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Simultaneous reduction of NOx emission and smoke opacity of biodiesel-fueled engines by port injection of ethanol
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Fuel 87 (2008) 1289–1296 www.fuelfirst.com
Simultaneous reduction of NOx emission and smoke opacity of biodiesel-fueled engines by port injection of ethanol Xingcai Lu *, Junjun Ma, Libin Ji, Zhen Huang School of Mechanical and Power Engineering, Shanghai Jiao Tong University, Shanghai, People’s Republic of China Received 3 January 2007; received in revised form 8 July 2007; accepted 10 July 2007 Available online 3 August 2007
Abstract In the present study, the detailed combustion characteristics and emissions of biodiesel-fueled engines with premixed ethanol by port injection were investigated. The experiments were carried out on a single-cylinder, four-stroke, natural aspirated direct injection engine at a fixed speed. The heat release analysis indicates that, with the introduction of ethanol fuel by port injection, the ignition timing of the overall combustion event delays remarkably, while the maximum heat release rate increases smoothly. At a leaner fuel/air mixture, the peak value of the heat release rate (HRR) increases slightly, the maximum in-cylinder gas pressure and temperature decrease, and the indicated thermal efficiency (ITE) deteriorates with the increase of ethanol proportion. While at a rich fuel/air mixture, with the increase of the ethanol proportion, the maximum HRR increases rapidly, the overall combustion event is completed at an earlier crank angle. Moreover, the maximum values of the HRR reach the peak point in a certain premixed ratio which ranges from 20% to 40%. Also, the ITE reaches the largest value at this operation point. Due to the introduction of the ethanol fuel by port injection, both the NOx emission and smoke opacity decrease to a very low level under overall operation conditions. During the experimental points of this test, NOx and smoke opacity simultaneously decrease about 35–85% compared to those of the neat biodiesel-fueled engines. 2007 Elsevier Ltd. All rights reserved. Keywords: Biodiesel; Ethanol; Dual fuel combustion engine; Combustion; Emission
1. Introduction Biodiesel is an alternative diesel fuel that can be produced from renewable feedstock such as vegetable oils, waste frying oils, and animal fats. Biodiesel can either be blended in any proportion with mineral diesel to create a biodiesel blend or be used in its pure form. It is an oxygenated, non-toxic, sulphur-free, biodegradable, and renewable fuel. Recently, a great deal of research has been conducted on the production of the biodiesel and its performance and emissions when it is used as an alternative fuel in diesel engines [1–7]. The above-mentioned studies have come to the same conclusion: when used in conventional diesel engine, biodiesel substantially reduces emis-
*
Corresponding author. Tel.: +86 21 34206039; fax: +86 21 34205553. E-mail address: [email protected] (X. Lu).
0016-2361/$ - see front matter 2007 Elsevier Ltd. All rights reserved. doi:10.1016/j.fuel.2007.07.006
sions such as carbon dioxide (CO2), volatile organic compounds (VOCs), unburned hydrocarbons, carbon monoxide, sulfur oxides (SOx), polycyclic aromatic hydrocarbons, nitrated polycyclic aromatic hydrocarbons, and particulate matter, while maintaining the same thermal efficiency level as that of the conventional diesel fuel. However, biodiesel increases nitrogen oxides (NOx) emission for about 10–13%, mostly NO and NO2, which are considered as ozone hazardous compounds. There is evidence that the increase in NOx emission of biodiesel is related to the advance in fuel injection timing. Boehman et al. [8] confirmed that the higher bulk modulus of compressibility of vegetable oils and their methyl esters leads to an advanced injection timing of about 1–4CA. Lee et al. [9] described the atomization characteristics of various mixing ratio of biodiesel using PDA system on a common rail system. The author found that the mean size of the droplets increased in accordance with the mixing
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ratio and concluded that the NOx emission increased because of oxygen in the biodiesel and a shorter ignition delay. Kegl [10–12] reported an experimental test and numerical simulation on the injection process and atomization characteristics of different biodiesel fuels at different temperatures. After that, the author hopes to improve the NOx emission by optimizing the injection timing. Szybist et al. [13] reported that no trends were seen between NOx and ignition delay or maximum cylinder temperature, or fuel-specific and that relationships were seen between NOx and maximum heat release rate and the timing of maximum heat release rate. Furthermore, McCormick et al. [14] 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 biodiesel chemical structure, specifically the fatty acid chain length, and the number of double bonds on emissions of NOx. NOx emission of biodiesel can be slightly decreased by some complicated strategies [15–19], but it still cannot meet the future emission standard. To reduce the NOx emission of biodiesel fuel significantly, a new effective method should be used. It is widely known that the oxygenated biofuels such as biodiesel and ethanol are biodegradable, nontoxic, renewable alternatives to imported mineral diesel, and their use not only creates new markets for domestic agricultural products, but also greatly reduces particulate emissions. The ethanol has been widely used as fuel, mainly in Brazil, or as a gasoline additive for octane improver and better combustion in USA [20]. Faced with the degradation of global environment and foreseeable future depletion of worldwide petrol reserves, recent research interest has turned to renewable ethanol fuel in diesel engines [21,22]. Through partial substitution of ethanol for diesel, both the NOx and smoke emissions of compression ignition engines were reduced. Based on the above researches, in this study, the ethanol, with lower boiling point and larger latent heat of evaporation, was injected into the intake pipe by an electronic injector, while the biodiesel was directly injected into the cylinder. By using this dual fuel combustion system, the author hopes to improve the smoke opacity and NOx emission simultaneously of biodiesel-fueled engines. The aim of this paper is to provide new data on the effect of the equivalence ratio and premixed ratio of fuel on the characteristics of the combustion and exhaust emissions of the ethanol/biodiesel dual fuel combustion system.
