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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