 150
RNG based k-ε model, as implemented in KIVA-3V, was
Crank Angle (degree)
used to describe turbulence. Since most direct injection
engines have some intake generated swirl, it was
200
assumed that the swirl ratio (R ) was unity. (b)
s
Motoring
This study investigates the effect of engine operating
Firing
conditions, including speed, load, injection timing and 150
boost level on engine performance and emissions. The ar)
b
speeds selected (800, 1500 and 2200 RPM) span the e (
ur 100
range of typical diesel engine operation. Two levels of ss
e
overall equivalence ratio (0.6 for high load operation and Pr
0.2 for low load operation) were considered. Two levels of 50
intake boost (3 bar and 4 bar) were investigated. Two
injection timings were studied: early, with fuel injection
from 25° BTDC to 5° BTDC and late, with fuel injection 0
0 500 1000 1500 2000
from 10° BTDC to 10° ATDC. Baseline engine operation
Volume (cc)
was defined at a speed of 1500 RPM, overall equivalence
ratio of 0.6, injection timing of 25° BTDC and intake Figure 2. Variation of volume-averaged cylinder
pressure with (a) crank-angle and (b) cylinder
pressure of 3 bar. Parametric studies were conducted by
volume for motoring and firing under baseline
varying each of the operating parameters one at a time
conditions.
from their baseline values. Before selected parametric
studies are described, the baseline case results are
Table 2 shows the values of IMEP and typical emissions
explored in detail.
like NO, unburned hydrocarbons (UHC) and CO at 150°
2
ATDC, when the exhaust valve opens in a typical direct
BASELINE ENGINE CALCULATION RESULTS
injection engine. It should be noted that UHC values
were based on the mass of methane left in the
Figure 2 depicts the variation of volume-averaged
combustion chamber at 150° ATDC. The effects of
pressure with crank angle and cylinder volume, for
crevice flows, blowby and absorption/desorption of fuel in
motoring and firing cases at 1500 RPM. Obviously, for
the oil film were not modeled.
the motoring case, the cylinder pressure is symmetrical
about the top-dead-center (TDC) position due to the
assumption of adiabatic conditions for the cylinder walls. Table 2. Engine performance and emissions for the
It can also be seen that after fuel injection begins, at 25° baseline case
BTDC, there is no discernible rise in cylinder pressure for
IMEP (bar) NO (ppm) UHC (ppmC) CO (ppm)
about 10° crank angles (CA) compared to the motoring 2
case. This is due to the ignition delay period, when the 15.83 3382 1692 65222
fuel and air are mixing with each other and pre-
combustion reactions are occurring. The numerical value
Although no validation of the predictions of exhaust
of ignition delay, defined as the time taken to burn 4% of
the total injected fuel, is 8.2° CA for this early injection emissions has been attempted as part of this study, the
values obtained lie in the general range for diesel
case. The combination of early injection timing and high
engines [11,22]. In any event, our interest is primarily
load results in very high peak cylinder pressure for the
firing case which occurs at about 4° ATDC. focused on their relative values as engine operating
parameters are changed.
4
Figure 3 shows the variation of mass-averaged directly from the injector. The pocket on the right, on the
temperature with crank angle for the baseline case. It side of the piston bowl, is due to the slow rolling of the
can be seen that there is a slight drop in temperature fuel along the side of the piston. Fig. 4 clearly shows the
during the ignition delay period, when colder fuel is being importance of fuel-air mixing in determining how soon the
injected into the cylinder, but chemical heat release has injected fuel is consumed.
not yet started taking place. After the delay period is
Figure 5 (a) shows the variation of different species
over, the average temperature in the cylinder increases
masses in the combustion chamber as a function of
rapidly to around 2350 K and then decreases during the
crank angle. Methane is injected in a slug profile from
expansion stroke when the piston starts moving
25° BTDC to 5° BTDC. After the ignition delay period is
downwards. It can also be seen from Fig. 3 that at the
over, fuel that has mixed with air during the delay period
time of fuel injection (25° BTDC), the average air
is rapidly consumed resulting in the premixed burning
temperature in the cylinder is about 1200 K.
phase. There is a slight increase in unburned methane
mass after the initial decline due to a reduction in the rate
2500
of burning following the initial premixed burn peak. After
Motoring
fuel injection ends at 5° BTDC, the unburned methane
Firing
mass continuously decreases due to consumption. The
2000
mass of CO in the cylinder is initially constant at a value
2
K) of 54.2 mg, which is the mass of CO in the intake air due
ure ( to the assumption of 10% EGR. Afte2r combustion starts,
erat 1500 CO2 concentration starts increasing and continues to
p
m increase till the end of the calculation, as fuel is being
e
T
continuously burned. It should be noted that CO is
2
1000
locally assumed to be in equilibrium with CO. This was
done to predict species needed for the NO calculations.
