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Influence of Early Fuel Injection Timings on Premixing and Combustion in a Diesel Engine Sanghoon Kook, Seik Park, and Choongsik Bae* Engine Laboratory, Department of Mechanical Engineering, Korea AdVanced Institute of Science and Technology, 373-1 GuSeong-Dong, YuSeong-Gu, Daejeon 305-701, Republic of Korea ReceiVed August 29, 2007. ReVised Manuscript ReceiVed NoVember 8, 2007
The influence of fuel injection timing on precombustion mixing of diesel fuel and air, combustion, and emissions at early-injection conditions similar to homogeneous charge compression-ignition (HCCI) engine conditions was investigated experimentally in an automotive-size compression-ignition engine and a constantvolume vessel. The injection timing was controlled by electronic fuel injection equipment. In-cylinder pressure measurements, engine-out emission measurements, and imaging of the spray development were used to analyze the ignition delay period and fuel distribution. The ignition delay period was measured over a wide range of injection timings as well as at various compression ratios and engine speeds. With advancing fuel injection timing, the ignition delay increased and the engine-out nitrogen oxides (NOx) decreased, suggesting that increased premixing time results in a lean mixture and low flame temperature. It was also found that the ignition delay period to decrease NOx emissions to a negligible level was almost the same under any compression ratio or engine speed. From the emissions measurements, a drastic decrease in smoke, hydrocarbon (HC), and carbon monoxide (CO) emissions was observed at specific early-injection timing. To clarify the source of the observed behavior, the diesel spray impinging on the surface of a combustion chamber was visualized in a constantvolume vessel simulating in-cylinder environments under early-injection conditions. The images show that diesel spray should target the bowl-lip area to enhance precombustion mixing, which would be desirable for reducing incomplete combustion products.
1. Introduction Precombustion mixing is a process of fuel-air mixture preparation that occurs before the onset of ignition in typical direct-injection diesel engines. The short ignition delay period allows a small amount of fuel to be mixed before combustion starts so that most diesel fuel is burned through a mixingcontrolled diffusion process which generates high levels of smoke and nitrogen oxides (NOx).1 Traditionally, it has been thought that advancing the injection timing would improve thermal efficiency but increase NOx emission because of the associated high pressure and temperature.2 Reducing flame temperature by employing exhaust gas recirculation (EGR) has been attempted to suppress NOx emissions, but this in turn has increased smoke emission due to the lack of oxidants.2,3 Recent studies, however, suggest that not only low NOx emissions but also low smoke emissions are achievable solely with extremely advanced injection, as the lean homogeneous mixture formed through the entire cylinder volume results in low flame temperature.4,5 This technique, often termed homogeneous charge compression-ignition (HCCI), allows an extended ignition delay period so that the mixing is substantially complete * Corresponding author: Ph +82-42-869-3044; Fax +82-42-869-5044; e-mail
[email protected]. (1) Dec, J. E. SAE Trans. 1998, 106, 1319–1348 (970873). (2) Wade, W. R. SAE Trans. 1980, 89, 1379–1398 (800335). (3) Plee, S. L.; Myers, J. P.; Ahmed, T. SAE Trans. 1981, 90, 3738– 3754 (811195). (4) Zhao, F.; Asmus, T. W.; Assanis, D. N.; Dec, J. E.; Eng, J. A.; Najt, P. M. Homogeneous Charge Compression Ignition (HCCI) Engines; SAE International: Warrendale, PA, 2003; pp 145–154. (5) Lee, K.; Lee, C.; Ryu, J.; Kim, H. Energy Fuels 2005, 19, 393– 402.
