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View Article OnlineView Journal | View Issue

Effects of Jatrop

aCentre for Energy Sciences, Faculty of En

Kuala Lumpur, Malaysia. E-mail: yewheng

Tel: +603 79674448bSchool of Mechanical Engineering, Univers

14300 Nibong Tebal, Penang, Malaysia

Cite this: RSC Adv., 2014, 4, 50739

Received 10th August 2014Accepted 3rd October 2014

DOI: 10.1039/c4ra08464k

www.rsc.org/advances

This journal is © The Royal Society of C

ha biodiesel on the performance,emissions, and combustion of a convertedcommon-rail diesel engine

Y. H. Teoh,*ab H. H. Masjuki,a M. A. Kalam,a M. A. Amalinaa and H. G. Howa

An experimental investigation into the effects of Jatropha biodiesel fuels on the engine performance,

emissions, and combustion characteristics of a single-cylinder high-pressure common-rail diesel engine

was performed under six different load operations (0.1, 0.2, 0.3, 0.4, 0.5, and 0.6 MPa). The test fuels

included a conventional diesel fuel and three different blends of Jatropha biodiesel fuel (JB10, JB30, and

JB50). The results revealed that the biodiesel blended fuels had a significant influence on the brake

specific fuel consumption (BSFC) at all of the engine load conditions examined. In general, the use of

Jatropha biodiesel blends resulted in a reduction in brake specific nitrogen oxide (BSNOx), brake specific

carbon monoxide (BSCO), and smoke emissions, regardless of the load conditions. A large reduction of

20.2% in BSNOx emissions and 69.5% in smoke opacity were found for the engine when it was fuelled

with the biodiesel blends. In terms of the engine combustion characteristics, a slightly shorter ignition

delay (ID) and faster combustion duration were found to occur with the use of biodiesel blends under all

loading operations. It was revealed that the peak apparent heat release rate (AHRR) for biodiesel blends

is lower during low load operation; the AHRR was found to be comparable to that of baseline diesel

during high-load operation. Finally, the vibration results demonstrated that the largest reduction, 11.3%,

in the root mean square (RMS) of acceleration in comparison with the baseline diesel was obtained with

JB50 at an engine load of 0.5 MPa.

1. Introduction

Energy plays a vital role in our daily life. In recent years, rapidgrowth in population, development, and industrialization hasled to a high demand for energy worldwide. This energy ispredominately derived from non-renewable sources, such asfossil fuels and coal. Unfortunately, these resources are niteand are forecast to be diminished in less than a hundred years.1

Some predictions have stated that they will be depleted in lessthan 45 years.2 This scenario has triggered concern worldwideover energy security and has inevitably affected and pressuredmany countries in the world into seeking alternativeapproaches to satisfy ever rising energy demands. Aside fromthis energy issue, the world is also currently managing globalwarming and air pollution. The combustion of fossil fuels in thetransportation sector is the primary source of greenhouse gasand pollutant emissions. Consequently, it is clear that the worldis confronted with the twin crises of fossil fuel depletion andenvironmental degradation.

gineering, University of Malaya, 50603,

teoh@gmail.com; Fax: +603 79674448;

iti Sains Malaysia, Engineering Campus,

hemistry 2014

Recently, biodiesel has been considered as a major substi-tute for fossil diesel worldwide.3 Biodiesel can be dened as themonoalkyl esters of long-chain fatty acids that are derived fromchemical reactions (transesterication) of renewable feed-stocks, such as vegetable oil or animal fats, and alcohol with acatalyst. The full life-cycle analysis, including the cultivationand production of oil, and subsequent conversion to biodiesel,revealed that the net carbon dioxide (CO2) emissions are rela-tively low, and the use of biodiesel appears to have a signicantpositive impact on rural economic potential.4

Globally, there are more than 350 oil-bearing crops identiedas prospective sources for the production of biodiesel.5 Ingeneral, food crops, such as corn, sugar, and vegetable oil, havebeen the primary source of biodiesel fuels for transportation.These sources are considered to be the rst generation of bio-diesel feedstock because they were the rst crops to be used inthe production of biodiesel. However, the extensive use of thesefuels has caused food prices to rise, food price volatility, and anaccelerated expansion of agriculture in the tropics. In order toavoid possible negative consequences of this, solutionsincluding exploiting non-edible oils are being considered. Non-edible oil resources are gaining attention worldwide becausethey are not suitable for human consumption, eliminatingcompetition between fuel and food sources and reducing therate of deforestation. Hence, they are more environmentally

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friendly and economically competitive with edible oils. Bio-diesel production from non-edible crops is regarded as thesecond generation of biodiesel feedstocks. These feedstocksinclude Jatropha curcas, Calophyllum inophyllum, Ceiba pentan-dra, Karanja, Neem, Jojoba, and rubber seed, among others.6

1.1. Engine performance and exhaust emission for biodieselblends

There is considerable research focused on engine performanceand emission characteristics when using edible biodiesel, and itis only recently that greater efforts have been establishedinvestigating non-edible biodiesel fuel. Chauhan et al. per-formed a study that investigated the effects of biodieselproduced from non-edible oil on a diesel engine equipped witha mechanical pump-line-nozzle fuel injection system.7 Theauthors investigated the effect of various Jatropha biodieselblends (5, 10, 20, and 30%) on engine performance and theexhaust emissions using an unmodied diesel engine, butwithout performing supporting combustion results. Theexperimental results indicated that the engine performancewith Jatropha biodiesel and its blends were comparable to theperformance of the engine when using diesel fuel. The brakethermal efficiency (BTE), HC, CO, CO2, and smoke emissionswere found to be lower, while the BSFC, BSEC, and NOx emis-sions were higher when using Jatropha biodiesel blends incomparison with diesel. The authors suggested that biodieselderived from non-edible oil, like Jatropha and its blends, couldbe used in a conventional diesel engine without anymodication.

