1. Exhaust Emissions and Combustion Characteristics of a Direct
Injection (DI) Diesel Engine Fueled with Methanol-Diesel Fuel
Blends at Different Injection Timings
Mustafa Canakci,*,†,‡ Cenk Sayin,§ and Metin Gumus§
Department of Mechanical Education, Kocaeli UniVersity, Izmit 41380, Turkey, AlternatiVe Fuels Research
and DeVelopment Center, Kocaeli UniVersity, Izmit 41040, Turkey, and Department of Mechanical
Education, Marmara UniVersity, Istanbul 34722, Turkey
ReceiVed May 26, 2008. ReVised Manuscript ReceiVed August 20, 2008
In the recent years, environmental concerns and depletion in petroleum resources have forced researchers
to concentrate on finding renewable alternatives to conventional petroleum fuels. Therefore, alcohols as renewable
and alternative energy sources for the diesel engines gain importance. For this reason, in this study, the
performance, exhaust emissions, and combustion characteristics of a single cylinder diesel engine have been
experimentally investigated under different injection timings when methanol-blended diesel fuel was used
from 0 to 15%, with an increment of 5%. The tests were conducted at three different injection timings (15°,
20°, and 25 °CA BTDC) by changing the thickness of advance shim. All tests were conducted at four different
loads (5, 10, 15, and 20 Nm) at constant engine speed of 2200 rpm. The experimental test results showed that
BSFC, BSEC, combustion efficiency, and NOx and CO2 emissions increased as BTE, rate of heat release,
peak cylinder pressure, smoke number, and CO and UHC emissions decreased with an increasing amount of
methanol in the fuel blend. In comparison to the values at the original injection timing (20 °CA BTDC), the
values at the retarded injection timing (15 °CA BTDC) of peak cylinder pressure, rate of heat release, combustion
efficiency, and NOx and CO2 emissions decreased, while smoke number and UHC and CO emissions increased
at all test conditions. On the other hand, The advanced injection timing (25 °CA BTDC), smoke number, and
UHC and CO emissions diminished and peak cylinder pressure, rate of heat release, combustion efficiency,
and NOx and CO2 emissions increased at all test conditions. In terms of BSFC, BSEC, and BTE, retarded and
advanced injection timings gave negative results in all fuel blends compared to original injection timing.
Introduction
Because of their fuel economy and high reliability, compres-
sion-ignition (CI) engines known as diesel engines have been
penetrating a number of markets around the world. The existing
CI engines operate with conventional diesel fuel derived from
crude oil. It is well-known that the world petroleum resources
are limited and the production of crude oil is becoming more
difficult and expensive. On the other hand, because of the
growing concern over possible adverse health effects caused
by diesel emissions, the pollutions including unburned hydro-
carbons (UHC), carbon monoxide (CO), nitrogen oxides (NOx),
and particulate matter [quantified as smoke number (SN)] have
been regulated by laws in many developed countries. In the
last few years, many studies on the IC engines have been carried
out with the aim of reducing exhaust emissions by changing
operating parameters, such as valve timing, injection timing,
and atomization rate. At the same time, depletion of fossil fuels
and environmental considerations have led to investigations on
renewable fuels, such as methanol, ethanol, hydrogen, and
biodiesel.1-5
Alcohols (ethanol and methanol) have been considered as
alternative fuels for diesel engines. Methanol is manufactured
from any material that can be decomposed into CO (or CO2)
and hydrogen. In this regard, it may be produced from the
sources, which are independent from petroleum. The primary
feedstocks for methanol production are natural gas, lignite coal,
and renewable sources, such as wood, agricultural materials
biomass, waste biomass, and municipal wastes.6-8
Diesel fuel is a complex mixture of a large number of
hydrocarbons (such as C3-C25 hydrocarbons). For this reason,
its fuel properties can change depending upon the proportion
of hydrocarbon types used in the fuel mixture. However,
methanol (CH3OH) is a simple compound. It contains an oxygen
atom, so that it can be viewed as a partially oxidized
hydrocarbon fuel. It has a lower heating value than diesel fuel;
therefore, much more fuel is needed to obtain the same
performance as that of a diesel-fueled engine. Its high stoichio-
metric fuel/air ratio, high oxygen content, and high H/C ratio
may be beneficial for improving the combustion and reducing
* To whom correspondence should be addressed. Telephone: +90-262-
3032285. Fax: +90-262-3032203. E-mail: mustafacanakci@hotmail.com.
† Department of Mechanical Education, Kocaeli University.
‡ Alternative Fuels Research and Development Center, Kocaeli University.
§ Marmara University.
(1) Durgun, O.; Ayvaz, Y. The use of diesel fuel-gasoline blends in
diesel engines. Proceedings of the 1st International Energy and Environment
Symposium, Turkey, 1996; pp 9105-9120.
(2) Yuksel, F.; Yuksel, B. Renewable Energy 2004, 29, 1181–1191.
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Energy & Fuels 2008, 22, 3709–3723 3709
10.1021/ef800398r CCC: $40.75 2008 American Chemical Society
Published on Web 09/27/2008
2. the soot and smoke in diesel engines. Methanol has a higher
latent heat of vaporization than diesel fuel. It extracts much
more heat as it vaporizes, and it can lead to a cooling effect on
the cylinder charge. Therefore, the cylinder gas temperature can
be decreased because of its cooling effect on the charge, and
emissions of NOx may be reduced. On the other hand, methanol
has poor ignition behavior because of its low cetane number,
high latent heat of vaporization, and high ignition temperature.
Thus, it can cause a longer ignition delay. The autoignition
temperature of methanol is higher than that of diesel fuel, which
makes it safer for transportation and storage. However, it has a
lower flash point than that of diesel fuel, which is a disadvantage
for safety.9-11
The possible benefits and shortcomings of methanol as a fuel
for CI engines are summarized above. Problems concerning the
use of methanol in diesel engines can be different, which are
briefly described below. The simplest method of using methanol
in a CI engine is to blend it with diesel. The most important
problem is the phase separation. Using a mixer inside the fuel
tank can prevent this problem. Moreover, an ignition improver,
such as diethyl ether, can be added to the blended fuel to
compensate for the cetane number.12
The exhaust emissions from diesel vehicles are very complex
mixtures containing many types of pollutants, which can be
found in different forms, such as particulate, semi-volatile, and
gaseous phases. Using methanol-blended diesel fuel can reduce
the air pollution. Therefore, many researchers have studied the
influence of this alternative fuel on the engine performance and
exhaust emissions of IC engines. Huang et al.,13 for instance,
used various blend rates of methanol-diesel fuels in the engine
tests. The results indicated that the increase of methanol content
decreased smoke number and CO and UHC emissions but
increased brake-specific fuel consumption (BSFC) and NOx
emissions.
Lin and Chao14 investigated the effect of the methanol-
containing additive (MCA) on the biological characteristics of
diesel exhaust emissions. The engine was tested on a series of
diesel fuels blended with five additive levels (0, 5, 8, 10, and
15% of MCA by volume). The result showed that the MCA
additive moderately lowered the toxicity levels of particle-
associated (SOF) samples but generally increased the vapor-
phase (XOC) associated toxicity.
Bayraktar15 studied the effect of methanol-blended diesel fuel
between 2.5 and 15 vol % on the engine performance and found
that the methanol could reduce the effective power and brake
thermal efficiency (BTE) to some degree and moderately
increase BSFC.
Cheng et al.16 investigated the performance and exhaust
emissions of a four-cylinder diesel engine operating on biodiesel
with methanol in either the blended or fumigation mode. They
compared the results to those operating on pure biodiesel and
pure diesel fuel. Experiments were performed on a four-cylinder
naturally aspirated direct injection (DI) diesel engine operating
at a constant speed of 1800 rpm with five different engine loads.
