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Performance and emission characteristics of a turpentine–diesel dual fuel engine
ARTICLE IN PRESS
Energy 32 (2007) 1202–1209 www.elsevier.com/locate/energy
Performance and emission characteristics of a turpentine–diesel dual fuel engine R. Karthikeyana,, N.V. Mahalakshmib a
Adhiparasakthi Engineering College, Melmaruvathur, Tamil Nadu, India I.C. Engines Division, Department of Mechanical Engineering, College of Engineering Guindy, Chennai, Tamil Nadu, India
b
Received 26 December 2005
Abstract This paper describes an experimental study concerning the feasibility of using bio-oil namely turpentine obtained from the resin of pine tree. The emission and performance characteristics of a D.I. diesel engine were studied through dual fuel (DF) mode. Turpentine was inducted as a primary fuel through induction manifold and diesel was admitted into the engine through conventional fueling device as an igniter. The result showed that except volumetric efficiency, all other performance and emission parameters are better than those of diesel fuel with in 75% load. The toxic gases like CO, UBHC are slightly higher than that of the diesel baseline (DBL). Around 40–45% smoke reduction is obtained with DF mode. The pollutant Nox is found to be equal to that of DBL except at full load. This study has proved that approximately 75% diesel replacement with turpentine is possible by DF mode with little engine modification. r 2006 Elsevier Ltd. All rights reserved. Keywords: Turpentine; Dual fuel; Alternate fuel; Emission analysis; Combustion analysis
1. Introduction Depleting petroleum reserves and increasing cost of the petroleum products demands an intensive search for new alternative fuels. Biofuels are proved to be the best substitutes for the existing petro fuels. But they require little engine modification or fuel modification. Generally biofuels are the oils obtained from the living plant sources. These oils may be obtained from resin and plant seeds. Plant oils are renewable and have low sulphur in nature. As the biofuels are more expensive than fossil fuels, the wide spread use of biofuel was restrained from its use in I.C. engines [1,2]. The use of vegetable oil in diesel engine has been identified well before the exploration of the other promising alternative fuel alcohol. But the problems associated with vegetable oil are the high viscosity and low volatility. These properties have an adverse effect on fuel injection system and may lead to heavy carbon deposits in the engine combustion chamber [2–4]. In the Corresponding author. Tel./fax: +91 4115 229639.
E-mail address: [email protected] (R. Karthikeyan). 0360-5442/$ - see front matter r 2006 Elsevier Ltd. All rights reserved. doi:10.1016/j.energy.2006.07.021
present work turpentine is used as an alternate fuel to diesel. This is a biofuel obtained from resin extracted from pine tree. Turpentine was used in early engines with out any modification. The abundant availability of petro fuels had stopped the usage of turpentine in I.C. engines. But the increasing cost of petro fuel prevailing today reopens the utility of turpentine in I.C. engines. Turpentine oil has low cetane number; this oil could be used in direct injection diesel engines in the form of the turpentine–diesel blends [5] and dual fuel (DF) mode. The specific objectives of this study are to analyze performance and emission characteristics of turpentine in direct injection diesel engine and to analyze its feasibility as a fuel in a D.I. diesel engine. From the author’s earlier study [5], it was identified that 20% turpentine and 80% diesel is the optimum blend in terms of performance and emission characteristics. In the present work a normal diesel engine was modified to work in a DF mode with turpentine and diesel as primary and pilot fuels, respectively. The resulting homogeneous mixture was compressed to a temperature below the self ignition point. The pilot fuel was injected through the standard injection
ARTICLE IN PRESS R. Karthikeyan, N.V. Mahalakshmi / Energy 32 (2007) 1202–1209
Nomenclature CO carbon monoxide UBHC unburned hydrocarbon
system and initiated the combustion in the primary fuel air mixture. The primary fuel has supplied most of the energy. Parameters like thermal efficiency, volumetric efficiency, smoke, UBHC, CO and Nox emissions were evaluated. Peak pressure, ignition delay, combustion duration and heat release rate were obtained from pressure crank angle data. The CO and UBHC emissions will be usually higher in duel fuel mode when gaseous fuels are used as primary fuels [6,7]. But the present work shows that these emissions can be considerably decreased due to the higher heating value of turpentine. Advantages of turpentine are: 1. It is a renewable fuel and biofuel; obtained from pine tree. 2. Self-ignition temperature—close to diesel fuel. 3. Boiling point—almost equal to diesel fuel. 4. Calorific value—slightly higher than diesel fuel. 5. Viscosity—almost equal to diesel fuel. 6. Offers 11–15% higher calorific value compared to other biofuels (bio-diesel and neat vegetable oil).
2. Structure of the program The structure of the program is as follows: 1. The whole test was conducted for the standard engine injection pressure and injection timing. 2. Induction manifold was extended outward to accommodate primary fuel spray system. 3. There were two separate fuel-metering systems provided with test rig to measure primary and pilot fuel consumption, respectively. 4. The first test was conducted using 100% low sulfur diesel fuel to establish diesel baseline (DBL) for emission, fuel consumption and performance. 5. The engine was then started and its no load rack position was locked. 6. First 25% load was applied, consequently speed of the engine decreases. 7. Turpentine was admitted through primary fuel-regulating device and the initial speed was restored. 8. Emission, fuel consumption and cylinder pressure were measured at each load. 9. 50%, 75% and 100% loads were applied one by one and steps 7 and 8 were repeated for each load.