An alternative potential non-edible biodiesel, investigated byOng et al.,8 is produced from the plant Calophyllum inophyllum.The authors investigated the performance and emissions of thisbiodiesel in a single-cylinder, mechanical pump-line-nozzledirect fuel injection diesel engine. The test results revealedthat the use of B10 biodiesel yields an improvement in engineperformance, with higher BTE and lower BSFC, in comparisonwith diesel fuel. Emissions, such as CO and smoke opacity, werereduced when using B10 but NOx emission were found to besomewhat increased when using B10 at all engine speeds. Theused of Kapok (Ceiba pentandra) oil methyl ester as the source ofnon-edible biodiesel in diesel engines was explored by Vedharajet al.9 They conducted an experiment in a single-cylinder, four-stroke, water-cooled diesel engine that was equipped with amechanical pump-nozzle injection system. The investigationrevealed that the BTE of the engine when using B25 was supe-rior to conventional diesel by 4%. The combustion and emis-sions of HC, CO, NOx, and smoke for B25 were all comparablewith the results produced when using diesel. In a performanceand emissions test on a single-cylinder, four-stroke, water-cooled Ricardo E6 engine using Mahua (Madhuca indica) bio-diesel, Raheman and Ghadge10 affirmed a reduction in smokeand CO emissions, with an increase in NOx emissions when thepercentage of Mahua biodiesel blends was increased. The meanBSFC was reported to be 4.3–41.4% higher for Mahua biodieselblends. The mean BTE for B100 was 10.1% lower than dieselfuel at full load condition.

50740 | RSC Adv., 2014, 4, 50739–50751

Biodiesel fuel in engines equipped with a high-pressurecommon-rail direct injection (CRDI) system has recentlybecome an interesting research topic. This is predominantlyowing to the fact that a majority of the diesel engines in usetoday employ this technology. An et al.11 experimentally inves-tigated pure waste cooking oil biodiesel and blends of 10, 20and 50% in a four-stroke, four-cylinder, water-cooled, high-pressure CRDI diesel engine under various loads. As would beexpected, the BSFC of the biodiesel was reported to be higher incomparison with diesel fuel. In fact, the largest increase in BSFC(28.1%) was found at a 10% load. It was observed that thecylinder pressure decreased slightly with the use of biodiesel atall engine loads. The CO emissions were found to be increasedwhen using a higher percentage biodiesel blend ratio and adecreased engine speed. However, an opposite trend wasobserved at higher engine loads. Tan et al.12 investigated theused of biodiesel in a multi-cylinder, four-stroke, diesel enginethat was equipped with a high-pressure common-rail fuelinjection system. In this study, the authors examined theimpact of biodiesel fuel solely on regulated and unregulatedemissions. In comparison to the baseline diesel fuel, the addi-tion of biodiesel demonstrated no obvious difference in NOx

emission, a reduction in HC, smoke, acetaldehyde and toluene,and an increase in CO, formaldehyde and acetone.

1.2. Purpose of study

Among the topics discussed above, many previous studies intobiodiesel fuel have examined the engine-out responses of aconventional mechanical pump-line-nozzle fuel injectionsystem. With this fuel injection system, the lower compress-ibility and the viscosity of the biodiesel will usually lead to anadvanced start of injection, resulting in higher NOx emissions.13

These effects could be eliminated by using common-rail fuelinjection technology, in which fuel pressurization is indepen-dent of injection timing.14 Aside from the problem of fuelinjection technology, there is an instability and uctuation inthe price of crude palm oil (edible oil) in Malaysia. To minimizedependency on consuming biodiesel fuel that is primarilysourced from crude palm oil, the biodiesel policy of theMalaysian government recommended the utilization of non-edible oils for the production of biodiesel. Jatropha curcas oilis one of the major non-edible, tree-borne feedstocks used forthe large scale production of biodiesel in Malaysia and southeast Asia; this is because it is well adapted to local climaticconditions and is available in surplus quantities across theregion.15,16 Consequently, there is strong motivation to investi-gate the impact of Jatropha biodiesel blends in an engineequipped with a high-pressure common-rail injection system,and to analyse the effects of this biodiesel on engine perfor-mance, emissions, combustion, and vibration characteristics.This experiment was performed under six different brake meaneffective pressure (BMEP) (0.1, 0.2, 0.3, 0.4, 0.5, and 0.6 MPa)and at a rated engine speed of 1600 rpm. Parameters includingBSFC, BTE, BSCO, BSNOx, EGT, smoke opacity, peak pressure,peak of heat release, and vibration analysis were investigatedand evaluated.

This journal is © The Royal Society of Chemistry 2014

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2. Material and methods2.1. Test fuels and operating conditions

In the present study, fossil diesel fuel was obtained incommercial form and Jatropha oil was obtained from Indonesia.Biodiesel production was conducted via the acid-estericationand alkali-transesterication process. Table 1 contains adescription of the key physicochemical properties of neatJatropha methyl ester (JME) in comparison with ASTM and ENstandards. It can be observed that the physicochemical prop-erties of the produced biodiesel was comprehensively measuredand benchmarked against the biodiesel standards based onASTM D6751 and EN14214. It appears that all physicochemicalproperties of JME are sufficient to meet the ASTM and EN bio-diesel standards.

In this study, a total of three different methyl ester blends,JB10 (10% biodiesel, 90% petroleum diesel), JB30 (30% bio-diesel, 70% petroleum diesel), and JB50 (50% biodiesel, 50%petroleum diesel) were prepared and tested. The key physico-chemical properties of the diesel and biodiesel blends are listedin Table 2. As can be observed, all of the physicochemicalproperties of the biodiesel blends satisfy the ASTM D7467 bio-diesel blend standards. Additionally, the experimental resultssuggested that blending with petroleum diesel substantially

Table 1 The fuel properties of neat JME

Properties Unit Limit (ASTM

Kinematic viscosity @ 40 �C mm2 s�1 1.9–6.0Density @ 15 �C kg m�3 880Acid number mg KOH g�1 0.5 max.Caloric value MJ kg�1 —Flash point �C 130 min.Pour point �C —Cloud point �C ReportCold lter plugging point �C —Oxidation stability @ 100 �C Hours 3 min.Cetane number — 47 min.Carbon wt% 77Hydrogen wt% 12Oxygen wt% 11