before the onset of combustion, making the combustion phasing decoupled from injection timing and dominated by the kinetics of the chemical reactions. It is important to figure out the required advancement of the injection timing for achieving complete premixing in a directinjection type HCCI engine. Researchers studied the effect of the injection timing advancement concerning the engine poweroutput and emissions.6,7 On the other hand, the visualization of the in-cylinder fuel distribution should provide the direct evidence of homogeneous mixture and be useful to clarify physical process of the engine combustion.8 The optical techniques such as laser-induced fluorescence (LIF), however, need sophisticated modification of the engine to employ the light source and camera, and more than that, it is too weak to run at high load and the simplified piston design is not realistic in many cases. The variation of fuel distribution with advancing injection timing is another important factor influencing precombustion mixing. This is because the engine piston position relative to the fuel being injected is changed by the injection timing. In the diesel engine with the typical pip-in-the-bowl piston, the initial splitting of the spray upon striking the bowl lip influences the air utilization in the squish area and deep in the bowl. This effect, often termed as a spray-targeting, has been studied and showed that the optimal targeting was existed achieving low (6) Akagawa, H.; Miyamoto, T.; Harada, A.; Sasaki, S.; Shimazaki, N.; Hashizume, T.; Tsujimura, K. SAE Trans. 2000, 108, 120–32 (1999-010183). (7) Hasegawa, R.; Yanagihara, H. SAE Trans. 2004, 112, 1070–1077 (2003-01-0745). (8) Kashdan, J. T.; Docquier, N.; Bruneaux, G. SAE Trans. 2005, 113, 1783–1799 (2004-01-2945).
10.1021/ef700521b CCC: $40.75 2008 American Chemical Society Published on Web 12/12/2007
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Table 1. Specifications of the Research Engine engine type bore stroke displacement injector
single-cylinder, direct-injection, four valves, 4-stroke 83 mm 92 mm 498 cm3 injector nozzle VCO nozzle numbers 14 nozzle diameter 0.100 mm included angle 70°
soot and CO emissions.9 More extensive works have been done by Opat et al. recently, and from the computational data, they reported optimal fuel split for emissions between fuel entering the squish region and the piston bowl in highly dilute lowtemperature combustion regime.10 The goal of this study is to further investigate the effects of early-injection timings on premixing and combustion. NOx emission that can be measured relatively easily has been suggested as an in-direct indicator of the complete premixing. In an experimental engine, the ignition delay period was measured over a wide range of injection timings and correlated quantitatively with NOx emissions. The sudden drop of smoke and HC and CO emissions at certain injection timing that was observed previously by the authors11 as well as by Opat et al.10 was captured again at more advanced injection conditions with a narrow included-angle injector. In order to provide experimental proof for the previous publications, the temporal diesel spray development was visualized in a constant-volume vessel simulating early-injection conditions, and then the influence of spray-targeting over various injection timings on engine-out emissions has been discussed. 2. Experiments Diesel Engine. A single-cylinder, direct-injection, four-stroke diesel engine (EngineTech, RSi-090) was used to study the influence of injection timing on diesel premixing, combustion and emissions. The specifications of the engine are summarized in Table 1. The engine was a typical automotive-size diesel engine with a bore of 83 mm and a stroke of 92 mm, yielding a displacement of 498 cm3. It was equipped with a common-rail fuel injection system capable of a maximum rail pressure of 135 MPa controlled by a programmable injector driver (TEMS, TDA3000H) and a pressure regulator (TEMS, PCV). For the experiments presented here, a 14hole, valve-covered orifice (VCO) nozzle injector was employed with 0.100 mm nozzle diameter. An included angle, the angle made by two center-of-the-spray-cone lines, was selected as 70° to fit for early-injection conditions.12,13 A schematic of the engine is shown in Figure 1. The engine is originally equipped with an extended piston and quartz piston crown to provide optical access, which were replaced by metal parts with the same geometry just for the performance tests in this study. The intake and exhaust plenums were installed to minimize flow fluctuations problematic in a single-cylinder engine. In-cylinder pressure was recorded with a piezoelectric pressure transducer (KISTLER, 6052A) every 0.2 crank angle degrees (CAD). The recorded in-cylinder pressures were ensemble-averaged over 138 engine cycles to calculate the work produced and apparent heat release rate. A smoke meter (AVL, 415) was used to measure the (9) Lee, S.; Reitz, R. D. SAE Trans. 2007, 115, 532–541, 9 (2006-010918). (10) Opat, R.; Ra, Y.; Gonzalez, D. M. A.; Krieger, R.; Reitz, R. D.; Foster, D. E.; Durrett, R. P.; Siewert, R. M. SAE Tech. Pap. Ser. 2006, 2007-01-0193, 1–21. (11) Kook, S.; Bae, C.; Miles, P. C.; Choi, D.; Bergin, M.; Reitz, R. D. SAE Trans. 2007, 115, 111–132 (2006-01-0197). (12) Kim, M. Y.; Kim, J. W.; Lee, C. S.; Lee, J. H. Energy Fuels 2006, 20, 69–76. (13) Kook, S.; Bae, C. SAE Trans. 2005, 113, 563–78 (2004-01-0938).