NOx EGT DBL DF
1203
nitric oxide exhaust gas temperature diesel baseline dual fuel
10. Then the emission and the performance of the DF mode were compared with the DBL readings.
3. Experimental set up Fig. 1 shows the schematic diagram of experimental set up. The test engine was Kirloskar TAF 1, single cylinder, constant speed (1500 rpm), direct injection, with a bore of 87.5 mm and a stroke of 110 mm (Table 1). The rated output of the engine was 4.4 kW at 1500 rpm. The compression ratio was 17.5:1 and the manufacturer’s recommended injection timing and injection pressure were 261 before TDC and 190 bar, respectively. The combustion chamber was the direct injection with a bowl-in piston. The engine was coupled to an eddy current dynamometer to provide a brake load and it is controlled by a control system provided in the control panel, which is also consisting a speed indicator and a load indicator. Two separate fuel-metering systems were provided to meter both primary fuel and pilot fuel. Fuel consumption of an engine was measured manually by a graduated burette. The primary fuel (turpentine) was allowed into the engine through primary fuel spray system, which consists of a throttle less gasoline carburetor, whose fuel flow rate is controlled manually by a fuel flow adjustment screw. The pilot fuel was admitted through conventional fuel system. Time taken for fuel consumption was measured with the help of digital stopwatch. An orifice meter attached with an anti pulsating drum measures air consumption of an engine with the help of U tube manometer. The anti pulsating drum fixed in the inlet side of an engine maintains a constant suction pressure, to facilitate constant airflow through the orifice meter. Exhaust emission from the engine was measured with the help of QRO TECH, QRO-402 gas analyzer and smoke intensity was measured with the help of Bosch smoke meter. Bosch smoke meter usually consists of a piston typesampling pump and a smoke level measuring unit. Two separate sampling probes were used to receive sampleexhaust gases from the engine for measuring emission and smoke intensity, respectively. A 2-in diameter filter paper was used to collect smoke samples from the engine, through smoke sampling pump for measuring Bosch smoke number. A k-type thermocouple and a temperature indicator were used to measure EGT. Combustion diagnosis was carried by means of a PCB Piezotronics HSM111A22 quartz pressure transducer fitted
ARTICLE IN PRESS R. Karthikeyan, N.V. Mahalakshmi / Energy 32 (2007) 1202–1209
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DIESEL AQUISITION SYSTEM
TURPENTINE
GAS ANALYSER
DYNAMOMETER
AIR BOX
ENGINE
DYNAMOMETER CONTROL Fig. 1. Experimental set up.
5. Error analysis
Table 1 Engine details 1.
2. 3. 4. 5. 6. 7. 8. 9. 10.
General details
Make Cubic capacity Bore Stroke Compression ratio Speed (constant speed) Rated power Dynamometer Pressure pick up
Single cylinder, water cooled, compression Ignition, constant speed engine Kirloskar TAF 1 661 cc 87.5 mm 110 mm 17.5:1 1500 rpm 4.4 KW Eddy current Piezotronics HSM111A22 quartz pressure transducer
on the engine cylinder head and a crank angle encoder was fixed on the output shaft of the engine. The pressure and crank angle signals were fed to a data acquisition card fitted with the pentium4 personal computer. The combustion parameters such as peak pressure, heat release rate, mean gas temperature and ignition delay were computed.
Errors and uncertainties in the experiments can arise from instrument selection, condition, calibration, environment, observation, reading and test planning. Uncertainty analysis is needed to prove the accuracy of the experiments. An uncertainty analysis was performed using the method described by J.P. Holman [Book entitled ‘Experimental techniques’]. Percentage uncertainties of various parameters like total fuel consumption, brake power, specific fuel consumption and brake thermal efficiency were calculated using the percentage uncertainties of various instruments given in the Table 2. Total percentage uncertainty of this experiment is ¼ Square root offðuncertainty of TFCÞ2 þ ðuncertainty of brake powerÞ2 þ ðuncertainty of specific fuel consumptionÞ2 þ ðuncertainty of brake thermal efficiencyÞ2 þ ðuncertainty of COÞ2 þ ðuncertainty of CO2 Þ2 þ ðuncertainty of UBHCÞ2 þ ðuncertainty of Nox Þ2
4. Instrumentation Table 2 provides the range, accuracy, measurement technique and percentage uncertainties of various instruments involved in this experiment for observing various parameters.
þ ðuncertainty of Bosch smoke numberÞ2 þ ðuncertainty of EGT indicatorÞ2 þ ðuncertainty of pressure pickupÞ2 g ¼ square root offð1Þ2 þ ð0:2Þ2 þ ð1Þ2 þ ð1Þ2 þ ð0:2Þ2
ARTICLE IN PRESS R. Karthikeyan, N.V. Mahalakshmi / Energy 32 (2007) 1202–1209
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Table 2 List of instruments and its range, accuracy, measurement technique and uncertainties
1.
2. 3. 4. 5. 6. 7. 8. 9. 10.
Instruments
Range
Accuracy
Measurement techniques
Percentage uncertainties
Gas analyzer
CO 0–10% CO2 0–20% UBHC 0–10,000 ppm Nox 0–5000 ppm BSN 0–10 0–900 1C 0–10,000 rpm 0–100 kg
70.02% 70.03% 720 ppm 710 ppm 70.1 71 1C 710 rpm 70.1 kg 70.1 cc 70.6 s 71 mm 70.1 kg 711
NDIR principle (Nondepressive infra red sensor)
70.2 70.15 70.2 70.2 71 70.15 70.1 70.2 71 70.2 71 70.1 70.2
Smoke level measuring instrument EGT indicator Speed measuring unit Load indicator Burette for fuel measurement Digital stop watch Manometer Pressure pickup Crank angle encoder
0–110 bar
Table 3 Estimated world production of crude resin, rosin and turpentine Production (Tonnes)
Crude resin
Rosin
Turpentine
World total production of which: China, People’s Rep. of Indonesia Russia Brazil Portugal India Argentina Mexico Honduras Venezuela Greece South Africa Viet Nam Others
976,000
71,7000
99,400
1993
570,000
43,0000
50,000
1993 1992 1993 1992 1994 1993 1991 1992 1993 1993 1993 1990
100,000 90,000 65,000 30,000 30,000 30,000 30,000 8000 7000 6000 2000 2000 6000
69,000 65,000 45,000 22,000 21,000 21,000 22,000 6000 5000 4000 1500 1500 4000
12,000 9000 8000 5000 4000 4000 4000 1000 800 600 200 200 600
þ ð0:15Þ2 þ ð0:2Þ2 þ ð0:2Þ2 þ ð1Þ2 þ ð0:15Þ2 þ ð1Þ2 g ¼ 3%. Using the calculation procedure, the total uncertainty for the whole experiment is obtained to be 73%. 6. Test fuels The test fuels for the present study were turpentine (primary fuel) and high-speed diesel with sulphur contents 0.04 ppm. The resource of turpentine oil in the world is shown in Table 3. Table 4 shows the physical and chemical properties of turpentine oil [8].