Table 2 The fuel properties of diesel fuel and biodiesel blends

Properties Unit Diesel fuel

Bio

Lim

Kinematic viscosity @ 40 �C mm2 s�1 3.34 1.9Density @ 15 �C kg m�3 851.9 858Acid number mg KOH g�1 0.12 0.3Caloric value MJ kg�1 45.31 35Flash point �C 71.5 52Pour point �C 1 NoCloud point �C 8 NoOxidation stability @ 100 �C Hours >40 6Cetane number — 52 47

This journal is © The Royal Society of Chemistry 2014

improved the nal biodiesel blend properties. In particularly,the kinematic viscosity of the biodiesel blends was reduced asthe proportion of petroleum diesel was increased in the blends.In additional, the resultant ash points for all the biodieselblends were relatively higher in comparison with petroleumdiesel and were suitable for use as a transportation fuel.However, the caloric value of all of the biodiesel blends waslower than that of petroleum diesel. Another key property thatsignicantly inuences engine performance, emissions, andcombustion characteristics is the cetane number of fuel. Ingeneral, it can be observed that all biodiesel blends have ahigher cetane number than petroleum diesel fuels.

The experiment was conducted under a constant speed of1600 rpm and with varying BMEP (i.e. 0.1, 0.2, 0.3, 0.4, 0.5, and0.6 MPa). These six test points were selected as the mostrepresentative of a wide variety of engine load ranges. Initially,diesel fuel was used as the baseline fuel for the basis ofcomparison. Following this, mixtures of diesel and methyl esterwith 10, 30, and 50% volumetric proportions were tested.Consequently, a total of 24 runs experimental conditions,including baseline diesel were tested in this study. When theengine was fuelled with methyl ester blended fuels, the engineran satisfactorily throughout the entire test, which was per-formed at room temperature, and had no starting difficulties.

D6751) Limit (EN 14214) JME Test method

3.5–5.0 4.42 D445860–900 882.7 D1270.5 max. 0.37 D66435 39.98 D240120 min. 178.5 D93— 4 D2500— 5 D2500— 1 D63716 min. 6.5 EN1411251 min. 58 D6890— 76.8 D5291— 11.8 D5291— 10.9 D5291

diesel blends

JB10 JB30 JB50it (ASTM D7467) Test method

–4.1 D445 3.62 3.73 3.9max. D127 858.2 864.2 871.3max. D664 0.17 0.23 0.27

D240 44.84 43.66 42.57D93 87.5 96.5 103.5

t specied D2500 0 0 0t specied D2500 8 7 6

EN14112 18.6 12.8 10.8min. D6890 53 54 56

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The tests were performed under steady-state conditions with asufficiently warmed exhaust gas and water coolant temperature.To enhance the accuracy of the study, each test point wasrepeated twice to produce average readings. The repeatabilitywas matched over 95% for each test. This indicates that theeffects on emissions and combustion characteristics can bereliably analysed from this test system.

Fig. 1 Cross section view of the re-entrant type combustion chamber.

2.2. Experimental system

2.2.1. Test engine. The engine used in this study was basedon a single-cylinder, water-cooled, four-stroke, naturally aspi-rated direct injection diesel engine (YANMAR TF-120E). Origi-nally, the engine was equipped with a regular fuel injectionsystem consisting of a mechanical type of pressure fuel pump(200 bar) timed by the camsha, and a mechanical fuel injectorwith an injection angle of 150� and four 0.26 mm diameterholes. The engine is naturally aspirated with amaximum outputof 7.5 kW. The original fuel injection timing is constant and setto 17� BTDC. The engine was originally equipped with amechanical governor to control the engine speed. Thecombustion chamber is u shaped with a diameter of 50.5 mmand a depth of 18 mm. The specications of the test engine arelisted in Table 3, and the cross section view of the combustionchamber with key dimensions is illustrated in Fig. 1.

2.2.2. Engine modications and fuel delivery system. Theoriginal injection system of the engine was disassembled and anew common-rail injection system was retrotted as illustratedin Fig. 2. The system was based on commercially availablecommon-rail diesel engine components. A second-generation,electronically controlled common-rail high-pressure injectionsystem was installed to replace the original mechanical typepump-line-nozzle injection system. The fuel pump was exter-nally driven using an electric motor running at 750 rpm to

Table 3 Characteristics of single-cylinder engine

Parameter Units

Displacement 638 cm3

Bore 92 mmStroke 96 mmCompression ratio 17.7 : 1Rated power 7.8 kWRated speed 2400 rpmD/Hbowl 2.81Combustion chamber Re-entrant type

Original fuel injection systemFuel injection type Mechanical cam driven

injectionNumber of injector nozzle holes 4Nominal injector nozzle diameter 0.26 mm

Retrotted fuel injection systemFuel injection type Electronically common-rail

injectionNumber of injector nozzle holes 5Nominal injector nozzle diameter 0.134 mm

Fig. 2 The retrofitted engine setup.

50742 | RSC Adv., 2014, 4, 50739–50751

maintain the required high-pressure levels in the fuel rail and toensure a stable line pressure with minimum uctuation. Thisinjection system enabled variable injection pressure and injec-tion timing. Independent control of the injection parameterswas achieved using a custom-built electronic control unit (ECU).The measured nozzle diameter was found to be approximately0.134 mm, with ve evenly spaced nozzle holes.