Figure 1. Schematic diagram of the research engine.
engine-out smoke emissions. NOx, HC, and CO emissions were measured using emissions analyzers (HORIBA, MEXA1500D). Samples were taken from the exhaust pipe and transferred to the analyzers through a heated sample line. Smoke and particles were removed from the sample by a filter prior to the analyzers. Engine Operating Conditions. The engine operating conditions are summarized in Table 2. The engine was operated at 800 and 1200 rpm under both motored and fired conditions. Two different compression ratios (14.6, 18.7) were applied to test the effect of in-cylinder temperature on ignition delay period. Fuel injection timings were widely controlled from the compression stroke to the intake stroke. The coolant temperature was set to 353 K. The intake air temperature was controlled and held constant at 433 K with electric heaters in the intake plenum. This temperature was high enough to assist the vaporization of diesel fuel even in the case of injection at bottom dead center (BDC).13 The common-rail pressure was set to 120 MPa. This high pressure injection could aggravate the spray wall-impingement problem. However, earlier study by the authors showed that the total heat release increased with increasing injection pressure possibly due to the improved atomization and better air-fuel mixing.13 Therefore, the injection pressure was maintained at a relatively high value in order to achieve as high a power output as possible, which was comparable to that in the conventional diesel engine. The fuel used was conventional diesel with a cetane number of 50 and lower heating value (LHV) of 42.7 MJ/kg. The quantity of fuel injected to the engine cycle was 9.56 mg, which corresponds to 0.308 of an average equivalence ratio at a given air induction condition. The temperature at the fuel tank was 313 K. Prior to obtaining in-cylinder pressure and exhaust emissions, the engine was motored for a minimum of 40 s in order to preheat the combustion chamber walls. Subsequently, the engine was fired for 10 s, and then data of pressure and emissions were acquired over 138 additional fired cycles. While injection timings were swept, the same motored in-cylinder pressure history was attained to maintain the same ambient conditions for all tests. With changing injection timing, this pressure history varies due to the changes in the cylinder wall temperatures. Therefore, for some tests, the motored period was extended from the minimum of 40 s to match the motored pressure traces before the injection event. The incylinder pressures obtained before the fuel injection event were repeatable within 0.3%. Constant-Volume Vessel. A constant-volume vessel was also used to examine the process of precombustion mixing. A schematic of the vessel and the camera field of view are shown in Figure 2. The vessel has a cylindrical inner space with a volume of 13 804 cm3. The optical access is provided by four circular windows 89 mm in diameter. Two windows were utilized here: one for the light source and the other for the camera. Spray images frozen by a fiberoptic strobe (EG&G Optoelectronics; MVS-2601), which had light duration of shorter than 100 ns, were acquired with a CCD camera (PCO CCD Imaging, SensiCam). Inside the vessel, the radial cut
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Figure 3. Effect of injection timing on indicated mean effective pressure (IMEP) and fuel conversion efficiency (ηfc) over various compression ratios (14.8, 18.7) and engine speeds (800, 1200 rpm). Fueling rate: 9.56 mg/engine cycle.
Figure 2. Schematic of constant-volume vessel showing the camera view field. The nozzle cap allows only one nozzle to be activated. Table 2. Engine Operating Conditions engine speed compression ratio injection timing coolant temperature air temperature at intake port common-rail pressure fuel
800, 1200 rpm 14.8, 18.7 50-200 CAD BTDC 353 K 433 K 120 MPa type cetane no. lower heating value quantity temperature
diesel 50 42.7 MJ/kg 9.56 mg per engine cycle (φaverage ) 0.308) 313 K
piston with the same bowl geometry used in the engine experiments was inserted. The axial (z) position of the piston was adjusted by controlling the adaptor. A distance between an injector tip and piston top was calculated considering injection timings. To isolate the influence of spray-spray interaction on the fuel distribution, the specially designed nozzle cap14 was used, allowing only one nozzle to be activated; the other fuel sprays were bypassed without flow disturbance. The vessel was pressurized with nitrogen at room temperature in order to simulate the conditions of early-injection timings.
