Improved NOx and Smoke Emission Characteristics of a Biodiesel

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Energy & Fuels 2008, 22, 3798–3805

Improved NOx and Smoke Emission Characteristics of a Biodiesel-Fueled Engine with the Port Fuel Injection of Various Premixed Fuels Xingcai Lu,* Libin Ji, Junjun Ma, and Zhen Huang School of Mechanical and Power Engineering, Key Laboratory for Power Machinery and Engineering of Ministry of Education, Shanghai Jiao Tong UniVersity, Shanghai 200240, The People’s Republic of China ReceiVed July 2, 2008. ReVised Manuscript ReceiVed September 3, 2008

This work investigated the regulated emission characteristics of a biodiesel-fueled direct-injection (DI) engine with the addition of n-heptane, dimethoxymethane (DMM), and ethanol from the intake port on a singlecylinder engine. The effects of physical and chemical properties of the premixed fuel, partial equivalence ratio of biodiesel, and premixed ratio on emissions were evaluated. The experimental results revealed that the overall heat release curve exhibited a three-stage combustion with the addition of n-heptane but only one-stage with the addition of other two compounds. A simultaneous reduction of NOx and smoke opacity was obtained with the premixed fuels, and ethanol premixing showed the most significant effects on the reduction of NOx and smoke opacity. For a fixed partial equivalence ratio of each premixed fuel, premixing n-heptane showed a larger NOx increase slope and premixing DMM showed a larger smoke opacity increase slope with the increase of the partial equivalence ratio of biodiesel. HC emission was mainly influenced by the premixed fuel properties but not the equivalence ratio, and ethanol premixing had larger HC levels than n-heptane and DMM premixed fuels. When the overall equivalence ratio was kept constant, both HC and CO emissions increased but NOx emissions decreased at first up to a critical premixed ratio. As the premixed ratio exceeded the critical value, NOx emissions started to increase gradually, CO emissions began to decrease, but HC emissions almost remained constant. Under the above conditions, smoke opacity always improved with the increase of the premixed ratio.

1. Introduction Modern diesel engines require a clean burning, stable fuel that performs well under a variety of operating conditions. To reduce the threat of global warming caused by fossil fuels, research on alternative renewable fuels has been gaining greater importance worldwide. Biodiesel is an alternative fuel that can be used directly in any existing, unmodified diesel engine. Many studies revealed that the combustion of biodiesel fuel in compression ignition engines in general results in lower particulate matter (PM), CO, SOx, and HC emissions compared to standard diesel fuel combustion where the engine efficiency is maintained at the same level or slightly improved.1 However, because of the differences in cetane number, oxygen content, and other chemical and physical properties between the biodiesel and diesel fuel, the NOx emissions of pure biodiesel-fueled engines increase about 11-13%. To meet the strict emission standard, such as EURO 5 and Tier 2 of U.S.A., for light- and heavy-duty diesel engines, it is necessary to reduce the NOx emissions of biodiesel-fueled engines significantly. Many researchers have conducted studies on the NOx formation mechanism, control strategies, and suppression methods for biodiesel engines. Boehman et al.2 revealed that the higher bulk modulus of compressibility of vegetable oils and their methyl esters lead to an advanced * To whom correspondence should be addressed. Telephone: +86-2134206039. Fax: +86-21-34206139. E-mail: [email protected]. (1) Fernando, S.; Hall, C.; Jha, S. NOx reduction from biodiesel fuels. Energy Fuels 2006, 20 (1), 376–382. (2) Boehman, A. L.; Morris, D.; Szybist, J.; Esen, E. The impact of the bulk modulus of diesel fuels on fuel injection timing. Energy Fuels 2004, 18, 1877–1882.

