Post on 27-Jun-2020
Proceedings of 2nd International Conference on Mechanical and Aeronautical Engineering
Held on 13th – 14th July 2016, in Bangkok, ISBN: 9788193137352
40
AN EXPERIMENTAL INVESTIGATION ON THE USE
OF NATURAL GAS-DIESEL FUEL MIXTURE ON
PERFORMANCE AND EMISSIONS OF A CI ENGINE
Y. Karagoz,
Automotive Division, Department of
Mechanical Engineering, Yildiz Technical
University, Yildiz, Besiktas, Istanbul, 34349,
Turkey
T. Sandalci,
Automotive Division, Department of
Mechanical Engineering, Yildiz Technical
University, Yildiz, Besiktas, Istanbul, 34349,
Turkey
O. Isin,
Automotive Division, Department of
Mechanical Engineering, Yildiz Technical
University, Yildiz, Besiktas, Istanbul, 34349,
Turkey
A.S. Dalkilic
Heat and Thermodynamics Division,
Department of Mechanical Engineering, Yildiz
Technical University, Yildiz, Besiktas, Istanbul,
34349, Turkey
Abstract- Number of diesel engines increases
nowadays. Especially, diesel vehicle usage rates tend
to increase in European Countries in recent years.
However, diesel engines are disadvantageous due to
their high NOx and soot emissions. Stringent
emission regulations force engine manufacturers to
design high technology (electronical fuel injection
systems and after treatment systems) engines. Diesel
engines are also used on heavy duty commercial
vehicles and agricultural vehicles. In this study, a
compression ignition engine which has mechanical
fuel system is converted into common-rail fuel
system by means of a self-developed ECU. Then, the
engine is modified to be operated with diesel and
natural gas fuels in dual-fuel mode. For this aim,
diesel fuel was injected into the cylinder, and
natural gas was injected into intake manifold. Both
of gas and diesel injectors were controlled with the
ECU. Energy content of sprayed gas fuel is modified
by the way that organises 0% (only diesel fuel),
15%, 40% and 75% of total fuel’s energy content.
All tests are made at 1500 rpm stable engine speed
and at full-load condition. Both NOx and soot
emissions are taken under control with 15% and
40% energy content rates in gas-fuel mixture
compared to only diesel condition, respectively.
However an increase is observed in CO emissions
with 15% natural gas fuel addition compared to
only diesel condition. Moreover, although smoke
emission is reduced with gas fuel addition, the
dramatic increase in NOx emissions could not be
prevented with 75% natural gas fuel addition. In
conclusion, an explosive type combustion
characteristic, similar with gasoline combustion, is
observed for rate of heat release.
Keywords- Diesel engine; Natural gas; Emissions;
NOx; Smoke; Engine performance.
I. Introduction
Fast extinction of fossil fuels and increasing oil prices
imposes engine manufacturers to work on a new
alternative fuel [1]. Due to the damage caused by
petroleum derived fuels to the environment and to human health; a study on alternative, reliable and
environmental fuels has become an inevitable
requirement [2]. Nowadays, an important part of the
energy requirements in the transport sector are still
covered by fossil fuels. In spite of Kyoto Protocol, CO2
emissions increased by 27 % between 1990-2004
years, while the amount of emitted CO2 emissions
stemming from transport increased by 37 % [3].
More rigorous emission regulations are requesting
nearly “0” NOx emissions [4]. Despite high efficiency
advantage of diesel engines, they have a disadvantage in terms of NOx and soot particles emissions [5]. NOx
causes photochemical formation of soot and acid rain
formation. The soot particle emission increases heart
and vascular related death rates, affects lung
development in children, it leads to a number of other
Proceedings of 2nd International Conference on Mechanical and Aeronautical Engineering
Held on 13th – 14th July 2016, in Bangkok, ISBN: 9788193137352
41
health problems [6]. Diesel engines use SCR and DPF,
respectively in order to reduce NOx and soot particle
emissions. However, owing to the high cost of catalyst
materials, orientation through alternative gaseous fuels
has become essential [7]. In parallel with these
developments, automotive sector has started to work on
the improvement of performance, exhaust emissions
and combustion characteristics with the use of
alternative fuels in engines currently in use. [1, 8].