3. Results and Discussion If the injected fuel quantity is fixed, the advance in injection timing increases the indicated mean effective pressure (IMEP) and fuel conversion efficiency (ηfc). This measure of engine efficiency is given by ηfc )
W mfQLHV
(1)
where W is produced engine work that can be calculated from the measured in-cylinder pressure and known volume at a given crank angle. mf is the mass of fuel injected per cycle, and QLHV is the lower heating value of the used diesel fuel. Figure 3 shows the variation in IMEP and ηfc over various injection timings for two compression ratios and engine speeds. IMEP and ηfc increase by around injection timings of 70-90 (14) Bae, C.; Kang, J. Int. J. Engine Res. 2006, 7, 319–334.
Figure 4. Effect of injection timing on in-cylinder pressure and apparent heat release rate traces at 1200 rpm and 14.8 compression ratio.
CAD BTDC. As shown in in-cylinder pressure and apparent heat release rate traces at fixed engine speed and compression ratio (Figure 4), all of the early injections produced a negative work due to the early ignition and combustion phasing. This negative work was decreased as injection timing was advanced by 70 CAD BTDC due to the extended ignition delay period and retarded combustion phasing. The extended ignition delay period also increased the portion of the premixed burning which preceded the peak in-cylinder pressure to the higher value. At more advanced injection timings, however, the IMEP and ηfc start to decrease. It was due to the decreased combustion temperature; the well-distributed fuel-air mixture decreased the local equivalence ratio (φ < 1). Decrease of the peak of the in-cylinder pressure and apparent heat release rate reinforced this view. It was notable that a small peak prior to the hightemperature heat release was more clear as injection timings were advanced, which was due to the low-temperature reaction.4,5
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Figure 7. Measured ignition delay periods over various injection timings, compression ratios, and engine speeds.
Figure 5. Effect of compression ratio and engine speed on in-cylinder pressure and apparent heat release rate traces at injection timing ) 70 CAD BTDC.
Figure 8. Engine-out NOx emission over various injection timings, compression ratios, and engine speeds.
Figure 6. Definition of the ignition delay period in this study. In-cylinder pressures of both motored and fired are shown along with the calculated apparent heat release rate and measured injection signal.
At fixed injection timing, the decrease in compression ratio increased IMEP and ηfc due to a reduced compression pressure, which released more energy later timings (less negative work). This view was supported by the in-cylinder pressure and heat release rate traces given in Figure 5. The higher engine speed increased IMEP and ηfc due to less time to lose heat energy to the cylinder walls, especially during the expansion stroke. The data in the following sections were acquired at the same operating conditions mentioned above. Precombustion Mixing Time. Precombustion mixing occurs during the ignition delay period. In the present study, the ignition delay period τid is defined as the time from start of injection (SOI) to start of combustion (SOC). As shown in Figure 6, SOC is determined by the crank angle at which the fired in-cylinder pressure crosses over the motored pressure. The fuel was injected into relatively hot ambient gas of the combustion chamber. Upon start of injection, the fuel entrained and mixed with hot ambient leading to fuel droplet vaporization. This vaporization cooling is shown as a lower fired in-cylinder pressure than the motored in-cylinder pressure. As the vapor
continued to mix with the ambient, fired in-cylinder pressure crossed over the motored in-cylinder pressure as the combustion started. The calculated apparent heat release shows that this crank angle correlates well with the beginning of rapid, hightemperature heat release. Also shown is the injection control signal measured by a high-current probe. Figure 7 shows the measured ignition delay period. Since the ignition delay period is influenced by ambient temperature and pressure,15 advanced injection before TDC results in a longer ignition delay period due to low air temperature and pressure. At fixed injection timing, a low compression ratio increased the ignition delay period for the same reason. In contrast, the increased ignition delay period at low engine speed was not due to low air temperature and pressure as those were matched by adjusting the engine warm-up time for each measurement cycle. (Details may be found in the engine operating condition section.). Even if the ignition delay periods in crank angle degrees were the same, more time was available for premixing at low engine speed; 1 CAD corresponds to about 0.21 ms at 800 rpm and 0.14 ms at 1200 rpm. Measured NOx emissions at the same operating conditions as Figure 3 are shown in Figure 8. The decreasing trends in NOx emissions with advancing injection timing are clearly seen. It was explained that more premixing time allowed more parcels of fuel-air mixture to reach a lower equivalence ratio, and hence the flame temperature was lower than the NOx formation limit. Notably, the injection timing that made the NOx emission negligible (less than 10 ppm here) existed, and those timings (15) Heywood, J. B. Internal Combustion Engine Fundamentals International Edition; McGraw-Hill: New York, 1988; pp 539–548.