injection timing of about 1-4 °CA and confirmed that the increase in NOx emission of biodiesel is related to the advance in fuel injection timing. McCormick et al.3 conducted an experimental study using a variety of real-world feedstock as well as pure (technical-grade) fatty acid methyl and ethyl esters to understand the impact of the biodiesel chemical structure, specifically the fatty acid chain length, and the number of double bonds on emissions of NOx. Agarwal et al.4 researched the effects of exhaust gas recirculation (EGR) on the NOx emissions level and brake-specific energy consumption of the biodiesel engines. Fernando1 reviewed the techniques that have been attempted to reduce NOx emissions of a biodiesel-fueled engine. Hess et al.5 proposed an interesting method to control the NOx formation during the combustion process. Several antioxidants, which are capable of terminating some kinds of radical reactions, were added to 20% soy biodiesel/80% diesel blends (B20) at a concentration of 1000 ppm antioxidants. The results showed that the addition reduced NOx emissions at a moderate level. Furthermore, a new combustion model named “low-temperature combustion” (LTC) was used by Zheng et al. to control the (3) McCormick, R. L.; Graboski, M. S.; Alleman, T. L.; Herring, A. M. Impact of biodiesel source material and chemical structure on emissions of criteria pollutions from a heavy-duty engine. EnViron. Sci. Technol. 2001, 35, 1742–1747. (4) Agarwal, D.; Sinha, S.; Agarwal, A. K. Experimental investigation of control of NOx emissions in biodiesel-fueled compression ignition engine. Renewable Energy 2006, 31, 2356–2369. (5) Hess, M. A.; Haas, M. J.; Foglia, T. A.; Marmer, W. N. Effect of antioxidant addition on NOx emissions from biodiesel. Energy Fuels 2005, 19, 1749–1754.

10.1021/ef800526e CCC: $40.75  2008 American Chemical Society Published on Web 10/04/2008

ImproVed Emissions of a Biodiesel-Fueled Engine

NOx emissions from the biodiesel engine.6 As result, simultaneous reduction of NOx and soot was achieved by more than 50% EGR at low-load conditions. According to the flame model of the direct-injection (DI) combustion process,7 it is very difficult to suppress the NOx and soot emissions to very low levels simultaneously because of the “trade-off” relationship between the NOx and soot emissions. Therefore, it is necessary to develop a new combustion method to simultaneously improve the NOx and soot emissions for biodiesel engines. Recently, a new combustion mode, named as the homogeneous charge compression ignition (HCCI) combustion, is widely researched worldwide.8-10 It features a leaner and homogeneous fuel/air mixture, compression autoignition, and low-temperature combustion. Consequently, it shows substantial reduction in both NOx and PM, while still providing diesel-like efficiency. However, some problems, including the control of ignition timing and combustion rate, still remain in the operation of HCCI engines. To solve these problems, partial HCCI or compound HCCI combustion was proposed by some researchers.11-14 This concept is a compromise to the potential fully premixed HCCI concept. One advantage of such configuration is that, by partially premixing the fuel, the non-premixed portion of the combustion event is reduced. In the present study, the partial HCCI or compound HCCI combustion mode was introduced into the biodiesel engines. By this method, the authors hoped to simultaneously reduce the NOx and soot emissions of the biodiesel engine. To achieve this target, a moderate amount of premixing fuel with lower boiling point, such as n-heptane, DMM, and ethanol, was injected into the intake port and used as a partial substitute of biodiesel. As a result, a leaner homogeneous fuel/air mixture could be formed because the premixing by port fuel injection allowed sufficient time for evaporation and air mixing. Near the top dead center (TDC), the biodiesel fuel was injected into the in-cylinder by the original diesel injector. The experimental investigations were carried out under various premixed quantities for different premixed fuels with the same DI biodiesel fuel and various DI biodiesel fuel quantities with the same premixed fuels from the intake port. Through this controllable premixed (6) Zheng, M.; Mulenga, M. C.; Reader, G. T.; Wang, M.; Ting, D. S.K.; Tjong, J. Biodiesel engine performance and emissions in low temperature combustion. Fuel 2008, 87, 714–722. (7) Flynn, P. F.; Durrett, R. P.; Hunter, G. L.; Loye, A. O. Z.; Akinyemi, O. C.; Dec, J. E.; Westbrook, C. K. Diesel combustion: An integrated view combining laser diagnostics, chemical kinetics, and empirical validation. SAE Tech Pap. 1999-01-0509, 1999. (8) Sjober, M.; Dec, J. E.; Cernansky, N. P. Potential of thermal stratification and combustion retard for reducing pressure-rise rates in HCCI engines, based on multi-zone modeling and experiments. SAE Tech. Pap. 2005-01-0113, 2005. (9) Kim, M. Y.; Kim, J. W.; Lee, C. S. Effect of compression ratio and spray injection angle on HCCI combustion in a small DI diesel engine. Energy Fuels 2006, 20 (1), 69–76. (10) Lu, X. C.; Ji, L. B.; Zu, L. L.; Hou, Y. C.; Huang, C.; Huang, Z. Experimental study and chemical analysis of n-heptane homogeneous charge compression ignition combustion with port injection of reaction inhibitors. Combust. Flame 2007, 149, 261–270. (11) Inagaki, K.; Fuyuto, T.; Nishikawa, K.; Nakakita, K. Dual-fuel PCI combustion controlled by in-cylinder stratification of ignitability. SAE Tech. Pap. 2006-01-0028, 2006. (12) Kim, D. S.; Lee, C. S. Improved emission characteristics of HCCI engine by various premixed fuels and cooled EGR. Fuel 2006, 85, 695– 704. (13) Kim, A. S.; Kim, M. Y.; Lee, C. S. Effect of premixed gasoline fuel on the combustion characteristics of compression ignition engine. Energy Fuels 2004, 18, 1213–1219. (14) Simescu, S.; Fiveland, S. B.; Dodge, L. G. An experimental investigation of PCCI-DI combustion and emissions in a heavy-duty diesel engine. SAE Tech. Pap. 2003-01-0345, 2003.

