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Investigating various effects of reformer gasenrichment on a natural gas-fueled HCCIcombustion engine
Sina Voshtani, Masoud Reyhanian, Mohammadali Ehteram,Vahid Hosseini*
Mechanical Engineering Department, Sharif University of Technology, Tehran, Iran
a r t i c l e i n f o
Article history:
Received 6 July 2014
Received in revised form
22 September 2014
Accepted 23 September 2014
Available online 17 October 2014
Keywords:
HCCI combustion
Multi-zone model
Reformer gas enrichment
Chemical kinetic model
* Corresponding author. Tel.: þ98 21 6616558E-mail addresses: [email protected]
(M. Ehteram), [email protected], vhosseihttp://dx.doi.org/10.1016/j.ijhydene.2014.09.10360-3199/Copyright © 2014, Hydrogen Ener
a b s t r a c t
Homogenous charge compression ignition (HCCI) combustion has the potential to work
with high thermal efficiency, low fuel consumption, and extremely low NOx-PM emissions.
In this study, zero-dimensional single-zone and quasi-dimensional multi-zone detailed
chemical kinetics models were developed to predict and control an HCCI combustion
engine fueled with a natural gas and reformer gas (RG) blend. The model was validated
through experiments performed with a modified single-cylinder CFR engine. Both models
were able to acceptably predict combustion initiation. The result shows that the chemical
and thermodynamic effects of RG blending advance the start of combustion (SOC), whereas
dilution retards SOC. In addition, the chemical effect was stronger than the dilution effect,
which was in turn stronger than the thermal effect. Furthermore, it was found that the
strength of the chemical effect was mainly dependent on H2 content in RG. Moreover, the
amount of RG and concentration of species (COeH2) were varied across a wide range of
values to investigate their effects on the combustion behavior in an HCCI engine. It was
found that the H2 concentration in RG has a more significant effect on SOC at lower RG
percentages in comparison with the CO concentration. However, in higher RG percentages,
the CO mass concentration becomes more effective than H2 in altering SOC.
Copyright © 2014, Hydrogen Energy Publications, LLC. Published by Elsevier Ltd. All rights
reserved.
Introduction
HCCI combustion is the auto-ignition of a premixed air/fuel
mixture without an external ignition source. The modern
history of HCCI engines began in 1979 with a two-stroke en-
gine known as the ATAC, developed by Onishi et al. [1], and
progressed in 1983 with a four-stroke engine, developed by
5; fax: þ98 21 66000021.m (S. Voshtani), [email protected] (V. Ho30gy Publications, LLC. Publ
Najt and Foster [2]. It was found that HCCI combustion is
affected mostly by chemical kinetics. HCCI combustion was
suggested as a new, preferable type to replace conventional SI
and CI combustion because of a low combustion temperature
and a highly diluted or lean mixture, which results in high
thermal efficiency and low NOx and PM emissions [3,4]. Sub-
sequently, it was found that HCCI combustion is highly sen-
sitive to parameters that affect the chemistry of combustion,
[email protected] (M. Reyhanian), [email protected]).
ished by Elsevier Ltd. All rights reserved.
i n t e rn a t i o n a l j o u r n a l o f h y d r o g e n en e r g y 3 9 ( 2 0 1 4 ) 1 9 7 9 9e1 9 8 0 919800
such as fresh charge properties and compression stroke
characteristics [5,6].
Using natural gas as a fuel has a lot of advantages,
including cleaner combustion with respect to carbon-related
substances. Several countries benefit from large resources of
natural gas, which they use widely to fuel their transportation
and generate power for other purposes. Natural gas is
composed mostly of methane with a high octane number and
is the only common fuel to exhibit relatively pure, single-stage
combustion. Other fuels have stronger low-temperature re-
actions, but methane molecules are able to resist decompo-
sition by free radicals in the compression stroke.
Natural gas combustion in HCCI combustion engines
needs high initial temperatures, achieved by high levels of
pre-heating or a high compression ratio. In addition, because
of the high rate of heat release, the mixture has to be heavily
diluted with EGR and/or excess air to prevent combustion
knocking. As a result, natural gas HCCI combustion is more
difficult to achieve in comparison to that of other liquid
fuels.
There are challenges in the implementation of an HCCI
combustion engine. First, HCCI combustion initiation depends
solely on chemical kinetics characteristics, and hence, it is
difficult to control HCCI combustion phasing to achieve high
efficiency. This leads to the second drawback of using an HCCI
engine: its limited operating range. Other challenges include
high levels of unburned hydrocarbon (HC) and carbon mon-
oxide (CO) emissions and difficulties with a cold-start [7].