2. Experimental system 2.1. Experimental apparatus A single-cylinder, four-stroke, natural aspirated direct engine was employed in the test, and the engine speed was fixed at 1800 rpm. 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 1, the experimental system is shown in Fig. 1. 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). CO, HC, and NOx emissions were measured by an analyzer (AVL Digas 4000). Smoke opacity was measured by a smoke meter (AVL 439). 2.2. Experimental procedure Ethanol was injected into the intake pipe by an electronic fuel injector at the location of approximately 0.35 m upstream to the intake port, so that the leaner homogenous ethanol/air mixture could be formed during the intake stroke and compression stroke. Near the top dead center (TDC), the biodiesel fuel, which was used to ignite the overall stratified fuel/air mixture, was directly injected into the combustion chamber by the original injection system. To insure the repeatability and comparability of the measurements for different operating conditions, the coolant-out temperature remained at 85 C and held to within ±2 C, while the oil temperature was kept at 90–95 C. The engine speed was 1800 rpm and held to within ±2 rpm. 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 0CA 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 zero-dimension combustion model.
Table 1 Specifications of the single-cylinder engine Bore · stroke Displacement (L) Combustion mode Inlet valve open Inlet valve close
98 · 105 0.782 Direct injection 16CA BTDC 52CA ABDC
Compression ratio Needle open pressure Advanced angle of injector open Exhaust valve open Exhaust valve close
18.5 24 MPa 9CA BTDC 66CA BBDC 12CA ATDC
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Fig. 1. Schematic of experimental system.
3. Experimental results and discussion In this paper, the premixed ratio (PI) is defined as the ratio of cycle energy of premixed fuel, i.e. ethanol, to total energy which includes premixed fuel and direct injected fuel. The PI can be calculated using the following formula: PI ¼
m_ E HuE 100% m_ BD HuBD þ m_ E HuE
where m_ E and m_ BD indicate the fuel consumption rate of ethanol and biodiesel, respectively. HuE and HuBD are the lower heating values of ethanol and biodiesel. When PI is equal to 0, it means that there is no premixed ethanol fuel and the engine runs with neat biodiesel by direct injection. In the following analysis, the ignition timing h1 is defined as the crank angle where in-cylinder pressure, separate from the compression curve. h2, is defined as the center point of the heat release curve; h3 is the crank angle of 90% of accumulated heat release. Fig. 2 gives the in-cylinder pressure, heat release rate (HRR), and the mean gas temperature for various premixed ratios in a fixed overall equivalence ratio (/). At a leaner fuel/air mixture (Fig. 2a), the ignition timing retards clearly, the peak values of the in-cylinder gas pressure and the maximum mean gas temperature decrease, while the maximum value of the heat release rate increases with the increase of the premixed ratio of ethanol. For a larger overall equivalence ratio (Fig. 2b), the combustion phasing still delays and the peak value of the heat release also increases substantially with the increase of the PI, but both the maximum values of the in-cylinder gas pressure and mean gas temperature increase gradually and attain to the peak point
at a specific PI. After that, both the maximum values of the pressure and temperature begin to decrease with the increase of the PI. Essentially, for ethanol/biodiesel dual fuel combustion system, the combustion event can be understood as the premixed ethanol being ignited by biodiesel. Since the larger latent heat of evaporation of ethanol fuel will lead to a noticeable temperature drop at the end of the compression stroke, the premixed ratio of ethanol will give significant impact on the whole combustion characteristics. Accordingly, the homogeneity of the total fuel/air mixtures and the percentage of the biodiesel fuel play an important role in the combustion phasing and the combustion rate. As a result, the duration of the premixed burn or diffusion burn, and the formation pathway of pollutants in the cylinder change substantially. During the following sections of this paper, the authors will discuss the detailed combustion characteristics and emissions versus premixed ratio of ethanol. Fig. 3 gives the comparison of h1, h2, and h3 for various premixed ratios at different equivalence ratios of total fuel. From these figures, some characteristics can be found as follows: (1) At each overall equivalence ratio, the combustion phasing retards clearly and the center point of heat release delays slightly with the port injection of ethanol fuel. (2) The combustion duration, which can be defined as the interval angle between the ignition timing (h1) and the point of 90% accumulated heat release (h3), prolongs substantially with the increase of the equivalence ratio of total fuel.
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Fig. 2. Effect of PI on combustion characteristics.
Fig. 3. h1, h2, and h3 of the biodiesel-fueled engine with port injection of ethanol in various equivalence ratios.
(3) The premixed burn duration (h2 h1) decreases with the adding of premixed ethanol fuel, but prolongs with the increase of the overall equivalence ratio.
(4) At leaner fuel/air mixture, such as shown in Fig. 3a and b, premixed ratio of ethanol has a moderate effect on the end point of the heat release, while at
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larger fuel/air mixture, such as shown in Fig. 3c and d, the PI of ethanol has an important impact on the end point of the heat release. Consequently, the diffuse burn duration (h3 h1) prolongs slightly in the lower engine load but significantly in the larger engine load. Fig. 4 illustrates the maximum heat release rate and its corresponding crank angle versus PI in various equivalence ratios of total fuel. It can be seen from the figure that the maximum value of the heat release rate (HRRmax) decreases with the increase of the equivalence ratio of total fuel, but the crank angle corresponding to the (HRRmax) advances with the increase of the overall equivalence ratio. This is reasonable because of the decrease of the premixed burn which is induced by the advanced ignition timing at larger engine load. In a lower equivalence ratio, the peak values increase slowly, and its crank angle also delays with the increase of the PI value. On the other hand, for large equivalence ratio of total fuel, the maximum heat release rate increases at a larger magnitude and its crank angle delays quickly with the increase of the PI value. Fig. 5 shows the maximum pressure rise rate and its corresponding crank angle as a function of PI in various equivalence ratios of total fuel. It is obvious that the maximum pressure rise rate increases slowly and attains to the peak point, and then begins to drop substantially with the increase of PI. It is of interesting to note that the largest value of maximum pressure rise rate is observed at PI varying about 25–40% in different equivalence ratios. But when the PI increases up to 50–60%, the maximum pressure rise rate drops to a lower level. Due to the ethanol introduction from the intake pipe, this leads to a noticeable temperature drop at the end of
Fig. 4. The maximum value of heat release rate and its crank angle as a function of PI.