Energy & Fuels, Vol. 22, No. 6, 2008 3799 Table 1. Physical and Chemical Properties of Biodiesel and Premixed Fuels biodiesel

molecular weight (g) density (g/mL) viscosity (mm2/s) boiling point (°C) autoignition (°C) latent evaporation heat (kJ/kg) (20 °C) low heating value (MJ/kg) cetane number oxygen content (%)

n-heptane

DMM

ethanol C2H5OH

0.875 5.78 >320 316

100 0.692 0.39 98 235 317

CH3O CH2OCH3 76 0.860 0.34 42 237 365

38.5

44.5

22.4

27

49.9 11

56 0

25-30 44.1

5-8 34.8

molecule formula

nC7H16

46 0.794 1.06 78 423 856

Table 2. Specifications of the Single-Cylinder Engine bore (mm) × stroke (mm) displacement (L) nozzle number and diameter (mm) inlet valve open (CA BTDC) inlet valve close (CA ABDC) rate power (kW) and speed (rpm)

98 × 105 0.782

compression ratio

needle open pressure (MPa) 5 × 0.24 fuel delivery advance angle (CA BTDC) 16 exhaust valve open (CA BBDC) 52 exhaust valve close (CA ATDC) 14 at 3200 maximum torque (N m) and speed (rpm)

18.5 24 9 66 12 45 at 1800

combustion system, the simultaneous reductions of NOx and smoke opacity from the biodiesel engine were verified and the factors influencing the combustion and emission characteristics were evaluated. 2. Experimental System and Premixed Fuels In this paper, the main injected fuel is biodiesel, which is made from soybean. The physical and chemical properties of biodiesel and premixed fuels are presented in Table 1. The research engine was based on a single-cylinder, four-stroke, natural aspirated DI diesel with 782 cc of piston displacement, and the engine speed was fixed at 1800 rpm in this work. The engine was coupled to an electrical eddy dynamometer through which load was applied by increasing the field voltage. The detailed engine specifications are shown in Table 2. The experimental system is shown in Figure 1. Premixed fuel was injected into the intake port by an electronic fuel injector at the location of approximately 0.35 m upstream to the intake port, so that the leaner homogeneous fuel/air mixture could be formed during the intake stroke and compression stroke. Near the top dead center (TDC), the biodiesel fuel was directly injected into the combustion chamber by the original diesel engine injection system. In this work, the injection amount of the premixed fuel was controlled by an additional universal ECU, which was synchronized with an engine encoder and various sensors, including engine speed sensor and oil and coolant temperature sensor. The measured parameters and their accuracies are shown in Table 3. The cylinder pressure was measured by a pressure transducer (Kistler model 6125A). The charge output from this transducer was converted to an amplified voltage using a charge amplifier (Kistler model 5015). Pressure data were recorded using a high-speed memory (Yokogawa GP-I). At each test point, the 1440 pulses per rotation (4 pulses per crank angle) from a shaft encoder on the engine crankshaft were used as the data acquisition clocking pulses to acquire the cylinder pressure data. In each operating condition, the cylinder pressures recorded at each crank angle were averaged over 50 consecutive cycles for the experiment. For all data presented, the 0 CA was defined as top dead center (TDC) at the compression stroke. According to the averaged in-cylinder gas pressure, the heat release curve at each operating point could be calculated by a zero-dimension combustion model.15

3800 Energy & Fuels, Vol. 22, No. 6, 2008

Lu et al.

