Experimental and Computational Assessment of Inlet Swirl Effects on a Gasoline Compression Ignition (GCI) Light-Duty Diesel Engine 2014-01-1299 Published 04/01/2014 Paul Loeper, Youngchul Ra, David Foster, and Jaal Ghandhi Univ of Wisconsin CITATION: Loeper, P., Ra, Y., Foster, D., and Ghandhi, J., "Experimental and Computational Assessment of Inlet Swirl Effects on a Gasoline Compression Ignition (GCI) Light-Duty Diesel Engine," SAE Technical Paper 2014-01-1299, 2014, doi:10.4271/2014-01-1299 Copyright © 2014 SAE International Abstract The light-medium load operating regime (4-8 bar net IMEP) presents many challenges for advanced low temperature combustion strategies (e.g HCCI, PPC) in light-duty, high speed engines In this operating regime, lean global equivalence ratios (Φ3% COV of IMEP) or pressure rise rates (>10 bar/ deg) were avoided Table shows baseline engine operating parameters for case experiments (fixed IMAP, Tin, and IMEP) Table In case experiments, inlet temperature and IMAP remained fixed at bar net IMEP Depending on engine speed, inlet temperature and IMAP differed in order to capture the full range of combustion phasing for the inlet swirl ratios investigated Figure Experimental cylinder pressure and heat release rates for bar net IMEP operation at 1300 RPM given changes in inlet swirl With Tin and IMAP fixed at 65C and 130 kPa, respectively, CA50 varies substantially (∼10 CAD over the inlet swirl range), and advances as inlet swirl is reduced The results of a similar experimental set at 2000 RPM is shown in Figure Note at 2000 RPM, due to reductions in both mixing time and the progression of autoignition chemistry, higher intake temperature and pressures were required to ensure gasoline autoignition For both engine speeds investigated in case experiments, the fuel injection strategy remained fixed, and consisted of two injections at an injection pressure of 500 bar Additionally, 70% of fuel mass was injected early during the intake stroke (−350° ATDC; also referred to as % premix) with the remaining 30% of fuel mass injected at −31° ATDC [26, 27] As shown in Figure 7, at 1300 RPM, the effects on combustion phasing (and pressure rise rates) due to varying inlet swirl levels are significant, and indicates in-cylinder temperature distribution as the dominant process For example, reducing the inlet swirl ratio from 2.2 to 1.5 results in the advancement of CA50 by over CAD (from 6.9° ATDC to 0.6° ATDC) This advanced, short combustion duration event, results in high pressure rise rates (10.8 bar/deg) and high NOx emissions (9.1 g/kg-FI), and indicates a hotter, more-thermally homogeneous mixture distribution prior to ignition, as described by Aceves et al [9, 11] In contrast, increasing inlet swirl from 2.2 to 3.5 was observed to retard combustion phasing (CA50 retards from 6.9° ATDC to 10.2° ATDC) Pressure rise rates are reduced as well (from 4.3 to 3.2 bar/deg), while combustion duration (in this case, CA10-75) increases from 8.1 to 10.3 CAD (an increase of 27%) In the case of increasing inlet swirl, these results agree with previous assessments of increased swirl ratios in LTC strategies (specifically HCCI) As inlet swirl is increased, heat Figure Experimental cylinder pressure and heat release rates for bar net IMEP operation at 2000 RPM given changes in inlet swirl With Tin and IMAP fixed at 80C and 150 kPa, respectively Volumetric efficiency effects are more apparent at the higher engine speed At Rs=1.5, cylinder pressure during compression is significantly lower causing CA50 to retard, relative to Rs=2.2 and 3.5 Given fixed IMAP and Tin at 2000 RPM, the combustion phasing variations due to changing inlet swirl are in the opposite direction as observed at 1300 RPM (see Figure and Figure 8) Using throttle plates in the intake runner to adjust inlet swirl levels effectively reduces the volumetric efficiency of the engine, and these effects figure more prominently at 2000 RPM For example, at 1300 RPM, reducing swirl from 2.2 to 1.5 results in a reduction of volumetric efficiency from 94.8% to 86.9% In contrast, at 2000 RPM, the same inlet swirl variation reduces volumetric efficiency from 94.8% to 74.4% This reduction results in a decrease in cylinder pressure (TDC pressure is reduced from 63.2 bar to 50.8 bar, a 19.6% reduction) causing CA50 at Rs=1.5 to retard 1.3 CAD (from 8° to 9.3° ATDC) Increasing Rs from 