Using reformer gas to control HCCI combustion phasing
has been previously suggested [8]. Reformer gas is amixture of
hydrogen, carbon monoxide, and some diluents that can be
produced on board the vehicle from anymain fuel by using an
on-board catalytic reformer [9,10]. In order to control HCCI
combustion phasing using RG, substantial engine modifica-
tion is not required. RG also removes the necessity of having
two fuels onboard, as one of the fuels can be converted to RG
using the fuel reformer.
The main effects of RG are to provide the necessary energy
to initiate combustion by thermodynamic and chemical
means. These lead to changes in engine operating parameters,
such as expanding flammability limits, increasing dilution
and EGR tolerance, reducing cold start emissions, and
improving efficiency. There are several reformer technologies
to produce hydrogen from a hydrocarbon fuel [11,12].
Several experimental studies have been done using RG as
an HCCI combustion enhancer/controller using various fuels
[8,13]. Shudo et al. examined the effect of variable blend
fractions of methanol reformer gas (MRG) and DME [14,15].
Eng et al. in an experimental chemical kinetics modeling
study examined the effects of POx RG on HCCI combustion of
n-heptane and iso-octane with two strategies of internal EGR
through exhaust re-breathing and external well-mixed EGR
[16]. Hosseini and Checkel [8,9,17e19] investigated the effects
of RG addition on HCCI combustion of a series of low- and
high-octane primary reference fuels (PRFs), natural gas, and n-
heptane. In their investigation, RG altered combustion timing
and expanded the operating region. In addition, they found
that at a constant l and EGR, increasing the RG concentration
increases NOx emissions and the rate of pressure change and
decreases HC and CO emissions.
The effects of the addition of hydrogen on HCCI combus-
tion characteristics have been investigated numerically.
Elkelawy et al. used a zero-dimensional thermodynamics
model to study the combustion enhancement of a natural gas
HCCI engine using hydrogen as an additive. They found that a
small quantity of hydrogen added to the natural gas enhanced
the auto-ignition process and resulted in advancement of SOC
[20]. Guo et al. also investigated the influence of hydrogen
enrichment on the combustion characteristic of n-heptane by
using a multi-zone model. They reported that hydrogen
addition retarded combustion phasing due to the dilution and
chemical effects, and the dilution effect is more dominant
because of the introduction of the methods of an artificial
inert component [21].
The idea of a single-zone model is that all properties in-
side the combustion chamber are considered to be uniform.
Not considering the temperature gradient developed during
compression in this model and neglecting boundary layer
zones leads to an over-prediction of the peak pressure, peak
temperature and NOx emission. In addition, a single zone
model cannot accurately predict combustion duration, heat
release rate and the rate of pressure change since the model
inherently predicts that all of the mixture ignites at one time.
The multi-zone model is used to overcome the limitations of
a single zone model by accounting for the spatial in-
homogeneity, especially between core regions and bound-
aries. Instead of considering the space inside a combustion
chamber as one big volume, it is divided into several small
volumes (zones or cells). The number of zones and zone
configurations has been varied depending on a compromise
between accuracy and computational efforts. Several studies
have already been done on a multi-zone model. A compre-
hensive model was developed by Fiveland and Assanis [22].
They could predict HCCI engine performance and emissions
by means of full-cycle simulation code. This model consists
of a core zone, a boundary layer zone and crevice zone.
Komninos proposed a multi-zone model to simulate HCCI
combustion [23]. In this model mass and heat transfer be-
tween zones considered. The combustion parameters pre-
dicted well with this model. Kongsereeparp and Checkel [10]
presented a new multi-zone model. In this model only work
and heat transfer considered between zones and there was
no mass transfer between them. The simulated inecylinder
pressure showed good agreement with experimental results.
Recently, Neshat and Khoshbakhti developed a new
comprehensive multi-zone model. This model contained
crevice zone, boundary layer zone, outer zones and core
zone. Mass and heat transfer considered between zones in
this model. Combustion parameters and emissions predicted
well with this model [24].
Proper operation of an HCCI combustion engine with
natural gas is difficult mainly because of the fuel resistance
to auto-ignition. Previous studies have shown that RG is
capable of reducing intake heating and lowering the
required compression ratio of natural gas-fueled HCCI
combustion [19]. The current study emphasizes the under-
standing of various effects of RG blending by developing a
method of using artificial inert components in the HCCI
combustion of natural gas using an in-house stand-alone
chemical kinetics model. In addition, the concentration of
Table 1 e Main components of the engine hardware inFig. 1.
