Effect of EGR and injection timing on combustion and emission characteristics of split injection strategy DI-diesel engine fueled with biodiesel

Fuel 90 (2011) 1884–1891

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Effect of EGR and injection timing on combustion and emission characteristics of split injection strategy DI-diesel engine fueled with biodiesel Donghui Qi a,⇑, Michael Leick b, Yu Liu c, Chia-fon F. Lee b a

Key Laboratory of Automotive Transportation Safety Technology Ministry of Communication, Chang’an University, Xi’an 710064, China Department of Mechanical Science and Engineering, The University of Illinois at Urbana-Champaign, Urbana, IL 61801, USA c School of Automotive Engineering, Jilin University, Changchun 130012, China b

a r t i c l e

i n f o

Article history: Received 3 September 2010 Received in revised form 22 December 2010 Accepted 10 January 2011 Available online 25 January 2011 Keywords: Biodiesel Exhaust gas recirculation Split injection strategy Combustion Emission

a b s t r a c t In this study, the effect of injection timing and EGR rate on the combustion and emissions of a Ford Lion V6 split injection strategy direct injection diesel engine has been experimentally investigated by using neat biodiesel produced from soybean oil. The results showed that, with the increasing of EGR rate, the brake specific fuel combustion (BSFC) and soot emission were slightly increased, and nitrogen oxide (NOx) emission was evidently decreased. Under higher EGR rate, the peak pressure was slightly lower, and the peak heat release rate kept almost identical at lower engine load, and was higher at higher engine load. With the main injection timing retarded, BSFC was slightly increased, NOx emission was evidently decreased, and soot emission hardly varied. The second peak pressure was evidently decreased and the heat release rate was slightly increased. Ó 2011 Elsevier Ltd. All rights reserved.

1. Introduction Today, air pollutants emitted from diesel vehicles, such as NOx and particulate matter (PM), have created serious air pollution problems in major cities around the world. Regulations and control measures aimed at lowering exhaust emissions from truck and bus diesel engines have been adopted in an effort to improve air quality in cities. However, the degree of improvement seems to be not very satisfactory, mostly because of the difficulties in removing NOx and the compromise between PM and NOx emissions. Over the past years, the investigations on diesel engines have expanded in the area of alternative fuels, among which biodiesel represents a very promising fuel. Pure biodiesel is a fuel composed of mono-alkyl esters of long chain fatty acids derived from vegetable oils or animal fats. Biodiesel is a sulfur-free, nontoxic, biodegradable, oxygenated, and renewable fuel and can be used in standard diesel engines with little or no engine or fuel system modifications. In comparison to diesel fuel, biodiesel has comparable energy density and cetane number, little sulfur and much oxygen [1–5]. However, high viscosity, high molecular weight, low volatility, etc. of biodiesel may in some cases lead to problems such as severe engine deposits, injector cooking, and piston ring sticking [6–9]. In general, biodiesel provides comparable fuel efficiency and horsepower. Using biodiesel instead of diesel fuel reduces emissions such as the overall life cycle ⇑ Corresponding author. Tel.: +86 29 82334464; fax: +86 29 82334476. E-mail address: [email protected] (D. Qi). 0016-2361/$ - see front matter Ó 2011 Elsevier Ltd. All rights reserved. doi:10.1016/j.fuel.2011.01.016

of carbon dioxide (CO2), PM, carbon monoxide (CO), sulfur oxides (SOx), volatile organic compounds (VOCs), and unburned hydrocarbons (HC). However, despite its potential benefits of lowering these pollutants, experiments indicated that the NOx emission will be increased with the use of biodiesel [10–17]. The above comparisons between diesel and biodiesel were based on the original setting of an engine using diesel fuel. The formulation of fuel composition can improve the biodiesel combustion performance and exhaust emissions. However, the results showed that it was difficult to acquire NOx emission neutral while improving other pollutant emissions simply by fuel reformulation [18–22]. Therefore, modification of engine parameters may be feasible to optimize the engine emissions due to the difference in chemical composition and combustion characteristics between diesel and biodiesel. These were related, for example, to the injection strategies, or to exhaust gas recirculation (EGR). Recently, a split injection strategy has been proposed as a means to reduce NOx emissions, and this allows the injection to be retarded to reduce NOx emissions without a significant penalty in soot levels. Kim et al. [23] studied the effect of split injection on exhaust emissions, soot particulates, and engine performance using an electrically controlled direct injection diesel engine fueled with neat biodiesel derived from soybean. The results showed that split injection reduces NOx emissions significantly without a significant increase in soot emissions. Decreasing soot, median particle diameter, and particle number concentration were seen in accordance with retarded injection timing for split injection. Zhang and