The results indicated reductions in CO2, NOx, and particulate
mass emissions and reductions in the mean particle diameter,
in both cases, compared to diesel fuel.
Chao et al.4 investigated the effect of the MCA on the
regulated and unregulated emissions from a diesel engine. The
engine was tested on a series of diesel fuels blended with five
additive levels (0, 5, 8, 10, and 15% of MCA by volume).
Results showed that MCA addition slightly decreased particulate
matter (PM) emissions but generally increased both UHC and
CO emissions.
Cheung et al.17 studied the effects of fumigation methanol
on the emissions of a diesel engine fueled with biodiesel as the
baseline fuel. The biodiesel used in this study was converted
from waste cooking oil. Experiments were performed on a four-
cylinder naturally aspirated diesel engine operating at a constant
speed of 1800 rpm for three engine loads. The results indicated
no significant change in BTE and CO2 emission, an increase in
both CO and UHC emissions, and a decrease in both NOx and
PM emissions.
The Lubrizol Corporation, in conjunction with the Lovelace
Respiratory Research Institute and several subcontracting
laboratories, recently conducted a health assessment of the
combustion emissions of Puri NOx diesel fuel emulsion
(diesel-water-methanol) in rodents. Combustion emissions
from either of two, 2002 model Cummins 5.9 L ISB engines
were diluted with charcoal-filtered air to expose concentrations
of 125, 250, and 500 mg of total PM/m3. The engines were
operated on a continuous, repeating, heavy-duty certification
cycle (U.S. Code of Federal Regulations, Title 40, Chapter I).
NO and PM were reduced when engines were operated on Puri
NOx versus California Air Resources Board diesel fuel under
these conditions.18
In a different study,19 a stabilized methanol-diesel blend was
developed and combustion characteristics and heat release
analysis were carried out in a CI engine. According to the
experimental results, increasing the methanol mass fraction in
the methanol-diesel fuel blends resulted in an increase in the
heat release rate at the premixed burning phase and shortened
the combustion duration at the diffusive burning phase.
Yao et al.20 investigated the controlling strategies of homo-
geneous charge compression ignition (HCCI) fueled by dimethyl
ether (DME) and methanol. The experimental work was carried
out on a modified single-cylinder diesel engine, which was fitted
with port injection of DME and methanol dual fuel. The results
showed that the exhaust gas recirculation (EGR) rate and DME
percentage are two important parameters to control the HCCI
combustion process.
For a diesel engine, fuel injection timing is a major parameter
that affects the combustion and exhaust emissions. The state of
air into which the fuel injected changes as the injection timing
is varied and, thus, ignition delay will vary. If the injection starts
earlier, the initial air temperature and pressure are lower;
therefore, the ignition delay will increase. If the injection starts
later (when the piston is closer to TDC), the temperature and
pressure are initially slightly higher and a decrease in ignition
delay results. Hence, variation in injection timing has a strong
(9) Kowalecwicz, A. Proc. Inst. Mech. Eng. 1993, 207, 43–52.
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Proc. Inst. Mech. Eng., Part D 2004, 218, 435–447.
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(15) Bayraktar, H. Fuel 2008, 87, 158–164.
(16) Cheng, C. H.; Cheung, C. S.; Chan, T. L.; Lee, S. C.; Yao, C. D.;
Tsang, K. S. Fuel 2008, 87, 1870–1879.
(17) Cheung, C. S.; Cheng, C.; Chan, T. L.; Lee, S. C.; Yao, C.; Tsang,
K. S. Energy Fuels 2008, 22, 906–914.
(18) Reed, M. D.; Blair, L. F.; Daly, I.; Gigliotti, A. P.; Gudi, R.;
Mercieca, M. D.; McDonald, J. D.; O’Callaghan, J. P.; Seilkop, S. K.;
Ronsko, N. L.; Wagner, V. O.; Kraska, R. C. Toxicol. Ind. Health 2006,
22, 65–85.
(19) Huang, Z. H.; Lu, H. B.; Jiang, D. M.; Zhang, J. Q.; Wang, X. B.
J. Automot. Eng. 2004, 218, 1011–1024.
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2046–2056.
3710 Energy & Fuels, Vol. 22, No. 6, 2008 Canakci et al.
3. effect on the engine performance and exhaust emissions,
especially on the BSFC, BTE, and NOx emissions, because of
changing the maximum pressure and temperature in the engine
cylinder.21,22
Several studies have shown that the injection timing affects
the engine performance and exhaust emissions of CI engines.
Parlak et al.23 studied the influence of injection timing on the
NOx emissions and BSFC of a low-heat rejection (LHR) IDI
diesel engine using diesel fuel. The tests were conducted with
variable loads at the engine speeds of 1000, 1400, 1800, and
2000 rpm and the static injection timing of 38°, 36°, 34°, and
32° crank angle (CA) before top dead center (BTDC). After
the load tests were conducted for the original diesel engine, the
same test order was adopted for the LHR engine. When the
LHR engine was operated with the injection timing of 38 °CA
BTDC, which is the original start of injection timing of the
engine, the NOx emission increased about 15%. However, when
the injection timing was retarded to 34 °CA in the LHR case,
some decreases in the exhaust emissions and BSFC were
observed compared to those of the original case.
Nwafor24 examined the effect of advanced injection timing
on the engine performance and exhaust emissions of natural
gas used as a primary fuel in a dual-fuel CI engine. The test
engine had an original (ORG) start of injection timing of 30
°CA BTDC. The injection was advanced by 3.5° (i.e., 33.5 °CA
BTDC). The results indicated that dual-fuel combustion pro-
duced higher HC emissions than that of pure diesel fuel.
Significant reductions in CO and CO2 emissions were obtained
when running the engine with the advanced injection timing.
On the other hand, advanced injection timing caused a slight
increase in BSFC and decrease in BTE.
Sayin and Canakci25 investigated the influence of injection
timing on the exhaust emissions and engine performance of a
DI diesel engine using ethanol-blended diesel fuel from 0 to
15%, with an increment of 5%. The original start of injection
timing of the engine was 27 °CA BTDC in that study. The tests
were performed at five different injection timings (21°, 24°, 27°,
30°, and 33 °CA BTDC). The experimental test results showed
that BSFC and emissions of NOx and CO2 increased as BTE
and emissions of CO and HC decreased with an increasing
amount of ethanol in the fuel mixture. When compared to the
results of original injection timing (27 °CA BTDC), NOx and
CO2 emissions increased and unburned HC and CO emissions
decreased for the retarded injection timings (21° and 24 °CA
BTDC) at the all test conditions. On the other hand, with the
advanced injection timings (30° and 33 °CA BTDC), HC and
CO emissions decreased and NOx and CO2 emissions increased.
In terms of BSFC and BTE, retarded and advanced injection
timings gave negative results for all engine speeds and loads
as compared to the original injection timing in all fuel blends.
From the literature review, the influence of injection timing
on the exhaust emissions, combustion characteristics, and
performance parameters of a diesel engine has not been clearly
studied when using methanol-blended diesel fuel in a CI engine.
Therefore, these topics need to be investigated to make up for
the deficiency in the literature. For this reason, in the present
study, the effects of both injection timing and methanol-blended
diesel fuel on the engine performance, exhaust emissions, and
combustion characteristics were experimentally investigated on
a single-cylinder CI engine.