Figure 1. Schematic of the experimental system.

CO, HC, and NOx emissions were measured by an analyzer (AVL Digas 4000). Smoke opacity was measured by a smoke meter (AVL 439).

3. Results and Discussion In this paper, the premixed ratio (PI) is defined as the ratio of the cycle energy of premixed fuel to the total energy, which includes premixed and direct-injected fuel. The PI can be calculated using the following formula: PI )

m ˙ 1Hu1 × 100% m ˙ 1Hu1 + m ˙ 2Hu2

where m ˙ 1 and m ˙ 2 are the fuel consumption rate for premixed fuel and biodiesel, respectively, and Hu1 and Hu2 are the lower heating values of premixed fuel and biodiesel, respectively. In the case of PI equal to 0, it means that there is no premixed fuel and the engine is running with neat biodiesel. During the experimental process, the air flow and fuel consumption rates of each test fuel were measured. Consequently, the overall equivalence ratio (φ), the partial equivalence ratio of premixed fuel (φ1), and the partial equivalence ratio of biodiesel (φ2) can be obtained using the following formula: ˙ 2AF2)/Gair φ2 ) (m φ1 ) (m ˙ 1AF1)/Gair φ ) φ1 + φ2 where Gair is the mass flow rate of intake air at each test point and AF1 and AF2 denote the stoichiometric A/F ratio of premixed fuel and biodiesel, respectively. 3.1. Effects of Premixed Fuel Properties. Figure 2 presents the comparison of the in-cylinder gas pressure, heat release rate, and the mean gas temperature under the same overall equivalence ratio and partial equivalence ratio for various premixed fuels. Because of the differences in physical and chemical properties, the combustion events display different characteristics. (1) With the port fuel injection (PFI) of the n-heptane, the overall combustion process exhibits a three-stage heat release, which includes the low-temperature reaction (LTR, the firststage combustion), high-temperature reaction (HTR, the secondstage combustion), and diffusion combustion (the third-stage

Table 3. Accuracies of the Measurements measured parameters

unit

measurement range

accuracy

engine speed engine torque cylinder pressure oil inlet temperature oil outlet temperature coolant inlet temperature coolant outlet temperature CO emission HC emission NOx emission fuel consumption

r/min Nm MPa °C °C °C °C % ppm ppm kg/h

0-10000 0-200 0-25 0-150 0-150 0-150 0-150 0-4.0 0-10000 0-4000 0-20

(1 (0.1 (0.0005 (1 (1 (1 (1 (0.01 (1 (1 (0.01

combustion). However, with the PFI of DMM and ethanol, the combustion events display a single-stage heat release. (2) Because of the heat release during the LTR and HTR of n-heptane, the initial temperature and pressure corresponding to the start of the third-stage combustion show a noticeable increase compared to those of the other two combustion events. (3) In comparison to the neat biodiesel engine, premixing ethanol always leads to the delaying of the ignition. With the premixing of DMM, the ignition timing delays at lower overall equivalence ratio but slightly advances in the larger overall equivalence ratio on the basis of the neat biodiesel engine. The remarkable difference in combustion characteristics is dominated by the physical and chemical properties of the premixed fuels. It is well-known that the n-heptane is a primary reference fuel with a cetane number of 56. The leaner n-heptane/ air mixture can be easily operated in a HCCI engine. The addition of premixed n-heptane activates the low- and hightemperature chemical reactions and results in a slightly faster initiation of the heat release. Hence, it prevents the sharp increase of the heat release in the premixed burn region of conventional biodiesel combustion. On the other hand, ethanol is a renewable biofuel with a super low cetane number and a larger latent heat value of vaporization. As a result, the combustion event of the biodiesel engine with port injection of ethanol occurs very late. This promotes the sharp heat release in the premixed burn region of the overall combustion event. It is very interesting to find that the combustion characteristics with the premixed of DMM exhibit another phenomenon, which (15) Heywood, J. B. Internal Combustion Engine Fundamentals; McGraw-Hill Book Company: New York, 1988.


