SAE Technical Paper

Abstract

Presently, urban transportation highly depends on the fossil fuels, but its rapid fluctuating economic issues and environmental consequences impose the variegation of energy sources. Hydrogen enriched compressed natural gas (HCNG) engines offer the potential of higher brake thermal efficiency with low emissions, which also satisfies the strict pollutant emission standards. The two-zone turbulent entrainment quasi-dimensional combustion model is developed to predict the combustion process of spark-ignited hydrogen enriched compressed natural gas-fueled engines. The fundamentals of thermodynamic process, turbulent flame propagation model and other sub-models like laminar burning velocity, adiabatic temperature and ignition lag model are introduced for the better accuracy. The experiments have been conducted for three different fuels; pure CNG, 20% HCNG, and 40% HCNG blends under MAP of 105 kPa for various excess air ratios (λ) and ignition timing (θ i). The three calibration coefficient of the model; Turbulent intensity coefficient C 2 , the Taylor length scale coefficient C 3 , and Ignition lag coefficient C ig are tuned to generate the pressure traces which closely resembled to experimental results. After comparing the numerical simulation results with the experiment's outcomes it is found that the predictive accuracy of the presented model is quite impressive , and it is well accepted for the extremely fuel lean conditions where issues of bad combustion become serious.

Full text

2018-01-1687 Published 10 Sep 2018
© 2018 SAE International. All Rights Reserved.
Study of Turbulent Entrainment Quasi-Dimensional
Combustion Model for HCNG Engines with Variable
Ignition Timings
Roopesh Kumar Mehra, Fanhua Ma, and Duan Hao Tsinghua University
Romualdas Juknelevičius Vilnius Gediminas Technical University
Citation: Mehra, R.K., Ma, F., Hao, D., and Juknelevičius, R., “Study of Turbulent Entrainment Quasi-Dimensional Combustion Model
for HCNG Engines with Variable Ignition Timings,” SAE Technical Paper 2018-01-1687, 2018, doi:10.4271/2018-01-1687.
Abstract
Presently, urban transportation highly depends on the fossil
fuels, but its rapid uctuating economic issues and environ-
mental consequences impose the variegation of energy
sources. Hydrogen enriched compressed natural gas (HCNG)
engines oer the potential of higher brake thermal eciency
with low emissions, which also satises the strict pollutant
emission standards. e two-zone turbulent entrainment
quasi-dimensional combustion model is developed to predict
the combustion process of spark-ignited hydrogen enriched
compressed natural gas-fueled engines. e fundamentals of
thermodynamic process, turbulent ame propagation model
and other sub-models like laminar burning velocity, adiabatic
temperature and ignition lag model are introduced for the
better accuracy. e experiments have been conducted for
three dierent fuels; pure CNG, 20% HCNG, and 40% HCNG
blends under MAP of 105kPa for various excess air ratios (λ)
and ignition timing (θi). e three calibration coecient of
the model; Turbulent intensity coecient C2, the Taylor length
scale coecient C3, and Ignition lag coecient Cig are tuned
to generate the pressure traces which closely resembled to
experimental results. Aer comparing the numerical simula-
tion results with the experiment’s outcomes it is found that
the predictive accuracy of the presented model is quite impres-
sive, and it is well accepted for the extremely fuel lean condi-
tions where issues of bad combustion become serious.
Introduction
There are numerous experimental researches of the
hydrogen enriched compressed natural gas (HCNG)
engines is performed during the past 20years, which is
focused on its performance and emission characteristics [1].
e results of these experimental research studies revealed that
hydrogen enrichment can improve efficiency of the CNG
engines as well as decrease the emissions by using lean burn
technology [2]. e blends of hydrogen and natural gas can
boost the combustion process eciently by recalibrating the
engine parameters. Moreover, it extents the lean-burn limit of
CNG engines very signicantly which results in decrease in
fuel consumption and harmful emissions [3]. e cycle-by-cycle
variation is also one of the parameters to evaluate the stability
of the engine. In the case of HCNG engines, the cycle-by-cycle
variation reduces signicantly as compared to the CNG eng ines,
which displays the benets of hydrogen enrichment [4].
Generally, the research, design and development of HCNG
engines have originated from CNG engines. is new fuel can
beused in CNG engines a er some modications in the indig-
enous design. Nevert heless, there is a need to recalibrating CNG
engines, especially for ignition timing, excess air ratio and
compression ratio. Hydrogen possesses higher caloric value
compared to CNG, thus it is necessary to shi the spark timing
near to TDC for achieving the optimum performance.
