1. Paper Number
Effect of EGR and Air Fuel Ratio on Gasoline HCCI
Combustion with Direct Injection
Jayneel Gajjar
Masters in Automotive Systems, Kettering University
Dr. Bassem Ramadan
Professor, Kettering University
ABSTRACT
HCCI gasoline direct injection is simulated using AVL
Fire. In this paper, charge stratification was achieved
using spray and premixed. Basically, charge stratification
was intended to get one spray during intake stroke which
allows the spray to mix well with the incoming air and
make a homogenous mixture and the second pilot spray
will be done during compression stroke. However, in order
to save simulation time, the spray during the intake stroke
was replaced with a premixed. The combustion
characteristics of HCCI engine was investigated by
changing EGR percentage and A/F ratio. Two different
combustion models i.e. Coherent Flame Model and
Characteristic Timescale model were experimented with
during the simulation. It was seen that combustion in
HCCI engine is characterized by high pressure rise
because HCCI has homogenously distributed self-igniting
spots which combust uniformly throughout the cylinder.
Since there is no spark to control the initiation of
combustion, using EGR or varying the A/F or fuel
concentration can help in controlling the combustion.
Deep Bowl is used which unlike gasoline engines uses
flat head piston. Swirl in the cavity of bowl will help to mix
the fuel sprayed during the second injection well in turn
reduce NOx Emission [1].
INTRODUCTION
HCCI engine combines the benefits of spark Ignition and
compression ignition where the charge is homogenously
mixed and auto ignited. It has higher thermal efficiency
and lower emissions. The problem with HCCI engine is
controlling the combustion rate and ignition time. Also
HCCI engines have a very low range of operation.
In order to control the combustion in gasoline HCCI
engine there are several methods including: air
preheating [2], supercharging [3], VCR [4], EGR [5] etc.
For controlling the ignition there are several other
methods like adopting dual fuel with different ignition
property [6], mixing peroxide additive to improve the
ignition property [7]. These methods of controlling add up
to the cost of vehicle and make the technology little
impractical in the market where cost plays a major factor.
Therefore, EGR and direct injection seems to be a good
strategy for control. Gasoline direct injection makes it
possible to control HCCI combustion [8,9,10].
ENGINE SPECIFICATION
Table 1 shows the engine specification for the engine simulated
in AVL Fire.
GEOMETRY
Two geometries were used for the study. Effect of EGR
was studied on deep bowl with bowl volume of 34.6 cc.
Figure 1 Sector mesh of Deep Bowl geometry used for
investigating the effect of EGR.
Sr.No Parameter Value
1 Bore 95 mm
2 Stroke 105 mm
3 Connecting Rod 198 mm
4 Volume per Cylinder 745 cc
5 Engine Speed 1400 RPM
6 Type
Naturally
Aspirated

2. The compression ratio is 11. A sector of bowl was meshed
using Pointwise. Since six injectors are being considered
and the bowl geometry is axisymmetric simulating a
sector saves lot a computation time instead of simulating
the whole cylinder. The total number of elements in the
sector is 44,640.
Since such deep bowls are not used for gasoline engines
and also the squish height is more. The bowl was made
shallow and compression ratio was changed to 14. The
total number of elements in the sector is 88,260.
Figure 2 Sector mesh for Shallow bowl used for investigating
effect of A/F ratio
COMBUSTION MODEL
As mentioned before two different combustion models
were examined. Coherent flame model (ECFM-3z) and
Characteristic time scale model.
Coherent flame model (ECFM-3z) is based on a flame
surface density transport equation and a mixing model
that candescribe inhomogeneous turbulent premixed and
diffusion combustion [11]. It basically divides a cell into 3
zones i.e. unmixed air, mixed air fuel and unmixed fuel. It
is used for gasoline as well as diesel engine simulation.
In order to test the model, 4 cases were run considering
the engine was naturally aspirated. The equivalence ratio
was set to 1 and EGR ratio of 10%, 20% & 50% were
used. Since charge stratification was used, a premixed
with equivalence ratio of 0.94 is set and the remaining fuel
of 0.6 mg was sprayed during the compression stroke at
40 deg BTDC.
The NOx emission model compatible with ECFM-3z is
Extended Zeldovich + Prompt + Fuel. The corresponding
soot model is Hiroyasu/Nagle/Strickland-Constable.
