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Impact of ethanol blending on particulate emissions from a spark-ignition direct-injection engine
Fuel 236 (2019) 1548–1558
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Full Length Article
Impact of ethanol blending on particulate emissions from a spark-ignition direct-injection engine
T
Stephen Sakaia,c, , David Rothamerb,c ⁎
a
University of Wisconsin-Madison, 1500 Engineering Dr, ERB 140, Madison, WI 53706, USA University of Wisconsin-Madison, 1500 Engineering Dr, ERB 127, Madison, WI 53706, USA c DOE Great Lakes Bioenergy Research Center, University of Wisconsin-Madison, USA b
ARTICLE INFO
ABSTRACT
Keywords: Gasoline Ethanol Particulate Soot Size distribution Combustion Direct injection
Particulate formation due to combustion of a wide range of ethanol-gasoline blends was investigated in a sparkignition direct-injection (SIDI) engine. This study is a follow-up to a previous study done by the authors in which particulate formation for ethanol-gasoline blends was examined under fully premixed combustion, eliminating physical effects of the fuel. In this study, fuel was injected directly into the cylinder to investigate how the physical properties of the blended fuels influence particulate formation. The engine was operated at a fixed load, phasing, and equivalence ratio while end of injection timing was varied. The results of this work show that increasing ethanol content leads to a decrease in engine-out particulate in spite of, sometimes significant, changes in fuel properties. However, it was also shown that particulate results can be affected by engine operating history which, if not taken into account, could have implications for research and real world applications.
1. Introduction In the last few decades, particulate matter (PM) emissions from internal combustion engines have become an area of increased interest and study. Initially, this was primarily focused on compression-ignition (CI) engines due to their high mass of PM emissions compared to sparkignition (SI) engines [1,2]. Fuel-injected SI engines have historically used port-fuel injection (PFI), where fuel is injected just upstream of the intake valves. Over the past two decades, spark-ignition direct-injection (SIDI) engines have gained popularity due to their ability to increase fuel economy and reduce greenhouse gas emissions [3,4]. Unlike PFI, SIDI engines inject fuel directly in-cylinder, which allows greater control over fuel delivery, leading to reduced cylinder-to-cylinder variability [4]. Currently, SIDI engines account for almost 50% of light-duty vehicles [5]. However, the reduction in greenhouse gas emissions comes at the price of higher PM emissions than PFI engines [4,6]. Renewable energy standards in the United States have resulted in most gasoline being blended with 10% ethanol, this blended fuel accounts for nearly all of the gasoline consumed in the United States today [7]. Combined with the EPA allowing the sale of E15 [8] and E85, the door is open for future increases in gasoline ethanol content. With focus shifting to the increased importance of PM emissions from SI engines and the ethanol content in gasoline increasing, it is important
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to understand how ethanol influences the sooting tendency of gasoline on a more fundamental level. Engine-out PM emissions from ethanol-blended gasoline are still not nearly as well understood as those for diesel or gasoline. Most gasolineethanol PM research and research in engines to date has focused on low ethanol blend percentages (< 20 vol%) and has shown varying results. Some studies saw little or no sensitivity to ethanol content up to 10% by volume, indicating that engine operating conditions play a more critical role in engine PM emissions at low blend levels [9–11]. Price et al. showed little change in PM for ethanol concentrations up to 30% but indicated a large reduction for E85 [12]. Other work has shown a consistent decrease in PM with ethanol content [13–15]. In contrast, Catapano et al. and Di Iorio and coworkers indicated that E50 and E85 made more particulate than neat gasoline [16,17]. Nearly all studies cite engine operation as a significant factor in the results due to the impact on parameters such as air-fuel mixture preparation, fuel vaporization, and wetting of surfaces. In a previous study by the authors [18], the particulate emissions from an SIDI engine fueled with premixed prevaporized (PMPV) gasoline-ethanol fuel blends were investigated. By premixing and prevaporizing the fuel, the physical property effects of the fuel were removed and the sooting propensity from a purely gas-phase chemistry standpoint was investigated. The results of that study indicated that the
Corresponding author. E-mail addresses: [email protected] (S. Sakai), [email protected] (D. Rothamer).
https://doi.org/10.1016/j.fuel.2018.09.037 Received 20 June 2018; Received in revised form 3 September 2018; Accepted 7 September 2018 0016-2361/ © 2018 The Authors. Published by Elsevier Ltd. This is an open access article under the CC BY-NC-ND license (http://creativecommons.org/licenses/BY-NC-ND/4.0/).
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Fig. 1. Distillation curves for gasoline (EEE) blended with ethanol for ethanol blend percentages (by volume) ranging from 0 to 100%.
Fig. 2. Ratio of enthalpy of vaporization for ethanol-gasoline blends to the enthalpy of vaporization of gasoline (EEE) as a function of vol% of ethanol in the mixture. Curve was calculated by combining data from [21] and experimental data.
sooting propensity of the fuel decreased linearly with increased ethanol content. As a companion to that study, this work has been designed to investigate the effects of the physical properties of the fuel blends on soot production in the same engine under similar load conditions for SIDI operation. In the current work, the effects of fuel physical properties on mixture preparation are examined by varying the timing of the injection event.
with the significant increase in vaporization enthalpy for high ethanol concentrations, makes this an important consideration for mixture preparation.
2. Background
3.1. Engine
Ethanol and gasoline have several significant chemical and physical property differences. When blending the fuels, these properties do not necessarily blend in a linear fashion. For example, three important physical property characteristics for gasoline/ethanol blends that influence mixture formation are the Reid vapor pressure (RVP), the distillation curve, and the enthalpy of vaporization. All of which behave non-linearly when the two fuels are blended. RVP is a good example of this non-linearity. Previous studies have shown that the measured RVP for ethanol-gasoline blends deviates significantly from the expected ideal mixture behavior [19–22]. The vapor pressure of ethanol-gasoline blends first increases with ethanol content, peaking between E10 and E20, then reduces at higher ethanol percentages until reaching the value of E100 for 100% ethanol. Ethanol-gasoline blends have highly non-ideal behavior. In fact, between E0 and E50, the RVP is higher than either base fuel. Fig. 1 shows ASTM D86 distillation data taken by the authors for ethanol-gasoline blends made from the fuel components used in this study. This data matches well with similar results in the literature [23]. Again, non-linear effects are seen as the distillation curve changes significantly with as little as 10% ethanol by volume. Of note is that the distillation curve is depressed below that of both base fuels at low volume recovery percentages for intermediate ethanol blends. Work done by Kar et al. [21] examined the change in vaporization enthalpy with the addition of ethanol to gasoline. Fig. 2 shows the ratio of the vaporization enthalpy for ethanol-gasoline blends to that for pure gasoline as a function of ethanol volume % in the blend. Kar found that not only was the change in enthalpy non-linear, it also showed inconsistent increases in enthalpy with ethanol. This non-linearity, combined
The engine used for these experiment was a single-cylinder engine configured to be representative of a modern spark-ignition direct-injection engine. The cylinder head featured a 4-valve pent-roof combustion chamber with a centrally-mounted spark plug and a sidemounted fuel injector. The fuel injector used in this work was a singlehole pressure-swirl type. Table 1 lists the specifications for the engine. It should be noted that all engine timings are listed with 0 crank angle degrees (CAD) referenced to top dead center (TDC) of the compression stroke, times before TDC are negative and times after TDC are positive. In-cylinder pressure was measured using a high-speed piezo-electric pressure transducer (Kistler 6125C). An average of 50 cycles of measured pressure was used to set the operating condition based on gross indicated mean effective pressure (IMEPg ) and location of 50% cumulative heat release (CA50). Pressure data were then acquired for 500
3. Experimental setup
Table 1 Engine geometric parameters.