Item# Description
1 Intake air mass flow meter
2 Intake air pulsation damping barrel
3 Throttle valve
4 Intake heater
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species (COeH2) in the RG composition was widely varied in
order to show the effect of RG composition on combustion
behavior in an HCCI engine. Therefore, zero-dimensional
single-zone and quasi-dimensional multi-zone chemical
kinetics models were developed. The models were validated
against the experimental results of a single-cylinder CFR
engine.
5 RG injector
6 Intake plenum
7 Fuel injector
8 CFR engine
9 Intake valve
10 Exhaust valve
11 Exhaust back pressure valve
12 To the building exhaust system
13 EGR pipe, insulated and water cooled
14 EGR valve
Experimental setup
A single-cylinder cooperative fuel research (CFR) engine was
modified to operate in HCCI combustion mode. Previous
studies have explained the details of the experimental setup
[9,19]. The intake system of the engine was equippedwith two
port fuel injectors of natural gas and RG. An insulated external
EGR line and an electrical air heater were implemented in the
intake system. The engine was supercharged to overcome the
low IMEP and high internal friction of the engine operating
under HCCI conditions.
A schematic of the engine and the main experimental
components is provided in Fig. 1 and Table 1. The operating
conditions are summarized in Table 2 as well.
Experiments were performed at steady state conditions.
The intake mixture temperature, engine speed, and
compression ratio were kept constant at 140 �C, 800 RPM, and
17.5, respectively. Natural gas was delivered from a high
pressure tank. Simulated RG was provided from a high pres-
sure cylinder. The RG composition was fixed at 75% hydrogen
and 25% carbonmonoxide by volume. The reformer gas blend
fraction, RGmass, was calculated using the mass flow rate of
the base fuel _mfuel, and the mass flow rate of RG _mRG, which is
defined by:
RGmass% ¼ _mRG
_mRG þ _mfuel� 100 (1)
The relative air to fuel ratio, l, was defined considering
both fuels. Keeping l constant, EGR was calculated using the
measured volumetric concentration of CO2 upstream and
downstream of the engine as follows:
EGR% ¼ CO2;Upstream
CO2;Downstream� 100 (2)
As l is kept constant during experiments, increasing the RG
mass fraction results in decreasing the base fuel flow rate.
Fig. 1 e Schematic of the engine lab hardware, describe in
Table 1.
Numerical investigation
Model description
A chemical kinetic model has been developed to investigate
the kinetic, thermodynamic, and dilution effects of RG
blending. The fraction of species (COeH2) in the RG composi-
tion was changed to show the effect of reactions on the
combustion behavior of natural gas in an HCCI engine. The
model simulates one closed cycle of an HCCI engine.
The mixture was assumed to be an ideal gas with lumped
properties in each zone and the pressure throughout the
chamber was considered to be uniform. The simulation was
performed from inlet valve close (IVC) to exhaust valve open
(EVO). During the intake and exhaust processes, the gas ex-
change sub-model was combined with the single-zone model
and was implemented to provide a reasonable estimation of
various parameters at intake valve closing conditions (initial
conditions). It should bementioned that a single zonemodel is
simply a one-zonemulti zonemodelwhich is firstly developed
to predict SOC and then implemented in iterative solving in
gas exchange process. Therefore, all the following simplifica-
tions and equation are also applicable to a single zone model.
When the temperature was less than 700 K, in order to
reduce the computational time, the composition and re-
actions were assumed to be frozen for each zone. The
boundary interactions between zones are work and heat ex-
changes, and the mass transfer between zones is not
Table 2 e CFR engine specifications and operationconditions.
Engine model Waukesha CFR
Engine type Water cooled at atmospheric
pressure, single cylinder
Displacement (cm3) 612
Bore � stroke (mm) 82.6 � 114.3
Connecting rod (cm) 24
Compression ratio 16 to 21
Engine speed (RPM) 800
Atmospheric pressure (kPa) 93.5 ± 0.7
Throttle Fully open
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considered. Moreover, the heat transfer sub-model includes
convection from gases to walls for the last zone, radiation to
walls, and conduction between zones.
As shown in Fig. 2, the core volume was fixed at the
center of the combustion chamber and surrounded by outer
concentric annular shells with a thickness of t. The number
of shells can be changed to give varying degrees of
accuracy.
Start of combustion definition
To predict of the start of combustion, two methods can be
employed: heat release analysis and the third pressure de-
rivative. In the former method, heat release calculation has
been developed by Heywood (see chapter 9 of Ref. [25]).