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Nomenclature ATDC BMEP BSFC CO CO2 DI ECM EGR HC

after top dead center brake mean effective pressure brake specific fuel consumption carbon monoxide carbon oxides direct injection electronic control module exhaust gas recirculation hydrocarbon

Boehman [24] studied the impact of biodiesel on NOx emissions in a 2.5 L common rail DI-diesel engine with different fuel injection strategies. They found that retarding injection timing under single injection conditions was the more effective approach to reduce the NOx emissions than using pilot injection with retarded main injection in terms of NOx and fuel consumption tradeoff. Under the low engine load condition, the pilot injection strategy led to substantially reduced NOx emissions. Others have studied the effects of combining biodiesel and EGR. The general conclusion from these studies was that combining EGR and biodiesel was an effective way to decrease NOx and/or PM. The majority of these studies have been performed on relatively small diesel engines. Tsolakis et al. [25] found that the use of biodiesel fuel could decrease the smoke and NOx from a single-cylinder engine equipped with EGR under certain engine conditions when compared to diesel. The retardation of the injection timing could result in further reduction of NOx at a cost of small increases of smoke and fuel consumption. Prandeep and Sharma [26] varied EGR levels and engine load on a single-cylinder engine and found that biodiesel produced more smoke at lower loads and less smoke at higher loads when compared to diesel fuel. Agarwal et al. [27] tested a two-cylinder engine equipped with EGR and biodiesel fuel and found that the 15% EGR and 20% biodiesel blend was the optimum combination. Saleh [28] studied the effect of EGR on the NOx emissions of twocylinder, naturally aspirated direct injection diesel engine fueled with jojoba methyl ester. They concluded that EGR is an effective technique for reducing NOx emissions with JME fuel especially in light-duty diesel engines. For all operating conditions, a better trade-off between HC, CO and NOx emissions can be attained within a limited EGR rate of 5–15% with very little economy penalty. The work in this paper was a continuation of our experimental and numerical investigations of the influence of neat biodiesel from soybean oil on the characteristics of a diesel engine [29]. Attention was focused on the determination of the injection timing and EGR rate for biodiesel with the aim to reduce all engine harmful emissions, especially NOx and PM.

2. Experimental section

k NOx PM SOx TDC VOCs

coefficient of light absorption of the smoke nitrogen oxide particulate matter sulfur oxides top dead center volatile organic compounds

piezoelectric injectors that had six nozzles each. Depending on the performance and emission requirements, the injection pressure can vary from 0.22 to 1.650 MPa. The number of injections can vary from 1 to 6 for every engine cycle, namely two pilot injections, two main injections, and two post injections. For all the studies done here, there were one pilot injection and one main injection involved. It is known that biodiesel is incompatible with materials such as rubber and brass, and a long-term usage of incompatible materials in a biodiesel engine will result in material degradation or corrosion. The fuel system was built to become full biodiesel compatibility, as all the fuel lines and fittings were carefully chosen to be fully biodiesel compatible. Table 2 shows the main properties of diesel and biodiesel. To determine the desired engine parameters, the engine comes with an Electronic Control Module (ECM), which controls the entire basic engine operating parameters such as the injection timings, injected fuel mass and EGR using Ford’s configurations. However, those parameters can be reconfigured by using ETAS INCA software, which is used for the development and calibration of the control and diagnostic parameters in the ECM. With INCA, data acquisition and real-time recording of many engine operating conditions present in the ECM can be realized. In other words, the internal calibrations of the ECM can be changed depending on the experimental setup. When starting the ECM, the engine runs with the default calibrations supplied by Ford. After the connection is established between INCA and ECM, the calibrations can be

Table 1 The specifications of test engine. Engine type

Ford Lion 6-cylinder 4-stroke direct injection

Bore (mm) Stroke (mm) Connecting rod length (mm) Compression ratio Total engine displacement (l) Induction system Fuel injection system Fuel injection type EGR

81 88 160 17.3:1 2.7 Variable geometry turbocharger Common rail (up to 1.65 MPa) Piezoelectric with six nozzles Water-cooled EGR pumps

2.1. Engine and data acquisition The engine used in this study was Ford Lion V6 diesel engine. The engine specifications can be found in Table 1. Although the engine was equipped with a variable geometry turbocharger, the turbocharger was not used. Instead, an independent air supply system was set up to simulate turbocharger conditions to provide intake air to the engine at a more precisely controlled pressure and temperature. The injection system of the engine was a common-rail system, which was capable of supplying a fuel pressure up to 1.65 MPa at high engine load and speed using highly accurate