Fig. 5. The maximum pressure rise rate and its corresponding crank angle versus PI in various equivalence ratios.
the compression stroke. As a result, more homogeneous fuel/air mixtures are obtained before the ignition. Consequently, both the maximum values of the heat release rate and pressure rise rate increase with the introduction of ethanol fuel. Since the overall combustion event occurs and is completed during the expansion stroke, when the premixed ratio of ethanol exceeds to a certain value, the ignition occurs at an excessively late angle. Though the maximum heat release rate is still very high, the diffusion burn rate decreases substantially (an excessive longer combustion duration) because induced by the expanding of the cylinder volume, the temperature and pressure begin to drop quickly. As a result, the maximum pressure rise rate begins to drop, too. For the same reason, the crank angle corresponding to the maximum pressure rise rate delays with the increase of the PI. Fig. 6 shows the effects of the premixed ratios on the maximum values of the in-cylinder gas pressure (pmax) and mean gas temperature (Tmax). In a lower equivalence ratio, both the maximum values of the pressure and temperature decrease slowly with the introduction of ethanol fuel due to the delaying of the combustion event. In a larger equivalence ratio of total fuel, both the maximum values of the temperature and pressure begin to increase slowly with the increase of the PI value. At a certain PI value, both Tmax and pmax attain to the largest. After that, pmax and Tmax drop slowly. In general, pmax and Tmax are dominated by many factors. The delaying of the combustion phasing will lead to the increase of the combustion rate, which has positive effect on the pressure rise. On the other hand, the combustion event which occurs and is completed during the expansion stroke will have negative effect on the pressure rise. As shown in Fig. 4, the maximum heat release rate increases at a small magnitude when the engine runs in a lower
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Fig. 6. The maximum in-cylinder pressure and mean gas temperature versus premixed ratio of ethanol.
equivalence ratio. Then, induced by the expansion process of combustion event, the maximum pressure and temperature begin to drop. For a specific PI in a large equivalence ratio, heat release quickly leads to a significant increase of the pressure. As a result, the largest pmax value is observed. Fig. 7 displays the indicated mean effective pressure (IMEP) versus PI in various equivalence ratios. In a smaller equivalence ratio, as such 0.45, the IMEP drops almost linearly with the increase of PI. In the equivalence ratio of 0.53 and above, the IMEP increases up to the peak point at a certain PI and then decreases with the increase of PI. Fig. 8 indicates the indicated thermal efficiency (ITE) as a function of PI in different equivalence ratios. For neat
Fig. 7. The indicated mean effective pressure versus PI in various equivalence ratios.
Fig. 8. The indicated thermal efficiency as a function of PI for different equivalence ratios.
biodiesel (NBD), the ITE decreases with the increase of the equivalence ratio, while the ITE shows a different tendency with the introduction of the premixed ethanol fuel. For the equivalence ratio of total fuel in 0.45, the ITE decreases significantly with the increase of PI. Once the equivalence ratio is larger than 0.45, the ITE begins to increase slowly up to the peak point at certain PI and then decrease remarkably. Furthermore, it should be noted that the increase magnitude of ITE also increases with the increase of the equivalence ratio. For example, the ITE increases from 43.7% to 46.2% at PI of 32% for overall equivalence ratio of 0.53, and the ITE increases from 41.2% to 47.5% at PI of 44% for equivalence ratio of 0.68. Moreover, it can be found from the figure that the larger PI will lead to uncompleted combustion in all equivalence ratios. Fig. 9 shows the emissions of the ethanol/biodiesel dual combustion system in various overall equivalence ratios. HC and CO emissions are very low for neat biodieselfueled engine. But both the HC and CO emissions show an increase tendency with the introduction of ethanol fuel. In the equivalence ratio of 0.45, HC emission increases almost linearly with the increase of PI. However, there is another tendency of HC versus PI in three other equivalence ratios. Under these operating conditions, HC emissions increase linearly with PI at lower PI values, and then increase very slowly to the peak point at certain PI in different equivalence ratios. After that, HC emissions almost maintain a constant level regardless of PI but dependant on the equivalence ratio. HC emissions of ethanol/biodiesel dual fuel combustion feature those of spark ignition engines and compression ignition engines. During the intake stroke and compression stroke, a part of ethanol is squeezed into the clearance of piston ring and dead zone of the combustion chamber. Due to their bad ignitibility, these mixtures are partially oxidized during the combustion process. Consequently, the uncompleted combustion products are discharged from the cylinder, leading to a high HC level. As a result, the HC emissions of ethanol/biodiesel dual fuel combustion are much higher than those of biodiesel engines. Furthermore, with the increase of combustion rate and a larger peak value of the in-cylinder gas temperature, part of HC
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Fig. 9. Effects of PI on emissions of the biodiesel-fueled engine with port injection of ethanol.