Additionally, hydrogen has a higher octane number, so that it
is impressing to raise the compression ratio of the base CNG
engine to atta in enhanced fuel economy. To achieve the desired
fuel/air ratio, some modications needed in the control strate-
gies according to the HCNG’s properties. However, the NOx
emission of HCNG is higher due to higher cylinder tempera-
ture, but this can belower by leaning the charge or EGR applica-
tion to meet the desire emission restrictions. Aer all, labora-
tory test results are benecial to optimize the base CNG engine
to works on the HCNG fuel for achiev ing the best performance
and emission characteristics [5]. Hann et al. [6] studied the
inuence of binary CNG mixture on the burn rate, engine
knock, and cycle-by-cycle variations. ey proposed a new
correlation for laminar burning speed of methane based on the
reaction chemica l kinetics ca lculation. In addition, this research
shows the association of various 0D/1D models and their
mutual inuences permit to forecast engine operation limiting
factors, which upli the computer aided engine development
activity signicantly. Tangoz etal. [7] experimentally investi-
gated the consequences of compression ratio on the perfor-
mance and emission characteristics of modied diesel engine
fueled by various HCNG blends. eir outcomes reveals that
the maximum brake torque and minimum BSFC were found
at compression ratio of 12.5 for each fuel blends and when
compression ratio was 9.6, hydrogen addition increases the
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2 STUDY OF TURBULENT ENTRAINMENT QUASI-DIMENSIONAL COMBUSTION MODEL FOR HCNG ENGINES
torque output of the engine. THC emission are found generally
below than Euro-VI standard and NOx emission increased due
to hydrogen enrichment. Afshari etal. [8] studied the perfor-
mance of SI engine fueled by HCNG blends with the help of
the quasi-dimensional combustion model. Additionally, they
also presented NOx and CO emission sub models to predict the
emissions traces. It is observed that the accuracy of presented
2-zone quasi-dimensional model is very impressing in case of
pressure and thermal eciency prediction but emission sub
models are not showing any close relation.
Nevertheless, experimental investigations are great time-
consuming process and require economy. In the past 25years,
due to the rapid development in the computer science eld, it is
possible to simulate the engine’s process by computer coding to
analy ze its engine’s performance with t he help of thermodyna mic
principles and classical modeling techniques [9]. is research
article presents a simulation performed on the t urbulent entrain-
ment quasi-dimensional combust ion model of hydrogen enriched
compressed natural-gas engines under various excess air ratios,
ignition timing and hydrogen enrichment ratios.
NumericalModel
e knowledge of physics, chemistry and mat hematics provides
us a platform to perform the engine’s thermodynamic process
through computer programing. e numerica l modeling along
with the experimental study is a common practice to analyze
and design the internal-combustion engines [10]. e results
of the two-zone quasi-dimensional combustion models are
more accurate, and have the ability to report a combustion
process very precisely, due to two-zone consideration (unburned
and burned zone). Verhelst and Sierens [11] have presented a
simulation code for expansion cycle of hydrogen-fueled engines
using quasi-dimensional model combustion model. Their
combustion model consists of two dierent dierential equa-
tions, rst for the entrainment mass burning rate and second
for the ful ly burned mass burning rate. Perini etal. [12] studied
performance characteristic of the SI engine fueled by hydrogen-
methane blends with a help of quasi-dimensional combustion
model. e forecasting capabilities a nd accuracy of their model
improved by using fur ther sub-models accounting for knock ing
intensity, formation and development of ame kernel as well as
detailed turbulence explanations.
Ma etal. [13] developed a two-zone quasi-dimensional
model of SI engine fueled by various hydrogen-methane
blends under dierent operating conditions. Aer validation,
they concluded that the precision of their model is satisfying
for extremely mixture lean conditions. Later, the same author
also presented a fractal geometry based quasi-dimensional
combustion model for HCNG engines. ey developed the
enhanced fractal dimension model considering the conse-
quences of the operating condition and hydrogen enrichment
ratio [14]. Ji etal. [15] developed and validated the quasi-
dimensional combustion model for forecasting the perfor-
mance of a hydrogen added methanol engine. ey analyzed
the model prediction accuracy under various engine param-
eters like dierent hydrogen volume fractions, equivalence
ratios, engine speeds and MAPs. eir outcomes reveal when
the engine is r unning using 4% hydrogen addition to metha nol,
and the equivalence ratio was lowered from 1.2 to 0.8, the
enhancement ratio of turbulent ame speed was increased
from 32.4% to 54.6%. In addition, the enhancement ratio of
brake thermal eciency raised from 3.8% to 25.7%.
StructureofTwo-Zone
ThermodynamicModel
It is presumed that the pressure, temperature and composition
of combustible gas mi xture (fuel + air) in the cyl inder of internal
combustion engines a re homogeneous in the single zone (zero-
dimensional) model. e use of zero dimensional model is ver y
usefu l for forecasting of heat release pat tern and the combustion
cycle is called as the heat-addition process. While multi-zone
models are capable to forecast the power output, emissions,
waste energy, fuel consumptions, and performance of the
engine. is is the reason behind why multi-zone models have
a various operations in broaden the every stage of designing of
engines. In the single zone and multi-zone models, there is a
dierence of unburned and burned zone. In these two zones,
the composition of mixture and temperature are dier but the
pressure is considered to beconstant in both the zones.
ere are two main sections of the presented predictive
numerical model; rst, the thermodynamic model, which is
able to replicate the cylinder pressure by using mass fraction
burned data. Second, turbulent entrainment combustion
model, which can forecast the above-introduced mass fraction
burned data while laminar burning velocity sub-model is used
for calculation of the entrainment rate of the unburned gas
mixture. Adiabatic ame temperature sub-models are very
useful for estimating the ignition kernel temperature, which is
used for calculating the enthalpy of the products equals to the
reactants, with the consideration of the chemical equilibrium.