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Parameter
Value
Unit
Bore Stroke Displacement Compression Ratio Connecting Rod Length Intake Valve Open Intake Valve Close Exhaust Valve Open Exhaust Valve Close I/E Valve Max. Lift
85.96 94.6 549 11.97 152.4 +350 −140 +150 −355 9.9
mm mm cm3 [–] mm CAD CAD CAD CAD mm
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Table 2 Properties of gasoline and ethanol fuels. Property RON Lower Heating Value Density @ 293 K Boiling Point Heat of Vaporization @ 298 K Reid Vapor Pressure Hydrogen/Carbon Ratio Oxygen/Carbon Ratio Stoichiometric Air-Fuel Ratio
Table 3 Engine operating parameters. EEE
E1001
Unit
97.2 42.8 0.744 T10 = 326 T90 = 434 349 61.3 1.845 0 14.6
108.6 26.9 0.789 351 931.1 15.9 3 0.5 8.97
MJ/kg kg/L K kJ/kg kPa [–] [–] [–]
Parameter
Value
Unit
Engine Speed IMEPg
2100 334
RPM kPa
CA50 Fuel Pressure Intake Temperature Exhaust Sample Temperature Exhaust Pressure Equivalence Ratio End of Injection Timing Ethanol Percentage
1 Values given are for neat ethanol. Ethanol used in this study contained 3.27% of hydrocarbon denaturant by volume. Property data from Haltermann Solutions.
8.0 11.0 45 250 101.5 0.98 −190, −220, −250, −280, −310 0, 20, 40, 60, 80, 100
CAD MPa °C °C kPa [–] CAD % Vol.
approximate areas swept by the piston during the injection event with EEE fuel. By changing the injection timing, the likelihood of fuel wetting various in-cylinder surfaces was varied. This included introducing potential piston or cylinder wall wetting and injecting into periods of increased or decreased charge motion. This allowed for the impact of fuel physical properties on PM emissions to be investigated under similar in-cylinder conditions. The operating condition parameters are listed in Table 3.
cycles and averaged. A MATLAB post-processor was used to calculate cumulative heat release, heat release rate, and mass averaged in-cylinder temperature. 3.2. Fuels In this study, denatured ethanol (E100, 3.27% denaturant) was blended into Tier II EEE gasoline, both supplied by Haltermann Products. These two were splash blended to create blends of 20, 40, 60, and 80% ethanol by volume (E20, E40, E60, and E80). The ethanol gasoline blends plus the neat base fuels (E100 and EEE) gave a total of 6 test fuels. Properties of the base fuels are shown in Table 2.
3.4. Particulate sampling system Engine out particle size distributions (PSDs) were measured using a particulate measurement system composed of a partial-flow dilution system (Dekati FPS 4000) and a scanning mobility particle sizer (SMPS). The SMPS utilizes an electrostatic classifier (EC, TSI model 3080), a differential mobility analyzer (LDMA, TSI model 3081), and a condensation particle counter (CPC, TSI model 3010). A diagram showing the exhaust sampling system relative to the engine is shown in Fig. 4. Exhaust was sampled at a location downstream of the exhaust surge tank. The dilution process is performed in two stages. The first uses filtered air heated to 400 °C to maintain near isothermal dilution at the sample temperature, which is approximately 250 °C, this is followed
3.3. Engine operating conditions For these experiments, the engine was operated at 2100 RPM with IMEPg , CA50, and equivalence ratio ( ) maintained at 334 kPa, 8.0 CAD, and 0.98, respectively. For each fuel, the end of injection (EOI) timing was swept from an early injection timing of −310 CAD to a late timing of −190 CAD in 30 degree increments. Fig. 3 shows the
Fig. 3. Piston location ranges for time between start and end of injection for EEE.
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Fig. 4. Experimental setup schematic showing exhaust sampling system.
Fig. 5. Comparison of (a) cylinder pressure and (b) and heat release for an end of injection of −190 CAD for all fuels.
by a second stage dilution with filtered air at ambient temperature. The overall dilution ratios for these experiments ranged from 40 to 50:1 with a minimum dilution ratio of 40:1. After dilution, the EC and LDMA classify particles according to mobility diameter (dm ) in a range from 7 to 300 nm after which the classified particle stream is sent to the CPC. A Horiba 6-gas emissions bench was used to measure undiluted gaseous emissions: hydrocarbons (HC), carbon monoxide (CO), carbon dioxide (CO2 ), nitrogen oxides (NOx ), oxygen (O2 ), as well as, diluted CO2 which was used to determine the dilution ratio.
4. Results and discussion 4.1. Matching experimental conditions For this work, the experimental results consist of EOI sweeps at a fixed load (constant IMEPg ), combustion phasing (constant CA50), equivalence ratio, and engine speed for each fuel. The goal is to isolate the change in ethanol content as the primary variable when performing the experiments, such that the impact of ethanol content on PM can be 1551
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Fig. 6. Comparison of (a) cylinder pressure and (b) and heat release for an end of injection of −280 CAD for all fuels.
4.2. Experimental repeatability Over the course of this investigation, a behavior was identified where the PSDs would shift after running for long periods of time, usually on the order of hours or days depending on the fuel and operating condition. At times, the PSD would increase, at others, it would decrease. Fig. 7 shows 5 sets of the EEE EOI −250 case taken over the course of 3 months time. For PSD results shown in this work, each trace shown represents the average of seven individual scans for that case. Markers are shown every 10th data point and error bars indicate a 95% confidence interval about the mean based on scan-to-scan variability. For the data shown in Fig. 7, the labels indicate the order in which the data were taken. The pressure and heat release results, shown in Fig. 8, demonstrate that the cases are well matched, thermodynamically. While data taken over months is shown in the figure, it was possible for this magnitude of variation to be seen over the course of several days of running engine. When examining these results, different methods were attempted to determine the cause for this long term variation in particulate emissions. These included maintenance operations such as changing the oil and spark plug, as well as, cleaning the fuel system with a fuel system cleaner (Techron, Chevron). None of these attempts eliminated the observed behavior. It should be noted that in previous work involving the exclusive use of gasoline fuels, fuel system cleaners were able to restore PM particle size distributions to normal, clean fuel injector conditions. It was eventually determined that operating the engine on E100 would bring the PSDs to minimum, repeatable levels. Fig. 9(a) shows PSDs taken for the EEE EOI −250 condition immediately before and after running E100. As shown in the figure, after running the engine for several hours on E100 and then switching back to EEE as the fuel, the PM emissions are significantly decreased. It is believed that there is some buildup of deposits in the fuel injector or incylinder which, while not affecting fuel flow rate or other measured parameters by a measurable amount, adversely impact particulate formation. This sensitivity was most noticeable for PSDs with lower magnitudes. These deposits appear to be soluble in ethanol but not in the fuel system cleaner used, suggesting that they are left by intermediate ethanol fuel blends. In order to keep results consistent and repeatable, an operating procedure was developed where the test condition for each fuel was separated by a test of E100 at the EOI −280 condition. Fig. 9(b) shows the results of multiple EEE EOI −250 cases run on different days using the new operating procedure. It is difficult
Fig. 7. Particle size distributions using EEE operated at the EOI −250 operating condition.
assessed for each injection timing. Fig. 5a shows pressure traces for all of the fuels for the EOI −190 case. The trace for each fuel represents the mean of 500 engine cycles. This case had the highest cycle-to-cycle variability of all of the cases as measured by the coefficient of variance (COV) of IMEPg . For this condition, all of the fuels had COV values in the range of 5–6%. Even so, the peak pressures are all within 5% (1 bar) of each other and the pressures at all other crank angles (CA) are also within this tolerance. The pressure traces match well due to the matching of load and CA50. Fig. 5b also shows the average in-cylinder temperature and heat release rate for the same conditions. In comparison, for more stable conditions, such as EOI −280, the variations between fuels are significantly reduced. This is illustrated in Fig. 6, where the average pressure and heat release rate are very well matched (within a few percent) during the combustion process.