emissions may be further oxidized. It can be found therefore that HC emissions almost maintain the same level at a wide range of PI in larger equivalence ratio. A CO emission versus PI is different from HC emission versus PI. In general, CO emissions always increase with the increase of PI, but decrease with the increase of the equivalence ratio of total fuel. A major technical obstacle of biodiesel engines is the higher NOx emission. However, it is very interesting to note that NOx emission decreases substantially with the introduction of ethanol fuel. For example, when the PI value ranges from 18% to 21%, NOx emission is suppressed about 75%, 50%, and 38% in overall equivalence ratios of 0.45, 0.62, and 0.68 respectively. When the PI value increases up to 55–60%, the NOx emission further decreases by 83%, 67%, and 55% in equivalence ratio of 0.45, 0.62, and 0.68, respectively. Due to the oxygen content in the biodiesel, the smoke opacity of biodiesel engines is much lower than that of conventional diesel engines. It can be found from Fig. 9 that the smoke opacity of ethanol/biodiesel dual fuel combustion engines shows a further remarkable decrease when compared to that of neat biodiesel combustion engines. In the case of PI values ranging from 18% to 21%, the smoke opacity reduces by 35%, 38%, 44%, and 51% in the equivalence ratio ranging from 0.45 to 0.68. When the PI values increase to 55–60%, the smoke opacity decreases about 75%, 83%, 83%, and 85% in four different equivalence ratios. Furthermore, it can found from Fig. 10 that the traditional PM-NOx trade-off relationship is destroyed by ethanol addition from the intake port. Both the smoke opacity and NOx emissions are simultaneously
Fig. 10. The relationship between the NOx and smoke opacity of the biodiesel-fueled engine with port injection of ethanol.
reduced for biodiesel-fueled engines with the port injection of ethanol fuel. This result can be attributed to the following reasons. First of all, due to the intense latent heat of evaporation for ethanol, the initial temperature at the bottom of the compression stroke may drop substantially. Secondly, the more homogenous mixture of the dual fuel combustion system will lead to the decrease of the local high temperature. Lastly, the oxygen content in ethanol and the homogenous mixture have positive effect on the smoke opacity. 4. Conclusions This paper has investigated the combustion characteristics and emissions of the biodiesel-fueled engine with port injection of ethanol. Some conclusions can be drawn from the above analysis:
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(1) With the increase of premixed ratio of ethanol, the combustion phasing delays obviously and the center point of heat release retards slightly. The end point of heat release shows little difference for different premixed ratio values in a lower equivalence ratio, but advances substantially for larger premixed ratio values in a larger equivalence ratio. (2) With the increase of the equivalence ratio of total fuel, the peak value of heat release decreases. But with the increase of ethanol fuel, the peak value of heat release increases slowly in a lower equivalence ratio and remarkably in a larger equivalence ratio. The maximum pressure rise rate increases up to a largest value in a premixed ratio of about 20–40% and then begins to decrease with the further increase of the premixed ratio. (3) In a lower equivalence ratio of total fuel, the indicated thermal efficiency decreases linearly with ethanol introduction, but it increases slowly and attains to the peak point at certain premixed ratio value and then decreases obviously in a large equivalent ratio. (4) CO and HC emissions of dual fuel combustion system are higher than those of neat biodiesel engines, but NOx and smoke emissions simultaneously decrease by 35–85%. The dual fuel combustion system has destroyed the conventional trade-off relationship between NOx and PM for traditional diesel engines. Acknowledgements This work was supported by the Natural Science Foundation of Shanghai (Grant No. 06ZR14045). Also, this work is financially supported by Young Scholar Foundation of Shanghai Jiaotong University (Grant No. 05DBX003). References [1] Labeckas G, Slavinskas S. The effect of rapeseed oil methyl ester on direct injection Diesel engine performance and exhaust emissions. Energ Convers Manage 2006;47:1954–67. [2] Shi XY, Pang XB, Mu YJ, He H, Shuai SJ, Wang JX, Chen H, Li RL. Emission reduction potential of using ethanol-biodiesel-diesel fuel blend on a heavy-duty diesel engine. Atmos Environ 2006;40:2567–74. [3] Knothe G, Sharp CA, Ryan TW. Exhaust emissions of biodiesel, petrodiesel, neat methyl esters, and alkanes in a new technology engine. Energ. Fuel 2006;20:403–8.