Among alternative fuels, alcohols, vegetable oil, LPG,
CNG, LNG, air gas, biogas and hydrogen has come to
the fore. Alternative fuels should be examined in terms of
criteria such as potential, source, fuel supply, safety,
toxicity, health hazard, engine performance, emissions,
storage and easy availability. Natural gas, providing
many of these criteria, is an important alternative fuel
that may be used as a substitute fuel in internal
combustion engines. Easy availability, having more
reserves than oil, lower costs, cleaner combustion
characteristics, lower vehicle emissions in addition to
the existence of distribution systems, make natural gas
an extremely convenient alternative fuel. 90-96% of natural gas is composed of CH4 (methane) gas. The rest
is composed of 2.411% C2H6 (ethane), 0.736% C3H6
(propane), 0.371% C4H10 (butane), 0.776% N2
(nitrogen), 0.164% C5H12 (pentane) and 0.085% CO2
(carbon dioxide) [9]. Moreover, composition of natural
gas varies depending on the country. A way to capture
the high emission standard in diesel engines is to
change diesel fuel with a cleaner, low carbon
alternative fuel [6]. Furthermore, although CNG is a
fossil fuel, a reduction of up to 25% can be observed in
greenhouse gases due to methane’s low C: H ratio,
[10]. Methane has low flame speed and narrow flammability
limits [7]. Reduction in CO2 emissions will be seen
compared to the same energy value due to its lower
carbon content compared to gasoline and diesel fuel
[11, 12]. Because of 540oC self -ignition temperature
and narrow flammability limits (5% to 15%) of CNG (a
clean fuel consisting largely of methane gas), ignition
is provided by pilot diesel injection in diesel engines
[4].
Biogas, produced by anaerobic fermentation of organic
matter, is an alternative energy source. It consists of methane mainly, and contains a high proportion of
CO2, and therefore, it is a fuel having low thermal
energy. Due to its high octane number, biogas can be
used to achieve higher efficiency in high compression
ratio engines. Therefore, biogas can be used in dual-
mode diesel engines easily. However, long ignition
delay, lean mixture at low loads, low flame speed and
low thermal efficiency problems are encountered due
to high CO2 content [13]. Due to the presence of CO2
and other gases, it has a lower energy content
compared to natural gas. So, the flame speed falls and
ignition delay period increases. Furthermore, CO
amount increases in studies with biogas compared to
natural gas and NOx amount decreases [14] finally.
The application of natural gas in vehicles is difficult
due to their requirement for a high pressure tank and
low range, but the use for buses is more suitable. The
use of gas fuel in city traffic is advantageous in terms
of low emissions and soot particles. In addition, through the use of natural gas in SI engines, if the
compression ratio is considered to be between 10 and
12, the inner pressure is around 30 bars, cylinder
pressure value reaches 90 bar, and spark plug gap
tension leads to nearly threefold at full load in a
supercharged engine. So as to overcome this problem,
it is more appropriate to ensure ignition by a small
amount of diesel pilot injection [12]. Moreover,
efficiency reduces due to the usage of natural gas at
partial load in SI engines, throttle existence in Otto
engines and low compression ratios [12]. Combustion of methane and diesel at dual mode takes
place in three modes: Firstly, pilot diesel combustion,
secondly, combustion of methane around the diesel
spray, and thirdly, the progress of the flame in methane
- air mixture. Extremely poor mixture, flame extinction
zone increase, and incomplete combustion for all three
sections at low loads can result with HC and CO
emission increase at partial loads [10].
HCCI (homogeneous charge compression ignition)
engines could be the solution in order to reduce both
NOx, as well as soot emissions, because they have very
low NOx emission and particulate emission which is almost 0. They carry superior features of diesel and
Otto engines. Since combustion having premixed and
lean mixture self-ignite with high compression rate
occurs, it has many advantages such as low emission,
high thermal efficiency and high output rate of heat
sources. However, the biggest disadvantage of them is
difficulty in controlling starting point of ignition in a
wide range of speed and load [4]. To overcome this
challenge of HCCI, the combustion phase has been
controlled with VVT and EGR. However, one of the
most effective methods in controlling combustion is to spray a small amount of pilot diesel [10].
Studies having the use of natural gas fuel in ICEs are
summarized in next paragraph as shown below.