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Fig. 8. Comparison of (a) cylinder pressure and (b) heat release for EEE EOI −250 cases in Fig. 7.
Fig. 9. PSDs for (a) EEE EOI −250 run immediately before and after E100 and (b) multiple EEE EOI −250 run with E100 in between. The data in each plot is from one day of running.
to determine from the current work how universal this problem may be, but is worth keeping in mind when considering data from studies comparing ethanol-gasoline blends at varying blend levels.
Fig. 11 shows images of spray plumes for EEE, E20, E50, and E100 at selected times after SOI. To obtain these images, fuel was injected into a quiescent nitrogen environment at 45 °C and 35 kPa, conditions similar to those at the time of injection for the engine experiments. As can be seen in the figure, as ethanol content is increased, the penetration rate of the spray is reduced. Examination of the spray structure shows that there is increased entrainment and dispersion of the spray plume, possibly due to increased flash boiling of the fuel. The increase in air entrainment would allow for improved mixing, especially in conditions where there is increased charge motion which would affect the spray pattern. It should be noted that, in the quiescent environment into which the fuel was injected for imaging, all of the fuels were able to travel a distance greater than the distance from the injector to the cylinder liner with significant amounts of liquid fuel still present. In cases where the ethanol content is higher, which require more fuel mass due to reduced fuel energy density, the increased amount of fuel injected
4.3. Effect of ethanol on the spray pattern Wetting of in-cylinder surfaces with fuel is a significant cause of particulate formation in SIDI engines [12,24,25]. Fig. 10 shows a representative EEE injection spray pattern near the end of injection (t = 0.8 inj , where inj is the injection duration) overlaid onto the EOI diagram of Fig. 3. The injector used in the image is same as that used in these experiments and the spray and engine components are approximately to scale. The relative size of the spray in-cylinder illustrates the potential for liquid fuel impingement at different EOI timings. Therefore, it is important to examine changes to the injector spray pattern and spray penetration as ethanol is introduced.
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at this operating condition which was determined in previous work [18,26]. The particulate in this PSD is believed to be from sources other than fuel-borne carbon as it has been shown to be independent of the fuel used. The NFB PSD shown was taken during premixed prevaporized operation where the air and fuel are completely premixed and the fuel is completely evaporated before entering the engine. Including this PSD allows for comparison to a condition in which the fuel is effectively non-sooting. The EOI −190 condition, shown in Fig. 12a, is the latest injection timing condition. Because of the late timing, it has the lowest expected in-cylinder charge motion due to the lower piston speed, the shortest incylinder residence time, and the highest probability of fuel impingement on the cylinder wall. As a result, mixture preparation is expected to be the poorest of all the EOIs tested. All fuels have their highest particulate emissions for this EOI of all tested. As can be seen in Fig. 12a, there is a clear trend of decreasing particle number with increasing ethanol content. This trend agrees well with previous results from premixed experiments in the same engine where the sole contributing factor to PM emissions was the fuel chemistry [18]. At the EOI −220 condition, Fig. 12b, the piston is near the bottom of its stroke, charge motion may be slightly greater than for EOI −190, and residence time for mixing and evaporation is increased relative to EOI −190. All fuels show decreased overall particle numbers at the earlier end of injection time. However, E40-E80 show very similar particulate levels, within the uncertainty bounds in some parts of the distributions. E100 has dropped to NFB levels of particulate. At EOI −250 CAD, Fig. 12c, the piston is near the middle of its stroke and charge motion is expected to be higher than the other cases. Due to these combined effects, spray impingement on the piston and cylinder wall is expected to be minimized. At this condition, the particulate levels for all of the fuels have reduced to levels which appear close to the NFB. This is an interesting result for several reasons, firstly that it is possible to achieve the same particulate levels as for premixed prevaporized operation with DI engine operation and secondly that all fuels are able to achieve this. In previous measurements for EEE without using the operating strategy of running with E100 between engine operating conditions, particulate number emissions were
Fig. 10. EEE fuel plume overlaid onto EOI timing piston window diagram.
would lead to higher likelihood of liquid fuel surviving in-cylinder. Additionally, the higher enthalpy of vaporization for ethanol would reduce in-cylinder temperatures reducing the evaporation rate of liquid droplets. 4.4. Particulate results The results for all of the tested fuels at constant EOI timings are shown in Fig. 12. Included in each plot is the non-fuel baseline (NFB) PSD. This PSD represents the minimum particulate level for this engine
Fig. 11. Backlit images of spray plumes for EEE, E20, E50, and E100 fuels injected into quiescent nitrogen at 45 °C and 35 kPa at selected times after start of injection.
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Fig. 12. Particulate size distributions for all fuels at EOI timings of (a) −190, (b) −220, (c) −250, and (d) −280 CAD.
generally higher than the NFB at this operating condition. This points to toward an interesting observation, that the particulate emissions at this apparently optimal injection timing must be very sensitive to the properties of the fuel spray, and that it only requires a small amount of injector deposits to alter the particle number emissions by orders of magnitude. In the case of EOI −280, the piston is in the top half of its stroke and is moving away from the fuel injector. Despite this, there is still a possibility for piston and cylinder wall wetting from the injector. At this condition, higher ethanol blends remain near the NFB, however, EEE and E20 show increased particulate levels. The EOI −310 injection timing results are shown in Fig. 13. For this injection time, the piston is still near the top of the cylinder for most of the fuel injection event. Charge motion is expected to be low but the incylinder residence time of the fuel is maximized. The piston being so close to the injector and combustion chamber increases the probability of fuel wetting of the piston and chamber surfaces. However, due to the limited amount of cylinder-liner exposed at this time, wetting of the
cylinder liner may be reduced or eliminated. At this condition, the particulate levels for all of the fuels have reduced to the NFB level. Figs. 14 and 15 show the data displayed by fuel. Viewing the data in this way allows the effect of EOI on each fuel to be seen more clearly. In general, advancing the injection timing has the effect of reducing the particulate level for all the fuels tested. This is a trend which has been shown elsewhere in the literature [12,27,28]. For higher ethanol blends, ethanol content > 40%, a consistent reduction is seen with advancing EOI until particulate levels reach the NFB by −280 CAD. As mentioned previously, as a general trend, increased ethanol reduces the fuel’s particulate formation. The results shown here imply that residence time and reduced sooting propensity are likely the dominant factors in the particulate formation for these fuels. If injected as early as −280 CAD, the residence time, combined with their reduced sooting propensity, allows these fuels to avoid making any particulate. In conditions where there is reduced time for evaporation and mixing (EOI −190 & −220), particulate emissions increase. In cases where surface wetting is expected
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higher local equivalence ratio and, as a result, higher particulate emissions. The cylinder liner temperature is the lowest temperature surface in the engine, which should be closer to the maintained coolant temperature of about 90°C than the temperature of the piston. This lower temperature will inhibit the vaporization of liquid fuel resulting in diffusive burning of liquid fuel or areas of high equivalence ratio leading to increased particulate. Also, because the flame travels radially from the center of the combustion chamber and is quenched at the cylinder liner, complete oxidation of the fuel in this region is less likely. Fig. 16 shows the data converted to indicated specific particle number (ISPN) in which each of the PSDs are integrated and shown on a per kilowatt-hour basis. Again, all of the previously described trends are visible. For each EOI, the total particle number decreases with ethanol content. The trend of total particulate for each fuel decreasing to minimum levels with earlier injection is plainly visible, as is the increase for EEE and E20 at EOI −280. 5. Conclusions In a previous study by the authors, it was observed that the low sooting tendency of ethanol was able to lower PM emissions under premixed prevaporized operation in an SIDI engine when blended with gasoline. The goal of the current study was to determine whether a similar effect would be seen under DI operation where the physical properties of the fuel may play a significant role. It should be noted that the data and results in this study are limited to the engine and operating condition used here and may not be applicable if the conditions of the experiment are changed significantly. However, the results due give some general indications of the behavior of ethanol/gasoline blends under relatively early injection conditions in SIDI engines. The data shown in this study indicate that, in cases where there is significant particulate, increased ethanol content, for ethanol blending percentages greater or equal to 20%, will reduce particulate levels under the moderate-load early-injection conditions tested in this study. This effect is consistent regardless of the injection timing. There is also a general trend shown of reduced particulate with earlier injection timing. The results from this study can be divided into two behaviors separated by fuel: EEE-E20 and E40-E100. In the latter, particulate levels consistently decrease with earlier injection timing until reaching NFB particulate levels at −280 CAD EOI. For the former two fuels, particulate levels decrease in a similar fashion but with an increase at EOI −280 CAD. The increase possibly caused by wetting of the cylinder liner for that condition and the increased sooting propensity of those fuels compared to the higher ethanol blends. The results of this study show that in situations where fuel wetting of the liner is possible and residence time is reduced, increased particulate levels can be seen even for a fuel with low sooting propensity such as ethanol. Over the course of this study, it was found that the PSD results were susceptible to what the authors believe to be ethanol soluble deposits which greatly affected particulate levels. This change in particulate level was present with no changes to engine operating parameters or experimentally measured parameters. This effect caused particulate levels to increase or decrease, apparently dependent on the fuel operating history. While the results shown here are very repeatable due to the operating methodology used, the issue with the aforementioned deposits could have implications for research applications where inconsistent results would present themselves without other obvious problems appearing. Real world applications where wide ranges of ethanol content fuels are used over longer periods of time, such as in flex-fuel vehicles, may also be susceptible to this problem.