[4] Sharp CA, Howell SA, Jobe J. The effect of biodiesel fuels on transient emissions from modern diesel engines, Part I regulated emissions and performance. SAE 2000-01-1967. [5] Sharp CA, Howell SA, Jobe J. The effect of biodiesel fuels on transient emissions from modern diesel engines, Part II unregulated emissions and chemical characterization. SAE 2000-01-1968. [6] Armas O, Herna´ndez JJ, Ca´rdenas MD. Reduction of diesel smoke opacity from vegetable oil methyl esters during transient operation. Fuel 2006;85:2427–38. [7] Lebedevas S, Vaicekauskas A, Lebedeva G, Makareviciene V, Janulis P, Kazancev K. Use of waste fats of animal and vegetable origin for the production of biodiesel fuel: Quality, motor properties, and emissions of harmful components. Energ Fuel 2006;20:2274–80. [8] Boehman AL, Morris D, Szybist J, Esen E. The impact of the bulk modulus of diesel fuels on fuel injection timing. Energ Fuel 2004;18:1877–82. [9] Lee CS, Park SW, Kwon S. An experimental study on the atomization and combustion characteristics of biodiesel-blended fuels. Energ Fuel 2005;19:2201–8. [10] Kegl B. Numerical analysis of injection characteristics using biodiesel fuel. Fuel 2006;85:2377–87. [11] Kegl B, Hribernik A. Experimental analysis of injection characteristics using biodiesel fuel. Energ Fuel 2006;20:2239–48. [12] Kegl B. Experimental investigation of optimal timing of the diesel engine injection pump using biodiesel fuel. Energ Fuel 2006;20:1460–70. [13] Szybist JP, Kirby SR, Boehman AL. NOx emissions of alternative diesel fuels: Comparative analysis of biodiesel and PT diesel. Energ Fuel 2005;19:1484–92. [14] McCormick RL, Graboski MS, Alleman TL, Herring AM. Impact of biodiesel source material and chemical structure on emissions of criteria pollutions from a heavy-duty engine. Environ Sci Technol 2001;35:1742–7. [15] Fernando S, Hall C, Jha S. NOx reduction from biodiesel fuels. Energ Fuel 2006;20:376–82. [16] Agarwal D, Sinha S, Agarwal AK. Experimental investigation of control of NOx emissions in biodiesel-fueled compression ignition engine. Renew Energ 2006;31:2356–69. [17] Nabi MN, Akhter MS, Shahadat MZ. Improvement of engine emissions with conventional diesel fuel and diesel–biodiesel blends. Bioresour Technol 2006;97:372–8. [18] Hess MA, Haas MJ, Foglia TA, Marmer WN. Effect of antioxidant addition on NOx emissions from biodiesel. Energ Fuel 2005;19:1749–54. [19] McCormick RL, Alvarez JR, Graboski MS. NOx Solutions for biodiesel. February 2003, NREL/SR-510-31465. [20] Poulopoulos SG, Samaras DP, Philippopoulos CJ. Regulated and unregulated emissions from an internal combustion engine operating on ethanol-containing fuels. Atmos Environ 2001;35:4399–43406. [21] Lu XC, Yang JG, Zhang WG, Huang Z. Effect of cetane number improver on heat release rate and emissions of high speed diesel engine fueled with ethanol-diesel blend fuel. Fuel 2004;83:2013–20. [22] Lu XC, Huang Z, Zhang WG, Li DG. The influence of ethanol additives on the performance and combustion characteristics of diesel engines. Combust Sci Technol 2004;176:1309–29.
Fuel 87 (2008) 1289–1296 www.fuelfirst.com
Simultaneous reduction of NOx emission and smoke opacity of biodiesel-fueled engines by port injection of ethanol Xingcai Lu *, Junjun Ma, Libin Ji, Zhen Huang School of Mechanical and Power Engineering, Shanghai Jiao Tong University, Shanghai, People’s Republic of China Received 3 January 2007; received in revised form 8 July 2007; accepted 10 July 2007 Available online 3 August 2007
Abstract In the present study, the detailed combustion characteristics and emissions of biodiesel-fueled engines with premixed ethanol by port injection were investigated. The experiments were carried out on a single-cylinder, four-stroke, natural aspirated direct injection engine at a fixed speed. The heat release analysis indicates that, with the introduction of ethanol fuel by port injection, the ignition timing of the overall combustion event delays remarkably, while the maximum heat release rate increases smoothly. At a leaner fuel/air mixture, the peak value of the heat release rate (HRR) increases slightly, the maximum in-cylinder gas pressure and temperature decrease, and the indicated thermal efficiency (ITE) deteriorates with the increase of ethanol proportion. While at a rich fuel/air mixture, with the increase of the ethanol proportion, the maximum HRR increases rapidly, the overall combustion event is completed at an earlier crank angle. Moreover, the maximum values of the HRR reach the peak point in a certain premixed ratio which ranges from 20% to 40%. Also, the ITE reaches the largest value at this operation point. Due to the introduction of the ethanol fuel by port injection, both the NOx emission and smoke opacity decrease to a very low level under overall operation conditions. During the experimental points of this test, NOx and smoke opacity simultaneously decrease about 35–85% compared to those of the neat biodiesel-fueled engines. 2007 Elsevier Ltd. All rights reserved. Keywords: Biodiesel; Ethanol; Dual fuel combustion engine; Combustion; Emission
1. Introduction Biodiesel is an alternative diesel fuel that can be produced from renewable feedstock such as vegetable oils, waste frying oils, and animal fats. Biodiesel can either be blended in any proportion with mineral diesel to create a biodiesel blend or be used in its pure form. It is an oxygenated, non-toxic, sulphur-free, biodegradable, and renewable fuel. Recently, a great deal of research has been conducted on the production of the biodiesel and its performance and emissions when it is used as an alternative fuel in diesel engines [1–7]. The above-mentioned studies have come to the same conclusion: when used in conventional diesel engine, biodiesel substantially reduces emis-
*
Corresponding author. Tel.: +86 21 34206039; fax: +86 21 34205553. E-mail address: [email protected] (X. Lu).