Table 2 Main properties of diesel and biodiesel. Properties

Diesel

Biodiesel

Specific gravity (kg/l) Lower heating value (MJ kg 1) Viscosity (cS t @40 °C) Cetane number Stoichiometric air–fuel ratio (kg/kg) Sulfur (ppm)

0.85 41.90 2.0–3.0 43–47 14.6 7.0–15.0

0.88 38.37 4.11 47.1 12.6 <3

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Computer #1 PCI-MIO-16E-4 Multi-purpose input/output board

Computer #2 ETAS INCA ES580 interface card

PC-AO-2DC Analog output board

Air pressure control valve

Air heater temperature control

Engine ECM

SCXI-1000 Signal conditioning chassis

SCXI-1100 32-channel multiplexer module

SCXI-1120 8-channel isolated amplifier and multiplexer module

SCXI-1180 Signal feedthrough module

Accelerator pedal position signal

SCXI-1303 Terminal block with CJC sensor

Thermocouples

Shaft Encoder quarter-degree & TDC signals

Steady-state pressure transducer

In-cylinder pressure transducers

Dynamometer speed & torque

Fig. 1. Schematic diagram of engine data acquisition and control system.

modified. In this study, the injection timing and EGR were varied. Fig. 1 shows the schematic diagram of engine data acquisition and control system. National Instruments LabView version 8.6 was used as the data acquisition and timing software. The engine temperatures and pressures were monitored with a multifunction data acquisition board. The necessary timing was controlled by 16 up/down 32-bit counter/timers. An optical shaft encoder was mounted to the crank shaft at the front of the engine to determine the engine timing and locate the top dead center (TDC) of the piston for cylinder #6. The shaft encoder was capable of providing a resolution of 0.25 crank angle degree and therefore able to output 2880 pulses for every engine cycle. An AVL pressure transducer was used to measure the in-cylinder pressure for cylinder #6. It had a bore of 3 mm and fitted into the glow plug hole through the use of a glow plug adaptor. The glow plug heating system was therefore disabled although the rest of the glow plugs were still in the engine. The pressure transducer was connected to a Kistler charge amplifier to amplifier its signal, which was then passed to the data acquisition board for future processing. The cylinder pressure data were recorded for 50 consecutive cycles and then averaged in order to eliminate the effect of cycle-to-cycle variations. On the basis of the in-cylinder pressure data, a zero dimensional, single-zone model was used to calculate the heat release rate and in-cylinder temperature, as described by Heywood [30].

2.2. NOx and smoke analyzer A MEXA-720 NOx analyzer from Horiba was used to measure the NOx concentration from the exhaust pipe. A NOx sensor was installed in the exhaust pipe and connected to the signal analyzer for NOx concentration readings. The analyzer was calibrated according to the operation manual before emission measurement. The smoke meter was operated by manually pumping in the exhaust through a short copper tube connected to the main exhaust pipe. A clean filter paper was placed in the smoke meter, and smoke would be collected on the paper when the exhaust gas entered the smoke meter due to pumping. During an experiment, the smoke meter was pumped for fifty times for each case. The filter paper was then converted to a digital image for future processing. It was noted that the emission tests were performed on the raw exhaust gas coming from the exhaust ports of the engine. No exhaust after treatment system was ever used. 3. Results and discussion 3.1. Test conditions Since the gas mixture in the engine cylinder has a longer residence time with high combustion temperatures under low speed and high load conditions and low speed and low load conditions

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400

BSFC (g/kWh)

represent the typical road-load conditions for automotive diesel engines, it is expected that the effect of EGR and injection timing of biodiesel on NOx emissions will be more significant under these conditions. Therefore, test conditions of 1500 r/min and 0.3 MPa of BMEP and 1500 r/min and 0.6 MPa of BMEP were selected as the engine operating conditions. The impact of biodiesel on engine performance, emissions and combustion under various EGR rate and injection timing conditions was studied. Due to the mechanical delay for the solenoid valve in the fuel injector to respond to the triggering signal sent from the ECM, there was always a delay of the actual fuel injection timing compared to the injection timing calibration value stored in the ECM. The indicated injection timings that represent the injection timing calibration values were used in this study. Detailed engine test conditions are given in Table 3.