Electro chemical sensor. k-type (Cr Al) thermocouple Magnetic pick up type Strain gauge type load cell
Magnetic pick up type
Table 4 Property comparison of turpentine with existing petro fuels
Formula Molecular weight Composition (% wt) Density (kg/m3) Specific gravity Pour point (1C) Boiling point (1C) Vapor pressure (kPa) Viscosity c St @ 30 1C Latent heat of vaporization (kJ/kg) Lower heating value (kJ/ kg) Flash point (1C) Auto ignition temperature (1C) Flammability limit (% volume) Stoichiometric air–fuel ratio Flame speed m/s Cetane number
Gasoline
Diesel
Turpentine
C4 to C12 105 C 88 H 15 780 0.78 40 30–220 48–103
C10H16 136 C 88.2 H 11.8 860–900 0.86–0.9
350
C8 to C25 200 C 87 H 16 830 0.83 23 180–340 o1 3–4 230
43,890
42,700
44,400
43 300–450
74 250
38 305
1.4
1.0
0.8
14.7
14.7
150–180 o1 2.5 285
4–6 40–55
7. Performance analysis The variation of SFC with brake power output is shown in Fig. 2. Since the heat content of the turpentine is higher than that of diesel fuel, a very minimum hike in fuel consumption is obtained with DF mode. In DF mode a notable increase in fuel consumption is obtained between 75% load and full load. The main reason for the increased fuel consumption is due to the beginning of knock at 75% load. This is clearly seen in P–y diagram (Fig. 6). Unusual hike of peak pressure, more fluctuation in cylinder pressure
ARTICLE IN PRESS R. Karthikeyan, N.V. Mahalakshmi / Energy 32 (2007) 1202–1209
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0.5
35 B.TH. Efficiency in %.
Diesel baseline Dual fuel mode
SFC in Kg/Kwhr
0.45
0.4
0.35
30 25 20 15 10
Diesel baseline
5
Dual fuel mode
0
0.3
25
0
0.25
50
75 Load in %
100
125
Fig. 4. Variation of brake thermal efficiency with brake power.
0.2 0
25
50
75 Load in %
100
500
125
450
Fig. 2. Variation of SFC with brake power.
Fuel consumption in Kg/hr
Pilot fuel quantity 0.8
Primary fuel quantity
EGT in Degrees
1
Diesel baseline Dual fuel mode
400 350 300 250 200 150
0.6
100 0
0.4
25
50 75 Load in %
100
125
Fig. 5. Variation of EGT with brake power.
0.2
0 Load 0% 25% 50% 75% 100% Fig. 3. Comparison of fuel consumption of pilot fuel and primary fuel.
and pinking noise are the indicators of knock occurrence. The abnormal combustion of charge at full load leads to the poor fuel utilization and increased fuel consumption. More admission of octane fuel (primary fuel) at higher load ranges increases the tendency of knock. However, in the other work [5], the authors have reported that with the help of a suitable diluent the knock can be controlled to a very large extent (Table 2). Fuel consumption of DF mode is clearly depicted in Fig. 3. It shows that almost constant fuel supply is maintained by pilot-fuel supply system irrespective of load but the primary-fuel (turpentine) consumption is gradually increases as load changes from 25% to 100%. The maximum diesel replacement by turpentine at 75% load is approximately 75% of total fuel consumption, after that, a sudden hike in fuel consumption was recorded due to abnormal combustion (knocking). The variation of brake thermal efficiency with brake power output is shown in Fig. 4. It is clear that the brake thermal efficiency of the DF engine is very close to DBL
upto 75% load, beyond which it starts decreasing from the DBL. The maximum efficiency of DF engine is 29%. This is obtained at 75% of full load. The reason for the decrease of brake thermal efficiency beyond 75% load is the occurrence of knock. More admission of octane fuel at the time of full load increases the tendency of knock. The abnormal combustion of charge at full load leads to poor fuel utilization and increased fuel consumption. This leads to poor break thermal efficiency between 75% load, and full load. However, at 75% load the break thermal efficiency of DF mode is better than that of DBL. This is due to higher heating value of the turpentine oil, more charge homogeneity at the end of the compression stroke and rapid burning of primary fuel. Usually in gaseous fuel DF mode, the thermal efficiency at lower loads will be lower than that of DBL due to sluggish combustion of more diluted fuel air mixture [7]. But in the present case, it is not so, instead, turpentine DF mode maintains the same brake thermal efficiency as obtained in the DBL. The variation of EGT with brake power output is shown in Fig. 5. The EGT of turpentine–diesel DF engine is lower than that of DBL upto 75% load beyond which it starts increasing more steeply than DBL. The occurrence of knock in between 75% and 100% load increases the fuel
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120
94
Cylinder pressure in bar
Volumetric efficiency in %
R. Karthikeyan, N.V. Mahalakshmi / Energy 32 (2007) 1202–1209
92 90 88
Diesel baseline Dual fuel mode
86 84 0
25
50
75 Load in %
100
125
100 80 60 40 20
Diesel baseline Dual fuel mode
0 320
Fig. 6. Variation of volumetric efficiency with brake power.
340
360
380
400
420
Crank angle in degrees
8. Combustion analysis Fig. 7 shows that P–y diagram of turpentine–diesel DF engine at 75% of full load. From the figure it is observed that the ignition delay of DF engine is considerably longer than that of DBL. This is due to injection of pilot fuel in the environment where the oxygen concentration is less and also the temperature of the mixture at the time of injection is lesser
Fig. 7. Variation of P–y with brake power at 75% load.
300 Net heat release rate in J/Degree
consumption. This unusual behavior of DF mode increases the EGT to a larger extent. More mass burning, uncontrolled combustion and steep increase of mean gas temperature are the major reasons for higher EGT at full load. Another reason for the increased exhaust gas temperature at full load is the higher heat release during diffused combustion stage, higher heat content of turpentine and decreased volumetric efficiency. Usually in all DF engines the primary fuel is inducted along with inlet air and evaporated inside the cylinder by observing heat from the cylinder wall and the gas inside the cylinder. This helps to maintain the cylinder temperature well below self-ignition temperature of primary fuel and resulting low EGT. Usually, fuel that has high latent heat of vaporization will offer low EGT when compared to that of DBL. The variation of volumetric efficiency with brake power output is shown in Fig. 6. The volumetric efficiency of turpentine–diesel DF engine is lower than that of DBL at all loads except at 0% load. The reason for the decreased volumetric efficiency is due to the induction of turpentine along with the inlet air. After induction, turpentine starts vapourizing by absorbing heat from the cylinder wall and reduces the space available for fresh air entry. Hence the volumetric efficiency reduces as the load increases or the percentage of turpentine admission increases. The drastic reduction of volumetric efficiency is observed at full load due to the presence of knock. The occurrence of knock will lead to increased fuel consumption, mass burning of fuel and increased mean gas temperature. This is the major reason for drastic reduction of volumetric efficiency in between 75% load and full load. At full load, approximately 8% drop in volumetric efficiency is obtained.