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Figure 9. Engine-out NOx emission over ignition delay period. The arrow shows the minimum ignition delay period to achieve near-zero NOx emission.
were more advanced for faster engine speed at a fixed compression ratio. At fixed engine speed, the high compression ratio needed more advanced injection for the negligible NOx emission since the ignition delay period was decreased. To assess the precombustion mixing time or ignition delay period more closely, NOx emissions were drawn again over the ignition delay period, as shown in Figure 9; data from Figures 7 and 8 were reproduced. Note that, regardless of compression ratio and engine speed, NOx emission decreases linearly and becomes near zero at around τid ) 6.5 ms. It was found that the minimum τid existed such that the NOx emission became negligible when the precombustion fuel-air mixing was almost completed. Since the NO formation through Zeldovich mechanism is highly correlated with the flame temperature,3 low NOx emission could be translated low flame temperature and so the well-mixed charge. This view should be considered carefully as load, engine speed, and intake temperature conditions were limited. At a higher load condition, a higher fueling rate is needed, and hence longer τid would be required. Even with the same fueling rate, faster engine speed would need more advanced injection for achieving 6.5 ms of τid; more than 100 CAD of τid was needed at 3000 rpm, and it was increased to 200 CAD at 5000 rpm, which was quite occasional in modern automotive diesel engines. The higher intake temperature of the presented condition than the conventional engine should also be considered since negligible NOx emission can be reached earlier at lower intake temperature. Though the conclusion was driven under limited conditions, it is worth noting that the premixing time was one of the dominant parameters affecting early-injection, low-NOx combustion. Engine-Out Emissions and Fuel Distribution. The incomplete combustion products at the same operating conditions as Figures 3 and 8 are shown in Figure 10. Several general observations may be made. Engine-out smoke emission decreases at advanced injection timings due to the lean mixture and low flame temperature. In contrast to smoke and NOx emissions (Figure 8), HC and CO emissions increased at advanced injection timings. HC and CO emissions are typically small for conventional, high-temperature diesel combustion. Because of high flame temperature and lean mixture, HC and CO are easily oxidized during the combustion process. On the other hand, early-injection, low-temperature combustion systems often exhibit high HC and CO emissions. The region of applicability of these low smoke and NOx combustion regimes within the load-speed operating map is often limited by these
Figure 10. Engine-out smoke, HC, and CO emissions over injection timing for various compression ratios and engine speeds.
high HC and CO emissions with an accompanying fuel economy penalty (see fuel conversion efficiency in Figure 3). In addition to these common features, more important behaviors were observed. With injection timing at 70 CAD BTDC, smoke, HC, and CO emissions are decreased more than 2-fold over emission levels measured with injection timing at 60 CAD BTDC. Smoke emission with injection timing at around 90 CAD BTDC is increased again to become comparable to the emission level with injection timing at 60 CAD BTDC. Note that NOx emission was already near zero at this injection timing (Figure 8). Moreover, HC and CO emissions show a maximum value with injection timing at around 100 CAD BTDC rather than at more advanced injection timings. It was conceivable that the premixing process was significantly influenced by other factors besides the increased ignition delay period. Fuel spray images implying fuel distribution in the cylinder provide more insight into the reason for the sudden reduction in smoke, HC, and CO emissions. The visualization of the liquid spray development impinging on the radial cut piston
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significant for the small bore diameters and high injection pressures.11 The injection timings are shown above the image columns, and the times after the start of the injection signal are shown on the left-hand side of each image column. With the injection timing at 60 CAD BTDC (Fa ) 4.9 kg/ m3), most of the spray penetrated into the bowl. In contrast, with injection timing at 70 CAD BTDC (Fa ) 3.6 kg/m3), Figure 11 shows that the spray splits off evenly into the squish area and deep in the bowl which would result in maximum air utilization and subsequently low smoke, HC, and CO emissions. With injection timing at 90 CAD BTDC (Fa ) 2.3 kg/m3), the spray clearly impacts the piston top, and some portion of the fuel is moving into the crevice volume that would result in high smoke, HC, and CO emissions. The fuel trapped in the crevice volume is known as a major source of HC emission in this type of lowtemperature combustion system.17 The computational results of the flow structure and fuel distribution using KIVA code explained previously that, in an automotive-size diesel engine, the squish flow induced by the rising piston pushes the fuel on the piston top back into the bowl at near TDC.11 However, at advanced injection timing far from TDC, the fuel on the piston top (or squish area) likely remained for enough time to affect the fuel-air mixing. 4. Conclusions
Figure 11. Spray development in a constant-volume vessel: injection quantity ) 11.5 mm3; injection pressure ) 120 MPa.