Combustion starts when the gross heat release, GHR, reaches
10% of maximum of GHR. In the latter method, SOC is defined
as the crank angle position in which the third derivative of the
pressured3P=dq3, exceeds a determined limit [26]:
d3P
dq3>d3P
dq3
�����lim
(3)
Examining experimental values of SOC against cylinder
pressure data revealed that a value of 5 kPa=CAD3 is an indi-
cator of SOC for the CFR engine.
The third derivative is numerically calculated using a first-
order Taylor series with second-order accuracy considering 5
points at each time interval with Newton's backward differ-
ence formula:
d3P
dq3ðiÞ ¼ 5PðiÞ � 18Pði� 1Þ þ 24Pði� 2Þ � 14Pði� 3Þ þ 3Pði� 4Þ
2h3
þ Oðh2Þ(4)
Governing equations
The one-dimensional energy conservation and ideal gas
equations are combined with the species conservation
Fig. 2 e Schematics of combustion chamber in multi-zone
model.
equation to calculate temperature, pressure, and species
concentration for each time-step. The details of original gov-
erning thermodynamic equations for this multi-zone model
can be found in Refs. [6,10]. In addition, the chemical kinetic
calculation is based on the CHEMKIN manual [27]. Finalized
equations are represented as follows:
dVk
dt¼ mkVcylPNz
k¼1 mkRkTk
"RkTk
Vcyl
dVcyl
dtþ Rk
dTk
dtþ Tk
dRk
dt
� RkTkPNzk¼1 mkRkTk
XNz
k¼1
�mkRk
dTk
dtþmkTk
dRk
dt
�#(5)
dPdt
¼ 1Vcyl
�XNz
k¼1
�mkRk
dTk
dtþmkTk
dRk
dt
�� 1Vcyl
dVcyl
dt
XNz
k¼1
ðmkRkTkÞ
(6)
Cpk
dTk
dt¼ �
XNs
i¼1
uk;idYk;i
dt� RkTk
Vcyl
dVcyl
dt� Tk
dRk
dt
þ RkTkPNzk¼1 mkRkTk
XNs
i¼1
�mkRk
dTk
dtþmkTk
dRk
dt
�(7)
dYk;i
dt¼ _uk;iMWi
rkþXNE
n¼1
_mn
m
�Ycyl
i � YInleti
�(8)
dmcyl
dt¼ dmk
dt¼ 0 (9)
Where Tk, Vk, mk, and Rk, are respectively the temperature,
volume,mass and average gas constant onmass basis. Cpk and
uk,i which are calculated from the NASA polynomials [28],
represent the average specific heat constant at constant
pressure and the total internal energy. Furthermore,Yk,i, uk,i,
MW,i and rk are the mass fraction of species, the rate of
chemical production, molecular weight and average density,
respectively. It must be noted that the k, i, cyl respectively
indicate the zone number, species number, and cylinder. Heat
transfer sub-modeling includes convective heat trans-
ferdQconvection, to the cylinder wall, heat conduction between
zones, dQconduction, and radiation from zones towalls, dQradiation.
dQk
dt¼ dQconvection
dtþ dQradiation
dtþ dQconduction
dt(10)
The convective heat transfer, hc, between the cylinder wall
and the boundary zone was calculated based on the modified
Woschni correlation for the HCCI engine [29,30].
hcðtÞ ¼ bHðtÞ�0:2PðtÞ0:8TðtÞ�:0:73nðtÞ0:8 (11)
where b is a factor to account for engine geometry, H in-
dicates the instantaneous height of the chamber, and n rep-
resents the contribution of the gas velocity during convection
heat transfer. To calculate the radiation emissivity of the
mixture, the effects of gray gas composition (CO2, H2O and
CH4), temperature, and pressure were taken in to account
[31]. The conduction heat transfer between zones is calcu-
lated by using Fourier's law (Equation (12)). To find temper-
ature gradient between zones, a linear variation of
temperature around zone boundaries is assumed. Total
i n t e r n a t i o n a l j o u r n a l o f h y d r o g e n en e r g y 3 9 ( 2 0 1 4 ) 1 9 7 9 9e1 9 8 0 9 19803
conductivity (Ktot) is the sum of laminar and turbulence
conductivity (Equation (13)).
dQconduction
dt¼ �ktotA
vTvx
(12)
ktot ¼ klam þ ktur (13)
The laminar conductivity is calculated from the mixture-
averaged properties of ideal gas transport theory, while the
turbulent conductivity is computed using Yang's work [32].