2.2 to 3.5 causes combustion phasing to advance CAD (from 8° to 7° ATDC) and results in higher pressure rise rates as well (3.9 to 5.2 bar/ deg) Although variations in combustion phasing were observed at both engine speeds, the results at 2000 RPM were less significant While results to be discussed for case (matched CA50 and Φ) will provide a better understanding of swirl effects at 2000 RPM, the effects of increasing inlet swirl at 2000 RPM given fixed IMAP and Tin follow results observed in conventional diesel combustion; that is, increased swirl enhances air-fuel mixing and shortens ignition delay Figure compares CA50, Φ, and volumetric efficiency as a function of inlet swirl between the two engine speeds investigated at bar net IMEP (recall, fueling rate was adjusted as necessary to maintain constant load) Interestingly, at 2000 RPM, the variation in CA50 over the swirl range considered is less than that observed at 1300 RPM Specifically, at 1300 RPM, CA50 varies almost 10 CAD (over the inlet swirl range investigated), as opposed to CAD at 2000 RPM Regardless of these opposing trends in CA50 at 1300 and 2000 RPM, NOx emission trends are similar, and appear to be dominated by (and show sensitivity to) local mixture concentrations, which influences ignition location within the combustion chamber, and subsequently, peak combustion temperatures (as will be seen in CFD results, shown Figure 14b and c) For example, although Rs=1.5 causes combustion phasing to retard at 2000 RPM, NOx emissions remain highest (as shown in Figure 10); similar to results at 1300 RPM Similar NOx trends throughout the swirl range investigated are observed, i.e., increasing NOx with reduced inlet swirl UHC and CO emissions both increase as inlet swirl is increased for both engine speeds Specific to UHC trends, increasing inlet air turbulence appears to result in overly lean regions and cooler temperatures within the combustion chamber, causing oxidation kinetics to quench (and corroborated by CFD results) The CO trends between engine speeds are similar as well; however, at 2000 RPM, a substantial reduction in CO was observed (more so than at 1300 RPM) when inlet swirl was increased from Rs=2.2 to 3.5 (277 to 131 g/kg-FI) CO emissions for this piston bowl geometry have been shown to be affected by the ability to promote CO oxidation in the squish region [28, 29]; this reduction may indicate sufficient local mixture enrichment in this region Figure In a comparison of case experimental results, CA50 exhibits more variation at 1300 RPM than at 2000 RPM Further, at 1300 RPM, CA50 advances as swirl ratio is reduced; in contrast, at 2000 RPM, CA50 advances (although less) with increased inlet swirl Intake throttling resulting from required swirl plate adjustments at Rs=1.5 and 3.5 reduces volumetric efficiency and effectively creates a globally-richer mixture concentration In order to analyze the in-cylinder combustion behavior with swirl ratio variation, numerical simulations were performed for the engine operation at 1300 rev/min Figure 11 compares predicted pressure and heat release rate profiles with measured data In the figure, predicted (or calculated) and experimental results are presented with solid and dashed lines, respectively It is seen that the change of pressures during the compression and expansion strokes with swirl ratio variation is well captured by the prediction The predicted ignition timings are in good agreement with experiments for all three swirl ratios, while pressure rise is slightly over-predicted for the cases with swirl ratios of 2.2 and 3.5 While experimental heat release is derived from cylinder pressure data, numerical calculations consist of chemical heat release only (and absence of wall heat transfer) Figure 11 Comparison of predicted and measured pressure and heat release rate profiles for engine operations at 1300 rev/min in Case-1 Figure 12 Comparison of predicted and measured IMEP and emissions for engine operations at 1300 rev/min in Case (a) IMEP and NOx emissions, (b) UHC and CO emissions Figure 10 Experimental comparison of emission levels at bar net IMEP given case operating conditions NOx trends are similar, regardless of engine speed, over the inlet swirl range investigated This behavior indicates NOx emission rates are dominated by local mixture concentrations, as opposed to phasing effects (which could increase