Figure 3 Pressure distribution for different EGR ratio and
equivalence ratio 1.0
It can be seen from Figure 3, that pressure rise very is
instantaneous. HCCI is characterized by instantaneous
rise in pressure because of many homogenously self-
ignited spots. The pressure is not expected to rise to
maximum over a single crank angle. The cycle replicates
an Otto cyclewhich is impractical for a real engine. There
are two parameters under ECFM 3z model to control the
rate of chemical reaction and auto ignition. The
parameters are as below -:
 Auto Ignition Model Parameter -: Default value
1. Inverse of this value is multiplied with the
ignition delay time. Therefore, value greater than
1 will reduce the ignition delay [11].
 Rate of Chemical Reaction -: Default value
10000 s. This parameter influences the rate of
reaction of fuel during premixed combustion and
auto ignition [11].
In order to study the effect of each parameter, one of
above parameter was changed and the other was
changed for ɸ=1 and EGR 10%. Below tabulated column
shows the value changed ineach parameter and red color
depicts no combustion.
Table 2 Summary of cases along with parameters changed
Cases Parameter Value
11 Auto Ignition Delay 0.5
12 Auto Ignition Delay 0.8
13 Reaction Time 7500
14 Reaction Time 8000
15 Reaction Time 8500
16 Reaction Time 9000
17 Reaction Time 11000
18 Reaction Time 20000
19 Reaction Time/AID 8500/1.5
20 Reaction Time/AID 7500/2
21 Reaction Time/AID 8500/1.75

3. The main parameter which would affectthe instantaneous
increase in pressure rise is the rate of chemical reaction.
From Table 1, it can be observed that on decreasing the
parameter there is no combustion for reaction time of
7500, 8000, 8500 because even though the temperature
reaches the ignition point for gasoline, since the rate of
reaction isdecreased the software retards the combustion
post TDC. Since the piston has already started motion
towards BDC the temperature and pressure decreases
and in turn leads to no ignition. For cases 16, 17 and 18
the combustion did occur. Cases 16 and 17 did not affect
the result. However, for reaction time of 20000 instead of
affecting the slope of pressure curve, it shifted the ignition
point little before than what was achieved for default
reaction time of 10000s. Hence, the combustion model
was changed to Characteristic time scale model.
Characteristic Time scale model isused for diesel engine.
Since in diesel engines significant part of combustion is
considered to be mixing controlled. This model takes into
account interaction between turbulence and chemical
reaction. HCCI involves compression ignition and the
shallow bowl with swirl is used. It makes it a case of
mixing controlled. Therefore, this model was tried.
Therefore, the effect of EGR and A/F ratio were
investigated using CTM. The corresponding NOx and
Soot Emission
EFFECTS OF EGR
The primary purpose of EGR is to reduce the NOx
emissions. NOx are formed at temperatures higher than
1800 K. EGR dilutes the oxygen present in the incoming
air and acts as an absorbent of heat. Thereby reducing
the adiabatic temperature. However, the use of EGR
leads to a trade-off in terms of soot emissions moreover it
exhausted more unburned hydrocarbons (20–30%)
compared to conventional engines [12].
With ɸ=1, compression ratio of 11 and EGR ratios of 10%,
20%, 30%, 40% and 50% with initial of naturally aspirated
engine the simulations were run.
Figure 4 Pressure distribution for different EGR ratios
From Figure 4, the first thing that can be observed is that
as percentage of EGR increases it reduces the peak
pressure in turn reduces peak temperature. The cases
with 0% EGR and 10% EGR were run with initial
temperature of 300K and it canbe observed that peak has
shifted to right of the graph indicating a little ignition delay.
As more EGR is added the initial temperature was for
20%, 30% and 40% was increased to 310K. As the initial
temperature increases the ignition can be seen to
advance.
From Figure5 below, the concentration of NOx decreases
exponentially with increasein EGR as expected. But there
is a limit up to which EGR can be introduced ensuring
combustion can occur. Since EGR reduces the adiabatic
temperature of combustion beyond a certain value of
EGR it will make the mixture incombustible. For gasoline
engines the preferable EGR percentage lies between
20%-30%. Also, increasing the EGR causes the
temperature to drop and causes the formation of soot
which also of concern when emission is taken in to
consideration. Therefore, the percentage of EGR is
always a tradeoff between NOx and Soot emission.