2.3. Test bed conguration and instrumentation

The ST-7.5 model 7.5 kW A.C. synchronous dynamometer wasused to provide loading to the engine and to maintain theengine speed. An airow metre turbine with 2 to 70 litres persecond (L s�1) measuring range was employed to measure theintake airow rate. A type K thermocouple was used to monitorthe exhaust gas temperature. The fuel ow rate was measured

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with a positive displacement gear wheel ow metre (model:DOM-A05H), which interfaced with a ow rate totalizer (model:ZOD-Z3). The test system was equipped with the necessarysensors for the analysis of combustion and fuel injectiontiming. In-cylinder gas pressure was measured using a Kistler6125B type pressure sensor. To mount the sensor in the enginehead through a water cooling jacket, a dedicated mountingsleeve was fabricated and installed, as can be observed on theright-hand side of Fig. 1. The charge signal output of the pres-sure sensor was converted to a low-impedance voltage signalusing a PCB model 422E53 in-line charge converter; this unitwas powered using a PCB model 480E09 signal conditioner. Toacquire the top dead centre (TDC) position and crank anglesignal for every engine rotation, a high-precision incrementalencoder with 720 pulses per revolution was used. To determineand verify the SOI timing, the injector current signal wasmeasured with a hall effect current sensor. To simultaneouslysample the cylinder pressure, injector current signal, andencoder signals, a computer equipped with a high-speedADLINK DAQ-2010 simultaneous sampling data acquisitioncard, which has 14 bits resolution, 2 MS per s�1 sampling rate,and four analog input channels, was used. The acquired datawere further processed and analysed with Matlab soware. Toeliminate cycle-to-cycle variation in each test, 100 consecutivecombustion cycles (or equal to 200 cranksha revolutions) ofpressure data were collected and an averaged was calculated. Toreduce noise effects, smooths data using SPAN as the number ofpoints used to compute each element was applied to thesampled cylinder pressure data. Combustion parameters, suchas peak pressure magnitude, peak pressure location, heatrelease rate, peak heat release rate location, and ID, were allcomputed using Matlab soware. For the exhaust emissionmeasurement, an AVL DICOM 4000 5-gas analyser was used tomeasure the concentrations of CO, CO2, and NOx. Opacity ofsmoke was measured using AVL DiSmoke 4000. All emissionswere measured during steady-state engine operation. Themeasurement range and resolution of both of the instrumentsare provided in Table 4. The experimental set-up is illustrated inFig. 3. The CO and NOx emissions were converted into brakespecic emissions by using the following equations accordingto SAE J177:

BSCO(g kW�1 h�1) ¼ 0.0580 � CO(ppm)

� exhaust mass flow rate(kg min�1)/brake power(kW) (1)

Table 4 Measuring components, ranges and resolution of the AVL DICO

Equipment Measurement principle Compo

Gas analyzer Non-dispersive infrared CarbonNon-dispersive infrared CarbonElectrochemical NitrogeCalculation Excess

Smoke opacimeter Photodiode detector Opacity

This journal is © The Royal Society of Chemistry 2014

BSNOx(g kW�1 h�1) ¼ 0.0952 � NOx(ppm)

� exhaust mass flow rate(kg min�1)/brake power(kW) (2)

To perform engine vibration measurements, an accelerom-eter (PCB model 603C01) with calibrated sensitivity to a95 mV g�1 and 50 g measurement range was used. This ruggedaccelerometer is capable of performing over a wide frequencyrange of 0.5–10 000 Hz. Engine vibration motion in the lateral(y) axis (or perpendicular to cylinder axis) was chosen formonitoring vibrations. To sense the magnitude of vibration inthis direction, the accelerometer was mounted on the enginebody with an adhesive mounting base. The output signal fromthe sensor was connected to a constant current single channelsignal conditioner (PCB model 480C02) with unity gain. In eachtest, engine-block vibration signals for a total of 100 consecutivecombustion cycles at 0.125� CA resolution were recorded andthe averaged RMS was calculated.

3. Calculation methods3.1. Engine performance

The engine performance in this work was evaluated based onBSFC and BTE. The BSFC and BTE were determined andcalculated according to the following equations:

BSFCðg kW�1 h�1Þ ¼fuel consumption

brake power(3)

BTEð%Þ ¼ brake power� 100

calorific value� fuel consumption(4)

3.2. Combustion analysis

Heat release rate (HRR) analysis is a useful approach to assessthe effects of fuel injection system, fuel type, engine designchanges, and engine operating conditions on the combustionprocess and engine performance.17 Given the plot of AHRRversus crank angle, it is easy to identify the start of combustion(SOC) timing, the fraction of fuel burned in the premixed mode,and differences in combustion rates of fuels.18 In the presentpaper, fuels with different types of methyl ester and blend ratioswere fuelled in an identical compression ignition engine;hence, the AHRR information is an important parameter ininterpreting engine performance and exhaust emissions. In thisstudy, the averaged in-cylinder pressure data of 100 successive

M 4000 gas analyzer and DiSmoke 4000 smoke analyzer

nent Measurement range Resolution

monoxide (CO) 0–10 vol% 0.01 vol%dioxide (CO2) 0–20 vol% 0.1 vol%n oxides (NOx) 0–5000 ppm 1 ppmair ratio (l) 0–9999 0.001(%) 0–100% 0.10%

RSC Adv., 2014, 4, 50739–50751 | 50743

Fig. 3 Schematic diagram of the experiment setup.

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cycles, acquired with a 0.125� crank angle resolution, were used

to compute the AHRR. The AHRR, given bydQdq

, at each crank

angle was obtained from the rst law of thermodynamics, and itcan be calculated by the following formula:

dQ

dq¼ g

g� 1P

dV

dqþ 1

g� 1V

dP

dq(5)

where, g ¼ specic heat ratio, P ¼ instantaneous cylinderpressure (Pa), and V ¼ instantaneous cylinder volume (m3).