or decrease mixing time) UHC emissions increase with increasing inlet swirl, which could indicate more crevice volume entrapment Interestingly, CO emissions for both engine speeds peak at Rs=2.2; at Rs=1.5, reduced heat transfer (and overall hotter charge) facilitates more-complete CO oxidation while at Rs=3.5, CO levels are reduced due to an enrichment of squish region mixture concentration (as will be shown in CFD results) The predicted IMEP matched the experimental values of ∼4bar, and the NOx emissions are in good agreement with measured data both in trend and quantitatively, as shown in Figure 12a UHC emissions are slightly over-predicted, while CO emissions are significantly over-predicted, as seen in Figure 12b The over-prediction of CO emissions is attributed to the underprediction of CO oxidation during the expansion stroke after the main ignition (CA >10° ATDC) The first explanation for this is that while numerical calculations showed higher peak heat release rates (Figure 11), cumulative heat release was predicted lower, thus resulting in lower CO oxidation during piston expansion Secondly, the underpredicted mixing of high temperature burned gases with unburned charge in the squish region could be another reason for the numerical results leading to higher CO emissions Except for the discrepancy in CO emissions, the numerical simulations predict the engine performance and emissions trends quite well For the baseline operating conditions (Rs=2.2 in Case 1), distributions of spray droplet, gas temperature and fuel equivalence ratio in the cylinder are shown for various crank angles in Figure 13 The snapshot plots provide a means for characterizing combustion behavior through temperature and equivalence ratio distributions along a characteristic plane; in this case, the spray axis It is seen in Figure 13a that a small fraction of spray droplets enter the squish region, and impinge on the piston-top surface resulting in a thin film of fuel on the wall It is also seen that the wall film fuel layer moves in the direction of in-cylinder swirl during compression Figure 13b shows the temperature distribution on the spray axis plane The (black) isothermal contour shown in the figure indicates T=1400 K locations Ignition is predicted to occur at around +4° ATDC in the middle of the re-entrant bowl region High temperature burned gases are mainly seen in the bowl bottom, re-entrant and squish regions and move towards the cylinder liner being convected by the reverse squish flow as the piston descends Local maximum gas temperature peaks at slightly above 2500 K around 10° ATDC (Figure 13d) Targeting the bowl lip region in the cylinder, the fuel spray is split between the bowl and squish regions Due to the interaction between the squish flow (in the direction of the bowl center) and the spray-induced flow, the fuel vapor is directed to the re-entrant region and flows along the bowl surface while being affected by swirl flow This helps form charge mixtures favorable for ignition in the bowl re-entrant region On the contrary, the localized wall film fuel forms locally rich regions near the piston-top surface in the squish region (clearly seen at TDC in Figure 13c), which burns after the first ignition occurs (+4° ATDC) Typically, upon ignition, locally rich mixtures serve to enhance combustion heat release; however, in the squish region where rich mixtures appear, cooling by evaporation and wall heat transfer serves to suppress reactions and increase ignition delay It is notable that the maximum equivalence ratio in the cylinder decreases gradually as more mixing of fuel and air evolves until the ignition timing (see ϕmax values in Figure 13c) Since the equivalence ratio is calculated using the reactants only, the equivalence ratio of a local lean/rich mixture approaches zero/ infinite, once ignition occurs This is why the local maximum equivalence ratio is seen to increase to ϕmax=2.46 at the timing of ignition (+4° ATDC) in Figure 13c Further mixing of burned gases and unburned mixtures (likely to be lean) increases the uniformity of the in-cylinder mixtures and maximum equivalence ratios fall on the lean side (+10° ATDC) Figure 13 In-cylinder distributions of spray droplets, gas temperature and equivalence ratio for the simulation baseline operating conditions (a) Spray drop distribution at various cranks angles before ignition