275 250 225 200
DBL
175 150 DF
125 100 75 50 25 0 340
350
360
370
380
390
400
410
420
Crank angle in degrees Fig. 8. Variation of heat release rate with brake power at 75% load.
than the temperature of the DBL operation. This is due to the evaporation of turpentine oil inside the cylinder by absorbing heat from the combustion chamber. The more ignition delay at the time of higher load range increases the peak pressure than that of DBL and provides more pressure fluctuation. This is the main reason for more waviness in the expansion period of P–y diagram. In other words this period depicts the abnormal combustion or knocking period. However, this is the usual behavior of high-octane fuel in high compression ratio engines. This can be controlled by proper selection of compression ratio and admission of knock suppressers. Fig. 8 compares heat release rate of turpentine–diesel DF engine with DBL at 75% of full load. It shows that DF mode offers higher ignition delay than DBL. Because in DF mode, pilot fuel is injected inside the primary fuel and air mixture, which is kept at a temperature lower than DBL mode at the same load condition. Since the no load rack position was sealed, the premixed stage of combustion is shorter and also occurs 31–41 later than DBL. Most of the heat release is occurred only during the diffusive combustion phase. Longer ignition delay provides the higher heat release during the
ARTICLE IN PRESS R. Karthikeyan, N.V. Mahalakshmi / Energy 32 (2007) 1202–1209
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diffusive combustion phase. Hence, the duration of heat release is also longer than DBL, which leads to the increase in EGT. The increase in combustion duration is mainly due to the slow combustion of injected fuel and combustion of inducted turpentine by flame propagation.
9. Emission analysis Fig. 9 compares the emission performance of turpentine–diesel DF engine with DBL at various load conditions. The CO emission of turpentine–diesel DF engine gradually increases from 0% load to 75% load. The reason behind increased CO emission at lighter loads is higher fumigation rate and scarcity of oxygen. The flame quenching and cooled layer of charge near the wall will also be the reason for the increased CO emission. The drastic increase of CO emission at higher loads is due to the occurrence of knock and higher fuel consumption. Knocking leads to poor fuel utilization and increased cylinder temperature, which will
CO in %
6
4
2
Bosch Smoke Number
0 4 3 2 1 0
NOx in PPM
2000 1500 1000 500 0
UBHC in PPM
400 Diesel baseline Dual Fuel mode
300 200
cause more CO emission. Approximately 35% of higher CO emission is observed at 100% load of DF engine. Drastic reduction of volumetric efficiency at full load due to abnormal combustion reduces oxygen availability and hence leads to the higher CO emission. The UBHC emission of turpentine–diesel DF engine gradually increases from 0% load to 75% load. The reason behind increased UBHC emission is higher fumigation rate and scarcity of oxygen. The flame quenching and cooled layer of charge near the wall will be the reason for the increased UBHC emission. Though this is a conventional behaviour of the DF engine, the drastic increase of UBHC emission at higher loads is due to occurrence of knock and higher fuel consumption. At lighter loads due to charge homogeneity and higher oxygen availability the UBHC emission level is less, where as at higher load ranges due to higher quantity of fuel admission, charge temperature and drastically reduced volumetric efficiency, UBHC emission increases. Approximately 45% of higher UBHC emission is noted at 100% load in DF engine. The NOx of turpentine–diesel DF engine closely follows the DBL upto 50% load, beyond which it starts increasing from DBL. The reason for the increased NOx is due to the higher heat release during second stage of combustion, higher combustion temperature and higher heat content of turpentine. Increased ignition delay of the DF engine, which promotes diffusive combustion, is also be another reason for increased NOx. A drastic increase in NOx level is found beyond 75% load due to the abnormal combustion. This leads to a very high gas temperature, increase of fuel consumption and poor utilization of heat. The presence of high gas temperature inside the cylinder will create a conducive ambient for reaction of atmospheric nitrogen with oxygen. This is the major reason for higher NOx emission at higher load ranges. The maximum NOx emission is approximately 21% at full load. The variation of smoke emission with load is shown in Fig. 8. The turpentine–diesel DF engine reduces the smoke to a larger extent. The reason for the reduced smoke emission is the availability of premixed and homogeneous charge inside the engine well before the commencement of combustion. Higher heat content of turpentine, higher combustion temperature, extended duration of combustion and rapid flame propagation are the other reasons for reduced smoke level. The pilot fuel was used in less quantities and this may be one of the reason for the reduction of smoke emission. However, at higher load ranges due to non-availability of sufficient air and abnormal combustion there is a visible white smoke emission. 10. Conclusions
100 0 0
25
50 75 Load in %
100
125
Fig. 9. Variation of emission parameters with brake power.
From the detailed test conducted with turpentine–diesel DF engine the following conclusions are arrived: 1. The brake thermal efficiency of DF engine at 75% load is 1–2% higher than that of DBL.
ARTICLE IN PRESS R. Karthikeyan, N.V. Mahalakshmi / Energy 32 (2007) 1202–1209
2. Increased SFC is reported at full load due to the presence of knock. 3. EGT and NOx are found lower than that of DBL up to 50% load beyond which they start increasing due to the occurrence of knock. 4. Approximately 35% of higher CO emission is reported at full load of DF engine. 5. 48% of higher UBHC emission and 45% reduced smoke are reported at full load of DF engine. 6. Maximum of 8% drop in volumetric efficiency is reported in DF engine at full load. 7. Most of the research papers submitted on the same topic earlier have reported that the DF operation at light loads are less efficient than its diesel counter part; But, this investigation proves that the brake thermal efficiency at lighter loads are equal to that of DBL. 8. At higher loads due to higher admission of octane fuel (primary fuel) the engine starts knocking. Hence the performance of DF engine beyond 75% load starts deteriorating. From the above experiment, it is concluded that 60–65% replacement of diesel with turpentine is quite possible within 75% load. Except increased CO and UBHC emissions, all other emission parameters like smoke and NOx and performance parameter like brake thermal efficiency are found superior to that of DBL.
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References [1] Mayer-Pitroff R. Biofuel: technical. Economic and environmental impacts. In: Elaeis special issue: proceedings of the 1995 PORIM international biofuel conference. Kuala Lumpur: Palm Oil Research Institute of Malaysia; 16–17 January 1995. p. 1–11. [2] Rickeard DJ, Thompson ND. A review of the potential for bio-fuels as transportation fuels. SAE paper 932778, 1993. [3] Masjuki H, Abdulmuin MZ, Sii HS. Indirect injection diesel engine operation on palm oil methyl esters and its emulsions. Proceedings of the institution of mechanical engineers part D. J Automobile Eng 1997;211:291–9. [4] Husna K, Mariani O, Hamzah AH. Biofuel: impact on environment. In: Elaeis special issue: proceeding of the 1995 PORIM international biofuel conference. Kuala Lumpur: Palm Oil Research Institute of Malaysia; 16–17 January 1995. p. 99–109. [5] Karthikeyan R, Mahalakshmi NV. Internal combustion engine division of ASME 2005 fall technical conference, ICEF2005. Ottawa, Canada: Crowne Plaza Hotel; 11–14 September 2005. [6] Senthil Kumar M, Ramesh A, Nagalingam B. Experimental investigations on a jetropa oil methanol dual fuel engine. SAE paper 2001-010153. [7] Senthil Kumar M, Ramesh A, Nagalingam B. An experimental comparison of methods to use methanol and jetropa oil in a compression ignition engine. Biomass Bioenergy Paper 2003;25: 309–18. [8] ACGIH. Industrial ventilation—a manual of recommended practice. Cincinnati, OH: American Conference of Governmental Industrial Hygienists; 1986.