Cordiner et al. [15] introduced diesel and natural gas
mixture (natural gas charged homogeneously) into
compression ignition engine and small amount of
diesel is injected near TDC. A heavy duty diesel engine
Proceedings of 2nd International Conference on Mechanical and Aeronautical Engineering
Held on 13th – 14th July 2016, in Bangkok, ISBN: 9788193137352
42
is converted to dual fuel operation in that study. By
taking into account of obtained experimental results,
they made a 3-D simulation with modified version of
Kiva 3V. Zhang et al. [16] tested influence of dissolved
methane in diesel fuel on engine performance and
emissions in a single cylinder, direct injection diesel
engine. As methane concentration increases, max heat
release rate decreases, and ignition delay increases.
Diesel fuel containing dissolved methane released less
NOx, and released smoke depends on methane
concentration. Mostafa et al. [17] observed the effect of
diesel injection timing, boost pressure and diesel fuel injection pressure on diesel-methane dual fuel in a
single cylinder research engine. They tested diesel-
methane dual fuel at 1500 rpm constant engine speed,
80% engine load (5.1 bar imep), between 250 - 350oCA
injection advance, between 200 - 1300 bar diesel
injection pressure and at 1.1 bar - 1.8 bar boost
pressure. Zhang et al. [18] studied on a six cylinder,
diesel piloted direct injection natural gas engine at idle
condition for 2 different injection intervals (0.7ms and
1ms) at 4 - 13oBTDC natural gas injection advance and
for 3 different pressures (15, 18 and 24 MPa). On the other hand, studies having the use of biogas and
LPG fuel in ICEs are summarized in next paragraph as
shown below.
Tira et al. [19] injected 60% CH4 and 40% CO2 by
volume into intake manifold. According to obtained
results, NOx, PM and smoke decreased, but
combustion stability has been ruined, CO and THC
emissions increased. That’s why synthetic gas is added
to liquid so as to increase combustion stability, and
consequently emissions improved. Bora and Saha [20]
studied on a single cylinder, four stroke, direct
injection, naturally aspirated diesel engine at full engine load and 16 different combinations (23, 26, 29
and 32oBTDC; 18, 17.5, 17 and 16 CR) with biogas-
diesel dual fuel. Best results are obtained
thermodynamically with 29oBTDC and 18 CR.
Saravanan and Nagarajan [21] determined optimum
injection advance, injection time, and hydrogen flow
rate by both manifold gas injection and port gas
methods. According to obtained results, brake thermal
efficiency increased by 21% at 75% load, NOx
decreased by 4%, smoke decreased by 45% and CO
decreased by 50%. Ma et al. [22] investigated effect of diesel and diesel-propane blends on fuel injection
timing in a single cylinder compression ignition
engine. Propane rate, maximum heat release rate,
premixed heat release, maximum cylinder gas
temperature and NOx emissions increased for same
engine speed, engine load and injection advance while
total combustion time, CO, HC and smoke emissions
reduced.
A single cylinder, naturally aspirated, mechanical
diesel fuel, prototype, compression ignition engine is
developed for this study in Yildiz Technical University
in partnership with the companies of Şahin Metal and
Erin Motor. It is converted into common rail diesel fuel
system as a first work. By the help of self-developed
hybrid-ECU, not only it can operate with only diesel
fuel when needed, but also it can operate with diesel-
natural gas fuel if necessary. Natural gas mixture is
sent from intake manifold in this engine, and ignition is supplied with diesel injection. In this way, effect of
natural gas on engine performance, emissions and
combustion characteristics is studied in detail at
constant engine speed and wide range (0% - 75%
natural gas on energy basis).
II. Experimental setup
Schematic diagram of the experimental setup is shown
in Fig. 1. Conversion of a single cylinder, mechanic
diesel fuelled, 4 stroke, and naturally aspirated engine
into a common-rail fuel system is performed as shown in this figure. An electromagnetic Bosch diesel injector
is installed concentric to cylinder axis, because
previous mechanic diesel injector was concentric to
cylinder axis. Pulverization angle of diesel injector is
the same as mechanical injector (145o) and reaches
mechanical injector with regards to pulverization flow
rate. By utilizing a Denso fuel pump and a fuel rail (for
common rail), diesel fuel is stabilized at 1000 bar
pressure in rail. Diesel fuel meets EN 590 standards.
Natural gas fuel, which is composed of above 90%
methane and other gases such as pentane, propane,
butane, nitrogen etc., was purchased from market at compressed form (200 bar) in steel tanks for this study.