Table 3 Engine operating condition (1500 r/min). BMEP

EGR rate (%)

Split injection

0.3 MPa

49 54

Pilot SOI: 14 (ATDC) Main SOI: 4, 2, 0, 2, 4 (ATDC)

0.6 MPa

38 43

Pilot SOI: 16 (ATDC) Main SOI: 4, 2, 0, 2, 4 (ATDC)

300 0.3MPa,49%EGR 0.3MPa,54%EGR 0.6MPa,38%EGR 0.6MPa,43%EGR

250 200 150

-6

-4

-2

0

2

4

6

Main Injection Timing (CAD ATDC) Fig. 2. BSFC under different main injection timing and EGR rate conditions.

NOx (ppm)

3.2. Performance and emissions characteristics

450 400 350 300 250 200 150 100 50 0

0.3MPa,49%EGR 0.3MPa,54%EGR 0.6MPa,38%EGR 0.6MPa,43%EGR

-6

-4

-2

0

2

4

6

Main Injection Timing (CAD ATDC) Fig. 3. NOx emissions under different main injection timing and EGR rate conditions.

2000

In-cylinder temperature (K)

Fig. 2 shows the BSFC of biodiesel with main injection timing under two EGR rate conditions. It can be seen that the BSFC is slightly increased with the retarded main injection timing. It is likely that the turbulence effect of the main injection is reduced when it is phased later in the cycle, which would explain the increase in BSFC observed for the later phased main injections [31]. It also can be seen that the BSFC is slightly increased with higher EGR rate under different main injection timings. Since EGR in a diesel engine displaces a unit of fresh air with an equal unit of burned exhaust products, it not only alters the A/F ratio, but causes a dilution effect. By reducing the oxygen concentration, the mixing time between the direct injected fuel and the fresh oxygen increases and reduces the burn rate once diffusion combustion starts, therefore, making stable combustion more difficult to achieve, and indicating a speed and power output decrease and BSFC increase [28]. Furthermore, at high engine load the temperature of the mixture of EGR and fresh air increases and the cylinder trapped mass decreases, which has a detrimental effect on the volumetric efficiency [32]. So this may result in an increase in BSFC. Fig. 3 shows the NOx emission of biodiesel with the main injection timing under two EGR rate conditions. It illustrates a universal rule that NOx always decreases with retarding main injection timing. The main reason is that it reduces combustion temperature in the cylinder, and moreover, it reduces residence time of high-temperature-burned gas in the combustion chamber where NOx actively forms. The setting of the latest retarded main injection timing has the lowest average cylinder temperature (seen in Fig. 4) and, therefore, the least NOx emission. Fig. 3 also indicates that, with the increase of the EGR rate by 5%, the NOx emission is decreased by about 50% under different main injection timings. It may be suggested that the EGR results in a temperature drop in the burning zone due to a dilution effect, thermal, and chemical effects. The EGR dilutes the oxygen concentration of the intake fluid. Concurrently, the EGR increases the specific heat capacity of the working fluid, thereby reducing the combustion temperature (seen in Fig. 5). Furthermore, the endothermic dissociation of the EGR constituents such as H2O may contribute to the reduction in combustion temperatures [32]. Thus, a 5% EGR rate increasing can be

350

1800 1600

0.6MPa, 43%EGR

1400 1200 1000

-4 CAD ATDC -2 CAD ATDC

800

0 CAD ATDC

600

2 CAD ATDC

400 200 0 -20

4 CAD ATDC -10

0

10

20

30

40

50

60

Crank angle (CAD) Fig. 4. In-cylinder temperatures under different main injection timing conditions.

used for high reduction efficiency of NOx (about 50%) without obvious increase of BSFC (less than 2%). Fig. 6 shows the soot emission of biodiesel with main injection timings under two EGR rate conditions. The split injection strategy exhibits less sensitivity for soot emission to main injection timing; furthermore, retarding main injection timing slightly reduces soot emissions at low engine load. The trend of soot emissions with the EGR rate are observed for biodiesel. It can be seen that the soot emission is slightly increased with higher EGR rate at high engine loads. At high load, the heat capacity increases as the concentrations of CO2 and H2O are substantially higher. Both these molecules have higher heat capacities than air, with the higher heat capacity of the mixture, more energy is required to pre-heat the incoming mixture, thus lowering the flame temperature and deterioration in diffusion combustion [33]. 3.3. Combustion characteristics The in-cylinder pressures of biodiesel under two EGR rate conditions are shown in Fig. 7. As seen in this figure, when the

D. Qi et al. / Fuel 90 (2011) 1884–1891

2000 1800 1600 1400 1200 1000 800 600 400 200 0 -20

Pme=0.6MPa

38% EGR

-10

0

43% EGR

10

20

30

40

50

60

Crank angle (CAD)

200 180 160 140 120 100 80 60 40 20 0

0.3MPa,49%EGR 0.3MPa,54%EGR 0.6MPa,38%EGR 0.6MPa,43%EGR

-6

-4

-2

0

2

4

6

Main Injection Timing (CAD ATDC) Fig. 6. Soot emissions under different main injection timing and EGR rate conditions.