It is very signicant parameter for the two-zone consideration.
The presented quasi-dimensional combustion model
supposed to bethat the combustion cylinder is fragmented into
two dierent zones named as unburned and burned zones. e
rst zone; unburned zone contains the fresh combustible gas
mixture or reactant species at constant temperature and
pressure. Whi le the second zone; the burned zone that comprises
of mixture of burned gases. It is assumed that there is no heat
transfer between these two-zones and gases are considered as
ideal gas in both the zones. Additionally the pressure distribu-
tion in the combust ion chamber is consta nt and temperature in
each zone is distinct has its uniform temperature.
e two-zone thermodynamic model is based on the
following assumptions.
1. e uniform air-fuel mixture inside the combustion
cylinder is supposed to besplit into two-zone known as
burned zone and unburned zone as displayed in the
Figure 1.
2. Both the zones are considered as homogeneous ideal
gas and have same properties and unburned zone
having the mixture of hydrogen and methane while
ignoring burned species.
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3
STUDY OF TURBULENT ENTRAINMENT QUASI-DIMENSIONAL COMBUSTION MODEL FOR HCNG ENGINES
3. e Pascal’s law is applied inside cylinder, and the
in-cylinder pressure at any instant are a function of
crank angle.
4. Similarly, temperature is also split into burnt and
unburned temperature in the two-zones, which are
constant in their self-zone. Indeed, there is no heat
leakage between the zones.
5. e piston crevice eects are not considered. e
analysis is based on the closed valve period.
e above-explained assumptions are considered as the
fundamental of the two-zone model. Aer re-arranging the
Equation of State, the First Law, and the equation of mass
conservation, the equations (1), (2) and (3) rst-order dier-
ential equations can beachieved. e ‘u’ and ‘b’ subscript
represents the unburned and burnt zone.
dT
dmcV
dP
d
dQ
d
u
upu
u
u
qq
q
=+
æ
è
ç
ö
ø
÷
1
(1)
dT
dmcP
dV
dRT RT
dm
d
R
cVdP
d
dQ
d
b
upu
bb uu
b
u
pu
u
u
qq q
qq
=--
()
é
ë
ê
-+
æ
è
ç
1
öö
ø
÷+ù
û
ú
VdP
d
q
(2)
dP
dc
cVcR
Rc Vc
RV
c
RPdV
d
dQ
vu
pu
u
vb u
bpu
u
vb
b
vb
b
qq
=
-+
+
æ
è
çö
ø
÷
ì
í
î
-
11
dd uu cTR
RTdm
d
c
c
c
R
R
c
bu vb b
u
b
u
b
vu
pu
vb
b
u
p
qq
+-
()
--
æ
è
çö
ø
÷
é
ë
êù
û
ú
+-
uu
u
dQ
d
æ
è
çö
ø
֟
ý
ï
þ
ï
q
(3)
Where, the value of V, cp and cv is calculated by empirical
formulas. e heat transfer is represented by the dQ/dθ, which
is calculated by the popular Woschni’s equations [10]. Aer
calculating the dmb/dθ, above equations can besolved by the
forth-order Runge-Kutta method. However, the turbulent
entrainment combustion model is described in details in the
separate section, which is solely responsible for the calculation
of dmb/dθ. e molar specic heat at the constant pressure can
becalculated by the equation (4) and (5).
Ca
aaaa
KT K
pm,=+ +++<<
()
1
23
2
4
3
5
41200 6000
qq qq
(4)
Ca
aa aa KT K
pm,
=+ +++<<
()
67 8
2
9
3
10
4
200 1200
qq qq
(5)
Where θ means T/1000 and constants of equation a1−a10
can becalculated by Ref. [16]. e equation (6) calculates the
molar specic heat of dierent mixture.
CCX
pm pm
ii,,
,
=å
(6)
e molar fraction of chemical species ‘i’ is represented
by Xi. Finally, previous mentioned complex equation (1), (2)
and (3) can bereduced to following equations for the compres-
sion and expansion stroke:
dP
dV
mc dT
d
dQ
d
p
qqq
=-
ì
í
î
ü
ý
þ
1 (7)
dT
dmc
dQ
dVdP
d
p
qq
q
=-
ì
í
î
ü
ý
þ
1 (8)
TurbulentEntrainment
CombustionModel
e ow pattern of the combustible mixture inside engine
combustion space is mostly turbulent. e combustion process
greatly aected by the turbulent structure of the ame front.
e turbulent entrainment combustion model is one of the
unique model for forecasting the combustion behavior of
unvarying charge, SI engines. e combustion process is
correlated with the laminar burning velocity and turbulent
characteristics (intensity and scales of turbulence) in the
combustion chamber of SI engine. In the case of HCNG
engines, the above-mentioned features can change with
hydrogen enrichment ratio and engine operating conditions.
