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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: voshtani.sina@gmail.co

(M. Ehteram), vhosseini@sharif.edu, 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), m_reyhani54@hotmail.com (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,

nian@mech.sharif.edu (M. Reyhanian), ehteram@mech.sharif.edusseini).

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

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 919802

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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