Figure 5 Changes in NOx concentration with percentage of
EGR.
EGR can be used for ignition control as well. For cases
with 20%, 30% and 40% EGR since they are same initial
temperature it can be observed from Figure 6 that EGR
retards the initiation of ignition. Therefore, it can be used
to prevents knocks ensuring no damage to piston.
Figure 6 displays the effects of EGR on the ignition point for
HCCI engine.
Also, engines do not run every time with ɸ=1. Most of the
time engines run lean. Particularly in case of HCCI
engine, it can be run really lean plus using a higher
compression ratio and charge stratification can ensure

4. combustion with lesser emission. When this is combined
with EGR, the emission reduces drastically.
EFFECT OF AIR FUEL RATIO
The effect of air fuel ratio was studied on shallow bowl
and two different compression ratios were used as
mentioned before. Using a higher compression ratio
provides better thermal efficiency and will also ensure
very lean mixture auto ignition. Also in order to achieve
charge stratification two different injection strategies
where used.
1. Type 1 -: Premixed and Injection at 320 deg CA
(40 deg BTDC)
2. Type 2 -: Injection 1 at 240 deg CA (120 deg
BTDC) & Injection 2 at 340 deg CA (20 BTDC)
Before running the cases with different charge
stratification techniques one thing needs to be ensured
that premixed charge does not cause knocking. For
ensuring this, cases with equivalence ratio of 0.5, 0.7, 0.9
& 1.0 were tried. Initially whole fuel was considered to be
premixed. But CTM model does not combust the mixture
if it does not identify spray. One of the reasons for this can
be because it is exclusively designed for diesel engines.
Therefore, 99% of fuel was kept premixed and 1% was
injected. For compression ratio 11, none of the
equivalence ratio combusted. However, on increasing the
compression ratio to 14, the mixture ignited for ɸ=0.5.
Therefore, ɸ was decreased and the least ɸ which
combusted was 0.3. However, least ɸ which can actually
burn for compression ratio of 14 is 0.4. Also combustion
occurred between 356 to 360 deg CA. Therefore, while
considering injection 2 itwas taken carethat injection gets
completed before 356 deg CA.
PART 1
Withcompression ratio 11 and type 1 charge stratification
strategy cases were run with equivalence ratio of 0.5 ,0.7
0.9 & 1.0. Inthese cases, premixed was 35% and injected
mass was remaining 65%.
Table 3 Summary of results for compression ratio 11 and type 1
charge stratification.
From table 2, it can be seen that with increase in the
amount of fuel the maximum pressure and temperature
inside the cylinder increases. The NOx emissions also
follow similar trends. Since NOx are generally formed at
temperatures above 1800K the values from case 2
onwards shoot up drastically. Other than case 1 where
NOx emission is 27 PPM none of the other cases tend to
be under the EPA acceptable limit. However, the soot
emissions are lower with maximum being for case 3 i.e.
25 PPM. Since the overall temperature in the cylinder is
high the soot production is much lower.
Injection speed used was 350 m/s. Since 65% of mass
was injected for higher equivalence ratio the injection time
was much higher than 2 ms. Typically for gasoline GDI
the injection timings are kept lower than 1 ms. Therefore,
for the case with compression ratio of 14 the premixed
was kept to 80% and injected mass was 20% in order to
ensure that injection timing is around 1ms.
PART 2
The same type 1 charge stratification was tried with
compression ratio of 14. The results are as below -:
Table 4 Summary of result for compression ratio 14 and type 1
charge stratification
As compared to results of compression ratio 11, the
maximum pressure, temperature and NOx emissions are
really high. It is as expected since the compression is
higher. However, the soot emissions are really lower
since overall temperature in the cylinder is much higher
than the temperature region where soot is produced.
On comparing the emissions for Part 1 and Part 2, the
NOx emission are higher for Part 2 sincethe compression
ratio is high the maximum temperature also increases
which in increases NOx.