3.3. Vibration analysis

In the present study, the engine vibrations of different types ofbiodiesel blends were compared with the baseline diesel engineat different engine loads. The averaged RMS of the accelerationsignal was determined using the following equation:

arms ¼ 1

n

Xn

j¼1

0@

ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiXNi¼1

ai=N

vuut1A

j

264

375 (6)

3.4. Statistical and equipment uncertainty analysis

In any experiment, errors and uncertainties can arise frominstrument selection, condition, calibration, environment,observation, reading, and test procedure. The measurementrange, accuracy, and percentage uncertainties associated withthe instruments used in this experiment are listed in Table 5.Uncertainty analysis is necessary to verify the accuracy of the

50744 | RSC Adv., 2014, 4, 50739–50751

experiments. Percentage uncertainties of various parameters,such as BSFC, BTE, BSCO, and BSNOx were determined usingthe percentage uncertainties of various instruments employedin the experiment. To compute the overall percentage uncer-tainty due to the combined effect of the uncertainties of variousvariables, the principle of propagation of errors is consideredand can be estimated as �4.3%. The overall experimentaluncertainty was computed as follows:

Overall experimental uncertainty¼ square root of [(uncertainty of

fuel flow rate)2 + (uncertainty of BSFC)2 + (uncertainty of BTE)2

+ (uncertainty of BSCO)2 + (uncertainty of BSNOx)2 + (uncer-

tainty of EGT)2 + (uncertainty of smoke)2 + (uncertainty of

pressure sensor)2 + (uncertainty of crank angle encoder)2] ¼square root of [(2)2 + (1.95)2 + (1.74)2 + (2.22)2 + (0.73)2 + (0.15)2

+ (1)2 + (1)2 + (0.03)2] ¼ �4.3%.

4. Results and discussions4.1. Engine performance characteristics

4.1.1. Brake specic fuel consumption. Fig. 4 illustrates theBSFC of the fuel samples tested with respect to various BMEPs.BSFC is dened as the ratio of the fuel consumption rate to thebrake power output. From the results, it was observed that, atthe BMEP of 0.6 MPa, baseline diesel had the lowest BSFC of270.9 g kW�1 h�1, followed by 280.0 g kW�1 h�1, 287.6 g kW�1

h�1, and 291.2 g kW�1 h�1 for the JB10, JB30, and JB50 blends,respectively. The higher BSFC of JB50 means that more fuel was

This journal is © The Royal Society of Chemistry 2014

Table 5 List of measurement accuracy and percentage uncertainties

Measurement Measurement range Accuracy Measurement techniques % Uncertainty

Load �120 N m �0.1 N m Strain gauge type load cell �1Speed 60–10 000 rpm �1 rpm Magnetic pick up type �0.1Time — �0.1 s — �0.2Fuel ow measurement 0.5–36 L h�1 �0.01 L h�1 Positive displacement gear wheel ow meter �2Air ow measurement 2–70 L s�1 �0.04 L s�1 Turbine ow meter �0.5CO 0–10% by vol �0.001% Non-dispersive infrared �1NOx 0–5000 ppm �1 ppm Electrochemical �1.3Smoke 0–100% �0.1% Photodiode detector �1EGT sensor 0–1200 �C �0.3 �C Type K thermocouple �0.15Pressure sensor 0–25 000 kPa �12.5 kPa Piezoelectric crystal type �1Crank angle encoder 0–12 000 rpm �0.125� Incremental optical encoder �0.03

ComputedBSFC — �7.8 g kW�1 h�1 — �1.95BTE — �0.5% — �1.74BSCO — �0.1 g kW�1 h�1 — �2.22BSNOx — �0.1 g kW�1 h�1 — �0.73

Fig. 4 BSFC with the Jatropha biodiesel blends compared with dieselfuel at various BMEP. Fig. 5 BTE with the Jatropha biodiesel blends compared with diesel

fuel at various BMEP.

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consumed to develop the same amount of power. This wasexpected because of the relatively low caloric value of JB50 incomparison with diesel, which was approximately 6% less thanthat of diesel fuel. These results are in agreement with thosereported by Kivevele et al.19

4.1.2. Brake thermal efficiency. Engine BTE is a product oftwo important efficiencies, namely, the mechanical efficiencyand the net indicated thermal efficiency. Alternatively, it can becalculated by dividing the brake power output by the totalenergy input delivered to the system. Owing to the effect ofvarious loss mechanisms, such as combustion inefficiency,exhaust blow down, heat transfer, ow, and mechanical fric-tion, the BTE of a real operating diesel cycle is usually under50%, and is oen far lower.20 Of these loss mechanisms, themagnitude of heat transfer losses constitutes a major fraction,and it varies with the mean piston speed as well as thecombustion characteristics of the fuel. The variation of BTEversus BMEP for various test fuels is recorded in Fig. 5. Ingeneral, it was observed that BTE for all tested fuels increasedwith an increase in BMEP. This is attributed to the twin effects

This journal is © The Royal Society of Chemistry 2014

of increased brake power and reduced wall heat loss at higherengine loads.7,21 Additionally, it was observed that with theaddition of JME in the blend, the BTE is slightly reduced and islower than for diesel fuel, except for JB50 at BMEP of 0.1, 0.4,and 0.5 MPa. Conversely, there is a marginal improvement of2% in BTE with JB50 at 0.5 MPa. This may be attributed to thefact that, at higher concentrations of methyl ester blend, theearly initiation of combustion and increase in peak pressuresresults in higher BTE.

4.2. Exhaust emissions characteristics

During combustion, CO emissions appear when the availableoxygen is insufficient to fully oxidize all of the carbon in the fuelto carbon dioxide. The use of oxygenated fuel, such as methylester, would be expected to improve the combustion quality,especially in fuel-rich regions, consequently reducing COemissions. The variation in BSCO emissions of the engine withdifferent engine loads and fuel types is illustrated in Fig. 6. Theresults suggest that the magnitude of BSCO emissions was

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Fig. 6 BSCOwith the Jatropha biodiesel blends compared with dieselfuel at various BMEP.