Energy 32 (2007) 1202–1209 www.elsevier.com/locate/energy
Performance and emission characteristics of a turpentine–diesel dual fuel engine R. Karthikeyana,, N.V. Mahalakshmib a
Adhiparasakthi Engineering College, Melmaruvathur, Tamil Nadu, India I.C. Engines Division, Department of Mechanical Engineering, College of Engineering Guindy, Chennai, Tamil Nadu, India
b
Received 26 December 2005
Abstract This paper describes an experimental study concerning the feasibility of using bio-oil namely turpentine obtained from the resin of pine tree. The emission and performance characteristics of a D.I. diesel engine were studied through dual fuel (DF) mode. Turpentine was inducted as a primary fuel through induction manifold and diesel was admitted into the engine through conventional fueling device as an igniter. The result showed that except volumetric efficiency, all other performance and emission parameters are better than those of diesel fuel with in 75% load. The toxic gases like CO, UBHC are slightly higher than that of the diesel baseline (DBL). Around 40–45% smoke reduction is obtained with DF mode. The pollutant Nox is found to be equal to that of DBL except at full load. This study has proved that approximately 75% diesel replacement with turpentine is possible by DF mode with little engine modification. r 2006 Elsevier Ltd. All rights reserved. Keywords: Turpentine; Dual fuel; Alternate fuel; Emission analysis; Combustion analysis
1. Introduction Depleting petroleum reserves and increasing cost of the petroleum products demands an intensive search for new alternative fuels. Biofuels are proved to be the best substitutes for the existing petro fuels. But they require little engine modification or fuel modification. Generally biofuels are the oils obtained from the living plant sources. These oils may be obtained from resin and plant seeds. Plant oils are renewable and have low sulphur in nature. As the biofuels are more expensive than fossil fuels, the wide spread use of biofuel was restrained from its use in I.C. engines [1,2]. The use of vegetable oil in diesel engine has been identified well before the exploration of the other promising alternative fuel alcohol. But the problems associated with vegetable oil are the high viscosity and low volatility. These properties have an adverse effect on fuel injection system and may lead to heavy carbon deposits in the engine combustion chamber [2–4]. In the Corresponding author. Tel./fax: +91 4115 229639.
E-mail address: [email protected] (R. Karthikeyan). 0360-5442/$ - see front matter r 2006 Elsevier Ltd. All rights reserved. doi:10.1016/j.energy.2006.07.021
present work turpentine is used as an alternate fuel to diesel. This is a biofuel obtained from resin extracted from pine tree. Turpentine was used in early engines with out any modification. The abundant availability of petro fuels had stopped the usage of turpentine in I.C. engines. But the increasing cost of petro fuel prevailing today reopens the utility of turpentine in I.C. engines. Turpentine oil has low cetane number; this oil could be used in direct injection diesel engines in the form of the turpentine–diesel blends [5] and dual fuel (DF) mode. The specific objectives of this study are to analyze performance and emission characteristics of turpentine in direct injection diesel engine and to analyze its feasibility as a fuel in a D.I. diesel engine. From the author’s earlier study [5], it was identified that 20% turpentine and 80% diesel is the optimum blend in terms of performance and emission characteristics. In the present work a normal diesel engine was modified to work in a DF mode with turpentine and diesel as primary and pilot fuels, respectively. The resulting homogeneous mixture was compressed to a temperature below the self ignition point. The pilot fuel was injected through the standard injection
ARTICLE IN PRESS R. Karthikeyan, N.V. Mahalakshmi / Energy 32 (2007) 1202–1209
Nomenclature CO carbon monoxide UBHC unburned hydrocarbon
system and initiated the combustion in the primary fuel air mixture. The primary fuel has supplied most of the energy. Parameters like thermal efficiency, volumetric efficiency, smoke, UBHC, CO and Nox emissions were evaluated. Peak pressure, ignition delay, combustion duration and heat release rate were obtained from pressure crank angle data. The CO and UBHC emissions will be usually higher in duel fuel mode when gaseous fuels are used as primary fuels [6,7]. But the present work shows that these emissions can be considerably decreased due to the higher heating value of turpentine. Advantages of turpentine are: 1. It is a renewable fuel and biofuel; obtained from pine tree. 2. Self-ignition temperature—close to diesel fuel. 3. Boiling point—almost equal to diesel fuel. 4. Calorific value—slightly higher than diesel fuel. 5. Viscosity—almost equal to diesel fuel. 6. Offers 11–15% higher calorific value compared to other biofuels (bio-diesel and neat vegetable oil).
2. Structure of the program The structure of the program is as follows: 1. The whole test was conducted for the standard engine injection pressure and injection timing. 2. Induction manifold was extended outward to accommodate primary fuel spray system. 3. There were two separate fuel-metering systems provided with test rig to measure primary and pilot fuel consumption, respectively. 4. The first test was conducted using 100% low sulfur diesel fuel to establish diesel baseline (DBL) for emission, fuel consumption and performance. 5. The engine was then started and its no load rack position was locked. 6. First 25% load was applied, consequently speed of the engine decreases. 7. Turpentine was admitted through primary fuel-regulating device and the initial speed was restored. 8. Emission, fuel consumption and cylinder pressure were measured at each load. 9. 50%, 75% and 100% loads were applied one by one and steps 7 and 8 were repeated for each load.