Fig. 13. Particulate size distributions for all fuels at EOI −310 CAD.
with sufficient residence time (EOI −280 & −310) there is no increase in particulate above the NFB. It should be noted that the result would likely be very different if the engine/surface temperatures were not warm, as in this case. Much of the energy for evaporation is provided to the fuel via heat transfer from in-cylinder surfaces [29]. At cold-start conditions, heat transfer is reduced due to lower temperatures. This would result in reduced fuel vaporization and increased particulate emissions [29–31]. Fuels with the lowest ethanol content, EEE and E20, follow a trend similar to the higher ethanol content blends for the later EOIs, but show an increase in particulate at EOI −280 before decreasing again at EOI −310. This is an interesting result because these two fuels do not appear to make particulate at EOIs immediately adjacent to EOI −280. For an EOI of −250 CAD, there is expected to be a appreciable amount of charge motion; tumble flow is expected in this case. The absence of particulate suggests that the charge motion is sufficient to prevent significant wetting of the cylinder wall and/or provide sufficient energy to vaporize and/or mix the liquid fuel to minimize particulate formation. At −310 CAD, the piston being near TDC will result in significant wetting of the piston and combustion chamber. Given the lack of particulate seen at this condition for all fuels, it can be assumed that the fuel is able to sufficiently vaporize from the piston and chamber surfaces and mix sufficiently with air before combustion occurs. Fan et al. showed that fuel impingement on a piston with a flat crown would result in rapid horizontal spreading of fuel creating rich areas on the opposite side of the combustion chamber on the piston and liner surfaces [32]. In the case of EOI −280, there is a higher likelihood of piston and cylinder wall wetting due to a reduced amount of charge motion and proximity of the piston to the injector. The piston shape in this engine is flat, making it possible to have similar impingement results as seen by Fan if there is impingement in this case. Results at −310 CAD show that piston wetting does not necessarily increase particulate emissions for a warm engine. Stanglemaier showed that the largest fraction of liquid fuel was able to evaporate when impingement was located on the piston and the least evaporation took place when liquid fuel was located on the exhaust side of the liner [33]. This implies that fuel impinging on the cylinder liner would result in an area of
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Fig. 14. Particulate size distributions for all EOI timings for (a) EEE, (b) E20, (c) E40, and (d) E60.
Fig. 15. Particulate size distributions for all EOI timings for (a) E80 and (b) E100.
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standards/ethanol-waivers-e15-and-e10 [accessed 26.10.2017]. [9] Butler AD, Sobotowski RA, Hoffman GJ, Machiele P. Influence of fuel PM Index and ethanol content on particulate emissions from light-duty gasoline vehicles [Technical Report]. SAE Technical Paper; 2015. [10] He X, Ireland JC, Zigler BT, Ratcliff MA, Knoll KE, Alleman TL, et al. Tester, Impacts of mid-level biofuel content in gasoline on sidi engine-out and tailpipe particulate matter emissions; 2011. [11] Storey JM, Barone TL, Thomas JF, Huff SP. Exhaust particle characterization for lean and stoichiometric DI vehicles operating on ethanol-gasoline blends [Technical Report]. SAE Technical Paper; 2012. [12] Price P, Twiney B, Stone R, Kar K, Walmsley H. Particulate and hydrocarbon emissions from a spray guided direct injection spark ignition engine with oxygenate fuel blends [Technical Report]. SAE Technical Paper; 2007. [13] Dutcher DD, Stolzenburg MR, Thompson SL, Medrano JM, Gross DS, Kittelson DB, McMurry PH. Emissions from ethanol-gasoline blends: a single particle perspective. Atmosphere 2011;2:182–200. [14] Karavalakis G, Short D, Russell RL, Jung H, Johnson KC, Asa-Awuku A, Durbin TD. Assessing the impacts of ethanol and isobutanol on gaseous and particulate emissions from flexible fuel vehicles. Environ Sci Technol 2014;48:14016–24. [15] Szybist JP, Youngquist AD, Barone TL, Storey JM, Moore WR, Foster M, Confer K. Ethanol blends and engine operating strategy effects on light-duty spark-ignition engine particle emissions. Energy Fuels 2011;25:4977–85. [16] Catapano F, Di Iorio S, Lazzaro M, Sementa P, Vaglieco B. Effect of ethanol blends on soot formation and emissions in a gdi optical engine. In: XXXVI Meeting of the Italian Section of the Combustion Institute. [17] Di Iorio S, Lazzaro M, Sementa P, Vaglieco BM, Catapano F. Particle size distributions from a DI high performance si engine fuelled with gasoline-ethanol blended fuels [Technical Report]. SAE Technical Paper; 2011. [18] Sakai S, Rothamer D. Effect of ethanol blending on particulate formation from premixed combustion in spark-ignition engines. Fuel 2017;196:154–68. [19] Andersen VF, Anderson J, Wallington T, Mueller S, Nielsen OJ. Vapor pressures of alcohol- gasoline blends. Energy Fuels 2010;24:3647–54. [20] Furey RL. Volatility characteristics of gasoline-alcohol and gasoline-ether fuel blends [Technical Report]. SAE Technical Paper; 1985. [21] Kar K, Last T, Haywood C, Raine R. Measurement of vapor pressures and enthalpies of vaporization of gasoline and ethanol blends and their effects on mixture preparation in an si engine. SAE Int J Fuels Lubricants 2008;1:132–44. [22] Pumphrey J, Brand J, Scheller W. Vapour pressure measurements and predictions for alcohol-gasoline blends. Fuel 2000;79:1405–11. [23] Andersen VF, Anderson J, Wallington T, Mueller S, Nielsen OJ. Distillation curves for alcohol- gasoline blends. Energy Fuels 2010;24:2683–91. [24] Myung C, Park S. Exhaust nanoparticle emissions from internal combustion engines: A review. Int J Automotive Technol 2012;13:9. [25] Stevens E, Steeper R. Piston wetting in an optical DISI engine: fuel films, pool fires, and soot generation [Technical Report]. SAE Technical Paper; 2001. [26] Hageman MD, Sakai SS, Rothamer DA. Determination of soot onset and background particulate levels in a spark-ignition engine. Proc Combust Inst 2015;35:2949–56. [27] Chen L, Braisher M, Crossley A, Stone R, Richardson D. The influence of ethanol blends on particulate matter emissions from gasoline direct injection engines [Technical Report]. SAE Technical Paper; 2010. [28] Maricq MM, Podsiadlik DH, Brehob DD, Haghgooie M. Particulate emissions from a direct-injection spark-ignition (DISI) engine [Technical Report]. SAE Technical Paper; 1999. [29] Price P, Stone R, OudeNijeweme D, Chen X. Cold start particulate emissions from a second generation DI gasoline engine [Technical Report]. SAE Technical Paper; 2007. [30] Chen L, Stone R, Richardson D. A study of mixture preparation and pm emissions using a direct injection engine fuelled with stoichiometric gasoline/ethanol blends. Fuel 2012;96:120–30. [31] Whelan I, Smith W, Timoney D, Samuel S. The effect of engine operating conditions on engine-out particulate matter from a gasoline direct-injection engine during cold-start [Technical Report]. SAE Technical Paper; 2012. [32] Fan Q, Deng J, Hu Z, Li L. An experimental study on spray characteristics after impingement on a piston surface through an asymmetrical multi-hole injector of a direct-injection spark ignition gasoline engine. Proc Inst Mech Eng Part D: J Automobile Eng 2012;226:399–409. [33] Stanglmaier RH, Li J, Matthews RD. The effect of in-cylinder wall wetting location on the HC emissions from SI engines [Technical Report]. SAE Technical Paper; 1999.