Equation (14) is used for calculating ktot.
ktur ¼ klam � PrlamPrtur
� mtur
mlam
(14)
where mtur=mlam is the ratio between the turbulent viscosity to
the laminar viscosity and is calculated from
mtur
mlam
¼ kyþn
�1� exp
�� 2akyþn
(15)
In this equation k is Von Karman constant and equals to
0.41, a is a constant and equals to 0.06. For simplicity,yþn is kept
constant and is estimated from:
yþn ¼ u*
mw
Zbore=20
rdyn (16)
Where u* is characteristic velocity which assumed to be pro-
portional to piston speed with a constant.
The GRI-Mech 3.0 mechanism [28] was selected for natural
gas combustion. This mechanism contains 325 elementary
chemical reactions, including their associated rate coefficient
expressions, and thermochemical parameters for 53 species.
No further modifications were made to the mechanism for RG
addition, as RG complements were already included in it.
The flowchart of the code is given in Fig. 3. At first, after
arbitrarily determining geometry and number of zone, the
initial condition is estimated by comparison with experi-
mental data. This arbitrary setting only occurs in the first
iteration of program. However, after the first iteration, the
average initial value (temperature and composition) is ob-
tained from gas exchange iterative process. To find IVC tem-
perature and species mass concentration for each zone,
distribution of mass and temperature are calculated based on
inhomogeneity scheme which is discussed in Ref. [33]. To
reduce the computational time, the chemical reactivity is
considered to be negligible ( _u ¼ 0) up to the crank angle in
which the temperature is less than 700 K and user-defined
time-step was fixed at 1� CA. After this crank angle up to
20� ATDC, the time step reduced to a fine value of 0.1�CA. Ineach time step, ideal gas law, energy and species conservation
equation solved simultaneously to calculate temperature,
volume, and composition. Hence, the initial values of matrix
in ODE system solver were assigned. In the ODE solver, the
rate of temperature, volume, pressure, and species concen-
tration calculated based on backward differentiation formulas
with variable order solver. Then, heat transfer subroutines are
called, and all equations solved again including the interac-
tion between zones to reach a unified pressure and accurate
volume, temperature and species concentration. Finally, the
computed values are used in next time step calculations as
initial assumptions, and the algorithm of the problem imple-
mented for all time steps until EVO.
Results and discussion
Model verification
Cylinder pressure traces were used for the purpose of model
verification. Fig. 4 shows a comparison of an experimental
pressure trace with selected operating conditions, including
single-zone and multi-zone modeling results.
From 100 cycle samples, the cycle with the median
maximum pressure was selected for the purpose of compar-
ison. Fig. 4 shows that the multi-zone model can accurately
estimate the pressure trace through an entire HCCI combus-
tion cycle; however, it overestimates expansion pressure.
Owing to the homogenous nature of a single-zone model, it
can accurately predict SOC; however following the exact
pressure trace is problematic. Increasing the number of zones
from 1 (SZM) to 10 improved the accuracy of the model in
predicting cylinder pressure. Fig. 5 compares the pressure
trace of the 10 zones multi-zone model with experimental
data for five different operation conditions. The model and
experiment are in good agreement in this figure. The specifi-
cation of this cases provided in Table 3 including RG, EGR, and
lambda to approve the verification in wide range of RG for
following investigation.
Furthermore, SOC was selected as a representative com-
bustion parameter to evaluate the model in different cases.
Fig. 6 indicates that the model has a reasonable capability in
predicting combustion behavior trends for natural gas-fueled
HCCI combustion for various cases. The error bars in Fig. 6
shows the uncertainty of the cyclic variation.
RG effects on combustion-timing
A previous investigation of RG blending shows that increasing
the RG blend in the mixture advanced SOC and reduced
combustion duration for the fuel with a high-octane number
[17].
Fig. 7 indicates the effect increasing RG from 0% to 80% has
on SOC using the multi-zone model. In the case of an EGR
fraction of 30% and l ¼ 1.9, when the RG is near 0%, the
combustion was in partial burning and misfiring mode,
meaning several zones of the outer layer or all of them did not
ignite at all as the SOC is far from IVC. This shows the capa-
bility of RG enrichment to improve natural gas HCCI com-
bustion without increasing the compression ratio or intake
heating.
Increasing the RG fraction progressively advanced com-
bustion timing and moved combustion characteristics from
misfiring and partial burning to knocking conditions, as the
SOC get further away from IVC, covering the full operating
range of the HCCI combustion engine in the desired operating
conditions.
Fig. 8 shows the effect of RG blending on the specific heat
ratio during the end of the compression stroke. An increase in
the ratio of specific heat led to higher temperatures during the
end of the compression stroke; therefore, the combustion took
Fig. 3 e Flowchart for multi-zone algorithm.