Self-developed ECU controls fuel metering control
valve of high pressure fuel pump and fuel rail pressure
control valve and solenoid injector. After required
modifications have been done, a Keihin CNG injector
was installed on intake manifold. Self-developed
hybrid ECU controls both diesel and gas injectors. A
miniature oval gear type flow meter(Biotech, VZS-
005-Alu model) measures diesel fuel consumption
while a New-flow brand named, TMF series hotwire
type mass flow meter measures natural gas consumption. Cooling water inlet and outlet
temperatures and also exhaust temperature are
measured by means of K type thermocouples. A Kistler
6052C pressure-sensor is used so as to measure
cylinder gas pressure. Crank angle was specified via an
incremental type encoder. A Kistler charge amplifier
and a Lecroy digital oscilloscope were other equipment
Proceedings of 2nd International Conference on Mechanical and Aeronautical Engineering
Held on 13th – 14th July 2016, in Bangkok, ISBN: 9788193137352
43
utilized during tests. The load-cell of eddy-current
dynamometer, turbine type diesel flow-meter,
hydrogen + methane flow meter, AVL Dicom 4000 and
AVL 415S were linked to USB type NI6215 data
acquisition card. By means of LabView software, a
program was developed, and NI card was attributed to
a personal computer.
Fig. 1 Schematic diagram of the experimental setup
A.1 Test Engine and dynomometer
Properties of test engine can be listed as follows:
single cylinder, 1.16 L, naturally aspirated, 4 stroke,
compression ignition engine. Test engine is developed
by Yildiz Technical University in partnership with
Şahin Metal and Erin Motor Company. API brand,
Eddy current type, water cooled, 40 kW dynamometer
was used to load test engine, and varying magnetic
field engine brake torque was controlled. During tests,
engine speed (rpm) of dynamometer was inspected
with convenient PID coefficients experimentally. Table 1 depicts specifications of the test engine and engine
dynamometer. We performed all tests in the laboratory
of Yildiz Technical University. Test-cell was adapted
for hydrogen and methane gas mixture usage.
Table 1 Technical specifications of the test dyno and
test engine
Engine manufacturer Erin-motor engine
Aspiration Natural
Number of cylinders 1
Bore × stroke [mm] 108 × 127
Cylinder volume [cm3] 1163
Compression ratio 14.7
Speed range min-max [rpm] 800 - 2700
Number of intake & exhaust valves
2 & 2
Cooling Water cooled
Dyno type & power [kW] Eddy current & 40
A.2 Gas Fuel Line
Fig. 2 illustrates schematic diagram of gas fuel system.
A high pressure type steel tank was used to store
natural gas at 200 bar gas pressure. A double stage
type, stainless steel gas pressure regulator was used to
decrease the gas fuel mixture pressure. Steel tank was
located outside the laboratory. In order to avoid
probable backfiring, a solenoid gas valve (that can also
be operated as shut-off valve) was installed on gas fuel
system. A quick-connect type equipment acting as a
check valve is set before the gas fuel has been injected into intake manifold. A relief-valve was placed on gas
fuel system to be able to discharge gas fuel out of
laboratory if overpressure is seen. Line pressure is
regulated by means of a line type pressure-regulator. A
rotameter and a thermal mass flowmeter were
calibrated according to natural gas mixture. Hydrogen
is injected into intake port by a Keihin CNG injector.
All equipment (such as fittings, valves and gas tube) of
hydrogen fuel system is 316 stainless steel. The line is
resistant to 350 bar gas pressure.
Fig. 2 Schematic diagram of the gas fuel system
A.3 The self-developed hybrid ECU Fig. 3 depicts schematic diagram of self-developed
hybrid electronic control unit. Self- developed hybrid
ECU was used to control both diesel and gas injectors.
Energy of diesel and gas injectors was supplied by a 12
Volt, DC type power-supply. The self-developed ECU
operates with a Pic microcontroller. Signals generated
by incremental encoder help ECU to control both gas
and diesel injectors. Two signal outputs of encoder
control both injectors: first one generates 360 pulses
per revolution and second one generates a single pulse
per revolution (called as zero signal). The zero signal determines reference position of piston. The injection
advance of gas injector was set at TDC (top dead
center) at the outset of intake stroke. The injection
advance of diesel injector was set to 28oBTDC (before
top dead center) during compression stroke. Two
Ardunio-Due microcontroller boards were used to
change injection duration of both injectors: first one
Proceedings of 2nd International Conference on Mechanical and Aeronautical Engineering
Held on 13th – 14th July 2016, in Bangkok, ISBN: 9788193137352
44
changes the injection duration of diesel injector, the
second one changes the injection duration of gas
injector according to obtained signals from encoder.