Fig. 16. Indicated specific particle number for all fuels at all EOI timings.
Acknowledgements This work was funded by the DOE Great Lakes Bioenergy Research Center (DOE BER Office of Science DE-FC02-07ER64494). Appendix A. Supplementary data Supplementary data associated with this article can be found, in the online version, athttps://doi.org/10.1016/j.fuel.2018.09.037. References [1] Kamimoto Takeyuki, Bae Myurng-hoan. High combustion temperature for the reduction of particulate in diesel engines [Technical Report]. SAE Technical Paper; 1988. [2] Shi JP, Harrison RM, Brear F. Particle size distribution from a modern heavy duty diesel engine. Sci Total Environ 1999;235:305–17. [3] U.S. EPA. Final rulemaking to establish light-duty vehicle greenhouse gas emission standards and corporate average fuel economy standards. Joint Technical Support Document; 2010. [4] Zhao Fuquan, Lai M-C, Harrington David L. Automotive spark-ignited direct-injection gasoline engines, Prog Energy Combust Sci 25; 1999: 437–562. [5] Davis SC, Williams SE, Boundy RG, Moore SA. 2016 vehicle technologies market report [Technical Report]. Oak Ridge, TN (United States): Oak Ridge National Laboratory (ORNL); 2017. [6] Hall D, Dickens C. Measurement of the number and size distribution of particles emitted from a gasoline direct injection vehicle [Technical Report]. SAE Technical Paper; 1999. [7] U.S. Energy Information Administration. Almost all U.S. gasoline is blended with 10% ethanol – Today in Energy – U.S. Energy Information Administration (EIA); 2016.https://www.eia.gov/todayinenergy/detail.php?id=26092 [accessed 27.10. 2017]. [8] U.S. Epa, Ethanol Waivers (E15 and E10); 2017.https://www.epa.gov/gasoline-
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Full Length Article
Impact of ethanol blending on particulate emissions from a spark-ignition direct-injection engine
T
Stephen Sakaia,c, , David Rothamerb,c ⁎
a
University of Wisconsin-Madison, 1500 Engineering Dr, ERB 140, Madison, WI 53706, USA University of Wisconsin-Madison, 1500 Engineering Dr, ERB 127, Madison, WI 53706, USA c DOE Great Lakes Bioenergy Research Center, University of Wisconsin-Madison, USA b
ARTICLE INFO
ABSTRACT
Keywords: Gasoline Ethanol Particulate Soot Size distribution Combustion Direct injection
Particulate formation due to combustion of a wide range of ethanol-gasoline blends was investigated in a sparkignition direct-injection (SIDI) engine. This study is a follow-up to a previous study done by the authors in which particulate formation for ethanol-gasoline blends was examined under fully premixed combustion, eliminating physical effects of the fuel. In this study, fuel was injected directly into the cylinder to investigate how the physical properties of the blended fuels influence particulate formation. The engine was operated at a fixed load, phasing, and equivalence ratio while end of injection timing was varied. The results of this work show that increasing ethanol content leads to a decrease in engine-out particulate in spite of, sometimes significant, changes in fuel properties. However, it was also shown that particulate results can be affected by engine operating history which, if not taken into account, could have implications for research and real world applications.
1. Introduction In the last few decades, particulate matter (PM) emissions from internal combustion engines have become an area of increased interest and study. Initially, this was primarily focused on compression-ignition (CI) engines due to their high mass of PM emissions compared to sparkignition (SI) engines [1,2]. Fuel-injected SI engines have historically used port-fuel injection (PFI), where fuel is injected just upstream of the intake valves. Over the past two decades, spark-ignition direct-injection (SIDI) engines have gained popularity due to their ability to increase fuel economy and reduce greenhouse gas emissions [3,4]. Unlike PFI, SIDI engines inject fuel directly in-cylinder, which allows greater control over fuel delivery, leading to reduced cylinder-to-cylinder variability [4]. Currently, SIDI engines account for almost 50% of light-duty vehicles [5]. However, the reduction in greenhouse gas emissions comes at the price of higher PM emissions than PFI engines [4,6]. Renewable energy standards in the United States have resulted in most gasoline being blended with 10% ethanol, this blended fuel accounts for nearly all of the gasoline consumed in the United States today [7]. Combined with the EPA allowing the sale of E15 [8] and E85, the door is open for future increases in gasoline ethanol content. With focus shifting to the increased importance of PM emissions from SI engines and the ethanol content in gasoline increasing, it is important
⁎
to understand how ethanol influences the sooting tendency of gasoline on a more fundamental level. Engine-out PM emissions from ethanol-blended gasoline are still not nearly as well understood as those for diesel or gasoline. Most gasolineethanol PM research and research in engines to date has focused on low ethanol blend percentages (< 20 vol%) and has shown varying results. Some studies saw little or no sensitivity to ethanol content up to 10% by volume, indicating that engine operating conditions play a more critical role in engine PM emissions at low blend levels [9–11]. Price et al. showed little change in PM for ethanol concentrations up to 30% but indicated a large reduction for E85 [12]. Other work has shown a consistent decrease in PM with ethanol content [13–15]. In contrast, Catapano et al. and Di Iorio and coworkers indicated that E50 and E85 made more particulate than neat gasoline [16,17]. Nearly all studies cite engine operation as a significant factor in the results due to the impact on parameters such as air-fuel mixture preparation, fuel vaporization, and wetting of surfaces. In a previous study by the authors [18], the particulate emissions from an SIDI engine fueled with premixed prevaporized (PMPV) gasoline-ethanol fuel blends were investigated. By premixing and prevaporizing the fuel, the physical property effects of the fuel were removed and the sooting propensity from a purely gas-phase chemistry standpoint was investigated. The results of that study indicated that the
Corresponding author. E-mail addresses: [email protected] (S. Sakai), [email protected] (D. Rothamer).
https://doi.org/10.1016/j.fuel.2018.09.037 Received 20 June 2018; Received in revised form 3 September 2018; Accepted 7 September 2018 0016-2361/ © 2018 The Authors. Published by Elsevier Ltd. This is an open access article under the CC BY-NC-ND license (http://creativecommons.org/licenses/BY-NC-ND/4.0/).