Energy & Fuels, Vol. 22, No. 6, 2008 3801

Figure 2. Effects of the premixed fuels on the combustion characteristics.

Figure 3. Effects of premixed fuels on the in-cylinder pressure rising rate.

is different from the above two trends. Table 1 shows that the cetane number of DMM is about 30 and the autoignition temperature is close to that of the n-heptane. According to the chemical kinetics, it is not easy for DMM to activate a lowtemperature reaction under the HCCI-operating conditions. The above analysis shows that premixing DMM from the intake port inhibits the ignition in the lower overall equivalence ratio but promotes the combustion event in the larger equivalence ratio. Figure 3 displays the in-cylinder pressure rise rate corresponding to the conditions in Figure 2. The maximum pressure rise rate of the biodiesel engine without any premixed fuel is about 0.4 MPa/°CA under two conditions. With the addition of n-heptane from the intake port, the maximum gas pressure rise rate is depressed significantly because of the distributed heat release during the overall combustion event. It should be noted that, once the n-heptane concentrations attain a certain value, the “knock combustion” may be observed because of the sharp heat release in the HTR. Under these conditions, the maximum pressure rise rate is significantly larger than that of neat biodiesel fuel without port fuel injection. With the port injection of

ethanol, as mentioned above, the ignition timing and combustion event occur during the expansion stroke and the crank angle corresponding to the maximum pressure rise rate is greatly delayed compared to that of the other three combustion events, while the maximum value of the pressure rise rate is almost the same level with that of the neat biodiesel. For PFI of DMM, the maximum pressure rise rate is obviously larger than that of the other three combustion events in a higher overall equivalence ratio. Also, the crank angle corresponding to this value advances on the basis of biodiesel fuel. Figure 4 illustrates the engine emissions in the same overall equivalence ratio and partial equivalence ratio for different premixed fuels. With the addition of premixed fuels, both the CO and HC emissions are much larger than those of the engine without premixed fuel. It can be found however that both the overall equivalence ratios and cetane number of premixed fuel have important effects on CO emissions. The hydrocarbon emissions increase with the PFI amount and do not exhibit the trend observed in CO previously discussed. It is very likely that premixed fuel is trapped in the crevices


Lu et al.

Figure 4. Effect of premixed fuels on the engine emissions.

Figure 5. Effect of premixed fuels on the indicated thermal efficiency. Figure 6. Definition of some combustion parameters.

during compression and that this fuel is outgassed late in the cycle when the temperatures are too low for reaction. This is similar to the situation encountered in spark-ignition engines. It can be found in Figure 4b that HC emissions increase with a large magnitude with the decrease of the cetane number of the premixed fuel but show a moderate sensitivity to the overall equivalence ratio. Figure 4c compares the NOx emissions of biodiesel engine with and without premixed fuels. It can be found that NOx emissions may be reduced with the addition of premixed fuels from the intake port and the reduced magnitudes increase with the decrease of the cetane number of the premixed fuel. In general, NOx emissions are found to be strongly dependent upon the combustion temperature. However, in the equivalence ratio of 0.6, the maximum in-cylinder gas temperature increases in the order: Tethanol < Tn-heptane, Tbiodiesel < TDMM. It can be seen from Figure 2 that, in the same overall equivalence ratio at 0.4, the maximum in-cylinder gas temperature increases in the order: Tbiodiesel < Tethanol < Tn-heptane < TDMM. It can be concluded then that the NOx emissions levels are not only dominated by the maximum gas temperature but also by the homogeneity of the fuel/air mixture, combustion mode, and physical properties of the premixed fuels. With the addition of premixed fuels from the intake port, the directly injected fuel mass decreases. As a result, the overall combustion event may be dominated by the