Fig. 3 Schematic diagram of the self-developed
hybrid ECU
A.4 Tail-pipe emission measurement
CO, THC and NOx emissions were measured by AVL
Dicom 4000 exhaust gas analyser, and the smoke
emission was measured by AVL 415S smoke analyser.
AVL Dicom 4000 measures CO and THC emissions
as %volume, NOx emissions as ppm. AVL 415S measures as FSN or mg/m3. Brake engine torque and
brake engine power were fixed at 75.7 Nm engine
torque and 1500 rpm engine speed, for that reason,
emission units were not converted into brake specific
emissions.
III. Modelling of Dual Fuel Engine
A single- cylinder, 4 -stroke, naturally aspirated, direct
injection engine is modelled via Boost program at dual
fuel mode. Since the diesel engine will be used in dual
mode, an injector is positioned in the inlet manifold in order to send natural gas fuel. 1-D engine model of
engine in Boost program is depicted in Fig. 4. On the
other hand, taking advantage of using Mat lab, model
of diesel injector fuel spray pattern has been created
and entered into Boost Program. As seen in Fig. 5,
obtained rate of injection model has been given in AVL
MCC injector model. Since results are gathered via a 1-
D program and high rate of natural gas is used, results
are found in an acceptable agreement. Moreover, at all
test points, difference between values like power, imep,
emission, bsfc does not exceed ±10% deviation.
Fig. 4 1-D engine model of gas fuel mixture+diesel
fuelled diesel engine with Boost software
Fig. 5 AVL MCC injector model (normalized rate of
injection) of diesel injector
IV. Data reduction
Data obtained during experiments are used to calculate
the parameters with equations below. A single-zone
and zero dimensional rate of heat release model was used in this study. The combustion process was
analyzed via heat release equation as described by
Krieger et al. [23]:
(1)
where is rate of heat release (J/oCA), is the ratio of specific heat (unitless), cp/cv can be selected from
Janaf tables or 1.35 value can be used for diesel heat
release analysis (in this study, Janaf tables are used); is the crank angle (degree); P is cylinder gas pressure
value (Pa) and V is cylinder volume (m3).
Value on load-cell is read and brake engine torque is
calculated by using this value. By measuring moment
arm length, load-cell is attached to as follows [24]:
(2)
where T is the brake engine torque (N), F is force (Newton) and b is moment arm length (meter).
Test engine was loaded by eddy-current type
dynamometer. Brake engine power was calculated by
using brake torque and angular speed [25]:
(3)
where Pb is brake engine power (kW), ω is the angular
speed of the engine (rps) and T is the brake engine
torque of engine (Nm).
Engine brake thermal efficiency value is calculated by
means of engine brake power, fuel consumption per
unit time and LHV values [26]:
(4)
Proceedings of 2nd International Conference on Mechanical and Aeronautical Engineering
Held on 13th – 14th July 2016, in Bangkok, ISBN: 9788193137352
45
where ηBT is the engine brake thermal efficiency value
(percentage), Pb is brake engine power value (kW), d
mass flow rate of consumed diesel fuel per unit time
(kg/s), NG is mass flow rate of consumed natural gas
fuel per unit time (kg/s), LHVd is the lower heating
value of diesel fuel (kJ/kg), LHVNG is the lower
heating value of natural fuel (kJ/kg).
Total brake specific fuel consumption of the engine
was calculated by means of equivalent diesel amount
of consumed hydrogen and methane according to lower heating values of natural gas and diesel. Diesel mass
flow rate is added to equivalent diesel fuel mass flow
rate of natural gas. Total bsfc was calculated based on
equivalent diesel amount as follows [27]:
(5)
where bsfc is the total brake specific fuel consumption
(g/kWh), d is mass flow rate of consumed diesel fuel
(g/h), NG is the equivalent diesel mass flow rate of
natural gas (g/h) and Pb is the brake engine power value (kW).
Total uncertainty analysis values and measurement
accuracies of equipment are calculated by using Kline
and McClintock method [28]:
[(
)
(
)
(
)
]
⁄
(6)
where WR is total uncertainty value, R is given
function, x1, x2,…,xn are independent variables and w1,
w2,…wn symbolize uncertainty value of each
independent variable.