Due to this assumption, the rate of increase of CO mass
2
500 decreases slightly at 10° ATDC when the temperatures in
-150 -100 -50 0 50 100 150
the cylinder are locally at their highest values (see Fig. 3).
Crank Angle (degree)
However, the equilibrium assumption does not have an
Figure 3. Variation of the mass-averaged temperature impact on the values of CO2 concentrations shown in
with crank-angle for the baseline case. Table 2, because by 150° ATDC the cylinder temperature
falls down considerably. Note that the CO ppm values are
Figure 4 (shown in color) shows the variation of fuel-air lower than CO2 ppm values by more than two orders of
equivalence ratio in a plane cut through the center of the magnitude.
sector mesh, for pure injection (non firing) and the
Figure 5 (b) shows the variation of NO concentration with
baseline firing case at 10° ATDC. The non-firing case
crank angle in parts per million (ppm). It is seen that NO
shows that the fuel moves along the surface of the piston formation is slow and peaks around 15° ATDC, when
bowl while the piston is moving up, and then rolls
cylinder temperatures are very high. The critical time
upwards as the piston starts its descent. The light blue,
period for NO formation has been shown to be when the
green and yellow portions in Fig. 4 (a) are those that are
burned gas temperatures are at a maximum, i.e.,
favorable for combustion as the fuel and air are mixed
between the start of combustion and shortly after peak
within the flammability limits. By the nature of the Eddy-
cylinder pressure has occurred [11]. Mixture that burns
Break-Up (EBU) model of combustion used in this work,
early in the combustion process is especially important
the local rate of burning is highest in regions that have
as it is compressed to a higher temperature, increasing
low turbulence time-scale k/ε, and fuel-air mixtures that
the NO formation rate. During the expansion stroke, the
are in close to stoichiometric proportions. The shape of
burned gas temperatures decrease as the cylinder gases
the piston bowl plays a very significant role in aiding the
expand. This, coupled with the mixing of high-
mixing of injected fuel with air. This fact was explored in
temperature gas with air or cooler burned gas, freezes
greater detail in an independent study [15]. The firing
the NO chemistry and its concentration remains
case shows that the regions where the fuel is consumed
essentially constant during the late expansion process.
are indeed the regions in the non-firing case where the
At very high temperatures, some NO dissociates into N
equivalence ratio ranges from 0.3 to stoichiometric. Two
and O and its concentration decreases. It can be seen
distinct pockets of rich fuel-air mixture can still be seen by from Fig. 5 (b) that all the NO forms in the 25° CA period
10° ATDC in the firing case. As shown later, by this time, from the start of combustion (about 15° BTDC). Both
more than 75% of the injected fuel has been consumed.
freezing of NO and decrease in its concentration from the
The pocket on the left, on top of the piston bowl, is due to
peak value have been observed experimentally.
the accumulation of the fuel as it hits the piston top
5
of stoichiometric. In a recent experimental study of NO
400
(a) formation in diesel engines using PLIF imaging, Dec and
350 Canaan [23] found that no measurable NO was formed
Injected Methane
Unburned Methane during the initial premixed burn; instead the first NO
300
Carbon dioxide appeared in a thin layer around the fuel jet periphery just
g)
m 250 after the diffusion flame was established. In this work
ss ( too, it is seen that the first NO is formed in the outer
a
M 200
s edges of the injected methane jet where the fuel-air
e
eci 150 mixture is slightly lean of stoichiometric.
p
S
100 Figure 7 (shown in color) shows temperature and NO
densities at 10° ATDC when NO concentrations are at
50 their peak. Fuel injection was stopped at 5° BTDC, so
the intake-generated jet structure had disappeared by
0
-20 0 20 40 60 80 100 120 140 160 180 now. Burned gas temperatures show values of 3200 K in
Crank Angle (degree) the regions near the injector where some fuel may still be
present. It can be seen that most of the combustion
5000 chamber, except the squish region on the outer edges,
(b)
has been engulfed by the flame. Since most of the
injected fuel has been consumed, only the dark blue φ
4000
contour line corresponding to a value of 0.1 is seen
across the combustion chamber. There are still pockets
m) 3000 of richer fuel air mixtures present in the cylinder though,
p
O (p as shown in Fig. 4 (b). Peak NO concentrations occur in
N 2000 regions that have temperatures in the range of 2500 K-
3000 K. NO concentrations do not peak in the highest
temperature regions, but in the regions where post-flame
1000
gases reside. This is due to the slow chemistry of NO
formation. Dec and Canaan [23] found that as much as
0 1/3 of all NO produced is formed in the hot post-
-20 0 20 40 60 80 100 120 140 160 180 combustion gases after combustion ends. They also
Crank Angle (degree) reported that NO is found throughout the portion of the
combustion chamber where the reacting fuel jet had
Figure 5. Variation of (a) species mass and (b) NO with
traveled. This is also seen in Fig. 7 (b).
crank-angle for the baseline case.