premixed burn but not the diffusion burn. This will have positive effects on the inhibition of the NOx emissions. Figure 4d compares the smoke opacity of the premixed combustion in two different equivalence ratios on the basis of the neat biodiesel engine. It is obvious that the engine smoke opacity can be improved substantially with the addition of premixed fuels. Because of the large difference in physical and chemical properties of the premixed fuels, the reduced magnitudes also show a large difference. Mechanisms of soot formation and oxidization in partial or full HCCI engines are influenced by many factors and are too complicated to be fully understood. The HCCI combustion, in itself, of premixed charge should produce low or no soot emissions; therefore, the smoke opacity with the addition of n-heptane is mainly produced in the diffusion combustion stage of the biodiesel. With the addition of DMM and ethanol fuel, the combustion efficiency of the local rich fuel/air mixture may be improved remarkably and the soot oxidation may be more complete because of the oxygen content in both of the premixed fuels. Figure 5 gives the comparison of the indicated thermal efficiency of biodiesel engine with and without premixed fuels. It can be seen from this figure that the biodiesel engine with port fuel injection of DMM shows a larger indicated thermal efficiency under the two operation conditions. While, with the port fuel injection of n-heptane and ethanol, the indicated thermal efficiency is lower than that of neat biodiesel fuel.


Figure 7. Effect of the partial equivalence ratio of biodiesel on the maximum gas temperature.

3.2. Effect of the Partial Equivalence Ratio of Biodiesel. The partial equivalence ratio of biodiesel fuel has important effects on the combustion and emission characteristics. Because the NOx emissions are found to be strongly dependent upon the peak value and the crank angle duration above 1800 K of the gas temperature, some parameters related to the NOx formation are defined in Figure 6. With the maximum gas temperature Tmax, θ1 is defined as the crank angle when the gas temperature attains to 1800 K on the rising side of the temperature curve, while θ2 is defined as the crank angle when


the gas temperature attains to 1800 K on the falling side of the temperature curve. δ is defined as the interval angle between θ1 and θ2. Figure 7 gives the maximum gas temperature as a function of the partial equivalence ratio of biodiesel with the same partial equivalence ratio of premixed fuels. It is obvious that the gas temperature rises with the increase of the directly injected fuel quantity. However, with the addition of DMM from the intake port, the gas temperature is much higher than that of the other two premixed fuels. Figure 8 shows θ1 and δ versus the partial equivalence ratio of biodiesel for different premixed fuels. According to the definition, θ1 can be used to denote the initial crank angle corresponding to the start of NOx production during the combustion process. In the lower equivalence ratio of biodiesel, θ1 occurs very late in the expansion stroke and then quickly moves to the TDC with the increase of the biodiesel fuel. Also, δ increases almost linearly with the increase of the biodiesel fuel quantity. However, with the addition of DMM, δ shows a longer crank angle compared to the other two premixed fuels, as can be seen in Figure 8b. Figure 9 displays the regulated emissions as a function of φ2 while keeping φ1 at a constant for different premixed fuels. With the increase of φ2, both the engine power and in-cylinder gas

Figure 8. θ1 and δ versus the partial equivalence ratio of biodiesel for different premixed fuels.

Figure 9. Regulated emissions as a function of φ2.


Lu et al.

Figure 10. Effects of the premixed ratio on engine emissions.

temperature increase. Therefore, it is well-understood that the CO emissions decrease linearly with the increase of φ2 because of the addition of n-heptane and DMM from the intake port. However, with the premixing of ethanol, CO emissions increase at first up to a critical φ2, i.e., 0.3 in this experiment. As the equivalence ratio exceeds this critical value, CO emissions start to decrease gradually. Increased CO concentrations at lower φ2 under these conditions may result from the incomplete oxidation of premixed ethanol. The hydrocarbon emissions lightly decrease with the increase of φ2 and exhibit another trend observed in CO previously discussed, as can be seen in Figure 9b. It can be found that premixing ethanol produces the highest HC levels compared to premixing DMM and n-heptane. This significant difference in CO emissions may be attributed to the fact that the autoignition temperature of ethanol is much higher than that of DMM and n-heptane. Figure 9c shows the NOx emissions versus φ2 at a fixed φ1 for different premixed fuels. The general tendency is that NOx emissions increase with the increase of φ2. However, from lower to medium φ2, premixing DMM shows larger NOx emissions than premixing ethanol and n-heptane, while premixing nheptane shows a larger increase slope and attains the largest NOx emissions levels at higher φ2. Figure 9d shows the smoke opacity as a function of φ2 at a fixed φ1 for different premixed fuels. At lower φ2 (less than 0.3), there is no obvious difference in smoke opacity. At medium-large φ2, premixing DMM exhibits the largest smoke opacity increase of the slope compared to the other two premixed fuels. This can be attributed to the following reasons: with the premixing DMM, the diffusion burn fraction dominates the overall combustion event and the ignition timing in the larger equivalence ratio occurs near the top dead center. However, the combustion with the addition of n-heptane is dominated by distributed heat release, and the combustion event with the addition of ethanol occurs very late. 3.3. Effect of the Premixed Ratio. Figure 10 shows the regulated emissions versus the premixed ratio of n-heptane, DMM, and ethanol in two overall equivalence ratios. For the