0016-2361/$ - see front matter 2007 Elsevier Ltd. All rights reserved. doi:10.1016/j.fuel.2007.07.006
sions such as carbon dioxide (CO2), volatile organic compounds (VOCs), unburned hydrocarbons, carbon monoxide, sulfur oxides (SOx), polycyclic aromatic hydrocarbons, nitrated polycyclic aromatic hydrocarbons, and particulate matter, while maintaining the same thermal efficiency level as that of the conventional diesel fuel. However, biodiesel increases nitrogen oxides (NOx) emission for about 10–13%, mostly NO and NO2, which are considered as ozone hazardous compounds. There is evidence that the increase in NOx emission of biodiesel is related to the advance in fuel injection timing. Boehman et al. [8] confirmed that the higher bulk modulus of compressibility of vegetable oils and their methyl esters leads to an advanced injection timing of about 1–4CA. Lee et al. [9] described the atomization characteristics of various mixing ratio of biodiesel using PDA system on a common rail system. The author found that the mean size of the droplets increased in accordance with the mixing
1290
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ratio and concluded that the NOx emission increased because of oxygen in the biodiesel and a shorter ignition delay. Kegl [10–12] reported an experimental test and numerical simulation on the injection process and atomization characteristics of different biodiesel fuels at different temperatures. After that, the author hopes to improve the NOx emission by optimizing the injection timing. Szybist et al. [13] reported that no trends were seen between NOx and ignition delay or maximum cylinder temperature, or fuel-specific and that relationships were seen between NOx and maximum heat release rate and the timing of maximum heat release rate. Furthermore, McCormick et al. [14] 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 biodiesel chemical structure, specifically the fatty acid chain length, and the number of double bonds on emissions of NOx. NOx emission of biodiesel can be slightly decreased by some complicated strategies [15–19], but it still cannot meet the future emission standard. To reduce the NOx emission of biodiesel fuel significantly, a new effective method should be used. It is widely known that the oxygenated biofuels such as biodiesel and ethanol are biodegradable, nontoxic, renewable alternatives to imported mineral diesel, and their use not only creates new markets for domestic agricultural products, but also greatly reduces particulate emissions. The ethanol has been widely used as fuel, mainly in Brazil, or as a gasoline additive for octane improver and better combustion in USA [20]. Faced with the degradation of global environment and foreseeable future depletion of worldwide petrol reserves, recent research interest has turned to renewable ethanol fuel in diesel engines [21,22]. Through partial substitution of ethanol for diesel, both the NOx and smoke emissions of compression ignition engines were reduced. Based on the above researches, in this study, the ethanol, with lower boiling point and larger latent heat of evaporation, was injected into the intake pipe by an electronic injector, while the biodiesel was directly injected into the cylinder. By using this dual fuel combustion system, the author hopes to improve the smoke opacity and NOx emission simultaneously of biodiesel-fueled engines. The aim of this paper is to provide new data on the effect of the equivalence ratio and premixed ratio of fuel on the characteristics of the combustion and exhaust emissions of the ethanol/biodiesel dual fuel combustion system.
2. Experimental system 2.1. Experimental apparatus A single-cylinder, four-stroke, natural aspirated direct engine was employed in the test, and the engine speed was fixed at 1800 rpm. 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 1, the experimental system is shown in Fig. 1. 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). CO, HC, and NOx emissions were measured by an analyzer (AVL Digas 4000). Smoke opacity was measured by a smoke meter (AVL 439). 2.2. Experimental procedure Ethanol was injected into the intake pipe by an electronic fuel injector at the location of approximately 0.35 m upstream to the intake port, so that the leaner homogenous ethanol/air mixture could be formed during the intake stroke and compression stroke. Near the top dead center (TDC), the biodiesel fuel, which was used to ignite the overall stratified fuel/air mixture, was directly injected into the combustion chamber by the original injection system. To insure the repeatability and comparability of the measurements for different operating conditions, the coolant-out temperature remained at 85 C and held to within ±2 C, while the oil temperature was kept at 90–95 C. The engine speed was 1800 rpm and held to within ±2 rpm. 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 0CA 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 zero-dimension combustion model.
Table 1 Specifications of the single-cylinder engine Bore · stroke Displacement (L) Combustion mode Inlet valve open Inlet valve close
98 · 105 0.782 Direct injection 16CA BTDC 52CA ABDC
Compression ratio Needle open pressure Advanced angle of injector open Exhaust valve open Exhaust valve close
18.5 24 MPa 9CA BTDC 66CA BBDC 12CA ATDC
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Fig. 1. Schematic of experimental system.
3. Experimental results and discussion In this paper, the premixed ratio (PI) is defined as the ratio of cycle energy of premixed fuel, i.e. ethanol, to total energy which includes premixed fuel and direct injected fuel. The PI can be calculated using the following formula: PI ¼
m_ E HuE 100% m_ BD HuBD þ m_ E HuE
where m_ E and m_ BD indicate the fuel consumption rate of ethanol and biodiesel, respectively. HuE and HuBD are the lower heating values of ethanol and biodiesel. When PI is equal to 0, it means that there is no premixed ethanol fuel and the engine runs with neat biodiesel by direct injection. In the following analysis, the ignition timing h1 is defined as the crank angle where in-cylinder pressure, separate from the compression curve. h2, is defined as the center point of the heat release curve; h3 is the crank angle of 90% of accumulated heat release. Fig. 2 gives the in-cylinder pressure, heat release rate (HRR), and the mean gas temperature for various premixed ratios in a fixed overall equivalence ratio (/). At a leaner fuel/air mixture (Fig. 2a), the ignition timing retards clearly, the peak values of the in-cylinder gas pressure and the maximum mean gas temperature decrease, while the maximum value of the heat release rate increases with the increase of the premixed ratio of ethanol. For a larger overall equivalence ratio (Fig. 2b), the combustion phasing still delays and the peak value of the heat release also increases substantially with the increase of the PI, but both the maximum values of the in-cylinder gas pressure and mean gas temperature increase gradually and attain to the peak point
at a specific PI. After that, both the maximum values of the pressure and temperature begin to decrease with the increase of the PI. Essentially, for ethanol/biodiesel dual fuel combustion system, the combustion event can be understood as the premixed ethanol being ignited by biodiesel. Since the larger latent heat of evaporation of ethanol fuel will lead to a noticeable temperature drop at the end of the compression stroke, the premixed ratio of ethanol will give significant impact on the whole combustion characteristics. Accordingly, the homogeneity of the total fuel/air mixtures and the percentage of the biodiesel fuel play an important role in the combustion phasing and the combustion rate. As a result, the duration of the premixed burn or diffusion burn, and the formation pathway of pollutants in the cylinder change substantially. During the following sections of this paper, the authors will discuss the detailed combustion characteristics and emissions versus premixed ratio of ethanol. Fig. 3 gives the comparison of h1, h2, and h3 for various premixed ratios at different equivalence ratios of total fuel. From these figures, some characteristics can be found as follows: (1) At each overall equivalence ratio, the combustion phasing retards clearly and the center point of heat release delays slightly with the port injection of ethanol fuel. (2) The combustion duration, which can be defined as the interval angle between the ignition timing (h1) and the point of 90% accumulated heat release (h3), prolongs substantially with the increase of the equivalence ratio of total fuel.