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signicantly governed by the engine load setting and biodieselblending ratio. High fuel-borne oxygen content in biodiesel fuelplays a key role in reducing CO emissions. In general, it isobserved that the reductions in BSCO emissions were obtainedwith the use of methyl ester in the blend. In fact, it consistentlydecreased with the increase in the biodiesel blending ratio. Thiswas mainly due to the oxygen content in biodiesel promotingmore complete combustion in the engine.4,22 In comparisonwith the diesel fuel, the BSCO emissions at a high engine load of0.6 MPa decreased by 11.9, 30 and 49.9% for JB10, JB30, andJB50, respectively. Another observation was that the BSCOemissions at low load condition were generally higher than athigh load conditions, regardless of the fuel used. This waslargely owing to the better air-fuel mixing process, as a result ofthe higher fuel injection pressure at higher engine loads,resulting from the use of the converted common-rail system,and consequently, decreasing the BSCO emissions. An alterna-tive explanation is that this occurred owing to the fact that theair-fuel ratio is too lean for complete combustion at low loadconditions, leading to the higher BSCO emissions.23 The relativeair-fuel ratio (l) for various engine loads and fuel types isillustrated in Fig. 7. In short, l is dened as the ratio of theactual air-fuel ratio to that of the stoichiometric air-fuel ratiorequired to completely burn the fuel delivered. The air-fuel ratio

Fig. 7 Relative air-fuel ratio with the Jatropha biodiesel blendscompared with diesel fuel at various BMEP.

50746 | RSC Adv., 2014, 4, 50739–50751

of the mixture affects the combustion phenomenon and thecompleteness of combustion, especially at the fuel lean zone. Infact, the general trend indicates that the variations in BSCOemissions were very similar to the variation in l values. Inaddition, it appears that even under high load conditions (i.e.,0.6 MPa of engine load), the l values were still above unity,indicating a lean combustion process. Additionally, the addi-tion of methyl ester in the blend creates a slightly richcombustion process. Consequently, the ID becomes shorter,combustion duration increases, and combustion getscompleted properly, leading to a further decrease in COemissions.

As discussed earlier, fuel injection technology in biodieselengine has signicant effects on NOx emissions. Unlike theconventional mechanical pump-line-nozzle injection system,the modied common-rail injection system in the present studyeliminated the common issue of advanced injection timingowing to the relatively higher viscosity of biodiesel. Hence, thecorrelation of other effects on the variation of NOx emissionswhen using biodiesel can be analysed more comprehensibly. Inautomotive exhaust emissions, the formation of NOx dependson the fuel type, fuel properties, and engine operating condi-tions.24 In the literature, most researchers have reported anincreases of NOx emissions with the use of methyl ester blendedfuel.12,25,26 The explanations given are primarily based on thehigher oxygen content, which results in a higher combustiontemperature that promotes a thermal NOx formation pathway.However, some researchers have reported the opposite trend,with lower NOx emissions when using methyl ester blendedfuel.27,28 This is in good agreement with the results obtainedthroughout this study. As illustrated in Fig. 8, the presence ofJME decreased the BSNOx relative to baseline diesel, but it didnot decrease further as the degree of JME blending increased.The largest recorded reduction in BSNOx was approximately20.2% for the JB30 blend at BMEP of 0.2 MPa. This can beattributed to the relatively lower caloric value of the methylester fuels being used and, consequently, reduced HRR in thepremix combustion region and lower peak combustiontemperature.29 Additionally, the results also suggest that furtherincreases in the methyl ester concentration to JB50 resulted in

Fig. 8 Variations in BSNOx emissions with different engine loads andfuel types.

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an increase in BSNOx across all engine loads. A similar resulthas been observed by Mueller et al.30 who reported that thehigher cetane number of biodiesel blends relative to diesel (seeTable 2) causes ignition to occur earlier in the cycle. This allowsthe combustion products to have a longer residence time athigh temperatures, which increases NOx emissions. Anotherpossible reason may be associated with the reduction in theheat dissipation by radiation as a consequence of the largereductions of soot emitted with the use of biodiesel (see Fig. 10),resulting in an increase in BSNOx emissions.31

Owing to the lean operation and higher expansion ratio ofthe diesel engine, the EGT is typically lower than a petrolengine. A higher EGT is unfavourable as this will deteriorateengine fuel economy by discharging some of the useful energyinto waste exhaust thermal energy, as well as causing thermaldamage to piston components. As illustrated in Fig. 9, ingeneral, the EGT increased with an increase in the BMEP for allof the fuels tested in this study. Additionally, the presence ofJME increased the EGT relative to that for baseline diesel, but itdid not increase further as the degree of JME blending rose toJB50. The highest achievable EGT for JB30, JB10, and baselinediesel at a BMEP of 0.6 MPa were 494.2 �C, 489.6 �C, and 487 �C,respectively. Additionally, the EGT is lower for higher blends of

Fig. 10 Smoke opacity value with the Jatropha biodiesel blendscompared with diesel fuel at various BMEP.

Fig. 9 EGT with the Jatropha biodiesel blends compared with dieselfuel at various BMEP.

This journal is © The Royal Society of Chemistry 2014

JB50 because of the improved combustion provided by thebiodiesel under all engine loading conditions. In fact, manyresearchers have also reported that the EGT is lower with theengine fuelled with biodiesel blended fuel compared to thebaseline diesel.11,32–34 In general, this phenomenon is causedprimarily by the lower caloric value and the existence ofchemically bound oxygen in biodiesel blends, which reducesthe total energy released and improves the combustion,respectively. The EGT was thereaer decreased.

Smoke is an unwanted by-product of combustion incompression ignition diesel engines, which is primarily formedthrough incomplete combustion of hydrocarbon fuel. Ingeneral, the smoke from the exhaust tailpipe is emitted visiblyin the form of dark black smoke. The composition of smokehighly depends on the type of fuel, engine operating conditions,and carbon residue of the fuel. The emission of smoke opacity isdemonstrated in Fig. 10 for different Jatropha biodiesel fuelblends. In relation to the effect of biodiesel content on thesmoke opacity, it was observed that the smoke opacity generallytended to decrease as the blending ratio of biodiesel in the fuelblend increased. It was observed that the maximum reductionwas 69.5% with JB50 at a BMEP of 0.5 MPa, while it was a 21.9%reduction with JB10 at a BMEP of 0.1 MPa when both werecompared with their corresponding baseline diesel. The

Fig. 11 In-cylinder pressure and AHRR versus crank angle for testedfuels at a BMEP of (a) 0.1 MPa and (b) 0.6 MPa.