In-cylinder pressure (MPa)

6 5

(a) Pme=0.3MPa

4 3 2

49% EGR 54% EGR

0.08

1 0 -60

-40

-20

0

20

40

60

Crank angle (CAD)

In-cylinder pressure (MPa)

6 5

(b) Pme=0.6MPa

Heat release rate (kJ/CAD)

Soot (a.u.)

Fig. 5. In-cylinder temperatures under different EGR rate conditions.

engine load, the two peak pressures are slightly decreased with higher EGR rate, and the corresponding crank angles are nearly in the same position of crank angle. Fig. 8 gives the heat release rate of biodiesel under two EGR rate conditions. It can be seen that the whole combustion process is composed of pilot heat release and main heat release. At lower engine load, the pilot heat release is hardly varied, but the main heat release is slightly retarded. At higher engine load, the peak of pilot heat release is lower, but that of main heat release is higher with higher EGR rate. For the small amount of fuel injected during pilot injection, there is sufficient oxygen available in the cylinder even at high EGR rate conditions. Hence, oxygen availability is not a limiting factor for the pilot combustion. As the EGR rate is increased, the intake gas temperature increases and the cylinder trapped mass decreases, which increase the bulk cylinder gas temperature. Having a higher bulk cylinder gas temperature can enhance the vaporization of the injected fuel and form a larger amount of premixed combustible mixture before the first stage of combustion. But the chemical reaction rate also slows down a little bit with the higher EGR rate. The above two effects compensate one another. When the EGR rate increases a second effect is more pronounced. Thus, the pilot heat release slightly decreases as the EGR rate is increased at high engine load. At high engine load, the continuous decrease in oxygen availability in the cylinder with the increase of the EGR rate deteriorates the main combustion process, which prolongs the ignition delay and the premixed portion of main combustion starts to rise significantly. So the peak of main heat release is increased with higher EGR rate. Fig. 9 shows the curves of mass fraction burned under two EGR rate conditions. It can be seen that the whole combustion phase is retarded with the EGR rate increasing. Table 4 gives the crank angle corresponding to certain percent of mass fraction burned. The difference of the crank angles between two EGR rate for 10% and 50% mass fraction burned are one or two crank angles, while that of 90% are three or four crank angles. This means that the former half of

4 3

49% EGR

(a) Pme=0.3MPa

54% EGR

0.06 0.05 0.04 0.03 0.02 0.01 0 0

20

40

60

Crank angle (CAD)

2

38% EGR 0.07

43% EGR

1 0 -60

0.07

-0.01 -20

-40

-20

0

20

40

60

Crank angle (CAD) Fig. 7. In-cylinder pressures under different EGR rate conditions.

injection is divided into two events, distinctive double peaks of incylinder pressure are found, the first peak corresponds to the pilot injection, and the second peak reflects the main injection. At lower engine load, the first peak pressure keeps almost the same with the EGR rate, but the second peak pressure is decreased and the corresponding crank angle is retarded with higher EGR rate. At higher

Heat release rate (kJ/CAD)

In-cylinder temperature (K)

1888

0.06

38% EGR

(b) Pme=0.6MPa

43% EGR

0.05 0.04 0.03 0.02 0.01 0 -0.01 -20

-10

0

10

20

30

40

50

Crank angle (CAD) Fig. 8. Heat release rate under different EGR rate conditions.