Experimental Apparatus and Procedure
The experiments were conducted on a single-cylinder, four-
stroke, DI, and naturally aspirated CI engine. Details of the engine
specification are shown in Table 1. The engine was loaded by an
electrical dynamometer rated at 10 kW and 380 V. The load on
the dynamometer was measured using a strain gauge load sensor.
An inductive pickup speed sensor was used to measure the speed
of the engine. The pressure-time history of the cylinder was
measured by a Kistler Model 6052B air-cooled piezo-quartz
pressure sensor, which was mounted on the cylinder head. The
signals were then passed onto a Kistler Model 5644A charge
amplifier. The crankshaft position was obtained using a crankshaft
angle sensor to determine the cylinder pressure as a function of
the crank angle. The crank angle signal was obtained from an angle-
generating device mounted on the main shaft. The signal of the
cylinder pressure was acquired for every 0.75 °CA, and the
acquisition process covered 100 completed cycles. The engine oil
temperature, coolant temperature, exhaust temperature, inlet air
temperature were measured using K-type thermocouples. CO, CO2,
and HC emissions were measured with an infrared gas analyzer
(Bilsa Mod 210) with an accuracy of (0.001%, (0.01%, and (1
ppm, respectively. NOx emissions were recorded using an electro-
chemical gas analyzer (Kane-May Qintox KM9106) with an
accuracy of (1 ppm. Smoke levels were obtained using a Bosch
system with an accuracy of (0.1%. The analyzers were calibrated
before the experiments. The emission data were expressed as
“brake-specific” basis (g/kWh), except for the Bosch smoke number.
Brake-specific emissions are the mass flow rate of the pollutant
divided by the engine power. Fuel consumption was quantified by
the combined container method. Pressure in the intake manifold
was determined by an inclined manometer. The accuracies of the
measurements and the uncertainties in the calculated results are
given in Table 2.
To prepare the methanol-blended fuel mixture, two fuels (euro-
diesel and methanol) were used. Euro-diesel was obtained from
TUPRAS Petroleum Corporation. Methanol, with a purity of 99%,
was purchased from a commercial supplier. The fuel properties are
shown in Table 3. The euro-diesel was blended with methanol to
obtain four different fuel blends from 0 to 15%, with an increment
of 5%. The fuel blends were prepared just before starting the
(21) Heywood, J. B. Internal Combustion Engines; Mc-Graw Hill: New
York, 1984; pp 546-547 and 882-383.
(22) Borat, O.; Balci, M.; Surmen, A. Internal Combustion Engines;
Gazi University Publishing: Turkey, 2000; pp 264-265 (in Turkish).
(23) Parlak, A.; Yasar, H.; Hasimoglu, C.; Kolip, A. Appl. Therm. Eng.
2005, 25, 3042–3052.
(24) Nwafor, O. M. I. Sadhana 2002, 27 (3), 375–382.
(25) Sayin, C.; Canakci, M. Energy ConVers. Manage., in press, doi:
10.1016/j.enconman.2008.06.007.
Table 1. Technical Specifications of the Test Engine51
engine type Lombardini 6 LD 400
cylinder number 1
bore 86 mm
stroke 68 mm
total cylinder volume 395 cm3
injector opening pressure 200 bar
number of nozzle hole 4
start of injection timing 20 °CA BTDC
compression ratio 18:1
maximum torque 18 Nm at 2200 rpm
maximum power 8 kW at 2000 rpm
Table 2. Accuracies of the Measurements and the Uncertainties
in the Calculated Results
measurements accuracy
load (2 Nm
speed (25 rpm
time (0.5%
temperatures (1 °C
calculated results uncertainty
power (2.55%
BSFC (2.60%
BSEC (2.60%
BTE (2.60%
DI Diesel Engine with Methanol-Diesel Fuel Blends Energy & Fuels, Vol. 22, No. 6, 2008 3711
4. experiment to ensure that the fuel mixture was homogeneous. A
mixer was also mounted inside the fuel tank to prevent phase
separation. All data were collected after the engine stabilized. Each
test was repeated 3 times. The values given in this study are the
average of these three results. The experimental setup is shown in
Figure 1.
The original start of injection timing of the engine was 20 °CA
BTDC. The thickness of one advance shim, located in connection
place between the engine and fuel pump was 0.25 mm, and adding
or removing one shim changed the injection timing by 5 °CA.
Experiments were carried out in three different injection timings
(15°, 20°, and 25 °CA BTDC) values. All tests were conducted at
four different loads (5, 10, 15, and 20 Nm) at the constant engine
speed of 2200 rpm. The values of engine oil temperature, mass
flow rate of air, exhaust temperature, and pollutants, such as SN,
CO, CO2, UHC, and NOx, were recorded during the experiments.
Calculation Methods
Heat release analysis can yield valuable information about the
effect of engine design changes, fuel injection system, fuel type,
and engine operating conditions on the combustion process and
engine performance.26 In this study, the cylinder pressure data were
used to evaluate the rate of heat release (ROHR), which is a
simplified thermodynamic model. The ROHR was calculated using
the first law analysis of thermodynamics. The ROHR at each crank
angle was determined by the following formula:
Q )
γ
γ - 1
(PdV) +
1
γ - 1
(VdP) + Qw (1)
where Q is the apparent heat release rate (J), γ is the ratio of specific
heats, which is calculated according to an empirical equation,27 P
is the cylinder pressure (Pa), V is the instantaneous volume of the
cylinder (m3), and Qw is heat-transfer rate (J) from the wall
calculated on the basis of the Hohenberg correlation,28 and the wall
temperature was assumed to be 723 K. For this calculation, the
contents of the cylinder were assumed to behave as an ideal gas
(air), with the specific heat being dependent upon temperature;
leakage through the piston rings was neglected.29
Because the chemical energy of the fuel is not fully released
inside the engine during the combustion process, it is useful to
define combustion efficiency. Thus, the combustion efficiency was
calculated for the tested fuels in the engine. The combustion energy
losses take into account the energy required to form NO, NO2, and
the energy lost owing to incomplete oxidation of CO to CO2 and
UHC fuel to CO2 and H2O. The enthalpy of formation for H2O
and CO2 from fuel oxidation was not calculated theoretically;
instead, the LHV was used. Each enthalpy of formation calculation
is mass-specific to the emission quantity measured during the engine
testing.30
The enthalpy of formation for the measured exhaust emissions
is used to calculate the combustion efficiency in accordance with
the equation below
ηcombustion )
LHVfuel - HNONO - HNO2
NO2 - HCOCO - LHVfuelUHC
LHVfuel
× 100
(2)
where ηcombustion is the combustion efficiency, LHVfuel is the lower
heating value of the fuel (kJ/kg of fuel), HNO is the enthalpy
formation of NO (kJ/g of NO), HNO2 is the enthalpy formation of
NO2 (kJ/g of NO2), HCO is the enthalpy formation of CO (kJ/g of
CO), NO is the exhaust emission level of NO (g/kg of fuel), NO2
is the exhaust emission level of NO2 (g/kg of fuel), CO is the
exhaust emission level of CO (g/kg of fuel), and UHC is the exhaust
emission level of UHC (kg/kg of fuel).
Results and Discussion
Methanol-blended diesel fuel can decrease the pollutant
emissions. However, to reach the emission reduction, it may
require some modification on the engine. The injection timing
has a significant effect on the performance and exhaust
emissions of a CI engine. Therefore, in this study, the effects
of injection timing and methanol-blended diesel fuel on the
engine performance, exhaust emissions, and combustion char-
acteristics were experimentally investigated on a single-cylinder
CI engine. The experimental conditions were selected as follows:
four engine loads (5, 10, 15, and 20 Nm), 2200 rpm constant
speed, and three injection timings (15°, 20°, and 25 °CA BTDC).