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combined effects of lower impurities, higher oxygen, and lowersulphur content of methyl ester fuels are believed to beresponsible for the decreased smoke opacity level.21

4.3. Combustion characteristics

To evaluate the effect of the biodiesel blending ratio on thecombustion characteristics, the cylinder pressures for 100consecutive combustion cycles were recorded and compared atvarious loads and at a constant engine speed of 1600 rpm. Thein-cylinder pressure, AHRR and injector current traces under aBMEP of 0.1 MPa (low load) and 0.6 MPa (high load) are illus-trated in Fig. 11. As can be observed, the variation of engine loadand biodiesel blending ratio had the greatest effect on thecombustion characteristics. Under low load conditions, thepremixed combustion process dominated. Conversely, thediffusion ame combustion process dominated at high loadconditions. In the case of lower engine load, the cylinder pres-sure proles for all of the tested fuels were comparable with thebaseline diesel. Additionally, the shi in the SOC timing wasconsistent with the change of biodiesel concentration in theblend. In fact, the combustion event was shied earlier towardTDC as the biodiesel concentration was increased. A smallreduction in peak pressure in the range of 0.6–1.2 bar wasobserved for the operation with the biodiesel blend fuels. Thiscan be attributed to the marginal decrease in the AHRR duringthe premixed combustion phase. In the case of the higherengine load, more signicant variations in terms of peak

Table 6 Crank angle position corresponding to certain percent mass fr

BMEP (MPa) Fuel typeStart of injection,SOI (�ATDC)

Start of combustionSOC (�ATDC)

0.1 D �7.000 �0.250JB10 �7.000 �0.250JB30 �7.000 �0.375JB50 �7.000 �0.375

0.2 D �7.000 �0.500JB10 �7.000 �0.500JB30 �7.000 �0.500JB50 �7.000 �0.625

0.3 D �7.000 �0.750JB10 �7.000 �0.750JB30 �7.000 �0.750JB50 �7.000 �0.875

0.4 D �7.000 �0.875JB10 �7.000 �0.875JB30 �7.000 �0.875JB50 �7.000 �1.000

0.5 D �7.000 �1.000JB10 �7.000 �1.000JB30 �7.000 �1.000JB50 �7.000 �1.125

0.6 D �7.000 �1.000JB10 �7.000 �1.125JB30 �7.000 �1.250JB50 �7.000 �1.625

50748 | RSC Adv., 2014, 4, 50739–50751

pressure were observed among all biodiesel blends and baselinediesel fuel. It was observed that JB50 achieved the highest peakpressure of 74.4 bar followed by JB30 (73.6 bar), JB10 (72.1 bar),and diesel (71.6 bar). The results suggest that adding biodieselin the blend caused increases in the peak pressure and shiedthe location of occurrence earlier toward the TDC point. Thiscan be attributed to the prominent advance in SOC timing,which caused the earlier rise of the AHRR and thus increasedthe in-cylinder gas pressure.

Another interesting observation that can be made from theAHRR diagram is the variations in ID. Mathematically, ID isdened as the crank angle interval measured from the start offuel injection timing to the start of combustion timing; this istypically determined from the fuel injector signal and AHRRdata, respectively. As summarized in Table 6, it was found that,in general, regardless of the engine load, most of the biodieselblends exhibited shorter ID than baseline diesel owing to theirrelatively higher cetane number. Similar trends of shorter IDswith biodiesel blends were also reported by Ozsezen et al.35 Asillustrated in Fig. 12, lines indicating the mass fraction burnedof 10% (CA10), 50% (CA50), and 90% (CA90) were marked.Empirically, 10% and 90% lines marked the start and end of themain combustion duration, respectively. The period betweenCA10 and CA90 was dened as the combustion duration andthis is typically measured in the unit of crank angle. From theresults presented in Table 6, it can be seen that the generaltrend that indicates a shorter combustion duration was

action burned for all tested fuels under various BMEPs

,ID (�CA)

Crank angle for certainpercent mass fractionburned (�ATDC)

Combustion duration(�CA)10% 50% 90%

6.750 3.000 4.875 17.375 14.3756.750 3.000 4.875 17.375 14.3756.625 2.750 4.625 17.125 14.3756.625 2.375 4.375 16.625 14.2506.500 2.875 6.000 20.000 17.1256.500 2.625 6.000 19.750 17.1256.500 2.625 6.000 19.500 16.8756.375 2.375 5.750 18.750 16.3756.250 2.250 5.750 20.625 18.3756.250 2.250 5.750 20.500 18.2506.250 2.125 5.500 20.375 18.2506.125 1.875 5.500 20.000 18.1256.125 2.250 7.250 23.750 21.5006.125 2.125 7.125 23.375 21.2506.125 2.125 7.125 23.375 21.2506.000 1.875 7.125 22.500 20.6256.000 2.125 9.000 26.625 24.5006.000 2.125 9.000 26.625 24.5006.000 2.125 8.875 26.250 24.1255.875 1.750 8.125 24.250 22.5006.000 2.000 9.375 27.625 25.6255.875 1.750 9.375 27.375 25.6255.750 1.500 8.875 27.000 25.5005.375 1.000 8.000 25.625 24.625

This journal is © The Royal Society of Chemistry 2014

Fig. 13 Variations in RMS of acceleration for diesel and biodieselblends at different engine loads.

Fig. 12 Variations in mass fraction burned for diesel and biodieselblends at a BMEP of 0.1 MPa.

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obtained with the addition of biodiesel in the blends for allengine loads. In fact, in comparison with the correspondingbaseline diesel, the JB50 blend promoted a faster combustionduration by an average of 0.125� CA and 2� CA at engine load of0.1 MPa and 0.5 MPa, respectively. The oxygen enrichment andimproved combustion process of the JB50 blend are postulatedto be the reason for the shorter combustion duration. Addi-tionally, it can be observed that the CA50 shied according tothe change in the blend of the biodiesel. Typically, CA50 wasused as a parameter that affected the ensemble heat releaseprole, and it was applied widely in a simulation of engineperformance when the Wiebe function was employed.36 In thepresent study, it can be observed that the CA50 timing occurredslightly earlier with biodiesel blended fuels under all operatingconditions. In fact, the largest shi in CA50 was found to be anadvance of 1.375� CA for JB50 in comparison with that ofbaseline diesel at a higher load of 0.6 MPa.