Accordingly, both the facts inuenced combustion process
and base of the turbulent entrainment combustion model.
Although the combustion modeling requires several
empirical relations.
e theory of the turbulent entrainment combustion
model begins with the assumptions proposed by Blizard and
Keck [17] and Tabaczynski etal. [18]. e large-scale turbu-
lence is supposed to comprise high dissipative regions of
vortex tubes, and their dia meter is characteriz ed by Kolmogrov
scale (η) and spacing is typied by the Taylor’s micro scale
(LT) which is actually represents the eddy radius [17]. L is
denoted by the integra l length scale and shows the overall size
of the turbulence eddy and this structure was successfully
FIGURE 1 Detailed view of two-zone combustion chamber.
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4 STUDY OF TURBULENT ENTRAINMENT QUASI-DIMENSIONAL COMBUSTION MODEL FOR HCNG ENGINES
demonstrated by Smith’s experiments years later [19]. e
turbulent str ucture is shown in the Figure 2, which is captured
by the schlieren’s high-speed photography system in the
combustion bomb experiments. e rate of propagation of
ame front in the spark region aer the ignition leg is headed
by the turbulent intensity ‘u’ and laminar burning speed ‘SL’.
e laminar burning is supposed to transpire over the Taylor
micro scale; therefore, burn-up of the combustible mixture
behind the ignition region is headed by the micro-scale and
laminar burning speed.
e entrainment rate of the unburned gas mixture is
represented by the following equation:
dm
dt
AS u
e
uf L
=+
()
¢
r
(9)
Where me is denoted by the entrainment of mass into the
ame front; ρu and Af, represents the density of gas mixture
in the unburned zone and area of the entrainment front. Aer
entrainment occurs, the unburned gas mixture is starting
burning at a rate directly proportional to the mass of the
unburned gas mixture within the entrainment front, and it
is given by the following equation:
dm
dt
mm
b
ebc
=-
()
/
t
(10)
t
cTL
LS=/
(11)
Where mb, LT, and τc are represented as mass burned,
Taylor micro scale, and characteristic time.
From the equation (9) and (10), we can get the
following equation:
t
c
bbe
dm
dt
dm
dt
dm
dt
2
2
0
++= (12)
e above mathematical expression (equation 12) oers
a burning law model, which can beattached to the thermo-
dynamic laws, turbulent eddy structure and turbulent
characteristics. Al l these are correlated to the hydrogen enrich-
ment ratio and engine operating conditions.
e following empirical relations calculate the three
turbulent scales:
Integral length scale L [18, 20]:
LCH
L0=´
(13)
LL uu
=
()
00
13
rr
/
/
(14)
Where subscript ‘0’ denoted by the ignition timing. ρu
and H represent the density of gas mixture and chamber
height respectively.
Taylor micro scale LT [17]:
LL
Tivinu
=
()
08 0
34
./
/
rr
(15)
Where the density of the gas mixture is ρin, the li of the
intake valve is Liv, and ‘u’ is turbulent intensity [18, 20, 21].
uCC
um0¢=´
(16)
¢
¢
=
()
uu uu
00
13
rr
// (17)
Where Cm denotes velocity of the piston movement.
AdiabaticFlame
TemperatureModel
ofHCNG
Because the temperature of the ignition kernel is estimated
as the adiabatic ame temperature, therefore the adiabatic
ame temperature sub-model is also a signicant parameter
for the two-zone quasi-dimensional combustion model. e
temperature produces by the complete combustion of the
gas mixture inside the combustion chamber is known as
adiabatic ame temperature when there is no work, heat
transfer and no change occur in the kinetic or potential
energy. Adiabatic ame temperature model is a sub model
of quasi-dimensional combustion main model, and it is a
very signicant parameter for two-zone consideration. In
this numerical study, adiabatic ame temperature is obtained
by the following equations:
HTpH Tp
reactantsi products ad
,,
()
=
()
(18)
1
10
+
=-
()
l
l
HcTTM
upad i/ (19)
Reordering the above two equations weget:
TT M
lcH
ad i
p
u
=++
()
10
l
(20)
Where, cp= specic heat of the reaction at constant
pressure, M =Mean molar mass, lo=Stoichiometric ratio,
λ=Excess air ratio.
e following equations are used for calcu lating the lower
heating value of the intake combustible gas mixture and
stoichiometric ratio:
Hxx
x
kJ kg
u=
+-
()
-
()
120000 50040 88
87
/ (21)
lx
x
034 48
43
87
=
-
-
. (22)
In the above equation ‘x’ is representing the hydrogen
volume fraction.
FIGURE 2 Details of the turbulence structure [1].
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5
STUDY OF TURBULENT ENTRAINMENT QUASI-DIMENSIONAL COMBUSTION MODEL FOR HCNG ENGINES
The combustion chemical equation of the hydrogen
natural gas mixture is shown by the equation (23) based on
the excess air ratio and the hydrogen volume fraction.
xH
xCH
xO N
xCOxHO
24
22
22
1
215376
12
3762
+-
()
+-
()
+
()
®
-
(
)
+-
()
+
l
l
..