The fuels were M0, M5, M10, and M15, indicating the content
of methanol in different volume ratios (e.g., M5 contains 5%
methanol and 95% diesel fuel in volume).
Engine Performance. Table 4 shows the BSFC, BSEC, BTE,
and combustion efficiency values and the percent changes in
these parameters compared to M0 (diesel fuels), respectively,
at different engine loads and injection timings.
Brake-Specific Fuel Consumption (BSFC). The BSFC is
defined as the ratio of the fuel consumption to the brake power.
As shown in Table 4, the change in BSFC relative to M0 is
3.94, 7.67, and 18.26% for M5, M10, and M15, respectively,
at 5 Nm and ORG injection timing. The results show that
increasing the methanol ratio in the fuel mixture leads to an
increase in BSFC. This behavior is attributed to the LHV of
methanol, which is distinctly lower than that of the diesel fuel.31
Therefore, the amount of fuel introduced to the cylinder for a
desired energy input has to be greater with the methanol fuel.
The BSFC decreases about 2.07 times as the engine load
increases from 5 to 20 Nm constant loads for M10. This decrease
in BSFC could be explained by the fact that, as the engine load
increased, the rate of increasing brake power was much more
than that of the fuel consumption.
As seen in the table, the minimum BSFC values were
obtained at ORG injection timing for all fuel blends. When the
injection timing was advanced and retarded compared to ORG
injection timing, the change in BSFC was measured by 13.58
and 22.22%, respectively, for M10 at 10 Nm. With advancing
(26) Ghojel, J.; Honnery, D. Appl. Therm. Eng. 2005, 25, 2072–2085.
(27) Brunt, M. F. J.; Rai, H.; Emtage, A. L. SAE Tech. Pap. 981052,
1998.
(28) Hohenberg, G. H. SAE Tech. Pap. 790825, 1979.
(29) Hayes, T. K.; Savage, L. D.; Sorenson, S. C. SAE Tech. Pap.
860029, 1986.
(30) Canakci, M. Bioresour. Technol. 2007, 98, 1167–1175.
(31) Korkmaz, I. A study on the performance and emission character-
istics of gasoline and methanol fuelled spark-ignition engines. Ph.D. Thesis,
Istanbul Technical University, Turkey, 1996; p 49.
Table 3. Properties of the Fuels Used in the Tests
methanol49 euro-diesel50
formula CH3OH C14.34H24.75
molecular weight (kg/kmol) 32 196.8
boiling temperature (°C) 64.7 287
density (g/cm3, at 20 °C) 0.79 0.83
flash point (°C) 11 78
autoignition temperature (°C) 470 235
lower heating value (MJ/kg) 20.27 42.74
cetane number 4 56.5
viscosity (mm2/s, at 25 °C) 0.59 3.35
stoichiometric air/fuel ratio 6.66 14.28
stoichiometric fuel/air ratio 0.15 0.07
heat of vaporization (MJ/kg) 1.11 0.27
3712 Energy & Fuels, Vol. 22, No. 6, 2008 Canakci et al.
5. injection timing, the ignition delay will be longer and the speed
of the flame will be shorter. These cause a reduction in engine
output power. Therefore, fuel consumption per output power
will increase. On the other hand, retarding injection timing
means later combustion pressure rises only when the cylinder
volume was expanding rapidly, and this reduces the effective
pressure to do work.32
Brake-Specific Energy Consumption (BSEC). The BSEC is
described as the product of BSFC and LHV. As shown in the
Table 4, the BSEC increases with methanol content. Minimum
BSEC was acquired at 5.05 MJ/kWh for M0, 5.08 MJ/kWh for
M5, 5.10 MJ/kWh for M10, and 6.02 MJ/kWh for M15 at 20
Nm load and ORG injection timing, respectively. It is well-
known that the LHV of the fuel affects the engine power. The
lower heat content of the methanol-diesel blend causes some
reductions in the engine power. In addition, the theoretical air/
fuel ratio of diesel fuel is about 2 times higher than that of
methanol, as shown in Table 3. For these reasons, the effective
power should decrease with the increase of the methanol amount
in the fuel mixture. Thus, the engine needs to consume more
heat to maintain the same amount of power output.33,34
The BSEC reduced as the engine load increased because of
noticeably diminishing BSFC for all fuel blends and injection
timing. The BSEC reduced by 14.70% as the engine load
increased from 10 to 20 Nm constant loads for M15 at retarded
injection timing. When the injection timing changed from ORG
injection timing, BSEC values increased because of the increase
in the energy requirement to sustain the same amount of power
output at ORG injection timing. The increments for the advanced
and the retarded injection timings were 9.92 and 13.29% for
M0 at 10 Nm, respectively.
Brake Thermal Efficiency (BTE). The BTE is defined as the
ratio of the brake power to fuel consumption and LHV. As
demonstrated in Table 4, in comparison to M0, the BTE
decreased by 3.84, 14.25, and 34.17% for M5, M10, and M15,
respectively, at 15 Nm and ORG injection timing. BTE indicates
the ability of the combustion system to accept the experimental
fuel and provides comparable means of assessing how efficient
the energy in the fuel was converted to mechanical output. From
the previous discussion, it could be concluded that, as the
methanol amount increases in the fuel blend, the BSFC
increases, because the LHV value of the blend decreases. As
mentioned above, BTE is a function of BSFC and LHV of the
blend for a constant effective power. It is clear that BSFC is
more effective than LHV with regard to increasing BTE.
Therefore, the BTE increased as the methanol content decreased
in the blended fuel for all injection timings.35
Increasing engine loads caused an increase in BTE values
because of a noticeable decline in BSFC for all fuel blends and
injection timings. The BTE increased by 16.46% as the engine
load increased from 10 to 20 Nm at constant loads for M5 at
advanced injection timing. The best results in terms of BTE
were obtained at ORG injection timing. Retarded or advanced
injection timing diminished BTE values by increasing BSFC.
When the injection timing was advanced and retarded in
comparison to ORG injection timing, BTE decreased by 19.44
and 31.02% for M10 at 20 Nm load, respectively.
Combustion Efficiency. The change in combustion efficiency
increased with an increasing methanol ratio in the fuel blend at
all engine loads and injection timings, compared to M0. As
shown in Table 4, the change in combustion efficiency increased
by 0.15, 0.24, and 0.60% for M5, M10, and M15, respectively,
at 20 Nm engine load and advanced injection timing relative to
M0. When methanol is added into diesel fuel, the fuel contains
more oxygen, which reduces CO and UHC emissions and
increases NOx emissions. These effects caused an increase in
combustion efficiency as shown in eq 2.
Combustion efficiency slightly increased with an increasing
engine load from 5 to 20 Nm for all test fuels because of better
volumetric efficiency and atomization rate. The maximum
combustion efficiency (99.88%) was obtained with advanced
injection timing, followed by ORG (99.81%) and retarded
injection timing (99.42%) for M15 at 20 Nm load. The increase
in combustion efficiency with advancing the injection timing
was attributed to increases in the NOx emissions and decreases
in CO and UHC emissions, as mentioned below.
Exhaust Emissions. The exhaust emissions measured were
CO, CO2, UHC, NOx, and smoke number (SN). Table 5 and
Figures 2-5 show the emission values and percent changes in
the emissions at different engine load and injection timing
compared to M0, respectively.
Carbon Monoxide (CO) Emissions. In general, while the
engine is running under fuel-rich mixture conditions, the exhaust
(32) Sayin, C.; Kilicaslan, I.; Canakci, M.; Ozsezen, N. Appl. Therm.