4.4. Vibration analysis

Vibration signals in an internal combustion engine are usuallyused as a diagnostic tool. They allow engine bearing to bemonitored for wear and overall engine knock detection. Thereare many sources of vibration in an engine, including pistonslap, faults in valves, knocking, burning pressure oscillation,torsional vibration, and the rotation of other engine acces-sories. This vibration is transmitted via a variety of paths andthen ultimately radiated acoustically to the surroundings.Exposure to excess vibration can accelerate wear and tear ofmechanical components and have an adverse impact on humancomfort. The combustion process in a diesel engine has aneffect on the engine vibration. The methyl ester fuel blendsinuenced the combustion process and consequently, the noiseand vibration. Fig. 13 is an illustration of RMS of the vibrationacceleration signal for all tested fuels, which was calculatedaccording to eqn (6). For each test, the average RMS of theacceleration signal for 100 successive engine combustion cycleswas considered. The general trend indicates that the variationsin RMS of acceleration are decreased with engine load.

This journal is © The Royal Society of Chemistry 2014

Additionally, the results also indicate that the RMS of acceler-ation is affected by biodiesel fuel blends. It was observed thatthe JB50 blend consistently resulted in the lowest RMS ofaccelerations in comparison with the baseline diesel under allloading conditions. It is interesting to note that the largestreduction of 11.3% in RMS of acceleration was obtained withJB50 at engine load of 0.5 MPa in comparison with the baselinediesel.

5. Conclusions

In the present study, the performance, emissions andcombustion characteristics of an engine fuelled with fossildiesel fuel and Jatropha biodiesel blends were investigated atengine loads of 0.1, 0.2, 0.3, 0.4, 0.5, and 0.6 MPa. The followingmain conclusions can be drawn from this investigation.

A prominent increase in BSFC was observed at all loadconditions when biodiesel fuel was used. Additionally, therewas a marginal improvement of 2% in BTE with the JB50 blendat an engine load of 0.5 MPa. In terms of exhaust emissions, itwas observed that the engine load had the greatest effect onBSCO emissions. In general, the BSCO emissions decreasedwith the increasing biodiesel blend ratio and engine load.Additionally, it was found that, in general, the BSNOx emissionsdecreased with increases in engine load and biodiesel blendingratio. The presence of JME decreased the BSNOx relative to thebaseline diesel, but the BSNOx did not decrease further as thedegree of JME blending increased. The largest reduction inBSNOx recorded was approximately 20.2% for the JB30 blend ata BMEP of 0.2 MPa. It is worth noting that smoke emissionsfrom the Jatropha biodiesel blends were lower than baselinediesel across all the engine loading conditions. It was observedthat the maximum reduction, in comparison with the corre-sponding baseline diesel, was 69.5% with JB50 at a BMEP of 0.5MPa, while a 21.9% reduction was found with JB10 at a BMEP of0.1 MPa.

In terms of the combustion characteristics, it was found that,at a lower engine load of 0.1 MPa, most of the biodiesel blendshave lower peak pressure in the range of 0.6–1.2 bar incomparison with the baseline diesel. In the case of higher

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engine loads, more signicant variations and higher peakcombustion pressures were observed among all biodieselblends and baseline diesel fuel. In fact, the location of occur-rence is shied earlier towards the TDC point. Furthermore, itwas observed that, in general, the peak AHRR for biodieselblends is lower at low load operation and is comparable at highload operation in comparison with baseline diesel. In addition,a slightly shorter ID and faster combustion duration were alsorevealed with the use of biodiesel blends across the engine loadoperations.

For vibration analysis, the results indicated that the RMS ofacceleration was affected by biodiesel fuel blends. It wasobserved that the largest reduction of 11.3% in the RMS ofacceleration was obtained with JB50 at an engine load of 0.5MPa in comparison with the baseline diesel.

Nomenclature and symbol

AHRR

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Apparent heat release rate

ASTM American society for testing and materials ATDC Aer top dead centre B10 10% biodiesel + 90% diesel fuel B100 100% biodiesel fuel BMEP Brake mean effective pressure BP Brake power BSCO Brake specic carbon monoxide BSFC Brake specic fuel consumption BSNOx Brake specic nitrogen oxides BTDC Before top dead centre BTE Brake thermal efficiency CA Crank angle CA10 Burn point of 10% CA50 Burn point of 50% CA90 Burn point of 90% CI Compression ignition CIB Calophyllum inophyllum CO Carbon monoxide CO2 Carbon dioxide D Fossil diesel DI Direct injection ECU Electronic control unit EGT Exhaust gas temperature GC-FID Gas chromatograph-ame ionization detector FC Fuel consumption HC Hydrocarbons HPCR High pressure common-rail HRR Heat release rate ID Ignition delay IDI Indirect injection JB Jatropha biodiesel JB10 10% Jatropha biodiesel + 90% diesel fuel JB30 30% Jatropha biodiesel + 70% diesel fuel JB50 50% Jatropha biodiesel + 50% diesel fuel JME Jatropha methyl ester KB Karanja biodiesel l Relative air-fuel ratio (Lambda) MPLN Mechanical pump-line-nozzle

dv., 2014, 4, 50739–50751

NA

Naturally aspirated NOx Nitrogen oxides PB Polanga biodiesel ppm Part per million rpm Revolution per minute RMS Root mean square TC Turbocharged TDC Top dead centre THC Total hydrocarbon

Acknowledgements

The authors would like to acknowledge the Ministry of HigherEducation (MOHE) of Malaysia and University of Malaya fornancial support through UMRG (grant number RG145-12AET),HIR grant (UM.C/HIR/MOHE/ENG/07), and PostgraduateResearch Grant (PPP) (grant number PG035-2012B).

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