.--
()
+-
()
-
()
15 1215
22
..
xN
xO
l
(23)
The mean molar mass and specific heat at constant
pressure is calculated by the following equations:
M
xx
xx
=- +-
()
+
()
-+ -
()
()
16 14 21532105 28
105476 215
l
l
..
/.
.. (24)
c
xc xc
xc
x
p
pC
Op
HO
pN
=
+
()
+-
()
+-
()
+-
()
-
(
12
376215
1215
22
2
,,
,
..
.
l
l
))
é
ë
ê
ê
ê
ù
û
ú
ú
ú
-+ -
()
é
ë
ù
û
c
xx
pO,
/.
..
2
105476 215
l
(25)
Heretofore, the equat ion (20) is used to determine the
adiabatic ame temperature of the HCNG blends with various
hydrogen enrichment ratios.
LaminarBurningVelocity
ModelofHCNG
e laminar burning velocity of the fuel is an important char-
acteristic, and also a key parameter in the two-zone quasi-
dimensional combustion model which directly aecting the
speed of the ame and gas mixture entrainment. Sarli and
Benedetto [22] have analyzed the reasonability of a Le
Chatelier’s Rule-like formula to acquire a correlation of the
laminar burning velocity of HCNG, which is valid both for
the intermediate and high hydrogen fractions at dierent
values of the equivalence ratio.
SR
xS xS
lLCH
lH lCH
_
__
//
f
ff
,
()
=
()
+-
() ()
1
1
24
(26)
Where Sl_H2(ϕ) and Sl_CH4(ϕ)are laminar burning velocity
of H2 and CH4. e following empirical relation can determine
laminar burning velocity of both the above chemical species:
SA
TY T
T
TT
TT
LFu
mub
bu
n
=
()
-
-
æ
è
çö
ø
÷
0
0
0
, (27)
TE
p
B
0=- -
æ
è
ç
ö
ø
÷
/In (28)
A
TF G
T
0
0
()
=-
æ
è
ç
ö
ø
÷
exp (29)
From the above equations:
YF,u=Fuel fraction in the unburned gas mixture
P=Pressure at which reaction occurs
Tu and Tb=Temperature of unburned and burnt zone
B, E, F, G, m, and n=Constants, which are determined
by the fuel type.
EngineandLaboratory
SetupforExperiments
e series of experiments have been conducted in the HCNG
engine resea rch and development laboratory, Tsinghua University,
Beijing, China. A heavy duty, naturally aspirated, 6-cylinder,
spark-ignition NG engine (model no. WP6NG240E5) manufac-
tured by Weifang Diesel Eng ine Company Limited, Chi na is used
for tests. e detailed technical specication of the engine is
displayed in t he Tab le 1. Customized jigs and  xtures are desig ned
for instal ling the engine on the test bench. An eddy cur rent dy na-
mometer was coupled to the engi ne to determine t he control over
the load and speed of the engine. e traces of the harmful emis-
sions are survei lled by the HORIBA-MEXA-710 0DEGR emission
recording system. e air-fuel ratio is controlled by the HORIBA
wide-range excess air ratio analyzer. To determine the engine
cylinder pressure, a piezo-electric high-pressure transducer
Kistler 6117B was used.
A Kistler 2613B crank angle encoder was installed for
recording the corresponding cranksha position with a step
of 0.1°CA. e multipurpose combustion analysis system
Kistler KiBox was used to capture pressure signals, which is
reconcilable for both the static and vehicle on road applica-
tions. An in-cylinder pressure data of 201cycles was recorded
for the result analysis. All the devices calibrated before
recording the experimental data. e schematic diagram of
the laboratory setup is shown in the Figure 3.
An online HCNG blending system was designed to
achieve the desired amount of mixture, which is precisely
explained in Figure 4. e pressure stabilizing hydrogen
enriched compressed natural gas tank was separated into two
chambers with a damping line used to upli the blend unifor-
mity [23]. e rate of ow of natural gas and hydrogen was
recorded by a Micro Motion ow meter that uses the principle
of Coriolis force for a direct measure of mass ow. An ALICAT
ow control valve was used to adjust the ow rate of the
TABLE 1 Technical specification of test engine.
Item Value
Manufacturer Weifang Diesel Engine
Company Limited, China
Engine type in-line 6 cylinders, spark
ignition
Aspiration method Turbocharged intercooled
Compression ratio 11.5:1
Bore (mm) 105
Stroke (mm) 130
Displacement Volume (in liters) 6.75
Rated power 177kW
Rated speed 2300r/min
Emission Standard Euro V
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6 STUDY OF TURBULENT ENTRAINMENT QUASI-DIMENSIONAL COMBUSTION MODEL FOR HCNG ENGINES
FIGURE 3 Test-setup in the HCNG engine R&D laboratory, Tsinghua University.
© SAE International
FIGURE 4 High accuracy on-line HCNG blends preparation system (Patented).