biodiesel engine without any premixed fuel, both the CO and HC emissions are very low. With the addition of n-heptane and DMM, CO emissions increase rapidly and attain the highest levels in a certain premixed ratio. After that, CO emissions begin to decrease linearly as the premixed ratio further increases. In the lower overall equivalence ratio, CO emissions with the premixing ethanol increase linearly in the overall range of premixed ratios. In the larger equivalence ratio, it shows the same trend as that of premixing n-heptane and DMM. HC emissions increase substantially at first up to a critical premixed ratio. As the premixed ratio exceeds the critical value, which depends upon the overall equivalence ratio and physicalchemical properties of premixed fuels, HC emissions almost maintain the same level regardless of the premixed ratio. However, it should be noted that premixing ethanol has a larger HC value than the other two premixing fuels. Figure 10c shows the NOx emissions versus the premixed ratio for different premixed fuels in the same overall equivalence ratio. With the addition of ethanol, NOx emissions always decrease. However, with the addition of the other two premixed fuels, the NOx emission versus premixed ratio shows another tendency. Under these conditions, the NOx emission decreases at first, with the premixed fuel amounting up to a critical equivalence ratio. As the equivalence ratio exceeds the critical value, NOx emissions start to increase gradually. Moreover, it can be seen from Figure 10d that, with the increase of the premixed ratio, the smoke opacity can be further improved. However, the addition of ethanol results in the largest smoke opacity improvement. 4. Conclusions This paper evaluates the effects of premixed fuels from the intake port on the biodiesel-engine-regulated emissions, conducted on a single-cylinder engine. Some conclusions can be drawn from this work: (1) The heat release curves of the n-heptane/biodiesel premixed combustion presents a three-stage combustion, namely, cool flame and hot flame reaction of n-heptane, as well as diffusion combustion of biodiesel, while


the heat release curves show only one-stage heat release with the port injection of DMM and ethanol. (2) A simultaneous reduction of NOx and smoke opacity can be obtained with the premixed fuels, but ethanol premixing shows the most significant effects in the reduction of NOxand smoke opacity compared to other kinds of premixed fuels. (3) Under the same overall equivalence ratio and partial equivalence ratio of premixed fuel, premixing DMM exhibits larger indicated thermal efficiency, premixing n-heptane shows lower indicated thermal efficiency, and premixing ethanol display comparable thermal efficiency when compared to the original neat biodiesel engine. (4) For a fixed premixed ratio, premixing n-heptane shows the larger NOx increase in the slope but premixing DMM shows a larger smoke opacity increase of the slope with the increase of the partial equivalence ratio of biodiesel. While the HC emissions seem to be mainly


determined by the premixed fuel properties but not the equivalence ratio, ethanol premixing has larger HC levels than premixing n-heptane and DMM. (5) For a fixed overall equivalence ratio, both HC and CO emissions remain stable but NOx emissions decrease at first up to a critical premixed ratio. As the premixed ratio exceeds the critical value, NOx and CO emissions start to increase gradually but HC emissions almost maintain the same level. Under the above conditions, smoke opacity always improves with the increase of the premixed ratio. Acknowledgment. This work was supported by the Research Fund for the Doctoral Program of Higher Education (Grant 20070248112) and the Nature Science Foundation of Shanghai (Grant 06ZR14045). EF800526E

Improved NOx and Smoke Emission Characteristics of a Biodiesel

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