V. Experimental procedure
All engine tests were carried out at the ESC (European stationary cycle) at 1500 rpm, which corresponds to C
engine speed for the compression ignition engine, at
different natural gas energy fractions. Obtained results
at 0%, 15%, 40% and 75% natural gas energy content
and 100% engine load conditions are compared with
each other both experimentally and theoretically. First
of all, the compression ignition engine was operated
with pure diesel fuel until it reaches to steady-state
operating conditions, and then different energy levels
of hydrogen and methane gas mixture fuel were tested.
Natural gas is sent as 0%, 15%, 40% and 75% of total
fuel energy. The diesel fuel was injected into cylinder and the injection advance of diesel fuel was kept at
28oBTDC during compression stroke. The gas fuel
(hydrogen and methane) was injected into intake
manifold at TDC during intake stroke. Backfiring, pre-
ignition or knock problems were not seen during
engine tests. Measurement accuracies and calculated
uncertainities are shown in Table 2.
Table 2 Measurement accuracies and calculated uncertainities
Parameter Device Accuracy
Engine torque Load cell ±0.05 Nm Engine speed Incremental
encoder ±5 rpm
Cylinder pressure
Kistler 6253C ±0.5 %
Diesel flow rate
Biotech VZS-005
±1 % (of reading)
NG flow rate New-Flow TMF
±1 % (F.S.)
CO AVL DiCom 4000
0.01 % Vol.
THC AVL DiCom 4000
1 ppm
NOx AVL DiCom 4000
1 ppm
Smoke AVL 415S 0.4 % Vol. Calculated results
Uncertainty value
Engine power ±0.28 %
Bsfc ±1.10 % (0% Natural gas) ±1.57 % (15% Natural gas) ±1.48 % (40% Natural gas) ±1.39 % (75% Natural gas)
VI. Results and discussion
Emissions, performance and combustion characteristics
of a hydrogen and methane enriched diesel engine at
1500 rpm engine speed and 100% engine load are
investigaed in this study. The test-bench and the CI
engine were adapted to operate with hydrogen and
methane fuel mixture. Sandalcı and Karagöz [29]
studied the effect of 0%, 16%, 36% and 46% hydrogen
energy enrichment on performance and emission
characteristics of a CFR compression ignition engine at
full engine load and 1300 rpm engine speed (equal to C engine speed of European Stationary Cycle). In another
study, Karagöz et al. [30] studied the effect of 30%
hydrogen enrichment on energy basis at 1100 rpm
constant engine speed (equal to “A” engine speed of
European Stationary Cycle) at different engine load
conditions (40%, 60%, 75% and 100%) on a CFR
Proceedings of 2nd International Conference on Mechanical and Aeronautical Engineering
Held on 13th – 14th July 2016, in Bangkok, ISBN: 9788193137352
46
compression ignition engine. Karagöz et al. [31, 32]
also investigated the effect of hydrogen addition on
spark ignition engines experimentally.
In this study, the effect of 0%, 15%, 40% and 75%
natural gas addition on energy basis is investigated at
1500 rpm constant engine speed (equal to C engine
speed of ESC) experimentally. Effect of natural gas
delivery into intake manifold by gas injectors on
emissions, performance and combustion characteristics
is surveyed in a prototype diesel engine. After required
modifications had been done, compression ignition
type engine was manufactured as a prototype with a mechanically controlled fuel system which is converted
into a common-rail fuel system. By means of self-
developed hybrid ECU which controls diesel and gas
injectors, energy content of gas and diesel fuel is
adjusted, and energy content increased to 75%. A small
amount of hydrogen is blended with methane (30% of
gas fuel in volume basis, energy value extremely low)
and is injected via pilot diesel pulverization in diesel
engine, consequently a considerable improvement is
achieved in emissions. Due to hybrid controlled ECU
technology, gas and diesel injectors are controlled. With simple modifications, actual mechanical diesel
engines are being converted. During the experiments,
natural gas energy content of the fuel mixture did not
exceed 75%, due to the fact that, after this value the
engine suffers from combustion stability.
Brake thermal efficiency can be described as the
percentage of brake power and fuel energy consumed
by the engine. It demonstrates how input energy is
converted into useful output energy efficiently [33].
The variation of the brake thermal efficiency according
to several natural gas energy contents is shown in Fig.
6. Although brake thermal efficiency value reduced with natural gas and diesel dual fuel compared to only
diesel fuel, brake thermal efficiency decreased with
increasing methane energy content. Brake thermal
efficiency value is 24.6%, 17.5%, 19.1% and 22.7%,
respectively, when natural gas is 0%, 15%, 40% and
75%. Reduction in brake thermal efficiency value is
8%, 22% and 29% compared to neat diesel fuel with
15%, 40% and 75% methane addition on energy basis.