NOx EGT DBL DF
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nitric oxide exhaust gas temperature diesel baseline dual fuel
10. Then the emission and the performance of the DF mode were compared with the DBL readings.
3. Experimental set up Fig. 1 shows the schematic diagram of experimental set up. The test engine was Kirloskar TAF 1, single cylinder, constant speed (1500 rpm), direct injection, with a bore of 87.5 mm and a stroke of 110 mm (Table 1). The rated output of the engine was 4.4 kW at 1500 rpm. The compression ratio was 17.5:1 and the manufacturer’s recommended injection timing and injection pressure were 261 before TDC and 190 bar, respectively. The combustion chamber was the direct injection with a bowl-in piston. The engine was coupled to an eddy current dynamometer to provide a brake load and it is controlled by a control system provided in the control panel, which is also consisting a speed indicator and a load indicator. Two separate fuel-metering systems were provided to meter both primary fuel and pilot fuel. Fuel consumption of an engine was measured manually by a graduated burette. The primary fuel (turpentine) was allowed into the engine through primary fuel spray system, which consists of a throttle less gasoline carburetor, whose fuel flow rate is controlled manually by a fuel flow adjustment screw. The pilot fuel was admitted through conventional fuel system. Time taken for fuel consumption was measured with the help of digital stopwatch. An orifice meter attached with an anti pulsating drum measures air consumption of an engine with the help of U tube manometer. The anti pulsating drum fixed in the inlet side of an engine maintains a constant suction pressure, to facilitate constant airflow through the orifice meter. Exhaust emission from the engine was measured with the help of QRO TECH, QRO-402 gas analyzer and smoke intensity was measured with the help of Bosch smoke meter. Bosch smoke meter usually consists of a piston typesampling pump and a smoke level measuring unit. Two separate sampling probes were used to receive sampleexhaust gases from the engine for measuring emission and smoke intensity, respectively. A 2-in diameter filter paper was used to collect smoke samples from the engine, through smoke sampling pump for measuring Bosch smoke number. A k-type thermocouple and a temperature indicator were used to measure EGT. Combustion diagnosis was carried by means of a PCB Piezotronics HSM111A22 quartz pressure transducer fitted
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DIESEL AQUISITION SYSTEM
TURPENTINE
GAS ANALYSER
DYNAMOMETER
AIR BOX
ENGINE
DYNAMOMETER CONTROL Fig. 1. Experimental set up.
5. Error analysis
Table 1 Engine details 1.
2. 3. 4. 5. 6. 7. 8. 9. 10.
General details
Make Cubic capacity Bore Stroke Compression ratio Speed (constant speed) Rated power Dynamometer Pressure pick up
Single cylinder, water cooled, compression Ignition, constant speed engine Kirloskar TAF 1 661 cc 87.5 mm 110 mm 17.5:1 1500 rpm 4.4 KW Eddy current Piezotronics HSM111A22 quartz pressure transducer
on the engine cylinder head and a crank angle encoder was fixed on the output shaft of the engine. The pressure and crank angle signals were fed to a data acquisition card fitted with the pentium4 personal computer. The combustion parameters such as peak pressure, heat release rate, mean gas temperature and ignition delay were computed.
Errors and uncertainties in the experiments can arise from instrument selection, condition, calibration, environment, observation, reading and test planning. Uncertainty analysis is needed to prove the accuracy of the experiments. An uncertainty analysis was performed using the method described by J.P. Holman [Book entitled ‘Experimental techniques’]. Percentage uncertainties of various parameters like total fuel consumption, brake power, specific fuel consumption and brake thermal efficiency were calculated using the percentage uncertainties of various instruments given in the Table 2. Total percentage uncertainty of this experiment is ¼ Square root offðuncertainty of TFCÞ2 þ ðuncertainty of brake powerÞ2 þ ðuncertainty of specific fuel consumptionÞ2 þ ðuncertainty of brake thermal efficiencyÞ2 þ ðuncertainty of COÞ2 þ ðuncertainty of CO2 Þ2 þ ðuncertainty of UBHCÞ2 þ ðuncertainty of Nox Þ2
4. Instrumentation Table 2 provides the range, accuracy, measurement technique and percentage uncertainties of various instruments involved in this experiment for observing various parameters.
þ ðuncertainty of Bosch smoke numberÞ2 þ ðuncertainty of EGT indicatorÞ2 þ ðuncertainty of pressure pickupÞ2 g ¼ square root offð1Þ2 þ ð0:2Þ2 þ ð1Þ2 þ ð1Þ2 þ ð0:2Þ2
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Table 2 List of instruments and its range, accuracy, measurement technique and uncertainties
1.
2. 3. 4. 5. 6. 7. 8. 9. 10.
Instruments
Range
Accuracy
Measurement techniques
Percentage uncertainties
Gas analyzer
CO 0–10% CO2 0–20% UBHC 0–10,000 ppm Nox 0–5000 ppm BSN 0–10 0–900 1C 0–10,000 rpm 0–100 kg
70.02% 70.03% 720 ppm 710 ppm 70.1 71 1C 710 rpm 70.1 kg 70.1 cc 70.6 s 71 mm 70.1 kg 711
NDIR principle (Nondepressive infra red sensor)
70.2 70.15 70.2 70.2 71 70.15 70.1 70.2 71 70.2 71 70.1 70.2
Smoke level measuring instrument EGT indicator Speed measuring unit Load indicator Burette for fuel measurement Digital stop watch Manometer Pressure pickup Crank angle encoder
0–110 bar
Table 3 Estimated world production of crude resin, rosin and turpentine Production (Tonnes)
Crude resin
Rosin
Turpentine
World total production of which: China, People’s Rep. of Indonesia Russia Brazil Portugal India Argentina Mexico Honduras Venezuela Greece South Africa Viet Nam Others
976,000
71,7000
99,400
1993
570,000
43,0000
50,000
1993 1992 1993 1992 1994 1993 1991 1992 1993 1993 1993 1990
100,000 90,000 65,000 30,000 30,000 30,000 30,000 8000 7000 6000 2000 2000 6000
69,000 65,000 45,000 22,000 21,000 21,000 22,000 6000 5000 4000 1500 1500 4000
12,000 9000 8000 5000 4000 4000 4000 1000 800 600 200 200 600
þ ð0:15Þ2 þ ð0:2Þ2 þ ð0:2Þ2 þ ð1Þ2 þ ð0:15Þ2 þ ð1Þ2 g ¼ 3%. Using the calculation procedure, the total uncertainty for the whole experiment is obtained to be 73%. 6. Test fuels The test fuels for the present study were turpentine (primary fuel) and high-speed diesel with sulphur contents 0.04 ppm. The resource of turpentine oil in the world is shown in Table 3. Table 4 shows the physical and chemical properties of turpentine oil [8].