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7
STUDY OF TURBULENT ENTRAINMENT QUASI-DIMENSIONAL COMBUSTION MODEL FOR HCNG ENGINES
hydrogen according to the ow rate of CNG and obtain the
target hydrogen fraction. is system has been validated for
varied hydrogen fractions with the test outcomes, and the
comparison shows that the fuel mixing system performs well
under all the conditions where the absolute error in hydrogen
fraction is found always less than 1.5% [24]. e rate of ow
of mass of air is resorded by the thermal type gas mass ow-
meter Toceil20N100114LI. is system is listed in the China
Patent Database number ZL00710175797.9.
Comparisonof
Experimentaland
PredictedResults
There are three major constant factors in the turbulent
entrainment combustion model section including Turbulent
intensity coecient C2, the Taylor length scale coecient C3,
and Ignition lag coecient Cig. ese xed quantities need to
becalibrated under denite operating conditions by matching
the simulated pressure and mass fraction burned prole
results to the experimental results. e outcomes reveal that
the value of Cig always stay at 1.52. For excess air ratio 1.23,
the coecient C2, stay at 1.05, 1.05 and 1.22 for CNG, 20%
HCNG and 40% HCNG respect ively while coecient C3 varies
from 1.6-2.88, 1.8-4.84, and 0.5-1.56 respectively for the same.
When excess air ratio increased to 1.30, the coecient C2,
stay/varies at/from 1.05, 1.08-1.26, and 1.01-1.29 for CNG, 20%
HCNG and 40% HCNG respectively and value of coecient
C3 of varies/stay at/from 1.2-3.11, 1.74, and 1.74 respectively
for the same.
To analyze the accuracy of the presented turbulent
entrainment quasi-dimensional combustion model, 31 opera-
tional condition was selected for both test and predicted
results, which is displayed in the Ta ble 2. e range of oper-
ating conditions is based on the ignition timing (θig), excess
air ratio (λ) and the hydrogen enrichment ratio (x) while speed
of the engine (n) and manifold absolute pressure (MAP) was
kept constant at 1600rpm and 105kPa respectively. Basically,
this research article explores the applicability of the proposed
model under a wide range of the ignition timing in order to
compare with the experimental results.
e predicted and experimented pressure proles and
mass fraction burned proles are illustrated in the appendix
A and B respectively, which is based on the ignition timing
22° and 26°CA bTDC. All other operating conditions with
results are displayed in the appendix section. It is observed
that the pressure and MFB data given by the proposed model
are approximately matched with the experimental data.
However, there is a possibility of recalibrating of the model
in the future, especially for the constants of the ame front
area. Because of a signicant dierence in crank angle for
maximum predicted and experimental pressure values.
Aer precisely comparing, the simulated and experi-
mental data of pressure and mass fraction burned (MFB); it
is found that the MFB traces are quite impressive as compared
to the pressure proles in all the cases. Generally during cali-
bration of the model coecients it is found that the dierence
in crank angle of maximum pressure (simulation and experi-
mental results) is quite signicant, the maximum value is
found 11 degrees while minimum 3 degrees for the operating
condition of B1 and C3 as shown in the appendix A and B in
the appendix section. e acceptable values of the dierence
in crank angle of maximum pressure are found for operating
condition of B3,C1, and C2. e maximum and minimum
relative error in indicated mean eective pressure are found
to 5.91% and 0.019% at the operating condition of B6 and C2
respectively. Presented quasi-dimension combustion model
shows very impressive results for the mass fraction burned
traces, which are matches well under all operating conditions
as illustrated in the appendix B. e error in rapid combustion
duration (θrd °CA) for simulated and experimented results are
2.9° and 1.3° in case of A3 and A4 for same engine speed, MAP
and λ, but the ignition timing changes to 22° to 26°CA bTDC,
hence more time available for the completeness of combustion
at A4. Moreover, when excess air ratio increases to 1.3,
combustion becomes smoother results very close match of
MFB prole for B3 and B4 with rapid combustion duration
TABLE 2 Operational conditions at 1600r/min and MAP
105kPa.
S. No. OC’s θig (°bTDC) xλ
1 A1 14
0% 1.23
2 A2 18
3 A3 22
4 A4 26
5 A5 30
6 B1 14
0% 1.30
7 B2 18
8 B3 22
9 B4 26
10 B5 30
11 B6 34
12 C1 14
20% 1.23
13 C2 18
14 C3 22
15 C4 26
16 C5 30
17 D1 14
20% 1.30
18 D2 18
19 D3 22
20 D4 26
21 D5 30
22 E1 14
40% 1.23
23 E2 18
24 E3 22
25 E4 26
26 E5 30
27 F1 14
40% 1.30
28 F2 18
29 F3 22
30 F4 26
31 F5 30
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8 STUDY OF TURBULENT ENTRAINMENT QUASI-DIMENSIONAL COMBUSTION MODEL FOR HCNG ENGINES
(θrd °CA) error of 0.7° and 0.6° respectively. e maximum
and minimum error in θrd 3.4° and 0.4°CA at operating condi-
tion of C1 and D1 respectively shown in the table inserted in
the appendix section.