Eng. 2004, 8, 1315–1324.
(33) Can, O.; Celikten, I.; Usta, N. Energy ConVers. Manage. 2004,
45, 2429–2440.
(34) Rakopolous, C. D.; Kyristis, D. C. Energy 2001, 26, 705–722. (35) Sayin, C.; Uslu, K. Int. J. Energy Res. 2008, 32, 1006–1015.
Figure 1. Experimental setup.
DI Diesel Engine with Methanol-Diesel Fuel Blends Energy & Fuels, Vol. 22, No. 6, 2008 3713
7. will contain a large amount of CO emission, because there is
not sufficient oxygen to convert all of the carbon atoms of the
fuel into CO2. Thus, the most important parameters that affect
CO emissions are an insufficient amount of air and an insuf-
ficient time in the cycle for complete combustion.36
Concerning the effect of different fuels on CO emissions, it
was uncovered that increasing the methanol ratio in the fuel
blend lessened CO emissions. In comparison to M0, the change
in CO emissions was 18.64, 32.20, and 38.98% for M5, M10,
and M15, respectively, at 5 Nm load and advanced injection
timing, as demonstrated in Figure 2a. Methanol is an oxygenated
fuel and leads to more complete combustion; hence, CO
emissions reduce in the exhaust. CO emission decreased
gradually when the engine load increased. When the engine load
increases, the combustion temperature increases and CO emis-
sions start to decrease.37 For example, in comparison to Figure
3b and 4b, it was seen that the percent change in CO emissions
diminished by 20.68% for M5 as the engine load increased from
10 to 15 Nm constant load at ORG injection timing. Parts a-c
of Figure 3 illustrate that the percent change in CO emissions
with different methanol blends at different injection timings for
10 Nm load. From these figures, it was concluded that advanced
injection timing decreased the CO emission by 8.44% and
retarded injection timing increased the CO emission by 6.82%
compared to ORG injection timing for M10, respectively. The
advanced injection timing produced the higher cylinder tem-
perature and increasing oxidation process between carbon and
oxygen molecules. These lead to a decrease in the percent
change in CO emissions.38
Unburned Hydrocarbon (UHC) Emissions. UHC emissions
consist of fuel that is incompletely burned. The term HC means
organic compounds in the gaseous state, solid HCs are part of
the particulate matter. Typically, UHCs are a serious problem
at low loads in CI engines. At low loads, the fuel is less apt to
impinge on surfaces; however, because of poor fuel distribution,
large amounts of excess air, and low exhaust temperature, lean
fuel-air mixture regions may survive to escape into the
exhaust.39
With regard to the effect of different methanol contents on
UHC emission, it was found that increasing the methanol ratio
in the fuel blend reduced UHC emissions. For instance, the UHC
emissions compared to M0 at ORG injection timing decreased
by 14, 24, and 40% for M5, M10, and M15, respectively, at 10
Nm load and retarded injection timing, as seen in Figure 3c.
When methanol was added to the diesel fuel, it provided more
oxygen for the combustion process and led to the improving
combustion. In addition, methanol molecules are polar and
cannot be absorbed easily by the nonpolar lubrication oil, and
therefore, methanol can lower the possibility of the production
of UHC emissions.40
UHC emissions lessened reasonably with increasing load,
which was the same trend as with CO. For example, in
comparison to Figure 3a to 5a, it was observed that the change
in UHC emissions diminished by 6.66% for M10 as the engine
load increased from 10 to 20 Nm constant loads at advanced
injection timing. Parts a-c of Figure 5 show the change in UHC
(36) Ganesan, V. Internal Combustion Engine; McGraw-Hill: New York,
1994; p 414.
(37) Abdel-Rahman, A. A. Int. J. Energy Res. 1998, 22, 483–513.
(38) Gumus, M. Renewable Energy, in press, doi: 10.1016/
j.renene.2008.02.005.
(39) Canakci, M. Idealized engine emissions resulting from the combus-
tion of isooctane supplemented with hydrogen. M.Sc. Thesis, Vanderbilt
University, Nashville, TN, 1996; pp 25-26.
(40) Alla, G. H.; Soliman, H. A.; Badr, O. H.; Rabbo, M. F. Energy
ConVers. Manage. 2002, 43, 269–277.
Table
5.
Emission
Results
of
the
Fuels
emissions
(g/kWh)
except
SN
CO
CO
2
NO
x
HC
SN
fuel
type
5
Nm
10
Nm
15
Nm
20
Nm
5
Nm
10
Nm
15
Nm
20
Nm
5
Nm
10
Nm
15
Nm
20
Nm
5
Nm
10
Nm
15
Nm
20
Nm
5
Nm
10
Nm
15
Nm
20
Nm
25
°CA
BTDC
M0
0.59
0.28
0.26
0.23
52.10
30.70
34.76
29.34
1.75
1.14
1.30
1.51
0.71
0.36
0.34
0.33
0.65
0.70
0.76
0.94
M5
0.48
0.21
0.19
0.16
54.49
34.78
39.22
36.44
1.82
1.23
1.49
1.92
0.62
0.30
0.27
0.26
0.64
0.64
0.65
0.80
M10
0.40
0.18
0.16
0.14
55.82
38.31
44.65
38.91
2.15
1.45
1.80
2.29
0.53
0.24
0.22
0.20
0.63
0.63
0.63
0.77
M15
0.36
0.16
0.15
0.11
60.43
41.05
48.87
50.94
2.54
1.75
2.02
2.55
0.45
0.18
0.16
0.16
0.56
0.58
0.57
0.70
20
°CA
BTDC
(ORG
Injection
Timing)
M0
0.71
0.33
0.27
0.23
46.54
25.14
26.44
23.72
1.43
0.78
0.81
1.06
0.80
0.40
0.37
0.35
0.71
0.76
0.83
0.96
M5
0.60
0.29
0.23
0.19
47.55
26.78
28.76
27.70
1.51
0.87
0.92
1.23
0.72
0.34
0.31
0.30
0.69
0.70
0.73
0.84
M10
0.53
0.24
0.19
0.17
48.99
30.13
32.44
30.83
1.67
0.94
1.09
1.54
0.65
0.30
0.28
0.25
0.67
0.69
0.70
0.82
M15
0.46
0.21
0.17
0.14
53.05
32.51
36.65
39.15
1.91
1.09
1.21
1.69
0.57
0.23
0.21
0.20
0.62
0.65
0.66
0.79
15
°CA
BTDC
M0
0.95
0.44
0.38
0.31
42.65
24.68
26.94
21.16
1.36
0.70
0.73
0.84
1.04
0.50
0.45
0.42
0.72
0.76
0.86
0.99
M5
0.82
0.40
0.33
0.26
43.13
25.13
29.18
24.35
1.44
0.76
0.81
0.96
0.86
0.43
0.38
0.35
0.71
0.72
0.77
0.88
M10
0.72
0.35
0.29
0.22
44.84
28.03
31.02
26.76
1.56
0.83
0.90
1.17
0.86
0.38
0.32
0.29
0.70
0.70
0.74
0.87
M15
0.63
0.29
0.26
0.17
48.03
31.37
36.13
34.28
1.66
0.95
1.02
1.22
0.77
0.30
0.26
0.23
0.65
0.67
0.71
0.82
DI Diesel Engine with Methanol-Diesel Fuel Blends Energy & Fuels, Vol. 22, No. 6, 2008 3715
8. emissions with different methanol blends at different injection
timings for 20 Nm load compared to M0. From these figures,
it was found that advanced injection timing caused a reduction
in UHC emission by 8.66% and retarded injection timing
boosted the UHC emission by 2.55% compared to ORG timing
for M15, respectively. Advancing the injection timing caused
an earlier start of combustion relative to the TDC. Because of
this, the cylinder charge, being compressed as the piston moved
to the TDC, had relatively higher temperatures and thus lowered
the UHC emissions.41-43
Nitrogen Oxides (NOx) Emissions. In a diesel engine, the fuel
distribution is non-uniform. The pollutant formation process is
strongly dependent upon the changes in the fuel with time
because of mixing. The oxides of nitrogen form in the high-
temperature burned region, which is non-uniform, and the
formation rates are highest in the regions closest to the
stoichiometric region.44
In this study, it was discovered that increasing the methanol
ratio in the blend raised NOx emissions. For example, as
presented in Figure 4b, the change in NOx emissions was
compared to M0 and showed that NOx augmented by 13.58,
34.56, and 49.38% for M5, M10, and M15, correspondingly,
at 15 Nm load and ORG injection timing. Methanol contains
34% oxygen, and its cetane number is lower than diesel fuel,
which boost peak temperature in the cylinder. On the other hand,
the lower heating value (LHV) of methanol is nearly 2 times
lower than diesel fuel and latent heat of vaporization of methanol
is about 4 times greater than diesel fuel, which decreases peak
temperature in the cylinder. However, as shown in Figure 6,
the exhaust temperature increased with an increasing methanol
ratio in the fuel mixture. It is clear from the figure that the cetane
number and oxygen content are more effective than LHV and
(41) Payri, F.; Bejanes, J.; Arregle, J.; Riesco, J. M. Oil Gas Sci. Technol.