The natural gas/air mixture is lean at diesel engine
operations. It is difficult for pilot fuel to ignite and
provide sufficient combustion of the mixture. Thus, the very lean mixture cannot be burned and is released
with exhaust, and this situation concludes with poor
fuel efficiency and lower BTE [34]. Slower burning
rate increases the heat loss during combustion, causing
a decrease in BTE due to the slower flame propagation
speed [34]. Cheenkachorn et al. [35] researched the
effect of dual fuel (diesel fuel and natural gas) between
1100 - 2000 rpm and resulted with a decrease in BTE
at all cycles for dual mode. It should be noted that the
results of Cheenkachorn et al. [35] are consistent with
result of this study in terms of BTE.
Fig. 6 Effect of different amount of natural gas
addition on brake thermal efficiency
NOx is one of the most hazardous emissions released
from diesel engine and it is composed of nitrogen
monoxide (NO) and nitrogen dioxide (NO2). The
formation of NO in the combustion zone is chemically
complex and two typical mechanisms are involved:
thermal mechanism (Zeldovich mechanism) and
prompt mechanism (Fenimore mechanism). According
to thermal mechanism, NO formation is affected by in-cylinder temperature and oxygen concentration. NO
formation occurs when temperature is above 1800 K.
Its formation rate increases exponentially with increase
of in-cylinder temperature [38, 42]. The prompt NO
formation is only observed under rich fuel conditions
where a reasonable amount of hydrocarbon is
anticipated to react with N2. The prompt NO has
relatively weak temperature dependence in comparison
with thermal NO [43, 44]. Under diesel engine
combustion conditions, thermal mechanism is believed
to be the predominant contributor to total NOX formation [45-47].
Effect of natural gas addition on NOx emissions is
shown in Fig. 7. NOx emissions increased with rising
natural gas energy content. NOx emissions increased
by 14.5%, 2.5% and 89.4% with 15%, 40% and 75%
natural gas addition compared to only diesel fuel,
respectively. On the other hand, in terms of NOx
emissions, obtained results via Boost software were
quite close to experimental results and had similar
trend curves. The specific heat capacity ratio of natural
gas is higher than that of air [34]. The addition of
Proceedings of 2nd International Conference on Mechanical and Aeronautical Engineering
Held on 13th – 14th July 2016, in Bangkok, ISBN: 9788193137352
47
natural gas increases the overall heat capacity of the in-
cylinder mixture, accordingly, mean temperature at the
end of compression stroke and during the overall
combustion process reduce. The lower combustion
temperature reduces NOX formation. Natural gas
injection reduces the amount of air and concentration
of oxygen in the cylinder charge, resulting with
reduction in oxygen availability for NOX formation.
On the other hand, greater intensity of heat release in
the premixed combustion stage increases the maximum
combustion temperature causing an increase in NOX
emissions [34]. No increase is observed in NOx emissions with NG addition until a definite ratio, but
NOx emissions increase with 75% NG addition.
Egusquiza et al. [37] studied effect of NG addition on
NOx emissions in a 4 cylinder diesel engine and found
out that NOx emissions depend on engine load and NG
quantity. Papagiannakis et al. [41] studied on a
naturally aspirated single cylinder engine at 2 different
engine speed and 4 different engine loads (between
20% and 80%). NOx emissions reduced at dual mode
for all working conditions.
Fig. 7 Effect of different amount of natural gas
addition on NOx emission
Pressure transducers opened to combustion chamber
are installed on cylinder head. In cylinder gas pressure
values obtained by these pressure transducers supply
information about ignition delay, diesel noise and in
cylinder pressure characteristics [38]. In cylinder
pressure values also provide data on indicator diagram, in cylinder temperature values, heat release rate, and
fuel burn rate.
Parameters as peak in-cylinder pressure and pressure
rise rate are relevant to engine noise, vibration and
service life [34]. Fig. 8 illustrates measured in-cylinder
pressure data for several natural gas energy values.