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Fig. 1. Distillation curves for gasoline (EEE) blended with ethanol for ethanol blend percentages (by volume) ranging from 0 to 100%.
Fig. 2. Ratio of enthalpy of vaporization for ethanol-gasoline blends to the enthalpy of vaporization of gasoline (EEE) as a function of vol% of ethanol in the mixture. Curve was calculated by combining data from [21] and experimental data.
sooting propensity of the fuel decreased linearly with increased ethanol content. As a companion to that study, this work has been designed to investigate the effects of the physical properties of the fuel blends on soot production in the same engine under similar load conditions for SIDI operation. In the current work, the effects of fuel physical properties on mixture preparation are examined by varying the timing of the injection event.
with the significant increase in vaporization enthalpy for high ethanol concentrations, makes this an important consideration for mixture preparation.
2. Background
3.1. Engine
Ethanol and gasoline have several significant chemical and physical property differences. When blending the fuels, these properties do not necessarily blend in a linear fashion. For example, three important physical property characteristics for gasoline/ethanol blends that influence mixture formation are the Reid vapor pressure (RVP), the distillation curve, and the enthalpy of vaporization. All of which behave non-linearly when the two fuels are blended. RVP is a good example of this non-linearity. Previous studies have shown that the measured RVP for ethanol-gasoline blends deviates significantly from the expected ideal mixture behavior [19–22]. The vapor pressure of ethanol-gasoline blends first increases with ethanol content, peaking between E10 and E20, then reduces at higher ethanol percentages until reaching the value of E100 for 100% ethanol. Ethanol-gasoline blends have highly non-ideal behavior. In fact, between E0 and E50, the RVP is higher than either base fuel. Fig. 1 shows ASTM D86 distillation data taken by the authors for ethanol-gasoline blends made from the fuel components used in this study. This data matches well with similar results in the literature [23]. Again, non-linear effects are seen as the distillation curve changes significantly with as little as 10% ethanol by volume. Of note is that the distillation curve is depressed below that of both base fuels at low volume recovery percentages for intermediate ethanol blends. Work done by Kar et al. [21] examined the change in vaporization enthalpy with the addition of ethanol to gasoline. Fig. 2 shows the ratio of the vaporization enthalpy for ethanol-gasoline blends to that for pure gasoline as a function of ethanol volume % in the blend. Kar found that not only was the change in enthalpy non-linear, it also showed inconsistent increases in enthalpy with ethanol. This non-linearity, combined
The engine used for these experiment was a single-cylinder engine configured to be representative of a modern spark-ignition direct-injection engine. The cylinder head featured a 4-valve pent-roof combustion chamber with a centrally-mounted spark plug and a sidemounted fuel injector. The fuel injector used in this work was a singlehole pressure-swirl type. Table 1 lists the specifications for the engine. It should be noted that all engine timings are listed with 0 crank angle degrees (CAD) referenced to top dead center (TDC) of the compression stroke, times before TDC are negative and times after TDC are positive. In-cylinder pressure was measured using a high-speed piezo-electric pressure transducer (Kistler 6125C). An average of 50 cycles of measured pressure was used to set the operating condition based on gross indicated mean effective pressure (IMEPg ) and location of 50% cumulative heat release (CA50). Pressure data were then acquired for 500
3. Experimental setup
Table 1 Engine geometric parameters.
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Parameter
Value
Unit
Bore Stroke Displacement Compression Ratio Connecting Rod Length Intake Valve Open Intake Valve Close Exhaust Valve Open Exhaust Valve Close I/E Valve Max. Lift
85.96 94.6 549 11.97 152.4 +350 −140 +150 −355 9.9
mm mm cm3 [–] mm CAD CAD CAD CAD mm
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Table 2 Properties of gasoline and ethanol fuels. Property RON Lower Heating Value Density @ 293 K Boiling Point Heat of Vaporization @ 298 K Reid Vapor Pressure Hydrogen/Carbon Ratio Oxygen/Carbon Ratio Stoichiometric Air-Fuel Ratio
Table 3 Engine operating parameters. EEE
E1001
Unit
97.2 42.8 0.744 T10 = 326 T90 = 434 349 61.3 1.845 0 14.6
108.6 26.9 0.789 351 931.1 15.9 3 0.5 8.97
MJ/kg kg/L K kJ/kg kPa [–] [–] [–]
Parameter
Value
Unit
Engine Speed IMEPg
2100 334
RPM kPa
CA50 Fuel Pressure Intake Temperature Exhaust Sample Temperature Exhaust Pressure Equivalence Ratio End of Injection Timing Ethanol Percentage
1 Values given are for neat ethanol. Ethanol used in this study contained 3.27% of hydrocarbon denaturant by volume. Property data from Haltermann Solutions.
8.0 11.0 45 250 101.5 0.98 −190, −220, −250, −280, −310 0, 20, 40, 60, 80, 100
CAD MPa °C °C kPa [–] CAD % Vol.
approximate areas swept by the piston during the injection event with EEE fuel. By changing the injection timing, the likelihood of fuel wetting various in-cylinder surfaces was varied. This included introducing potential piston or cylinder wall wetting and injecting into periods of increased or decreased charge motion. This allowed for the impact of fuel physical properties on PM emissions to be investigated under similar in-cylinder conditions. The operating condition parameters are listed in Table 3.
cycles and averaged. A MATLAB post-processor was used to calculate cumulative heat release, heat release rate, and mass averaged in-cylinder temperature. 3.2. Fuels In this study, denatured ethanol (E100, 3.27% denaturant) was blended into Tier II EEE gasoline, both supplied by Haltermann Products. These two were splash blended to create blends of 20, 40, 60, and 80% ethanol by volume (E20, E40, E60, and E80). The ethanol gasoline blends plus the neat base fuels (E100 and EEE) gave a total of 6 test fuels. Properties of the base fuels are shown in Table 2.
3.4. Particulate sampling system Engine out particle size distributions (PSDs) were measured using a particulate measurement system composed of a partial-flow dilution system (Dekati FPS 4000) and a scanning mobility particle sizer (SMPS). The SMPS utilizes an electrostatic classifier (EC, TSI model 3080), a differential mobility analyzer (LDMA, TSI model 3081), and a condensation particle counter (CPC, TSI model 3010). A diagram showing the exhaust sampling system relative to the engine is shown in Fig. 4. Exhaust was sampled at a location downstream of the exhaust surge tank. The dilution process is performed in two stages. The first uses filtered air heated to 400 °C to maintain near isothermal dilution at the sample temperature, which is approximately 250 °C, this is followed
3.3. Engine operating conditions For these experiments, the engine was operated at 2100 RPM with IMEPg , CA50, and equivalence ratio ( ) maintained at 334 kPa, 8.0 CAD, and 0.98, respectively. For each fuel, the end of injection (EOI) timing was swept from an early injection timing of −310 CAD to a late timing of −190 CAD in 30 degree increments. Fig. 3 shows the
Fig. 3. Piston location ranges for time between start and end of injection for EEE.
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Fig. 4. Experimental setup schematic showing exhaust sampling system.
Fig. 5. Comparison of (a) cylinder pressure and (b) and heat release for an end of injection of −190 CAD for all fuels.
by a second stage dilution with filtered air at ambient temperature. The overall dilution ratios for these experiments ranged from 40 to 50:1 with a minimum dilution ratio of 40:1. After dilution, the EC and LDMA classify particles according to mobility diameter (dm ) in a range from 7 to 300 nm after which the classified particle stream is sent to the CPC. A Horiba 6-gas emissions bench was used to measure undiluted gaseous emissions: hydrocarbons (HC), carbon monoxide (CO), carbon dioxide (CO2 ), nitrogen oxides (NOx ), oxygen (O2 ), as well as, diluted CO2 which was used to determine the dilution ratio.