Additionally, to validate the reasonability of the proposed
model, some other parameters of the combustion were
observed, which includes maximum pressure (Pmax), the crank
angle of maximum pressure (θPmax), rapid combustion
duration (θrd) (10%~90% MFB), the crank angle of 50% MFB
(θ50%), indicated mean eective pressure (Pi) and relative error
of IMEP. e distinctions between the predicted and experi-
mental maximum pressure (Pmax) values are quite close but
their corresponding crank angle (θPmax), values are not
matched enough as it can beclearly seen in the Appendix
section. Although the rapid combustion duration (θrd)
(10%~90% MFB), the crank angle of 50% MFB (θ50%) error is
less than 5°CA and the relative error of IMEP is less than 6%.
There are two main factors inf luencing the precision of
computing: rstly, few assumptions are not compatible with
the test operations, for example, no heat transfer occurs
between burned and unburned zones, and fully chemical
reaction; secondly, the few errors in the empirical relations
are not able to calculate the specic heat, adiabatic tempera-
ture of the ame and laminar burning speed. e above-
mentioned restrictions are very important when considering
the precision.
Conclusion
e turbulent entrainment quasi-dimensional combustion
model of HCNG engine has been presented, which is accept-
able for various hydrogen enrichment ratios, excess air ratios
and ignition timings. e two-zone model is the main model
while the turbulent entrainment model is used as supporting
model to simulate the combustion process to generate the
MFB proles under various operating conditions. While other
sub models are helping the main model to generate the predic-
tive pressure prole.
Validating the acceptability of the proposed turbulent
entrainment quasi-dimensional combustion model, 31 opera-
tional condition have been chose for comparing the predicted
and test results. e key parameters, which are analyzed in
this presented model, are Pmax, θPmax, θrd (10%~90% MFB),
θ50%, Pi and relative error of IMEP. Aer analyzing the results
of the proposed model, it can beconcluded that the accuracy
of the model is convincing by the ample extent. Consequently,
the oered model in this research article can simulate the
performance of the SI HCNG engines numerically under the
various ignition timing, excess air ratio and hydrogen enrich-
ment ratio. However, this model gives impressive results of
simulated MFB proles for the values of coecients C2, C3
and Cig that are already explained in the rst paragraph of
comparison of experimental and predicted results section.
e predictive accuracy of the presented two-zone quasi-
dimensional model of HCNG engine is quite impressive in
terms of MFB matched proles while pressure traces are also
acceptable for simulated and experimental results. Moreover,
the high values of dierence in the crank angle of maximum
pressure (simulated and experimented) is point of concern in
the future and there is a possibility of more explorations in
this area.
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12. Perini, F., Paltrinieri, F., and Mattarelli, E., “A Quasi-
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Contact Information
Fanhua Ma and Roopesh K. Mehra
HCNG Engine R&D Laboratory
State Key Laboratory of Automobile Safety and Energy,
Tsinghua University
ma@tisnghua.edu.cn (Fanhua Ma)
luops15@mails.tsinghua.edu.cn (Roope sh KM.)
roopeshmehra.ind@gmail.com (Roopesh K M.)
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10 STUDY OF TURBULENT ENTRAINMENT QUASI-DIMENSIONAL COMBUSTION MODEL FOR HCNG ENGINES
Appendix
A: Comparison between Experimental and Predicted Pressure under 12 Operating Conditions.
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11
STUDY OF TURBULENT ENTRAINMENT QUASI-DIMENSIONAL COMBUSTION MODEL FOR HCNG ENGINES