Case
Equivalence
Ratio
Maximum
Pressure (MPa)
Maximum
Temperature
(K)
Nox
Emission
(PPM)
Soot
Emission
(PPM)
1 0.5 5.25 1542.8 27 15
2 0.7 6.32 1915.51 142.2 9
3 0.9 6.45 2142.59 585 25
4 1 8.26 2284.45 942 11
Case
Equivalence
Ratio
Maximum
Pressure
(MPa)
Maximum
Temperature
(K)
Nox
Emission
(PPM)
Soot
Emission
(PPM)
1 0.5 7.22 1603.56 35 15
2 0.7 9.28 2035.43 480 1
3 0.9 10.76 2325.14 1912 0.6
4 1 11.25 2409.4 2518 5.5

5. Figure 7 Comparison of CO2 and CO emission for compression
ratio 11 and compression ratio 14 with type 1 charge
stratification
From Figure 7, it can be seen that for CO2 emissions are
almost comparable for both Part 1 and Part 2 up to
equivalence of 0.9 after which the CO2 emissions fall for
compression ratio of 11 and rises for compression ratio
14. Possibly the reason can be that since the temperature
is higher it supports oxidization of CO to CO2. However,
the CO emissions show opposite trend i.e. the emissions
are higher for compression ratio 11 than for compression
ratio 14. Since same injection strategy is used but
different fuel ratios are used comparison of the emissions
of Part 1 and Part 2 cannot be 100% correlated.
PART 3
For compression ratio 14 since the emissions are higher
another charge stratification strategy i.e. was tried. For
type 2 strategy, late injection for second spray was tried.
One of the reasons for using late injection was that once
the fuel is injected the temperature and pressure at that
crank angle will be sufficient enough to vaporize the
spray. So there will not be any liquid droplet travelling to
wall and interacting with it. Also, instead of premixed the
first injection was done in compression stroke itself at 240
deg CA. Injection 1 had 40% fuel and Injection 2 had 60%
except for ɸ=0.9 & 1 where the percentages were
reversed. For injection 2 with 60% fuel will cause the
injection duration to almost overlap with piston reaching
TDC and combustion occurring even through the
injection. In order to avoid it the percentages were
reversed for ɸ=0.9 & 1. For the same equivalence ratios
mentioned above the results are as follows -:
Table 5 Summary of results for compression ratio 14 and type 2
charge stratification.
As seen from Figure 5, the trends for pressure
temperature and emissions are almost similar to previous
cases. However, NOx emission do not follow the trend for
ɸ=0.9 & 1. The possible reason can be that since both the
sprays are injected in compression stroke, the
combustion mixture is little cold as compared to ɸ<=0.8.
Therefore, overall temperature is not as high as it can go.
Hence the NOx emissions are lower. For ɸ=1 the soot is
higher since the amount fuel spray ismuch higher mixture
did not get enough time to combust which can lead to few
regions inside cylinder remaining cold leading to higher
soot.
The two injection strategy causes little ignition delay as
seen from Figure 8. However, it pushes the ignition more
towards TDC which is desired. There is a compromise in
terms of maximum pressure.
The NOx emission for type 2 are lower than type 1 for all
ɸ except for ɸ=0.4. Also, the soot emissions are higher as
compared to type 1 they are below the limit. So
compression ratio 14 and type 2 strategy was chosen to
study the combined effect of air fuel ratio and EGR.
Figure 8 Comparison of Pressure distribution for Type 1 and
Type 2 for compression ratio 14 & ɸ=0.9
Case
Equivalence
Ratio
Maximum
Pressure
(MPa)
Maximum
Temperature
(K)
Nox
Emission
(PPM)
Soot
Emission
(PPM)
1 0.4 6.31 1381.7 46 3
2 0.5 7.18 1561.2 82 3
3 0.6 7.94 1704.8 150 3.55
4 0.7 8.65 1842.6 219 3.66
5 0.8 9.44 1994.6 513 2.8
6 0.9 10.02 2096.9 472 0.8
7 1 10.3 2120.57 437 37

6. Figure 9 Comparison of NOx and Soot emissions for
Compression ratio 14 with Type 1 and 2 strategies.
COMBINED EFFECT OF A/F RATIO AND EGR
The EGR percentage for gasoline engine generally varies
between 20-30%. Since higher compression ratio of 14 is
used higher limit of EGR can be used. Therefore, using
30% EGR with compression ratio of 14 and type 2
strategy the results are -:
Table 6 Summary of results for 30% EGR for compression ratio
of 14 and type 2 charge stratification
As seen from Table 6, the NOx drops exponentially when
EGR is used. For lower ɸ the NOx emissions are almost
zero. Also, the maximum temperature inthe cylinder does
not increase beyond 1677.8 K which inhibits NOx
formation. Although lower temperature supports soot
formation , soot emissions are well under control.