60

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D. Qi et al. / Fuel 90 (2011) 1884–1891

1.2

In-cylinder pressure (MPa)

Mass fraction burned

1

7 (a) Pme=0.3MPa

0.8 0.6 0.4 0.2

Biodiesel (49% EGR)

0

Biodiesel (54% EGR)

-0.2 -20

0

20

40

6

(a) 0.6MPa, 38%EGR

5

-4 CAD ATDC -2 CAD ATDC 0 CAD ATDC 2 CAD ATDC 4 CAD ATDC

4 3 2 1 0 -90

60

-60

In-cylinder pressure (MPa)

Mass fraction burned

(b) Pme=0.6MPa

0.8 0.6 0.4 Biodiesel (38% EGR) 0.2

Biodiesel (43% EGR)

0 -0.2 -20

0

30

60

90

7

1.2 1

-30

Crank angle (CAD)

Crank angle (CAD)

-10

0

10

20

30

40

50

60

6

(b) 0.6MPa, 43% EGR

5 4 3

-4 CAD ATDC -2 CAD ATDC 0 CAD ATDC 2 CAD ATDC 4 CAD ATDC

2 1 0 -90

Crank angle (CAD)

-60

-30

0

30

60

90

Crank angle (CAD)

Fig. 9. Mass fraction burned under different EGR rate conditions.

Table 4 Crank angle corresponding to certain percent mass fraction burned under two EGR rates. Pme (MPa)

EGR rate (%)

Crank angle for certain percent mass fraction burned (ATDC) 10%

50%

90%

0.3

49 54

10 11

14 15

31 34

0.6

38 43

8 9

14 16

29 33

combustion duration is almost the same for different EGR rates, but the later half is longer for higher EGR rate. Fig. 10 shows the in-cylinder pressure of biodiesel under varied main injection timings at high engine load. The findings are generally the same as at low engine load. It can be observed that the first peak pressures maintain similar values, and the second peak pressure is reduced with the retarded main injection timing. The main reason is that, due to the identical start of pilot injection timing, the first peak pressure and its corresponding crank angle keep almost the same. With the start of main injection timing retarded, the main combustion process is thus shifted away from TDC in the expanding stroke, which results in some efficiency loss (or high BSFC, as illustrated in Fig. 2) and the lower peak pressure. Fig. 11 show the heat release rate of biodiesel under varied main injection timings. It indicates that the peaks of the heat release rate are significantly affected by main injection timing. The ignition delay of the pilot combustion is increased, and the start of main combustion is retarded with the main injection timing. The peak heat release rate of pilot combustion is increased with retarded main injection timing. This caused a longer ignition delay,

Fig. 10. In-cylinder pressures under different main injection timing conditions.

which is the time for fuel vaporization and mixing. As a result, the retarded main injection may cause the formation of a more homogeneous mixture. Fuel quantity during pilot injection is small, making the mixture too lean to burn in some regions of the combustion chamber. This lean and homogeneous mixture causes a lower peak heat release rate, longer burn duration of pilot combustion, and a lower in-cylinder pressure and temperature at the start of main combustion, which cause the longer ignition delay of main combustion. The peak heat release rate of main combustion is increased when the main injection timing is retarded. Remaining fuel–air mixture from pilot combustion may be increased with retarded main injection timing. Less fuel from pilot injection is burned because of the lower in-cylinder temperature. Residual fuel from pilot combustion is burned during main combustion, increasing the peak heat release rate of main combustion. Fig. 12 shows the curves of mass fraction burned under different main injection timing conditions for two EGR rates. Table 5 gives the crank angle corresponding to certain percent of mass fraction burned. It can be seen that the crank angles corresponding to different percent mass fraction burned move backwards with the main injection timing being retarded. The former half of combustion duration is almost the same for different main injection timing conditions, but the later half is longer when the main injection timing is retarded after 2 ATDC.

4. Conclusions In this study, the combustion and exhaust emissions of a Ford Lion V6 split injection strategy direct injection diesel engine were measured for biodiesel at the different EGR rates and main injection timings. From the current study, the following conclusions can be drawn.

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D. Qi et al. / Fuel 90 (2011) 1884–1891 Table 5 Crank angle corresponding to certain percent mass fraction burned under different main injection timings.

Heat release rate (kJ/CAD)

0.07 0.06

(a) 0.6MPa,38%EGR

-4 CAD ATDC -2 CAD ATDC 0 CAD ATDC 2 CAD ATDC 4 CAD ATDC

0.05 0.04 0.03

EGR rate (%)

0.01

0

10

20

30

40

Heat release rate (kJ/CAD)

50%

90%

4 6 8 10 15

11 13 14 16 22

27 29 29 33 40

43

4 2 0 2 4

5 7 9 10 13

12 13 16 17 19

29 30 33 35 39

60

50

Crank angle (CAD) 0.09 0.08 0.07

(b) 0.6MPa, 43%EGR

-4 CAD ATDC -2 CAD ATDC 0 CAD ATDC 2 CAD ATDC 4 CAD ATDC

0.06 0.05 0.04 0.03 0.02 0.01 0 -0.01 -20

-10

0

10

20

30

40

50

60

Crank angle (CAD) Fig. 11. Heat release rate under different main injection timing conditions.