Figure 8 shows the variation of methane mass fraction
Figure 6 (shown in color) shows values of temperature burned and its rate of burning with crank angle. As
and NO density in a plane cut through the center of the described before, under the baseline operating
sector mesh at 5° BTDC. This is the time by which all the conditions, it takes about 8° CA before the ignition delay
fuel has been injected and combustion has begun in is over. Till then the injected fuel and air have time to mix
earnest. The fuel is injected at an angle of 45° with and by the time the ignition delay period is over, the rate
respect to the horizontal, but due to the flow-field created of burning becomes quite high. The first peak in Fig. 8
by the upward motion of the Mexican-hat shaped piston (b) is reached when the transition from full kinetics to the
bowl, it is forced to strike the piston bowl very close to the EBU model occurs. After the transition, the burning rate
axis of the cylinder. The fuel then spreads along the drops slightly, but then continues to rise as the fuel
surface of the piston. This results in fuel rich mixtures injection process enhances turbulence and hence the
along the surface of the piston and leaner mixtures rate of mixing. However, as the injection process
towards the cylinder head. This fact can be seen in Fig. 6 continues, the in-cylinder turbulence is not able to mix the
(b) which shows contours of equivalence ratios. The red, fuel with air, and so the burning rate starts falling. This
yellow, green, light blue and dark blue lines denote φ of trend continues after fuel injection ends at 5° BTDC. The
5.0, 2.0, 1.0, 0.5 and 0.1, respectively. As already mass fraction burned curve, shown in Fig. 8 (a), depicts
discussed before, the reaction rates, and temperatures the S-shaped profile that is characteristic of combustion
are highest for near stoichiometric mixtures. Due to heat in both diesel and spark-ignition engines. It can be seen
release, the temperature of the outer edges of the gas jet that not all the injected fuel burns by the time the
increases to around 2700 K. NO starts forming at around calculation was terminated at 180° ATDC. The constants
8° BTDC, see Fig. 5 (b). It is interesting to note that for in the EBU model probably need to be increased during
regions in which NO initially forms, not only are the expansion when cylinder densities fall.
temperatures high, but the fuel-air mixture is slightly lean
6
7
injection). Full load engine operation was assumed with
1
(a) an intake pressure of 3 bar. It is seen that the pressure
starts rising earlier for the slowest speed, followed by the
0.8 medium and then the fastest speed. Figure 9 (a) clearly
b
X shows that the ignition delay period is longest for the
d,
e highest speed. However, the delay measured in seconds
n 0.6
Bur actually decreases with speed. This is because the
on piston moves up more per unit time with increasing
acti 0.4 speed, leading to an increase in pressure and
Fr
s temperature of air inside the cylinder, both of which
s
a
M 0.2 decrease ignition delay. Also, in-cylinder turbulence
increases with engine speed as the flow velocity past the
intake valve increases.
0
-20 0 20 40 60 80 100 120 140 160 180
250
Crank Angle (degree) 800 RPM (a)
1500 RPM
0.06 200 2200 RPM
(b)
θ/dXb0.05 e (bar) 150
urned, d0.04 Pressur 100 SOI
B
n
ctio0.03 50
a
Fr
ss 0.02 0
a
M -50 -40 -30 -20 -10 0 10 20 30 40 50
of
e 0.01 Crank Angle (degree)
at
R
0 250
-25 -20 -15 -10 -5 0 5 10 15 20 25 30 35 40 800 RPM (b)
Crank Angle (degree) 1500 RPM
200 2200 RPM
Figure 8. Variation of (a) methane mass fraction burned
and (b) rate of mass fraction burned with bar) 150
crank-angle for the baseline case. e (
ur
s
es 100
Results of the baseline case have shown that the ignition- Pr
combustion-pollutant formation prediction methodology
50
used in this work provides physically meaningful results.
In the following sections, various engine operating
0
parameters are changed one by one and the results
0 100 200 300 400 500
compared to the baseline case.
Volume (cc)
EFFECT OF ENGINE SPEED Figure 9. Variation of volume-averaged cylinder
pressure with (a) crank-angle and (b) cylinder
Engine speed has a significant impact on the engine's volume at three different engine speeds.
performance and emissions. With higher speed, the
ignition delay period changes as more crank angles are Table 3 shows the values of ignition delay measured both
traversed in the same time. The changed delay affects in crank angles and time for the three engine speeds
combustion, and hence pressure rise and pollutant examined. As the ignition delay in crank angles
formation. Figure 9 shows the variation of volume- increases, it takes longer for the fuel-air mixture to start
averaged pressure with crank angle and cylinder volume, burning, and consequentially the pressure rise is lower.
at three engine speeds. The medium speed corresponds This leads to a lower IMEP as the speed increases.
to the baseline case and was discussed in the previous Table 4 shows the values of IMEP and typical pollutant
section. The injection timing was kept fixed for all the concentrations at 150° ATDC.
speeds at the baseline value of 25° BTDC (early
8