4. Results and discussion 4.1. Matching experimental conditions For this work, the experimental results consist of EOI sweeps at a fixed load (constant IMEPg ), combustion phasing (constant CA50), equivalence ratio, and engine speed for each fuel. The goal is to isolate the change in ethanol content as the primary variable when performing the experiments, such that the impact of ethanol content on PM can be 1551
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Fig. 6. Comparison of (a) cylinder pressure and (b) and heat release for an end of injection of −280 CAD for all fuels.
4.2. Experimental repeatability Over the course of this investigation, a behavior was identified where the PSDs would shift after running for long periods of time, usually on the order of hours or days depending on the fuel and operating condition. At times, the PSD would increase, at others, it would decrease. Fig. 7 shows 5 sets of the EEE EOI −250 case taken over the course of 3 months time. For PSD results shown in this work, each trace shown represents the average of seven individual scans for that case. Markers are shown every 10th data point and error bars indicate a 95% confidence interval about the mean based on scan-to-scan variability. For the data shown in Fig. 7, the labels indicate the order in which the data were taken. The pressure and heat release results, shown in Fig. 8, demonstrate that the cases are well matched, thermodynamically. While data taken over months is shown in the figure, it was possible for this magnitude of variation to be seen over the course of several days of running engine. When examining these results, different methods were attempted to determine the cause for this long term variation in particulate emissions. These included maintenance operations such as changing the oil and spark plug, as well as, cleaning the fuel system with a fuel system cleaner (Techron, Chevron). None of these attempts eliminated the observed behavior. It should be noted that in previous work involving the exclusive use of gasoline fuels, fuel system cleaners were able to restore PM particle size distributions to normal, clean fuel injector conditions. It was eventually determined that operating the engine on E100 would bring the PSDs to minimum, repeatable levels. Fig. 9(a) shows PSDs taken for the EEE EOI −250 condition immediately before and after running E100. As shown in the figure, after running the engine for several hours on E100 and then switching back to EEE as the fuel, the PM emissions are significantly decreased. It is believed that there is some buildup of deposits in the fuel injector or incylinder which, while not affecting fuel flow rate or other measured parameters by a measurable amount, adversely impact particulate formation. This sensitivity was most noticeable for PSDs with lower magnitudes. These deposits appear to be soluble in ethanol but not in the fuel system cleaner used, suggesting that they are left by intermediate ethanol fuel blends. In order to keep results consistent and repeatable, an operating procedure was developed where the test condition for each fuel was separated by a test of E100 at the EOI −280 condition. Fig. 9(b) shows the results of multiple EEE EOI −250 cases run on different days using the new operating procedure. It is difficult
Fig. 7. Particle size distributions using EEE operated at the EOI −250 operating condition.
assessed for each injection timing. Fig. 5a shows pressure traces for all of the fuels for the EOI −190 case. The trace for each fuel represents the mean of 500 engine cycles. This case had the highest cycle-to-cycle variability of all of the cases as measured by the coefficient of variance (COV) of IMEPg . For this condition, all of the fuels had COV values in the range of 5–6%. Even so, the peak pressures are all within 5% (1 bar) of each other and the pressures at all other crank angles (CA) are also within this tolerance. The pressure traces match well due to the matching of load and CA50. Fig. 5b also shows the average in-cylinder temperature and heat release rate for the same conditions. In comparison, for more stable conditions, such as EOI −280, the variations between fuels are significantly reduced. This is illustrated in Fig. 6, where the average pressure and heat release rate are very well matched (within a few percent) during the combustion process.
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Fig. 8. Comparison of (a) cylinder pressure and (b) heat release for EEE EOI −250 cases in Fig. 7.
Fig. 9. PSDs for (a) EEE EOI −250 run immediately before and after E100 and (b) multiple EEE EOI −250 run with E100 in between. The data in each plot is from one day of running.
to determine from the current work how universal this problem may be, but is worth keeping in mind when considering data from studies comparing ethanol-gasoline blends at varying blend levels.
Fig. 11 shows images of spray plumes for EEE, E20, E50, and E100 at selected times after SOI. To obtain these images, fuel was injected into a quiescent nitrogen environment at 45 °C and 35 kPa, conditions similar to those at the time of injection for the engine experiments. As can be seen in the figure, as ethanol content is increased, the penetration rate of the spray is reduced. Examination of the spray structure shows that there is increased entrainment and dispersion of the spray plume, possibly due to increased flash boiling of the fuel. The increase in air entrainment would allow for improved mixing, especially in conditions where there is increased charge motion which would affect the spray pattern. It should be noted that, in the quiescent environment into which the fuel was injected for imaging, all of the fuels were able to travel a distance greater than the distance from the injector to the cylinder liner with significant amounts of liquid fuel still present. In cases where the ethanol content is higher, which require more fuel mass due to reduced fuel energy density, the increased amount of fuel injected
4.3. Effect of ethanol on the spray pattern Wetting of in-cylinder surfaces with fuel is a significant cause of particulate formation in SIDI engines [12,24,25]. Fig. 10 shows a representative EEE injection spray pattern near the end of injection (t = 0.8 inj , where inj is the injection duration) overlaid onto the EOI diagram of Fig. 3. The injector used in the image is same as that used in these experiments and the spray and engine components are approximately to scale. The relative size of the spray in-cylinder illustrates the potential for liquid fuel impingement at different EOI timings. Therefore, it is important to examine changes to the injector spray pattern and spray penetration as ethanol is introduced.
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at this operating condition which was determined in previous work [18,26]. The particulate in this PSD is believed to be from sources other than fuel-borne carbon as it has been shown to be independent of the fuel used. The NFB PSD shown was taken during premixed prevaporized operation where the air and fuel are completely premixed and the fuel is completely evaporated before entering the engine. Including this PSD allows for comparison to a condition in which the fuel is effectively non-sooting. The EOI −190 condition, shown in Fig. 12a, is the latest injection timing condition. Because of the late timing, it has the lowest expected in-cylinder charge motion due to the lower piston speed, the shortest incylinder residence time, and the highest probability of fuel impingement on the cylinder wall. As a result, mixture preparation is expected to be the poorest of all the EOIs tested. All fuels have their highest particulate emissions for this EOI of all tested. As can be seen in Fig. 12a, there is a clear trend of decreasing particle number with increasing ethanol content. This trend agrees well with previous results from premixed experiments in the same engine where the sole contributing factor to PM emissions was the fuel chemistry [18]. At the EOI −220 condition, Fig. 12b, the piston is near the bottom of its stroke, charge motion may be slightly greater than for EOI −190, and residence time for mixing and evaporation is increased relative to EOI −190. All fuels show decreased overall particle numbers at the earlier end of injection time. However, E40-E80 show very similar particulate levels, within the uncertainty bounds in some parts of the distributions. E100 has dropped to NFB levels of particulate. At EOI −250 CAD, Fig. 12c, the piston is near the middle of its stroke and charge motion is expected to be higher than the other cases. Due to these combined effects, spray impingement on the piston and cylinder wall is expected to be minimized. At this condition, the particulate levels for all of the fuels have reduced to levels which appear close to the NFB. This is an interesting result for several reasons, firstly that it is possible to achieve the same particulate levels as for premixed prevaporized operation with DI engine operation and secondly that all fuels are able to achieve this. In previous measurements for EEE without using the operating strategy of running with E100 between engine operating conditions, particulate number emissions were
Fig. 10. EEE fuel plume overlaid onto EOI timing piston window diagram.
would lead to higher likelihood of liquid fuel surviving in-cylinder. Additionally, the higher enthalpy of vaporization for ethanol would reduce in-cylinder temperatures reducing the evaporation rate of liquid droplets. 4.4. Particulate results The results for all of the tested fuels at constant EOI timings are shown in Fig. 12. Included in each plot is the non-fuel baseline (NFB) PSD. This PSD represents the minimum particulate level for this engine
Fig. 11. Backlit images of spray plumes for EEE, E20, E50, and E100 fuels injected into quiescent nitrogen at 45 °C and 35 kPa at selected times after start of injection.