2006, 2, 247–258.
(42) Pukrakek, W. W. Engineering Fundamentals of the Internal
Combustion Engine; Simon and Schuster Co.: New York, 1997; pp 278.
(43) Ajav, E. A.; Singh, B.; Bhattacharya, T. K. Biomass Bioenergy
1998, 15 (6), 493–502. (44) Agarwal, A. K. Prog. Energy Combust. Sci. 2007, 33, 233–271.
Figure 2. Changes in the emissions relative to M0 at 5 Nm load.
3716 Energy & Fuels, Vol. 22, No. 6, 2008 Canakci et al.
9. latent heat of vaporization with regard to increasing peak
temperature in the cylinder. Therefore, the concentration of NOx
increased as the methanol content was increased in the fuel
blend.45
Unlike CO and UHC emissions, NOx emissions increased
with an increasing engine load. For example, in comparison to
Figure 3c and 4c, it was detected that NOx emissions increased
by 5.71% for M5 as the engine load augments from 10 to 15
Nm constant loads at the retarded injection timing. Parts a-c
of Figure 4 demonstrate the percent change in NOx emissions
with different methanol blends at different injection timings for
15 Nm load. As shown in the figures for M15, retarded injection
timing reduced in NOx emission by 9.66% and advanced
injection timing increased in NOx emission by 6.32% compared
to ORG injection timing, correspondingly. When the injection
timing was retarded, it was observed that NOx emissions
decreased for all fuel mixtures. Retarding the injection timing
decreased the peak cylinder pressure because more of fuel
burned after TDC. Lower peak cylinder pressures resulted in
lower peak temperatures. As a consequence, the NOx concentra-
tion diminished.35
Smoke Number (SN). Soot is formed in the cylinder, from
heavy hydrocarbons in the gas phase, which condense and
coalesce in the oxygen-deficient regions in the very rich core
of the fuel sprays.46 Regarding the effect of different methanol
contents on SNs, it was observed that increasing methanol ratio
in the blend reduced SNs. The change in SNs compared to M0
implied that they diminished by 14.89, 18.80, and 25.53% for
M5, M10, and M15, respectively, at 20 Nm load and advance
injection timing as illustrated in Figure 5a. The presence of
atomic-bound oxygen in methanol satisfies positive chemical
control over soot formation. The tendency to generate soot by
(45) Nwafor, O. M. I.; Rice, G.; Ogbonna, A. I. Renewable Energy 2000,
21, 433–444.
(46) Challen, B.; Baranescu, R. Diesel Engine Reference Book, 2nd ed.;
Butterworth and Heinemann Publishing: Oxford, U.K., 1999; p 480.
Figure 3. Changes in the emissions relative to M0 at 10 Nm load.
DI Diesel Engine with Methanol-Diesel Fuel Blends Energy & Fuels, Vol. 22, No. 6, 2008 3717
10. the fuel dense region inside a diesel diffusion flame sheath is
reduced, so that soot-free spray combustion could be achieved.33
The formation of smoke is most strongly dependent upon
the engine load. As the load increases, more fuel is injected,
and this increases the formation of smoke.39 The results obtained
in this study supported this statement. For instance, in com-
parison to Figure 4c and 5c, it was seen that the change of SNs
increased by 5.71% for M10 as the engine load increased from
5 to 15 Nm at retarded injection timing. Advancing the injection
timing reduced the smoke emissions. The earlier injection led
to higher temperatures during the expansion stroke and more
time in which oxidation of the soot particles occurred.46 Parts
a-c of Figure 3 present the percent change in SNs at different
injection timings for 10 Nm load. As seen from these figures
for M15, advanced injection timing lowered in SNs by 2.67%
and retarded injection timing raised in SNs by 2.63% compared
to ORG timing for M15, respectively.
Carbon Dioxide (CO2) Emissions. CO2 emission is produced
by complete combustion of fuel. As illustrated in the Figure
3a, when the methanol amount increased in the fuel mixture,
the percent change in CO and UHC decreased. The percent
change in CO2 had an opposite behavior when compared to the
CO concentrations, and this was due to improving the combus-
tion process as a result of the oxygen content in the methanol.
The maximum increase in the change of CO2 was observed at
24.19, 32.61, and 73.61% for M5, M10, and M15, respectively,
compared to M0 at 20 Nm engine load and advanced injection
timing. In this study, CO2 emissions increased with the advanced
injection timing for all fuel mixtures. As shown parts a-c of
Figure 4 for M10, advanced injection timing increased the
change in CO2 by 5.79% and retarded injection timing dimin-
ished in CO2 by 7.55% compared to ORG and retarded injection
timing, respectively.
Combustion Analysis. Figures 7 and 8 show the cylinder
gas pressure, and Figures 10 and 11 demonstrate the rate of
heat release for different fuel blends and ORG injection timing
at 20 and 10 Nm loads, respectively. Figures 9 and 12 illustrate
the cylinder gas pressure and rate of heat release for M0 and
M15 at different injection timing and 20 Nm load, correspond-
Figure 4. Changes in the emissions relative to M0 at 15 Nm load.
3718 Energy & Fuels, Vol. 22, No. 6, 2008 Canakci et al.
11. ingly. The measured start of combustion and ignition delay for
each fuel is shown in Table 6 at different injection timing.
Peak Cylinder Gas Pressure. To analyze the cylinder gas
pressure, the pressure data of 100 cycles with a resolution 0.75
°CA was averaged. Figure 7 shows the cylinder gas pressure
with respect to the crank angle at 20 Nm load and ORG injection
timing. As seen in the figure, peak cylinder gas pressure slightly
decreased with the increase of the methanol supplement rate.
The peak cylinder pressure occurred at 7.96 (at 3.20 °CA
ATDC), 7.86 (at 3.28 °CA ATDC), 7.78 (at 3.32 °CA ATDC),
and 7.77 MPa (at 3.44 °CA ATDC) for M0, M5, M10, and
M15 at 20 Nm load and ORG injection timing, respectively.