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place earlier. The ratios of specific heats for hydrogen and
carbon monoxide as diatomic molecules are approximately
1.41 and 1.40, respectively. The specific heat ratio for natural
gas is 1.32. Consequently, because of the thermodynamic ef-
fect, adding RG increases the specific heat ratio of the mixture
and advances the combustion phasing.
The participation of RG in reactions changes the chemical
kinetics of combustion in an HCCI engine. The chemical ef-
fects of RG blending in natural gas-fueled HCCI combustion
were investigated in a way similar to the analysis done by Guo
and Neil [21] for n-heptane.
Instead of H2 and CO in RG composition, two dummy inert
species named FH2 and FCO, which have the same thermo-
dynamic properties of H2 and CO, were added. These inert
species do not contribute in the reactions, meaning the dif-
ference between combustion phasing with inert species and
with real RG composition demonstrates the chemical effect of
RG addition.
Fig. 9 was the result of modeling efforts in three different
cases. In the reference case (dot-dashed black line), real RG
with complete chemical and thermodynamic properties was
used. As indicated before, increasing the RG fraction advanced
combustion timing considerably. For the case of dummy
species of FCO and FH2, which remove chemical effects,
combustion phasing by RG addition shows no real change
from the baseline. This shows that combustion phasing is
solely a chemical effect of RG. The difference between the two
cases shown reveals the degree to which SOC is altered by the
chemical effect of RG addition, the most significant of the ef-
fects of RG. This result is in contrast with the effect of H2
addition in normal heptane-fueled HCCI engine, which was
studied by Guo and Neil [21]. They showed that the addition of
Fig. 5 e Comparison of experimental and calculated
Fig. 4 e Comparison of experimental and numerical
simulation pressure traces to show the influence of zone
count at 21% EGR and 20% RG.
i n t e r n a t i o n a l j o u r n a l o f h y d r o g e n en e r g y 3 9 ( 2 0 1 4 ) 1 9 7 9 9e1 9 8 0 9 19805
hydrogen retards the combustion phasing of a normal
heptane-fueled HCCI engine because of dilution and chemical
effects, with the dilution effect being more significant. In
addition, to independently identify the dilution and thermo-
dynamic effects, the RG was replaced by two dummy species,
FDH2 and FDCO, which have the same thermodynamic prop-
erties of the fuel. These inert species do not contribute in the
reactions and therefore both thermal and chemical effects of
the real RG were negated. The dilution effect is indicated Fig. 9
with a solid (blue) line.
The pre-ignition chemistry of RG enrichment in natural
gas-fueled engines was investigated by considering the main
reactions and active radicals. Methane auto-ignition initiates
with two chain reactions, R52 and R118, which are dependent
on temperature range [34]. Fig. 10 indicates that the rate of
these reactions increased with RG enrichment. However, the
rate of R52 overcomes the rate of R118. It can be inferred that
increasing the concentration of H2 and CO as third bodies in
R52 conducts the auto-ignition reaction path of methane.
pressure trace for examined cases in Table 3.
Fig. 7 e Effect of RG blend fraction on SOC for HCCI engine
at 30% EGR and l ¼ 1.9.
Table 3 e Comparison of experimental and numericalsimulation for SOC prediction at four considered cases.
Start of combustion
Case Multi-zone model Experiment
1:
0@ l ¼ 3:95
EGR ¼ 0%RG ¼ 0%
1A 0.1CA 0.6CA
2:
0@ l ¼ 3:69
EGR ¼ 0%RG ¼ 10%
1A �4.6CA �4.4CA
3:
0@ l ¼ 2:93
EGR ¼ 22%RG ¼ 0%
1A 2.4CA 3.0CA
4:
0@ l ¼ 2:85
EGR ¼ 21%RG ¼ 20%
1A �5.0CA �4.7CA
5:
0@ l ¼ 1:9
EGR ¼ 30%RG ¼ 30%
1A �2.7CA �2.5CA
i n t e rn a t i o n a l j o u r n a l o f h y d r o g e n en e r g y 3 9 ( 2 0 1 4 ) 1 9 7 9 9e1 9 8 0 919806
H2 and CO have been implemented as third bodies with a
positive enhancement factor in the 52nd reaction of the GRI-
mech3 mechanism according to Table 4, where A, b, and E
are Arrhenius constants used to calculate the reaction rate.
Furthermore, RG enrichment causes the production of
active radicals. This is investigated by the following reactions,
initiated by R52: R38 (hydroxyl production), R53, R11, and R98.