Maximum cylindrical pressure values are 82.6 bar,
88.1 bar, 98.2 bar and 104.2 bar for 0% natural gas,
15% natural gas, 40% natural gas and 75% natural gas,
respectively. The peak cylindrical pressure increases by
6.6% at 15% natural gas content compared to 0%
natural gas, and it increases by 18.8% and 26.6% for
40% and 75% natural gas energy contents,
respectively. The reason of this increase is rapid heat
release of premixed mixture near TDC [51]. High
flame speed of natural gas increases maximum cylinder
gas pressure values based on fast combustion of
combustible mixture already in cylinder [52].
Combustion noise increased because of high flame
speed of natural gas depending on increase in hydrogen and methane gas mixture rate. Increasing gas fuel
quantity converts combustion phase into an explosive
type combustion characteristic. Lounici et al. [53]
reported higher peak in-cylinder pressures at dual fuel
operation conditions parallel to this study.
Fig. 8 Variation of the in-cylinder pressure with crank
angle according to several natural gas energy fractions
at a 1500 rpm constant engine speed Heat release rate is used for combustion analysis which
supplies data about in-cylinder combustion
characteristic. To obtain detailed information about in-
cylinder combustion characteristic, heat release rate
should be interpreted together with emission and
performance. So, effect of gas fuel energy content
variation on heat release rate is also studied. The effect
of different level of natural gas (0%, 15%, 40% and
75%) addition on rate of heat release at 1500 rpm
engine speed and 100% engine load is shown in Fig. 9.
The peak heat release rate is 39.7 J/oCA for 0% natural
gas, and as 50.5 J/oCA, 31.6 J/oCA, and 37.7 J/oCA for 15%, 40% and 75% natural gas energy contents,
respectively. The peak heat release rate increases by
27.2% at 15% natural gas content compared to 0%
natural gas, but it decreases by 20.3% and 5% for 40%
and 75% natural gas energy contents, respectively.
Proceedings of 2nd International Conference on Mechanical and Aeronautical Engineering
Held on 13th – 14th July 2016, in Bangkok, ISBN: 9788193137352
48
There are four phases of conventional diesel engine
combustion: ignition delay, premixed or rapid
combustion phase, mixing controlled combustion
phase, and late combustion phase [38]. During dual
fuel combustion mode, most of diesel fuel is replaced
by natural gas. Since the ignition delay is longer, there
is a few or no mixing controlled combustion.
Combustion characteristic changes totally when energy
content of gas fuel reaches at 75%, difference between
premixed combustion and mixing controlled
combustion phases is no longer recognizable.
Papagiannakis and Hountalas [54] found similar results with this study in terms of heat release rate.
Fig. 9 Variation of the heat release rate with crank
angle according to several natural gas energy fractions
at a 1500 rpm constant engine speed
VII. Conclusion Engine performance regarding with brake specific fuel
consumption and brake thermal efficiency, and
emissions of CO, THC, smoke and NOx were tested in
this experimental study. Combustion characteristic
related with cylinder gas pressure and heat release rate
was analysed on a four stroke, water cooled, naturally
aspirated, single cylinder compression ignition engine
at 1500 rpm engine speed, 100% engine load and
different natural gas energy levels (0%, 15%, 40% and
75%).
Obtained results are summarized briefly as shown
below: a. NOx emissions increased by 14.5%, 2.5%
and 89.4%, respectively with 15%, 40% and
75% natural gas addition compared to only
diesel fuel.
b. Combustion characteristic changes totally
when energy content of gas fuel reaches at
75%, difference between premixed
combustion and mixing controlled
combustion phases is no longer recognizable.
c. Numerical results were obtained by Boost
program. Majority of them were in good
agreement with experimental ones within
acceptable limits.
d. Validation of the all findings was performed
with the results from literature in detail.
Acknowledgements
This research was supported by the Yıldız Technical
University Scientific Research Projects Coordination Department. Project Number: 2014-06-01-DOP05.
Also, the authors are indebted to Şahin Metal A.Ş. and
Erin Motor for test apparatus and equipment donation.
Nomenclature
BSFC Brake specific fuel consumption
CI Compression ignition
CNG Compressed natural gas
CO Carbon monoxide
CO2 Carbon dioxide
DPF Diesel particulate filter ECU Electronic control unit
H2 Hydrogen
HC Hydrocarbons
HCCI Homogenous charge compression
ignition
He Helium
ICEs Internal combustion engines
LNT Lean NOx Trap
LPG Liquefied petroleum gas
N2 Nitrogen
NG Natural gas
NOx Oxides of nitrogen O2 Oxygen molecule
SCR Selective catalytic reduction
SI Spark ignition
THC Total unburned hydrocarbons
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