Lowering the cetane number by methanol addition was respon-
sible for the increase in the ignition delay. The increase in the
ignition delay would burn more fuel in the premixed burning
phase. Because of this, the rate of pressure rise increased and
peak cylinder gas pressure diminished.21
When Figure 7 is compared to Figure 8, it is seen that the
cylinder gas pressure increased with an increasing engine load.
Experimental results show that the increase in the cylinder
Figure 5. Changes in the emissions relative to M0 at 20 Nm load.
Figure 6. Exhaust gas temperatures at different engine loads and ORG
injection timing.
DI Diesel Engine with Methanol-Diesel Fuel Blends Energy & Fuels, Vol. 22, No. 6, 2008 3719
12. pressure was approximately 5.45% for M10 when the engine
load increased from 10 to 20 Nm. The fuel consumption per
unit time was measured between 0.42 and 0.81 g/s at these test
conditions (10 and 20 Nm load at ORG injection timing for
M10, respectively). Because of increasing fuel consumption per
unit time with increasing engine load, this behavior provided
an increase in the maximum cylinder gas pressure. The locations
of maximum cylinder gas pressure approached TDC at 20 Nm
engine load because starting the fuel injection occurred earlier
than that of 10 Nm.
Figure 9 shows a comparison of the changes in the cylinder
gas pressures with respect to the crank angle obtained for M0
and M15 at different injection timings and 20 Nm load. The
peak cylinder gas pressure was obtained at 8.47 (at 1.68 °CA
ATDC), 7.96 (at 3.20 °CA ATDC), and 7.18 MPa (at 5.08 °CA
ATDC) for advanced ORG and retarded injection timings,
respectively. The trend was such that, as injection started earlier,
peak pressures became higher, which applied to all fuel blends.
Also, the peak pressures occurred earlier with advancing
injection timings.42,47
Rate of Heat Release (ROHR). As illustrated in the Figure
10, ROHR decreased with the increase of the methanol amount
in the fuel blend. The maximum ROHR was obtained at 31.13
(at 8.93 °CA BTDC), 29.86 (at 8.90 °CA BTDC), 27.33 (at
8.86 °CA BTDC), and 26.95 (at 8.85 °CA BTDC) kJ/deg for
M0, M5, M10, and M15 at 20 Nm load and ORG injection
timing, respectively. Because methanol does not evaporate as
easily as diesel fuel, the ignition delay increases with an
increasing methanol ratio, as seen in Table 6. The increase in
the ignition delay may cause more fuel to be burned in the
premixed burned phase and an increase in the ROHR. Con-
versely, the LHV of methanol is lower than diesel fuel, which
reduces ROHR. It is apparent from the figure that the LHV of
(47) Ozsezen, A. N.; Canakci, M.; Sayin, C. Energy Fuels 2008, 22,
1297–1305.
Figure 7. Cylinder gas pressure versus CA at 20 Nm load and ORG injection timing.
Figure 8. Cylinder gas pressure versus CA at 10 Nm load and ORG injection timing.
3720 Energy & Fuels, Vol. 22, No. 6, 2008 Canakci et al.
13. methanol was more significant than the heat of vaporization
with regard to ROHR.26
Figures 10 and 11 show the ROHR for different fuel blends
and ORG injection timing at 20 and 10 Nm loads, respectively.
As seen in the figures, ROHR increased with the rise in the
engine load because of the increase in the quantity of fuel
injected. Figure 12 demonstrates a comparison of the changes
in the cylinder gas pressures with respect to the crank angle
obtained for M0 and M15 at different injection timings and 20
Nm load. The ignition delay in a diesel engine is defined as the
time between the start of fuel injection and the start of
combustion. As seen in Table 6, the ignition delay was raised
with advanced injection timing for M0 and M15 because the
fuel was injected earlier in the combustion chamber. This led
to greater accumulation of the fuel in the ignition delay period
and an increasing premixed heat release. This is the reason of
increasing ROHR. Retarding injection timing led to a lower
accumulation of fuel and poor combustion. However, both
physical and chemical reactions must take place before a
significant fraction of chemical energy in the fuel is released.
These reactions need a finite time to occur. However, as ignition
delay proceeds, the in-cylinder pressures and temperatures
decrease and reduce the favorable conditions for ignition. The
most favorable timing for ignition lies in between these two
conditions.48
Conclusions
In this study, the influence of injection timing on the
combustion characteristics, engine performance, and exhaust
emissions of a diesel engine have been experimentally inves-
tigated using methanol-blended diesel fuel. The results indicated
(48) Kumar, M. S.; Ramesh, A.; Nagalinman, B. Biomass Bioenergy
2003, 25, 309–318.
(49) MERCK. Product Specification, Germany, 2006.
(50) TUPRAS. Product Specification, Turkey, 2005 (in Turkish).
(51) Lombardini. Engine Technical Specification, Turkey, 2000 (in
Turkish).
Figure 9. Cylinder gas pressure versus CA at different injection timing and 20 Nm load.
Figure 10. Rate of heat release versus CA at 20 Nm load and ORG injection timing.
DI Diesel Engine with Methanol-Diesel Fuel Blends Energy & Fuels, Vol. 22, No. 6, 2008 3721
14. that NOx emissions increased, smoke number and CO and
unburned HC emissions decreased by methanol addition, and
CO2 emissions increased because of the improved combustion.
Using methanol-blended diesel fuels, smoke number and CO
and unburned HC emissions reduced 7-32, 15-85, and
13-48%, while CO2 and NOx emissions increased 8-22 and
6-74%, respectively, depending upon the engine test conditions.
Increasing the amount of methanol in the fuel mixture produced
higher peak temperature in the cylinder. This effect increased
NOx emissions.
The results confirmed that increasing the ratio of methanol
in the fuel blend leads to a decrease in the BSFC, BSEC, and
BTE. This is probably of the result of the LHV of methanol,
which is distinctly lower than that of the diesel fuel. The peak
cylinder pressure and ROHR decreased with an increasing
methanol ratio in the blend. A lower cetane number with
methanol supplement leads to the increase in the ignition delay.
Increasing the ignition delay caused the deteriorating combus-
tion, and the peak cylinder pressure decreased. The increase in
the ignition delay increased the rate of pressure rise and reduced
the peak cylinder gas pressure.
In terms of injection timing, the test results demonstrated that,
with advancing the injection timing, smoke number and CO
and unburned HC emissions decreased while NOx and CO2
emissions increased. When the injection timing was advanced,
CO emission decreased because of the improving reaction
between fuel and oxygen. This caused an increase in the CO2
emissions. Advancing the injection timing caused an earlier start
of combustion relative to the TDC. Because of this, the cylinder
charge, being compressed as the piston moved to the TDC, had
relatively higher temperatures and, thus, lowered the smoke
number and UHC emissions and increased NOx emissions. With
the advancing injection timing, the best results were obtained
for the UHC and CO emissions at 20 Nm and the best results
were gained for the smoke number at 5 Nm load. At these
conditions, the smoke number and CO and UHC emissions were
found as 62%, 0.06%, and 12 ppm, for M15, respectively. On
the other hand, retarding the injection timing at 5 Nm load
Figure 11. Rate of heat release versus CA at 10 Nm load and ORG injection timing.
Figure 12. Rate of heat release versus CA at different injection timing and 20 Nm load.
3722 Energy & Fuels, Vol. 22, No. 6, 2008 Canakci et al.
15. presented the minimum results of NOx and CO2 emissions. At
these working conditions, NOx and CO2 emissions were
measured as 5 ppm and 5.45%, respectively.