Fig. 11 indicate the rate of two reactions, R38 and R53,
which compete to consume H2. R38 plays the role of enhancer
in the ignition process to produce more active radicals, while
R53 plays an inhibiting role. Fig. 11 shows that by increasing
RG, the rate of R38 increases more than that of R53, which
results in increased production of OH by this reaction. More-
over, according to Table 4, the rate of R98, which is the main
reaction consuming OH, greatly increases. It is found that R38
as a chain-branching reaction is essential to produce OH and
cause the methane molecule to break.
RG composition effect
Many studies have focused on producing RG with H2 content
of 99.9% for the purpose of fuel cell applications. However, for
application in IC engines, the type and condition of reformer
Fig. 6 e A comparison of SOC predicted by SZM and
experimental results of cases indicated in Table 3.
can be more flexible as both H2 and CO are flammable gases
that contain chemical energy. For example, replacing RG 50/50
with RG 75/25 (75% H2 and 25% CO) increased the mixture'sresistance to auto-ignition and limited the maximum RG
blending fraction [9]. This is due to more auto-ignition resis-
tance of H2 in comparison to CO. Hence in this section, to
identify the effect of RG on SOC, the composition of RG is
widely varied.
Fig. 12 shows that, because of the greater strength of the H2
chemical effect, increasing H2% advanced combustion timing
for H2 fractions above 10% and a constant RG mass fraction.
RG enrichment for the mixture containing less than 3% H2
retards SOC due to the dilution effect of CO in RG. In fact,
because of the small amount of H2, it cannot intensify the
radicals producing reactions. In this case, RG addition with
large amounts of CO (about 20 times higher Cp) causes com-
bustion to retard. From 3% to almost 10%H2 in RG, SOC did not
change with increasing RG in the total mixture. In other
words, in this interval, increasing H2 in the combustion
mixture advanced SOC due to the chemical effect whereas CO
addition, due to dilution effect, retarded combustion timing.
HCCI engine at constant EGR and l The effect of the H2 and
COmass fraction in the combustion mixture on SOC has been
independently investigated in Fig. 13. This figure indicates
Fig. 8 e Effect of RG on specific heat ratio for CNG-based
HCCI engine at 0% EGR and l ¼ 1.9.
Fig. 9 e Chemical, thermodynamic and dilution effects of
RG blending on start of combustion.
Fig. 10 e Effect of RG enrichment on the rate of initiation
chain reaction, R52 and R118, of natural gas auto-ignition.
Fig. 11 e Effect of following reaction of R53 on SOC by
producing active radicals.
i n t e r n a t i o n a l j o u r n a l o f h y d r o g e n en e r g y 3 9 ( 2 0 1 4 ) 1 9 7 9 9e1 9 8 0 9 19807
constant values of SOC in terms of the H2 and CO mass frac-
tion. For example, 20% H2 and 10% CO means 70% CH4 in the
mixture. In this case, the estimated SOC is between 12.7 and
13.2 CAD before TDC.
Table 4 e Main chain reaction of methane auto ignition mecha
GRI-mech3
Reaction A
52. CH4(þM) ¼> H þ CH3(þM) 1.39E þ 16
H2
H2O
CH4
CO
CO2
C2H6
AR
Enhanced b
Enhanced b
Enhanced b
Enhanced b
Enhanced b
Enhanced b
Enhanced b
118. HO2 þ CH3 ¼> O2þCH4 1.00E þ 12
38. H þ O2 ¼> O þ OH 2.65E þ 16
53. H þ CH4 ¼> CH3þH2 6.60E þ 08
11. O þ CH4 ¼> OH þ CH3 1.02E þ 09
98. OH þ CH4 ¼> CH3þH2O 1.00E þ 08
As previously mentioned, RG has three effects on com-
bustion timing, kinetic, thermodynamic, and dilution, of
which the kinetic effect is most dominant. Consequently,
increasing the H2 mass fraction up to 10% causes the com-
bustion to advance. However, for an H2 mass fraction of more
than 10%, combustion timing slightly changed. This fact
shows that when the kinetic effect of H2 reaches a saturation
limit, the effect of increasing H2 on combustion timing
weakens.
Briefly, Fig. 13 also indicates that:
- For each constant mass concentration of CO, SOC
advanced by increasing the mass concentration of H2. It
proves that the H2 enrichment always increases the reac-
tivity of a reaction in the chemical kinetic mechanism of
methane.