has already been tried in order to show the fuel distributions.16 In the present study, tests were conducted over more various injection timings. Figure 11 illustrates the liquid spray development determined from CCD camera imaging. The radial cut piston is aligned for a vertical cut plane of the spray axis, and its axial distance between the nozzle tip and piston top is adjusted, considering the piston position at the injection event. Unlike the continuously moving piston in a real engine, the axial position was fixed at the position of the injection timing. However, this limitation could be justified as the piston movement was less than 2 mm during the fuel injection event, just 2.4% of the height of an image column. The in-cylinder density condition in an engine was simulated by using nitrogen gas at room temperature, and hence the captured images do not account vaporization process. Besides, the engine flows such as strong swirl flow that characterizes modern automotive diesel engines are not considered, but this could be justified since the traditional mechanism by which excessive swirl is thought to impede mixing by limiting jet penetration is not believed to be (16) Dingle, P. J.; Lai, M. D. Diesel Common Rail and AdVanced Fuel Injection Systems; SAE International: Warrendale, PA, 2005; p 29.
The influence of injection timings on precombustion mixing, combustion, and emissions were studied experimentally in a single-cylinder automotive-size direct-injection diesel engine. The engine was equipped with a narrow included angle (70°), small nozzle (0.100 mm) injector incorporated with a highpressure common-rail system. The diesel spray was visualized in a piston-inserted constant-volume vessel. From the analysis of ignition delay period and images of spray development, the major findings can be summarized as follows: The ignition delay period providing long precombustion mixing time to reduce NOx emissions to a negligible level was almost the same for any compression ratios and engine speeds. According to this measure, the minimum time required for the completely premixed charge was estimated as “6.5 ms” at least under the limited operating conditions in this study: less than 3 bar IMEP load, 14.8 and 18.7 compression ratios, and 800 and 1200 rpm engine speeds. The diesel spray should target the bowl-lip area to enhance precombustion mixing as would be desired for reducing incomplete combustion products. With a bowl-in type automotive-size piston, maximum air utilization can be achieved by a careful selection of the injection timing, making the spray impact the bowl lip. The spray splits off both into the squish area and deep in the bowl, resulting in a sudden decrease of smoke, HC, and CO emissions. Acknowledgment. The authors appreciate the support for this research provided by the CERC (Combustion Engineering Research Center) and Future Vehicle Technology Development Corps.
Nomenclature
ABBREVIATIONS ATDC BDC
after top dead center bottom dead center
(17) Christensen, M.; Johansson, B.; Hultqvist, A. SAE Tech. Pap. Ser. 2001, 2001-01-1893, 1–12.
Premixing and Combustion in a Diesel Engine BTDC CAD CR EGR HCCI IMEP LHV rpm SOI SOC TDC
before top dead center crank angle degrees compression ratio exhaust gas recirculation homogeneous charge compression-ignition indicated mean effective pressure lower heating value revolutions per minute start of injection start of combustion top dead center
Energy & Fuels, Vol. 22, No. 1, 2008 337 VCO
valve covered orifice
Greek Symbols ηfc W mf QLHV τid Fa EF700521B
fuel conversion efficiency produced work mass of fuel injected per cycle lower heating value ignition delay period ambient density