The ORG injection timing gave the best results for BSFC,
BSEC, and BTE compared to the other injection timings. When
the injection timing was advanced, the ignition delay was longer
and this led to a reduction in engine output power. Thus, fuel
consumption per output power was increased. On the other hand,
retarded injection timing diminished the peak cylinder pressure
because more fuel was burned at near TDC, and this result
increased fuel consumption per output power. Also, advanced
injection timing boosted peak cylinder pressure and ROHR
because of the increase in ignition delay.
Acknowledgment. This study was supported by the grant from
the Scientific Research Project Commission of Marmara University
(Project BSE-075/131102).
Nomenclature
ATDC ) after top dead center
BTDC ) before top dead center
BTE ) brake thermal efficiency (%)
BSEC ) brake-specific energy consumption (MJ/kWh)
BSFC ) brake-specific fuel consumption (g/kWh)
CA ) crank angle
CI ) compression ignition
CO ) carbon monoxide
CO2 ) carbon dioxide
UHC ) unburned hydrocarbon
LHV ) lower heating value (kJ/kg)
(LHV)b ) lower heating value of a given component in the fuel
blend (kJ/kg)
NOx ) nitrogen oxides
ppm ) particulate per million
rpm ) revolution per minute
TDC ) top dead center
EF800398R
Table 6. Combustion Characteristics of the Fuels at Different
Injection Timing
M0 M5 M10 M15
20 °CA BTDC (ORG injection timing), 20 Nm
start of injection (°BTDC) 20 20 20 20
start of combustion (°BTDC) 9.27 8.92 8.34 8.01
ignition delay (°) 10.73 11.08 11.16 11.99
15 °CA BTDC, 20 Nm
start of injection (°BTDC) 15 15 15 15
start of combustion (°BTDC) 5.98 5.67 5.34 5.01
ignition delay (deg) 9.02 9.33 9.66 9.99
25 °CA BTDC, 20 Nm
start of injection (°BTDC) 25 25 25 25
start of combustion (°BTDC) 12.60 11.61 10.51 9.12
ignition delay (deg) 12.40 13.39 14.49 15.88
20 °CA BTDC (ORG Injection Timing), 10 Nm
start of injection (°BTDC) 20 20 20 20
start of combustion (°BTDC) 5.97 5.54 5.24 4.91
ignition delay (deg) 14.03 14.46 14.76 15.09
15 °CA BTDC, 10 Nm
start of injection (°BTDC) 15 15 15 15
start of combustion (°BTDC) 2.68 2.27 2.04 1.52
ignition delay (deg) 12.32 12.73 12.96 13.48
25 °CA BTDC, 10 Nm
start of injection (°BTDC) 25 25 25 25
start of combustion (°BTDC) 9.27 8.421 6.76 5.37
ignition delay (deg) 15.73 16.59 18.24 19.63
DI Diesel Engine with Methanol-Diesel Fuel Blends Energy & Fuels, Vol. 22, No. 6, 2008 3723
![Exhaust Emissions and Combustion Characteristics of a Direct
Injection (DI) Diesel Engine Fueled with Methanol-Diesel Fuel
Blends at Different Injection Timings
Mustafa Canakci,*,†,‡ Cenk Sayin,§ and Metin Gumus§
Department of Mechanical Education, Kocaeli UniVersity, Izmit 41380, Turkey, AlternatiVe Fuels Research
and DeVelopment Center, Kocaeli UniVersity, Izmit 41040, Turkey, and Department of Mechanical
Education, Marmara UniVersity, Istanbul 34722, Turkey
ReceiVed May 26, 2008. ReVised Manuscript ReceiVed August 20, 2008
In the recent years, environmental concerns and depletion in petroleum resources have forced researchers
to concentrate on finding renewable alternatives to conventional petroleum fuels. Therefore, alcohols as renewable
and alternative energy sources for the diesel engines gain importance. For this reason, in this study, the
performance, exhaust emissions, and combustion characteristics of a single cylinder diesel engine have been
experimentally investigated under different injection timings when methanol-blended diesel fuel was used
from 0 to 15%, with an increment of 5%. The tests were conducted at three different injection timings (15°,
20°, and 25 °CA BTDC) by changing the thickness of advance shim. All tests were conducted at four different
loads (5, 10, 15, and 20 Nm) at constant engine speed of 2200 rpm. The experimental test results showed that
BSFC, BSEC, combustion efficiency, and NOx and CO2 emissions increased as BTE, rate of heat release,
peak cylinder pressure, smoke number, and CO and UHC emissions decreased with an increasing amount of
methanol in the fuel blend. In comparison to the values at the original injection timing (20 °CA BTDC), the
values at the retarded injection timing (15 °CA BTDC) of peak cylinder pressure, rate of heat release, combustion
efficiency, and NOx and CO2 emissions decreased, while smoke number and UHC and CO emissions increased
at all test conditions. On the other hand, The advanced injection timing (25 °CA BTDC), smoke number, and
UHC and CO emissions diminished and peak cylinder pressure, rate of heat release, combustion efficiency,
and NOx and CO2 emissions increased at all test conditions. In terms of BSFC, BSEC, and BTE, retarded and
advanced injection timings gave negative results in all fuel blends compared to original injection timing.
Introduction
Because of their fuel economy and high reliability, compres-
sion-ignition (CI) engines known as diesel engines have been
penetrating a number of markets around the world. The existing
CI engines operate with conventional diesel fuel derived from
crude oil. It is well-known that the world petroleum resources
are limited and the production of crude oil is becoming more
difficult and expensive. On the other hand, because of the
growing concern over possible adverse health effects caused
by diesel emissions, the pollutions including unburned hydro-
carbons (UHC), carbon monoxide (CO), nitrogen oxides (NOx),
and particulate matter [quantified as smoke number (SN)] have
been regulated by laws in many developed countries. In the
last few years, many studies on the IC engines have been carried
out with the aim of reducing exhaust emissions by changing
operating parameters, such as valve timing, injection timing,
and atomization rate. At the same time, depletion of fossil fuels
and environmental considerations have led to investigations on
renewable fuels, such as methanol, ethanol, hydrogen, and
biodiesel.1-5
Alcohols (ethanol and methanol) have been considered as
alternative fuels for diesel engines. Methanol is manufactured
from any material that can be decomposed into CO (or CO2)
and hydrogen. In this regard, it may be produced from the
sources, which are independent from petroleum. The primary
feedstocks for methanol production are natural gas, lignite coal,
and renewable sources, such as wood, agricultural materials
biomass, waste biomass, and municipal wastes.6-8
Diesel fuel is a complex mixture of a large number of
hydrocarbons (such as C3-C25 hydrocarbons). For this reason,
its fuel properties can change depending upon the proportion
of hydrocarbon types used in the fuel mixture. However,
methanol (CH3OH) is a simple compound. It contains an oxygen
atom, so that it can be viewed as a partially oxidized
hydrocarbon fuel. It has a lower heating value than diesel fuel;
therefore, much more fuel is needed to obtain the same
performance as that of a diesel-fueled engine. Its high stoichio-
metric fuel/air ratio, high oxygen content, and high H/C ratio
may be beneficial for improving the combustion and reducing
* To whom correspondence should be addressed. Telephone: +90-262-
3032285. Fax: +90-262-3032203. E-mail: mustafacanakci@hotmail.com.
† Department of Mechanical Education, Kocaeli University.
‡ Alternative Fuels Research and Development Center, Kocaeli University.
§ Marmara University.
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Energy & Fuels 2008, 22, 3709–3723 3709
10.1021/ef800398r CCC: $40.75 2008 American Chemical Society
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