Electro chemical sensor. k-type (Cr Al) thermocouple Magnetic pick up type Strain gauge type load cell
Magnetic pick up type
Table 4 Property comparison of turpentine with existing petro fuels
Formula Molecular weight Composition (% wt) Density (kg/m3) Specific gravity Pour point (1C) Boiling point (1C) Vapor pressure (kPa) Viscosity c St @ 30 1C Latent heat of vaporization (kJ/kg) Lower heating value (kJ/ kg) Flash point (1C) Auto ignition temperature (1C) Flammability limit (% volume) Stoichiometric air–fuel ratio Flame speed m/s Cetane number
Gasoline
Diesel
Turpentine
C4 to C12 105 C 88 H 15 780 0.78 40 30–220 48–103
C10H16 136 C 88.2 H 11.8 860–900 0.86–0.9
350
C8 to C25 200 C 87 H 16 830 0.83 23 180–340 o1 3–4 230
43,890
42,700
44,400
43 300–450
74 250
38 305
1.4
1.0
0.8
14.7
14.7
150–180 o1 2.5 285
4–6 40–55
7. Performance analysis The variation of SFC with brake power output is shown in Fig. 2. Since the heat content of the turpentine is higher than that of diesel fuel, a very minimum hike in fuel consumption is obtained with DF mode. In DF mode a notable increase in fuel consumption is obtained between 75% load and full load. The main reason for the increased fuel consumption is due to the beginning of knock at 75% load. This is clearly seen in P–y diagram (Fig. 6). Unusual hike of peak pressure, more fluctuation in cylinder pressure
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0.5
35 B.TH. Efficiency in %.
Diesel baseline Dual fuel mode
SFC in Kg/Kwhr
0.45
0.4
0.35
30 25 20 15 10
Diesel baseline
5
Dual fuel mode
0
0.3
25
0
0.25
50
75 Load in %
100
125
Fig. 4. Variation of brake thermal efficiency with brake power.
0.2 0
25
50
75 Load in %
100
500
125
450
Fig. 2. Variation of SFC with brake power.
Fuel consumption in Kg/hr
Pilot fuel quantity 0.8
Primary fuel quantity
EGT in Degrees
1
Diesel baseline Dual fuel mode
400 350 300 250 200 150
0.6
100 0
0.4
25
50 75 Load in %
100
125
Fig. 5. Variation of EGT with brake power.
0.2
0 Load 0% 25% 50% 75% 100% Fig. 3. Comparison of fuel consumption of pilot fuel and primary fuel.
and pinking noise are the indicators of knock occurrence. The abnormal combustion of charge at full load leads to the poor fuel utilization and increased fuel consumption. More admission of octane fuel (primary fuel) at higher load ranges increases the tendency of knock. However, in the other work [5], the authors have reported that with the help of a suitable diluent the knock can be controlled to a very large extent (Table 2). Fuel consumption of DF mode is clearly depicted in Fig. 3. It shows that almost constant fuel supply is maintained by pilot-fuel supply system irrespective of load but the primary-fuel (turpentine) consumption is gradually increases as load changes from 25% to 100%. The maximum diesel replacement by turpentine at 75% load is approximately 75% of total fuel consumption, after that, a sudden hike in fuel consumption was recorded due to abnormal combustion (knocking). The variation of brake thermal efficiency with brake power output is shown in Fig. 4. It is clear that the brake thermal efficiency of the DF engine is very close to DBL
upto 75% load, beyond which it starts decreasing from the DBL. The maximum efficiency of DF engine is 29%. This is obtained at 75% of full load. The reason for the decrease of brake thermal efficiency beyond 75% load is the occurrence of knock. More admission of octane fuel at the time of full load increases the tendency of knock. The abnormal combustion of charge at full load leads to poor fuel utilization and increased fuel consumption. This leads to poor break thermal efficiency between 75% load, and full load. However, at 75% load the break thermal efficiency of DF mode is better than that of DBL. This is due to higher heating value of the turpentine oil, more charge homogeneity at the end of the compression stroke and rapid burning of primary fuel. Usually in gaseous fuel DF mode, the thermal efficiency at lower loads will be lower than that of DBL due to sluggish combustion of more diluted fuel air mixture [7]. But in the present case, it is not so, instead, turpentine DF mode maintains the same brake thermal efficiency as obtained in the DBL. The variation of EGT with brake power output is shown in Fig. 5. The EGT of turpentine–diesel DF engine is lower than that of DBL upto 75% load beyond which it starts increasing more steeply than DBL. The occurrence of knock in between 75% and 100% load increases the fuel
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Cylinder pressure in bar
Volumetric efficiency in %
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92 90 88
Diesel baseline Dual fuel mode
86 84 0
25
50
75 Load in %
100
125
100 80 60 40 20
Diesel baseline Dual fuel mode
0 320
Fig. 6. Variation of volumetric efficiency with brake power.
340
360
380
400
420
Crank angle in degrees
8. Combustion analysis Fig. 7 shows that P–y diagram of turpentine–diesel DF engine at 75% of full load. From the figure it is observed that the ignition delay of DF engine is considerably longer than that of DBL. This is due to injection of pilot fuel in the environment where the oxygen concentration is less and also the temperature of the mixture at the time of injection is lesser
Fig. 7. Variation of P–y with brake power at 75% load.
300 Net heat release rate in J/Degree
consumption. This unusual behavior of DF mode increases the EGT to a larger extent. More mass burning, uncontrolled combustion and steep increase of mean gas temperature are the major reasons for higher EGT at full load. Another reason for the increased exhaust gas temperature at full load is the higher heat release during diffused combustion stage, higher heat content of turpentine and decreased volumetric efficiency. Usually in all DF engines the primary fuel is inducted along with inlet air and evaporated inside the cylinder by observing heat from the cylinder wall and the gas inside the cylinder. This helps to maintain the cylinder temperature well below self-ignition temperature of primary fuel and resulting low EGT. Usually, fuel that has high latent heat of vaporization will offer low EGT when compared to that of DBL. The variation of volumetric efficiency with brake power output is shown in Fig. 6. The volumetric efficiency of turpentine–diesel DF engine is lower than that of DBL at all loads except at 0% load. The reason for the decreased volumetric efficiency is due to the induction of turpentine along with the inlet air. After induction, turpentine starts vapourizing by absorbing heat from the cylinder wall and reduces the space available for fresh air entry. Hence the volumetric efficiency reduces as the load increases or the percentage of turpentine admission increases. The drastic reduction of volumetric efficiency is observed at full load due to the presence of knock. The occurrence of knock will lead to increased fuel consumption, mass burning of fuel and increased mean gas temperature. This is the major reason for drastic reduction of volumetric efficiency in between 75% load and full load. At full load, approximately 8% drop in volumetric efficiency is obtained.
275 250 225 200
DBL
175 150 DF
125 100 75 50 25 0 340
350
360
370
380
390
400
410
420
Crank angle in degrees Fig. 8. Variation of heat release rate with brake power at 75% load.
than the temperature of the DBL operation. This is due to the evaporation of turpentine oil inside the cylinder by absorbing heat from the combustion chamber. The more ignition delay at the time of higher load range increases the peak pressure than that of DBL and provides more pressure fluctuation. This is the main reason for more waviness in the expansion period of P–y diagram. In other words this period depicts the abnormal combustion or knocking period. However, this is the usual behavior of high-octane fuel in high compression ratio engines. This can be controlled by proper selection of compression ratio and admission of knock suppressers. Fig. 8 compares heat release rate of turpentine–diesel DF engine with DBL at 75% of full load. It shows that DF mode offers higher ignition delay than DBL. Because in DF mode, pilot fuel is injected inside the primary fuel and air mixture, which is kept at a temperature lower than DBL mode at the same load condition. Since the no load rack position was sealed, the premixed stage of combustion is shorter and also occurs 31–41 later than DBL. Most of the heat release is occurred only during the diffusive combustion phase. Longer ignition delay provides the higher heat release during the
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diffusive combustion phase. Hence, the duration of heat release is also longer than DBL, which leads to the increase in EGT. The increase in combustion duration is mainly due to the slow combustion of injected fuel and combustion of inducted turpentine by flame propagation.