(a) 0.6MPa,38%EGR

The authors wish to express their deep thanks to the Special Fund for Basic Scientific Research of Central Colleges, Chang’an University, No. CHD2010JC020 and the colleagues in the Engine and Fluid Lab. of the University of Illinois at Urbana and Champaign for their help on the engine test.

1 0.8 -4 CAD ATDC -2 CAD ATDC 0 CAD ATDC 2 CAD ATDC 4 CAD ATDC

0.6 0.4 0.2 0 -20

-10

0

10

20

30

40

50

References 60

Crank angle (CAD) 1.2 1

(b) 0.6MPa, 43%EGR

0.8 -4 CAD ATDC -2 CAD ATDC 0 CAD ATDC 2 CAD ATDC 4 CAD ATDC

0.6 0.4 0.2 0 -20

With the main injection timing retarded, the BSFC was slightly increased, NOx emission was decreased. The split injection strategy exhibited less sensitivity for soot emission. The first peak pressure kept identically, the second peak pressure was evidently decreased and the two peak heat release rates were slightly increased. In general, higher EGR rate and retarded main injection timing are effective methods to reduce NOx emission for split injection strategy DI-diesel engine fueled with biodiesel without more penalties of soot emission and BSFC. Acknowledgements

1.2

Mass fraction burned

10% 4 2 0 2 4

0 -10

Crank angle for certain percent mass fraction burned (ATDC)

38

0.02

-0.01 -20

Mass fraction burned

Main injection timing (ATDC)

-10

0

10

20

30

40

50

60

Crank angle (CAD) Fig. 12. Mass fraction burned under different main injection timing conditions.

With EGR rate increasing, the BSFC and soot emissions were slightly increased, and NOx emission was evidently decreased. At low engine load, the peak pressure was decreased except the first peak pressure at low engine load. The two peak heat release rates kept the same. At high engine load, the first peak heat release rate was lower, but the second one was higher with the increase of the EGR rate.

[1] Sharma YC, Singh B, Upadhyay SN. Advancements in development and characterization of biodiesel: a review. Fuel 2008;87:2355–73. [2] Carraretto C, Macor A, Mirandola A, Stoppato A, Tonon S. Biodiesel as alternative fuel: experimental analysis and energetic evaluations. Energy 2004;29:2195–211. [3] Hu ZY, Tan PQ, Yan XY, Lou DM. Life cycle energy, environment and economic assessment of soybean-based biodiesel as an alternative automotive fuel in China. Energy 2008;33:1654–8. [4] Candeia RA, Silva MCD, Filho JRC, Brasilino MGA, Bicudo TC, Santos IMG, et al. Influence of soybean biodiesel content on basic properties of biodiesel–diesel blends. Fuel 2009;88:738–43. [5] Qi DH, Geng LM, Chen H, Bian YZH, Liu J, Ren XCH. Combustion and performance evaluation of a diesel engine fueled with biodiesel produced from soybean crude oil. Renew Energy 2009;34:2706–13. [6] Sarin A, Arora R, Singh NP, Sarin R, Malhotra RK, Kundu K. Effect of blends of Palm–Jatropha–Pongamia biodiesels on cloud point and pour point. Energy 2009;34:2016–21. [7] Geller DP, Goodruml JW. Effects of specific fatty acid methyl esters on diesel fuel lubricity. Fuel 2004;83:2351–6. [8] Joshi RM, Pegg MJ. Flow properties of biodiesel fuel blends at low temperatures. Fuel 2007;86:143–51. [9] Knothe G, Steidley KR. Kinematic viscosity of biodiesel fuel components and related compounds, influence of compound structure and comparison to petrodiesel fuel components. Fuel 2005;84:1059–65. [10] Lin BF, Huang JH, Huang DY. Experimental study of the effects of vegetable oil methyl ester on DI diesel engine performance characteristics and pollutant emissions. Fuel 2009;88:1779–85. [11] Ramadhas AS, Muraleedharan C, Jayaraj S. Performance and emission evaluation of a diesel engine fueled with methyl esters of rubber seed oil. Renew Energy 2005;30:1789–800. [12] Boehman AL, Song J, Alam M. Impact of biodiesel blending on diesel soot and the regeneration of particulate filters. Energy Fuels 2005;19:1857–64. [13] Ferreira SL, dosSantos AM, deSouza GR, Polito WL. Analysis of the emissions of volatile organic compounds from the compression ignition engine fueled by