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Fig. 2. Effect of PI on combustion characteristics.
Fig. 3. h1, h2, and h3 of the biodiesel-fueled engine with port injection of ethanol in various equivalence ratios.
(3) The premixed burn duration (h2 h1) decreases with the adding of premixed ethanol fuel, but prolongs with the increase of the overall equivalence ratio.
(4) At leaner fuel/air mixture, such as shown in Fig. 3a and b, premixed ratio of ethanol has a moderate effect on the end point of the heat release, while at
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larger fuel/air mixture, such as shown in Fig. 3c and d, the PI of ethanol has an important impact on the end point of the heat release. Consequently, the diffuse burn duration (h3 h1) prolongs slightly in the lower engine load but significantly in the larger engine load. Fig. 4 illustrates the maximum heat release rate and its corresponding crank angle versus PI in various equivalence ratios of total fuel. It can be seen from the figure that the maximum value of the heat release rate (HRRmax) decreases with the increase of the equivalence ratio of total fuel, but the crank angle corresponding to the (HRRmax) advances with the increase of the overall equivalence ratio. This is reasonable because of the decrease of the premixed burn which is induced by the advanced ignition timing at larger engine load. In a lower equivalence ratio, the peak values increase slowly, and its crank angle also delays with the increase of the PI value. On the other hand, for large equivalence ratio of total fuel, the maximum heat release rate increases at a larger magnitude and its crank angle delays quickly with the increase of the PI value. Fig. 5 shows the maximum pressure rise rate and its corresponding crank angle as a function of PI in various equivalence ratios of total fuel. It is obvious that the maximum pressure rise rate increases slowly and attains to the peak point, and then begins to drop substantially with the increase of PI. It is of interesting to note that the largest value of maximum pressure rise rate is observed at PI varying about 25–40% in different equivalence ratios. But when the PI increases up to 50–60%, the maximum pressure rise rate drops to a lower level. Due to the ethanol introduction from the intake pipe, this leads to a noticeable temperature drop at the end of
Fig. 4. The maximum value of heat release rate and its crank angle as a function of PI.
Fig. 5. The maximum pressure rise rate and its corresponding crank angle versus PI in various equivalence ratios.
the compression stroke. As a result, more homogeneous fuel/air mixtures are obtained before the ignition. Consequently, both the maximum values of the heat release rate and pressure rise rate increase with the introduction of ethanol fuel. Since the overall combustion event occurs and is completed during the expansion stroke, when the premixed ratio of ethanol exceeds to a certain value, the ignition occurs at an excessively late angle. Though the maximum heat release rate is still very high, the diffusion burn rate decreases substantially (an excessive longer combustion duration) because induced by the expanding of the cylinder volume, the temperature and pressure begin to drop quickly. As a result, the maximum pressure rise rate begins to drop, too. For the same reason, the crank angle corresponding to the maximum pressure rise rate delays with the increase of the PI. Fig. 6 shows the effects of the premixed ratios on the maximum values of the in-cylinder gas pressure (pmax) and mean gas temperature (Tmax). In a lower equivalence ratio, both the maximum values of the pressure and temperature decrease slowly with the introduction of ethanol fuel due to the delaying of the combustion event. In a larger equivalence ratio of total fuel, both the maximum values of the temperature and pressure begin to increase slowly with the increase of the PI value. At a certain PI value, both Tmax and pmax attain to the largest. After that, pmax and Tmax drop slowly. In general, pmax and Tmax are dominated by many factors. The delaying of the combustion phasing will lead to the increase of the combustion rate, which has positive effect on the pressure rise. On the other hand, the combustion event which occurs and is completed during the expansion stroke will have negative effect on the pressure rise. As shown in Fig. 4, the maximum heat release rate increases at a small magnitude when the engine runs in a lower
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Fig. 6. The maximum in-cylinder pressure and mean gas temperature versus premixed ratio of ethanol.
equivalence ratio. Then, induced by the expansion process of combustion event, the maximum pressure and temperature begin to drop. For a specific PI in a large equivalence ratio, heat release quickly leads to a significant increase of the pressure. As a result, the largest pmax value is observed. Fig. 7 displays the indicated mean effective pressure (IMEP) versus PI in various equivalence ratios. In a smaller equivalence ratio, as such 0.45, the IMEP drops almost linearly with the increase of PI. In the equivalence ratio of 0.53 and above, the IMEP increases up to the peak point at a certain PI and then decreases with the increase of PI. Fig. 8 indicates the indicated thermal efficiency (ITE) as a function of PI in different equivalence ratios. For neat
Fig. 7. The indicated mean effective pressure versus PI in various equivalence ratios.