- As a result of altering the RG composition, it could be
inferred that a lower H2 fraction of less than 10% has a
substantial effect on controlling SOC. However, for a higher
fraction of H2 in the RG composition, the amount of CO is
critical and it should be precisely fixed to control com-
bustion timing. Finally, it was shown that not only the
amount of RG in the mixture is essential, but also pre-
dicting and fixing the appropriate fraction of species in RG
composition is necessary.
nism (GRI-mech3).
b E
�0.5 536.0
y 2.000E þ 00
y 6.000E þ 00
y 3.000E þ 00
y 1.500E þ 00
y 2.000E þ 00
y 3.000E þ 00
y 7.000E � 01
0.0 0.0
�0.7 17041.0
1.6 10840.0
1.5 8600.0
1.6 3120.0
Fig. 12 e Effect of H2 mole fraction variation on SOC (�ATDC)
for different amount of RG blending in CNG-based HCCI
engine at 0% EGR and l ¼ 1.9.
i n t e rn a t i o n a l j o u r n a l o f h y d r o g e n en e r g y 3 9 ( 2 0 1 4 ) 1 9 7 9 9e1 9 8 0 919808
Conclusion
The main concern with HCCI combustion engine application
is the lack of direct method to control ignition timing. One
solution to this problem is reformer gas enrichment with
various properties and different ratios. Therefore, in this
study, single-zone and multi-zone thermodynamic-kinetic
models have been developed to investigate the various effect
of RG blending on combustion alteration in a natural gas-
fueled HCCI engine. The numerical results also agreed with
measured data collected in a wide range of conditions. This
investigation of the combustion behavior of an HCCI engine
has shown acceptable levels of precision in calculating the
start of combustion.
Experimental preparation has been implemented for a CFR
engine to separately deliver CNG and RG into an engine. This
also operates at steady state conditions with an intake heater
and supercharged intake pressure. Moreover, numerical
Fig. 13 e Effect of different H2 and CO mass fraction in
combustion mixture on SOC for CNG-based.
calculation of combustion initiation was performed using two
distinct methods, heat release analysis and third pressure
derivative. A multi-zone model with 10 zones has been
properly verified with measured pressure, and it was well-
suited to investigate the effect of varied RG fractions on SOC.
The effect of RG blending on HCCI combustion was
examined. The RG% was varied to measure its impact on
combustion timing. Chemical kinetic, thermodynamic, and
dilution effects on SOC arising from the addition of RG were
investigated. It was shown that with the method of dummy
species manipulation, the chemical effect was stronger than
thermal effect, which in turn was stronger than the dilution
effect.
As H2 and CO have been implemented as third bodies with
a positive enhancement factor in the 52nd reaction of the GRI-
mech3 mechanism, it can be inferred that increasing the
concentration of H2 and CO as third bodies in R52 conducts the
auto-ignition reaction path of methane. Furthermore, Re-
actions R38 and R53 compete to consume H2. R38 plays the
role of enhancer in the ignition process to producemore active
radicals, while R53 plays an inhibiting role. By increasing RG,
the rate of R38 increases more than that of R53, which results
in increased production of OH by this reaction.
Various RG% blending ratios and a wide range of CO and H2
mass fractions in the combustion mixture were examined.
When H2 was lower than 10%, the amount of H2 in the com-
bustion mixture had a significant effect on SOC and the effect
of CO mass fraction could be ignored. However, after a satu-
rated level of H2, the CO mass fraction significantly altered
SOC.
Nomenclature
A Arrhenius coefficient
B bore diameter
C0p specific heat constant
Ei activating energy of ith reaction
F/A fuel-air ratio
G Gibb's free energy
GHR gross accumulative heat release _J_
H0k absolute enthalpy of the kth species
Hf enthalpy of formation
IVC intake valve closing
Kc equilibriumconstant based on species concentration
Kp equilibrium constant based on pressure
MWk molecular weight of the kth species
N engine speed (rpm)
NRHR net rate of heat release _J/CAD_
Q total heat transfer into/out of a system chamber
R ratio of connecting rod length to crank radius
Ru universal ideal gas constant on molar basis
RG reformer gas
S0k absolute entropy of the kth species
SOC start of main combustion
T in-cylinder temperature
TW wall temperature
U total internal energy inside a combustion chamber
Xk mole fraction of the kth species
Yk mass fraction of the kth species
aik the ith thermodynamics coefficient of the kth species
i n t e r n a t i o n a l j o u r n a l o f h y d r o g e n en e r g y 3 9 ( 2 0 1 4 ) 1 9 7 9 9e1 9 8 0 9 19809
dp3Lim threshold limit for 3rd derivative of pressure trace
CP average specific heat constant at constant pressure
(on mass basis)
CV average specific heat constant at constant volume
(on mass basis)
h convective coefficient
kfi forward rate constant of ith reaction
kri reverse rate constant of ith reaction
mcyl mass inside a cylinder
rc compression ratio
t time
uk total internal energy of the kth species
_uk rate of production of the kth species
q crank angle position
4 equivalent ratio
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