D. Qi et al. / Fuel 90 (2011) 1884–1891

[14]

[15]

[16]

[17]

[18]

[19]

[20]

[21]

[22]

diesel–biodiesel blend and diesel oil using gas chromatography. Energy 2008;33:1801–6. He C, Ge YS, Tan JW, You KW, Han XK, Wang JF, et al. Comparison of carbonyl compounds emissions from diesel engine fueled with biodiesel and diesel. Atmos Environ 2009;43:3657–61. Sahoo PK, Das LM, Babu MKG, Arora P, Singh VP, Kumar NR, et al. Comparative evaluation of performance and emission characteristics of Jatropha, karanja and polanga based biodiesel as fuel in a tractor engine. Fuel 2009;88:1698–707. Lee SW, Herage T, Young B. Emission reduction potential from the combustion of soy methyl ester fuel blended with petroleum distillate fuel. Fuel 2004;83:1607–13. Ozsezen AN, Canakci M, Turkcan A, Sayin C. Performance and combustion characteristics of a DI diesel engine fueled with waste palm oil and canola oil methyl esters. Fuel 2009;88:629–36. Qi DH, Chen H, Lee CH, Geng LM, Bian YZH. Experimental studies of a naturally aspirated, DI diesel engine fuelled with ethanol–biodiesel–water microemulsions. Energy Fuels 2010;24:652–63. Lin CY, Lin HA. Diesel engine performance and emission characteristics of biodiesel characteristics of biodiesel produced by the peroxidation process. Fuel 2006;85:298–305. Qi DH, Chen H, Matthews RD, Bian YZH. Combustion and emission characteristics of ethanol–biodiesel–water micro-emulsions used in a direct injection compression ignition engine. Fuel 2010;89:958–64. Qi DH, Chen H, Geng LM, Bian YZH, Ren XCH. Performance and combustion characteristics of biodiesel–diesel–methanol blend fuelled engine. Appl Energy 2010;87:1679–86. Keskin A, Gürü M, Altıparmak D. Biodiesel production from tall oil with synthesized Mn and Ni based additives: effects of the additives on fuel consumption and emissions. Fuel 2007;86:1139–43.

1891

[23] Kim MY, Yoon SH, Lee CS. Impact of split injection strategy on the exhaust emissions and soot particulates from a compression ignition engine fueled with neat biodiesel. Energy Fuels 2008;22:1260–5. [24] Zhang Y, Boehman AL. Impact of biodiesel on NOx emissions in a common rail direct injection diesel engine. Energy Fuels 2007;21:2003–12. [25] Tsolakis A, Megaritis A, Wyszynski ML, Theinnoi K. Engine performance and emissions of a diesel engine operating on diesel–RME (rapeseed methyl ester) blends with EGR (exhaust gas recirculation). Energy 2007;32:2072–80. [26] Pradeep A, Sharma RP. Use of HOT EGR for NOx control in a compression ignition engine fuelled with bio-diesel from Jatropha oil. Renew Energy 2007;32:1136–54. [27] Agarwal D, Sinha S, Agarwal AK. Experimental investigation of control of NOx emissions in biodiesel-fueled compression ignition engine. Renew Energy 2006;31:2356–69. [28] Saleh HE. Effect of exhaust gas recirculation on diesel engine nitrogen oxide reduction operating with jojoba methyl ester. Renew Energy 2009;34:2178–86. [29] Stringer VL, Cheng WL, Lee CF, Hansen AC. Combustion and emissions of biodiesel and diesel fuels in direct injection compression ignition engines using multiple injection strategies. SAE Paper No. 2008-01-1388; 2008. [30] Heywood JB. Internal combustion engine fundamentals. New York: McGraw Hill; 1988. [31] Arjan Helmantel, Valeri Golovitchev. Injection strategy optimization for a light duty diesel engine in medium load conditions with high EGR rates. SAE Paper, 2009-01-1441; 2009. [32] Yoon Seung Hyun, Suh Hyun Kyu, Lee Chang Sik. Effect of spray and EGR rate on the combustion and emission characteristics of biodiesel fuel in a compression ignition engine. Energy Fuels 2009;23:1486–93. [33] Raffael Schubiger, Andrea Bertola, Konstantinos Boulouchos. Influence of EGR on combustion and exhaust emissions of heavy duty DI-Diesel engines equipped with common-rail injection systems. SAE Paper 2001-01-3497; 2001.