CONCLUSION
1. Apart from helping in reducing NOx, EGR helps to
control the ignition by providing a little ignition delay.
Also it ensures that charge does not auto ignite way
before TDC by reducing the adiabatic temperature
which can cause higher NOx.
2. Type 2 charge stratification strategy which involved 2
sprays inthe compression stroke provided better NOx
emissions for compression ratio 14 as compared to
type 1. However, the soot emissions were higher.
3. The second injection helps in controlling the initiation
of combustion.
4. A/F ratio affects the ignition; rich mixture gets little bit
delayed ignition as compared to the lean mixture.
5. Effect of EGR, A/F ratio, injection strategy aid in
providing really low NOx which can meet stringent
emission regulations for future as well.
REFERENCES
1. Zhi Wang, JianXin Wang, Shi Jin Shuai and Qing Jun
Ma. Effects of Spark Ignition and stratified charge on
Gasoline HCCI Combustion with DirectInjection. SAE
2005-01-0137
2. Jialin Yang, Todd Culp and Thomas Kenney.
Development of a Gasoline Engine System Using
HCCI Technology – The concept and the Test
Results. SAE 2002-01-282
3. Jari Hyvonen, Goran Haraldsson and Bengt
Johansson. Supercharging HCCI to Extend the
Operating Range in a Multi-Cylinder VCR-HCCI
Engine. SAE 2003-01-3214
4. Magnus Christensen, Andres Hultqvist and Bengt
Johansson. Demonstrating the Multi Fuel Capability
of a Homogenous Charge Compression Ignition
Engine with Variable Compression Ratio. SAE 1999-
01-3679
5. Aaron Oakley, Hua Zhao and Nicos Ladommatos.
Dulution Effects on the Controlled Auto Ignition (CAI)
Combustion of Hydrocarbon and Alcohol Fuels. SAE
2001-01-3606
6. Jan Ola Olsson, Olof Erlandsson. Experiments and
Simulation of a Six Cylinder Homogenous Charge
Compression Ignition (HCCI) Engine. SAE 2000-01-
2867
7. J.A.Eng, W.R Leppard and T.M. Sloane. The Effects
of DI Tertiary Butyl Peroxide (DTBP) Addition to
Gasoline on HCCI Combustion. SAE 2003-01-3170
8. Tomonori Urushihara, Koji Hiraya, Akihiko Kakuhou
and Teryuki Itoh. Expansion of HCCI Operating
Region by the Combination of Direct Fuel Injection.
Negative Valve Overlap and Internal Fuel
Reformation. SAE 2003-01-0749
9. A Fuerhapter, W.f. Piock, G.k.Fraidi.CSI – controlled
Auto Ignition – the best solution for the fuel
consumption – versus emission trade-off. SAE 2003-
01-0754
10. A Fuerhapter, W.f. Piock, G.k.Fraidi, E.Unger. The
new AVL CSI Engine ~ HCCI Operation on a multi
cylinder gasoline engine. SAE 2004-01-0551
11. Combustion Module. Documentation AVL Fire v2014.
12. Jaffar Hussain, K. Palaniradja, N. Alagumurthi, R.
Manimaran. Effect of Exhaust Gas Recirculation
(EGR) on Performance and Emission characteristics
of a Three Cylinder Direct Injection Compression
Ignition Engine
CONTACT
Jayneel Gajjar – gajj7051@kettering.edu
Case
Equivalence
Ratio
EGR
Maximum
Pressure
(MPa)
Maximum
Temperature
(K)
Nox
Emission
(PPM)
Soot
Emission
(PPM)
1 0.5 30.0% 5.65 1260.81 0.0295 12
2 0.7 30.0% 6.38 1430.22 0.0587 23
3 0.9 30.0% 7.46 1639.53 0.172 34
4 1 30.0% 7.35 1677.8 0.22 52

7. APPENDIX
Figure 10 Series of images shows the spray cloud for Type 2
charge stratification for Injection 1 with ɸ=0.8 & 40% fuel injected
Figure 11 Series of Images shows the spray cloud for Type 2
charge stratification Injection 2 with ɸ=0.8 & 60% fuel sprayed.

8. Figure 12 Series of Images shows spray cloud for Type 1
charge stratification of Injection 2 with 20% fuel injection & ɸ=1.