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12 STUDY OF TURBULENT ENTRAINMENT QUASI-DIMENSIONAL COMBUSTION MODEL FOR HCNG ENGINES
B: Comparison between Experimental and Predicted Mass Fraction Burned Profiles under 12 Operating Conditions
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13
STUDY OF TURBULENT ENTRAINMENT QUASI-DIMENSIONAL COMBUSTION MODEL FOR HCNG ENGINES
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14 STUDY OF TURBULENT ENTRAINMENT QUASI-DIMENSIONAL COMBUSTION MODEL FOR HCNG ENGINES
C: Dierences between Predicted and Experimental Results
S. No. Operational Condition
Pmax θPmax Dierence of θPmax θrd Error θrd θ50% Pi
Relative Error Pi
(MPa) (°CA) (°CA) (°CA) (°CA) (°CA) (MPa)
1 A1 (Exp) 4.62094 373 8 34.4 1 376.3 1.18 5.45%
2 A1 (Pre) 4.61467 381 35.4 378.2 1.25
3 A2 (Exp) 5.30331 373 832.6 2.2 372 .1 1.26 3.99%
4 A2 (Pre) 5.30311 381 34.8 373.8 1.31
5 A3 (Exp) 5.87224 374 5 34.5 2.9 367 1.30 4.71%
6 A3 (Pre) 5.87624 379 37. 4 368.7 1.36
7 A4 (Exp) 5.93927 373 639.4 1.3 366.3 1.31 5.24%
8 A4 (Pre) 5.93586 379 40.7 369.2 1.38
9 A5 (Exp) 6.51764 369 742.2 1.9 362.4 1.42 1.54%
10 A5 (Pre) 6.52024 376 44.1 363.5 1.44
11 B1 (Exp) 3.98576 378 11 33.4 0.4 382 1.14 0.065%
12 B1 (Pre) 3.97812 389 33.8 383.1 1.21
13 B2 (Exp) 5.18505 376 630.3 0.9 372.4 1.22 0.064%
14 B2 (Pre) 5.18332 382 31.2 373.6 1.30
15 B3 (Exp) 5.68205 375 4 33.5 0.7 367.5 1.26 0.063%
16 B3 (Pre) 5.68851 379 34.2 368.8 1.34
17 B4 (Exp) 5.96875 372 537.7 0.6 366.1 1.29 0.056%
18 B4 (Pre) 5.96748 377 38.3 3 67.3 1.36
19 B5 (Exp) 6.93201 369 5 38.4 1.2 367.6 1.37 0.058%
20 B5 (Pre) 6.93537 374 39.6 368.4 1.45
21 B6 (Exp) 7.15349 366 7 44.5 0.4 359.2 1.44 0.019%
22 B6 (Pre) 7.1 797 8 373 44.9 359.7 1.47
23 C1 (Exp) 5.14046 378 430.4 3.4 375.1 1.22 5.41%
24 C1 (Pre) 5.13961 382 33.8 379.1 1.28
25 C2 (Exp) 5.45312 376 437. 2 3 367.5 1.23 5.91%
26 C2 (Pre) 5.45219 380 40.2 369.5 1.31
27 C3 (Exp) 6.51538 373 339.4 1.2 364.7 1.36 2.27%
28 C3 (Pre) 6.51460 376 40.6 365.7 1.39
29 C4 (Exp) 7.20708 369 6 48.3 2.2 361.7 1.39 4.48%
30 C4 (Pre) 7.20571 373 51.1 363.8 1.45
31 C5 (Exp) 7.33705 367 6 48.9 1 35 7.4 1.42 3.26%
32 C5 (Pre) 7.33418 373 49.9 358.9 1.47
33 D1 (Exp) 4.70964 380 5 33.8 0.3 3 74.4 1.18 0.064%
34 D1 (Pre) 4.70866 385 34.1 374.9 1.19
35 D2 (Exp) 5.13254 376 6 34.6 1.1 370.7 1.22 1.92%
36 D2 (Pre) 5.13316 382 35.7 371.6 1.25
37 D3 (Exp) 6.04937 371 7 36.2 0.6 368.4 1.29 0.061%
38 D3 (Pre) 6.04900 378 36.8 369.1 1.29
39 D4 (Exp) 5.96875 372 639.1 1.2 363.7 1.29 1.42%
40 D4 (Pre) 5.97975 378 40.3 365.4 1.30
41 D5 (Exp) 7. 0745 5 368 7 42.7 2.9 361.9 1.42 5.86%
42 D5 (Pre) 7.0 72 0 2 375 45.6 364.5 1.34
43 E1 (Exp) 4.47485 380 9 35.4 1.4 373.4 1.15 2.27%
44 E1 (Pre) 4.47349 389 36.8 375.5 1.18
45 E2 (Exp) 6.16324 371 842.2 1.7 366.1 1.32 4.71%
46 E2 (Pre) 6.15160 379 43.9 369.2 1.26
47 E3 (Exp) 6.58384 372 54 7.3 3 362.1 1.34 4.42%
48 E3 (Pre) 6.59578 377 50.3 364.2 1.29
49 E4 (Exp) 6.99809 369 651.3 1 358.8 1.38 2.69%
50 E4 (Pre) 6.99824 375 52.3 359.4 1.42
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15
STUDY OF TURBULENT ENTRAINMENT QUASI-DIMENSIONAL COMBUSTION MODEL FOR HCNG ENGINES
S. No. Operational Condition
Pmax θPmax Dierence of θPmax θrd Error θrd θ50% Pi
Relative Error Pi
(MPa) (°CA) (°CA) (°CA) (°CA) (°CA) (MPa)
51 E5 (Exp) 7.52694 365 952.4 0.7 356.6 1.44 1.97%
52 E5 (Pre) 7.52110 374 53.1 3 57. 3 1.47
53 F1 (Exp) 4.49848 378 7 32.8 1 .1 370.3 1.17 3.22%
54 F1 (Pre) 4.51090 385 33.9 370.6 1.20
55 F2 (Exp) 5.47236 374 739.3 1.1 369 1.22 1.83%
56 F2 (Pre) 5.47231 381 40.4 369.7 1.24
57 F3 (Exp) 5.78813 371 8 44.3 2.1 363.2 1.25 2.19%
58 F3 (Pre) 5.78843 379 46.4 365.7 1.27
59 F4 (Exp) 6.65806 370 651 2.2 358 1.36 4.83%
60 F4 (Pre) 6.65692 376 53.2 359.8 1.29
61 F5 (Exp) 6.97362 369 6 53.5 1.9 356 1.39 4.04%
62 F5 (Pre) 7.00084 375 55.4 358.4 1.34
(Continued)
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