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Fig. 12. Particulate size distributions for all fuels at EOI timings of (a) −190, (b) −220, (c) −250, and (d) −280 CAD.
generally higher than the NFB at this operating condition. This points to toward an interesting observation, that the particulate emissions at this apparently optimal injection timing must be very sensitive to the properties of the fuel spray, and that it only requires a small amount of injector deposits to alter the particle number emissions by orders of magnitude. In the case of EOI −280, the piston is in the top half of its stroke and is moving away from the fuel injector. Despite this, there is still a possibility for piston and cylinder wall wetting from the injector. At this condition, higher ethanol blends remain near the NFB, however, EEE and E20 show increased particulate levels. The EOI −310 injection timing results are shown in Fig. 13. For this injection time, the piston is still near the top of the cylinder for most of the fuel injection event. Charge motion is expected to be low but the incylinder residence time of the fuel is maximized. The piston being so close to the injector and combustion chamber increases the probability of fuel wetting of the piston and chamber surfaces. However, due to the limited amount of cylinder-liner exposed at this time, wetting of the
cylinder liner may be reduced or eliminated. At this condition, the particulate levels for all of the fuels have reduced to the NFB level. Figs. 14 and 15 show the data displayed by fuel. Viewing the data in this way allows the effect of EOI on each fuel to be seen more clearly. In general, advancing the injection timing has the effect of reducing the particulate level for all the fuels tested. This is a trend which has been shown elsewhere in the literature [12,27,28]. For higher ethanol blends, ethanol content > 40%, a consistent reduction is seen with advancing EOI until particulate levels reach the NFB by −280 CAD. As mentioned previously, as a general trend, increased ethanol reduces the fuel’s particulate formation. The results shown here imply that residence time and reduced sooting propensity are likely the dominant factors in the particulate formation for these fuels. If injected as early as −280 CAD, the residence time, combined with their reduced sooting propensity, allows these fuels to avoid making any particulate. In conditions where there is reduced time for evaporation and mixing (EOI −190 & −220), particulate emissions increase. In cases where surface wetting is expected
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higher local equivalence ratio and, as a result, higher particulate emissions. The cylinder liner temperature is the lowest temperature surface in the engine, which should be closer to the maintained coolant temperature of about 90°C than the temperature of the piston. This lower temperature will inhibit the vaporization of liquid fuel resulting in diffusive burning of liquid fuel or areas of high equivalence ratio leading to increased particulate. Also, because the flame travels radially from the center of the combustion chamber and is quenched at the cylinder liner, complete oxidation of the fuel in this region is less likely. Fig. 16 shows the data converted to indicated specific particle number (ISPN) in which each of the PSDs are integrated and shown on a per kilowatt-hour basis. Again, all of the previously described trends are visible. For each EOI, the total particle number decreases with ethanol content. The trend of total particulate for each fuel decreasing to minimum levels with earlier injection is plainly visible, as is the increase for EEE and E20 at EOI −280. 5. Conclusions In a previous study by the authors, it was observed that the low sooting tendency of ethanol was able to lower PM emissions under premixed prevaporized operation in an SIDI engine when blended with gasoline. The goal of the current study was to determine whether a similar effect would be seen under DI operation where the physical properties of the fuel may play a significant role. It should be noted that the data and results in this study are limited to the engine and operating condition used here and may not be applicable if the conditions of the experiment are changed significantly. However, the results due give some general indications of the behavior of ethanol/gasoline blends under relatively early injection conditions in SIDI engines. The data shown in this study indicate that, in cases where there is significant particulate, increased ethanol content, for ethanol blending percentages greater or equal to 20%, will reduce particulate levels under the moderate-load early-injection conditions tested in this study. This effect is consistent regardless of the injection timing. There is also a general trend shown of reduced particulate with earlier injection timing. The results from this study can be divided into two behaviors separated by fuel: EEE-E20 and E40-E100. In the latter, particulate levels consistently decrease with earlier injection timing until reaching NFB particulate levels at −280 CAD EOI. For the former two fuels, particulate levels decrease in a similar fashion but with an increase at EOI −280 CAD. The increase possibly caused by wetting of the cylinder liner for that condition and the increased sooting propensity of those fuels compared to the higher ethanol blends. The results of this study show that in situations where fuel wetting of the liner is possible and residence time is reduced, increased particulate levels can be seen even for a fuel with low sooting propensity such as ethanol. Over the course of this study, it was found that the PSD results were susceptible to what the authors believe to be ethanol soluble deposits which greatly affected particulate levels. This change in particulate level was present with no changes to engine operating parameters or experimentally measured parameters. This effect caused particulate levels to increase or decrease, apparently dependent on the fuel operating history. While the results shown here are very repeatable due to the operating methodology used, the issue with the aforementioned deposits could have implications for research applications where inconsistent results would present themselves without other obvious problems appearing. Real world applications where wide ranges of ethanol content fuels are used over longer periods of time, such as in flex-fuel vehicles, may also be susceptible to this problem.
Fig. 13. Particulate size distributions for all fuels at EOI −310 CAD.
with sufficient residence time (EOI −280 & −310) there is no increase in particulate above the NFB. It should be noted that the result would likely be very different if the engine/surface temperatures were not warm, as in this case. Much of the energy for evaporation is provided to the fuel via heat transfer from in-cylinder surfaces [29]. At cold-start conditions, heat transfer is reduced due to lower temperatures. This would result in reduced fuel vaporization and increased particulate emissions [29–31]. Fuels with the lowest ethanol content, EEE and E20, follow a trend similar to the higher ethanol content blends for the later EOIs, but show an increase in particulate at EOI −280 before decreasing again at EOI −310. This is an interesting result because these two fuels do not appear to make particulate at EOIs immediately adjacent to EOI −280. For an EOI of −250 CAD, there is expected to be a appreciable amount of charge motion; tumble flow is expected in this case. The absence of particulate suggests that the charge motion is sufficient to prevent significant wetting of the cylinder wall and/or provide sufficient energy to vaporize and/or mix the liquid fuel to minimize particulate formation. At −310 CAD, the piston being near TDC will result in significant wetting of the piston and combustion chamber. Given the lack of particulate seen at this condition for all fuels, it can be assumed that the fuel is able to sufficiently vaporize from the piston and chamber surfaces and mix sufficiently with air before combustion occurs. Fan et al. showed that fuel impingement on a piston with a flat crown would result in rapid horizontal spreading of fuel creating rich areas on the opposite side of the combustion chamber on the piston and liner surfaces [32]. In the case of EOI −280, there is a higher likelihood of piston and cylinder wall wetting due to a reduced amount of charge motion and proximity of the piston to the injector. The piston shape in this engine is flat, making it possible to have similar impingement results as seen by Fan if there is impingement in this case. Results at −310 CAD show that piston wetting does not necessarily increase particulate emissions for a warm engine. Stanglemaier showed that the largest fraction of liquid fuel was able to evaporate when impingement was located on the piston and the least evaporation took place when liquid fuel was located on the exhaust side of the liner [33]. This implies that fuel impinging on the cylinder liner would result in an area of
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Fig. 14. Particulate size distributions for all EOI timings for (a) EEE, (b) E20, (c) E40, and (d) E60.
Fig. 15. Particulate size distributions for all EOI timings for (a) E80 and (b) E100.
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Fig. 16. Indicated specific particle number for all fuels at all EOI timings.
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