9. Emission analysis Fig. 9 compares the emission performance of turpentine–diesel DF engine with DBL at various load conditions. The CO emission of turpentine–diesel DF engine gradually increases from 0% load to 75% load. The reason behind increased CO emission at lighter loads is higher fumigation rate and scarcity of oxygen. The flame quenching and cooled layer of charge near the wall will also be the reason for the increased CO emission. The drastic increase of CO emission at higher loads is due to the occurrence of knock and higher fuel consumption. Knocking leads to poor fuel utilization and increased cylinder temperature, which will
CO in %
6
4
2
Bosch Smoke Number
0 4 3 2 1 0
NOx in PPM
2000 1500 1000 500 0
UBHC in PPM
400 Diesel baseline Dual Fuel mode
300 200
cause more CO emission. Approximately 35% of higher CO emission is observed at 100% load of DF engine. Drastic reduction of volumetric efficiency at full load due to abnormal combustion reduces oxygen availability and hence leads to the higher CO emission. The UBHC emission of turpentine–diesel DF engine gradually increases from 0% load to 75% load. The reason behind increased UBHC emission is higher fumigation rate and scarcity of oxygen. The flame quenching and cooled layer of charge near the wall will be the reason for the increased UBHC emission. Though this is a conventional behaviour of the DF engine, the drastic increase of UBHC emission at higher loads is due to occurrence of knock and higher fuel consumption. At lighter loads due to charge homogeneity and higher oxygen availability the UBHC emission level is less, where as at higher load ranges due to higher quantity of fuel admission, charge temperature and drastically reduced volumetric efficiency, UBHC emission increases. Approximately 45% of higher UBHC emission is noted at 100% load in DF engine. The NOx of turpentine–diesel DF engine closely follows the DBL upto 50% load, beyond which it starts increasing from DBL. The reason for the increased NOx is due to the higher heat release during second stage of combustion, higher combustion temperature and higher heat content of turpentine. Increased ignition delay of the DF engine, which promotes diffusive combustion, is also be another reason for increased NOx. A drastic increase in NOx level is found beyond 75% load due to the abnormal combustion. This leads to a very high gas temperature, increase of fuel consumption and poor utilization of heat. The presence of high gas temperature inside the cylinder will create a conducive ambient for reaction of atmospheric nitrogen with oxygen. This is the major reason for higher NOx emission at higher load ranges. The maximum NOx emission is approximately 21% at full load. The variation of smoke emission with load is shown in Fig. 8. The turpentine–diesel DF engine reduces the smoke to a larger extent. The reason for the reduced smoke emission is the availability of premixed and homogeneous charge inside the engine well before the commencement of combustion. Higher heat content of turpentine, higher combustion temperature, extended duration of combustion and rapid flame propagation are the other reasons for reduced smoke level. The pilot fuel was used in less quantities and this may be one of the reason for the reduction of smoke emission. However, at higher load ranges due to non-availability of sufficient air and abnormal combustion there is a visible white smoke emission. 10. Conclusions
100 0 0
25
50 75 Load in %
100
125
Fig. 9. Variation of emission parameters with brake power.
From the detailed test conducted with turpentine–diesel DF engine the following conclusions are arrived: 1. The brake thermal efficiency of DF engine at 75% load is 1–2% higher than that of DBL.
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2. Increased SFC is reported at full load due to the presence of knock. 3. EGT and NOx are found lower than that of DBL up to 50% load beyond which they start increasing due to the occurrence of knock. 4. Approximately 35% of higher CO emission is reported at full load of DF engine. 5. 48% of higher UBHC emission and 45% reduced smoke are reported at full load of DF engine. 6. Maximum of 8% drop in volumetric efficiency is reported in DF engine at full load. 7. Most of the research papers submitted on the same topic earlier have reported that the DF operation at light loads are less efficient than its diesel counter part; But, this investigation proves that the brake thermal efficiency at lighter loads are equal to that of DBL. 8. At higher loads due to higher admission of octane fuel (primary fuel) the engine starts knocking. Hence the performance of DF engine beyond 75% load starts deteriorating. From the above experiment, it is concluded that 60–65% replacement of diesel with turpentine is quite possible within 75% load. Except increased CO and UBHC emissions, all other emission parameters like smoke and NOx and performance parameter like brake thermal efficiency are found superior to that of DBL.
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References [1] Mayer-Pitroff R. Biofuel: technical. Economic and environmental impacts. In: Elaeis special issue: proceedings of the 1995 PORIM international biofuel conference. Kuala Lumpur: Palm Oil Research Institute of Malaysia; 16–17 January 1995. p. 1–11. [2] Rickeard DJ, Thompson ND. A review of the potential for bio-fuels as transportation fuels. SAE paper 932778, 1993. [3] Masjuki H, Abdulmuin MZ, Sii HS. Indirect injection diesel engine operation on palm oil methyl esters and its emulsions. Proceedings of the institution of mechanical engineers part D. J Automobile Eng 1997;211:291–9. [4] Husna K, Mariani O, Hamzah AH. Biofuel: impact on environment. In: Elaeis special issue: proceeding of the 1995 PORIM international biofuel conference. Kuala Lumpur: Palm Oil Research Institute of Malaysia; 16–17 January 1995. p. 99–109. [5] Karthikeyan R, Mahalakshmi NV. Internal combustion engine division of ASME 2005 fall technical conference, ICEF2005. Ottawa, Canada: Crowne Plaza Hotel; 11–14 September 2005. [6] Senthil Kumar M, Ramesh A, Nagalingam B. Experimental investigations on a jetropa oil methanol dual fuel engine. SAE paper 2001-010153. [7] Senthil Kumar M, Ramesh A, Nagalingam B. An experimental comparison of methods to use methanol and jetropa oil in a compression ignition engine. Biomass Bioenergy Paper 2003;25: 309–18. [8] ACGIH. Industrial ventilation—a manual of recommended practice. Cincinnati, OH: American Conference of Governmental Industrial Hygienists; 1986.