Fig. 8. The indicated thermal efficiency as a function of PI for different equivalence ratios.
biodiesel (NBD), the ITE decreases with the increase of the equivalence ratio, while the ITE shows a different tendency with the introduction of the premixed ethanol fuel. For the equivalence ratio of total fuel in 0.45, the ITE decreases significantly with the increase of PI. Once the equivalence ratio is larger than 0.45, the ITE begins to increase slowly up to the peak point at certain PI and then decrease remarkably. Furthermore, it should be noted that the increase magnitude of ITE also increases with the increase of the equivalence ratio. For example, the ITE increases from 43.7% to 46.2% at PI of 32% for overall equivalence ratio of 0.53, and the ITE increases from 41.2% to 47.5% at PI of 44% for equivalence ratio of 0.68. Moreover, it can be found from the figure that the larger PI will lead to uncompleted combustion in all equivalence ratios. Fig. 9 shows the emissions of the ethanol/biodiesel dual combustion system in various overall equivalence ratios. HC and CO emissions are very low for neat biodieselfueled engine. But both the HC and CO emissions show an increase tendency with the introduction of ethanol fuel. In the equivalence ratio of 0.45, HC emission increases almost linearly with the increase of PI. However, there is another tendency of HC versus PI in three other equivalence ratios. Under these operating conditions, HC emissions increase linearly with PI at lower PI values, and then increase very slowly to the peak point at certain PI in different equivalence ratios. After that, HC emissions almost maintain a constant level regardless of PI but dependant on the equivalence ratio. HC emissions of ethanol/biodiesel dual fuel combustion feature those of spark ignition engines and compression ignition engines. During the intake stroke and compression stroke, a part of ethanol is squeezed into the clearance of piston ring and dead zone of the combustion chamber. Due to their bad ignitibility, these mixtures are partially oxidized during the combustion process. Consequently, the uncompleted combustion products are discharged from the cylinder, leading to a high HC level. As a result, the HC emissions of ethanol/biodiesel dual fuel combustion are much higher than those of biodiesel engines. Furthermore, with the increase of combustion rate and a larger peak value of the in-cylinder gas temperature, part of HC
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Fig. 9. Effects of PI on emissions of the biodiesel-fueled engine with port injection of ethanol.
emissions may be further oxidized. It can be found therefore that HC emissions almost maintain the same level at a wide range of PI in larger equivalence ratio. A CO emission versus PI is different from HC emission versus PI. In general, CO emissions always increase with the increase of PI, but decrease with the increase of the equivalence ratio of total fuel. A major technical obstacle of biodiesel engines is the higher NOx emission. However, it is very interesting to note that NOx emission decreases substantially with the introduction of ethanol fuel. For example, when the PI value ranges from 18% to 21%, NOx emission is suppressed about 75%, 50%, and 38% in overall equivalence ratios of 0.45, 0.62, and 0.68 respectively. When the PI value increases up to 55–60%, the NOx emission further decreases by 83%, 67%, and 55% in equivalence ratio of 0.45, 0.62, and 0.68, respectively. Due to the oxygen content in the biodiesel, the smoke opacity of biodiesel engines is much lower than that of conventional diesel engines. It can be found from Fig. 9 that the smoke opacity of ethanol/biodiesel dual fuel combustion engines shows a further remarkable decrease when compared to that of neat biodiesel combustion engines. In the case of PI values ranging from 18% to 21%, the smoke opacity reduces by 35%, 38%, 44%, and 51% in the equivalence ratio ranging from 0.45 to 0.68. When the PI values increase to 55–60%, the smoke opacity decreases about 75%, 83%, 83%, and 85% in four different equivalence ratios. Furthermore, it can found from Fig. 10 that the traditional PM-NOx trade-off relationship is destroyed by ethanol addition from the intake port. Both the smoke opacity and NOx emissions are simultaneously
Fig. 10. The relationship between the NOx and smoke opacity of the biodiesel-fueled engine with port injection of ethanol.
reduced for biodiesel-fueled engines with the port injection of ethanol fuel. This result can be attributed to the following reasons. First of all, due to the intense latent heat of evaporation for ethanol, the initial temperature at the bottom of the compression stroke may drop substantially. Secondly, the more homogenous mixture of the dual fuel combustion system will lead to the decrease of the local high temperature. Lastly, the oxygen content in ethanol and the homogenous mixture have positive effect on the smoke opacity. 4. Conclusions This paper has investigated the combustion characteristics and emissions of the biodiesel-fueled engine with port injection of ethanol. Some conclusions can be drawn from the above analysis:
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(1) With the increase of premixed ratio of ethanol, the combustion phasing delays obviously and the center point of heat release retards slightly. The end point of heat release shows little difference for different premixed ratio values in a lower equivalence ratio, but advances substantially for larger premixed ratio values in a larger equivalence ratio. (2) With the increase of the equivalence ratio of total fuel, the peak value of heat release decreases. But with the increase of ethanol fuel, the peak value of heat release increases slowly in a lower equivalence ratio and remarkably in a larger equivalence ratio. The maximum pressure rise rate increases up to a largest value in a premixed ratio of about 20–40% and then begins to decrease with the further increase of the premixed ratio. (3) In a lower equivalence ratio of total fuel, the indicated thermal efficiency decreases linearly with ethanol introduction, but it increases slowly and attains to the peak point at certain premixed ratio value and then decreases obviously in a large equivalent ratio. (4) CO and HC emissions of dual fuel combustion system are higher than those of neat biodiesel engines, but NOx and smoke emissions simultaneously decrease by 35–85%. The dual fuel combustion system has destroyed the conventional trade-off relationship between NOx and PM for traditional diesel engines. Acknowledgements This work was supported by the Natural Science Foundation of Shanghai (Grant No. 06ZR14045). Also, this work is financially supported by Young Scholar Foundation of Shanghai Jiaotong University (Grant No. 05DBX003). References [1] Labeckas G, Slavinskas S. The effect of rapeseed oil methyl ester on direct injection Diesel engine performance and exhaust emissions. Energ Convers Manage 2006;47:1954–67. [2] Shi XY, Pang XB, Mu YJ, He H, Shuai SJ, Wang JX, Chen H, Li RL. Emission reduction potential of using ethanol-biodiesel-diesel fuel blend on a heavy-duty diesel engine. Atmos Environ 2006;40:2567–74. [3] Knothe G, Sharp CA, Ryan TW. Exhaust emissions of biodiesel, petrodiesel, neat methyl esters, and alkanes in a new technology engine. Energ. Fuel 2006;20:403–8.
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