Spray axis planes are plotted for reference (b) Gas temperature distributions in the spray axis plane Iso-contour lines are for 1400 K (c) Equivalence ratio distributions in the spray axis plane Iso-contour lines are for f=0.5 and local maximum equivalence ratio at each crank angle is indicated, as well (d) Profiles of average and local maximum gas temperatures in the cylinder enhanced mixture richness in the squish region before ignition indicating the potential for improved CO oxidation kinetics (note Φ=0.5 iso-contour) As was seen from the pressure profiles (see Figure and Figure 11), the ignition timings were −2, +4 and + 7° ATDC for swirl ratios of 1.5, 2.2 and 3.5, respectively The swirl ratio change affected the ignition location and the size of ignition region, as shown in Figure 14c The ignition location is pushed towards the bottom of the bowl in the case of Rs=1.5, while the ignition region extends to the bowl lip region for the case of Rs=3.5 Due to earlier ignition, burned gas temperatures are much higher and the high temperature area is much wider in the case of Rs=1.5 than the other swirl ratio cases at 10° ATDC Together with longer time that burned gases reside in high temperatures, these contribute to the significant increase of NOx emissions in the case of Rs=1.5 (see Figure 12a) Due to altered fuel distribution, i.e., increased fuel amount entering the squish region with a lower swirl ratio), the NOx emissions distribution and their level are significantly affected For the lowest swirl ratio, NOx emissions are mainly formed in the squish region near the cylinder liner, while local maximum NOx concentration occurs at the bottom of the bowl in the highest swirl ratio case Enhanced uniformity of fuel/air mixtures in the case of Rs=3.5 reduces the burned gas temperatures, and thus suppresses NOx formation (Figure 14d) Figure 14 Comparison of in-cylinder behavior among the three swirl ratio cases in Case (a) Spray drop distribution at −15° ATDC, (b) equivalence ratio distributions at crank angle 1-degree before the ignition timings, (c) gas temperature distributions in the spray axis plane at ignition timings and 10° ATDC Local maximum gas temperatures at 10° ATDC are also indicated (d) Distributions of NOx at +15° ATDC The effects of swirl ratio variation on combustion behavior and emissions can be explained by comparison of in-cylinder distributions of spray droplets, equivalence ratio, gas temperatures and NOx emissions, as shown in Figure 14 With a reduced swirl ratio of 1.5, the spray penetrate further and more fuel enter the squish region forming a larger wall film (Rs=1.5 in Figure 14a) Reduced swirl flow reduces mixing of fuel vapor and air, which increases fuel stratification both in the bowl and squish regions Mixing of fuel vapor and air is enhanced by increased swirl, resulting in more uniform mixtures, which can be seen from the equivalence ratio distribution before ignition Figure 14b shows equivalence ratio distributions in the spray axis plane one crank angle degree before the ignition timings for the three swirl ratio cases It is seen that, in the case of Rs=1.5, rich regions are more localized and the local maximum equivalence ratio of the fuel/air mixtures is the highest (ϕmax=1.15) among the cases Increased stratification with lower swirl ratio tends to advance ignition timings In addition, as observed in both experiments and computations, CO emissions decreased when increasing inlet swirl from 2.2 to 3.5; CFD results indicate Experiments - Case Considering the effect that opening or closing the swirl plates had on volumetric efficiency, the case experimental objectives were to match equivalence ratios for the inlet swirl range investigated (while maintaining fixed IMEP and Tin) In addition, data were also collected at bar net IMEP Results at bar net IMEP, 1300 RPM, are shown in Figure 15 Figure 15 Swirl effects at bar net IMEP, 1300 RPM experiments, with matched equivalence ratios As a result of matching Φ (primarily through increasing IMAP), Rs=3.5 combustion advances indicating pressure enhancement of autoignition kinetics CA50 at Rs=1.5 remains well advanced Rs=1.5 retained the significant combustion phasing advancement, with respect to Rs=2.2 and 3.5, as was seen in case 1; both pressure rise rates (11.4 bar/deg) and NOx emissions (12.7 g/kg-FI) remained high for Rs=1.5 Adjustments in IMAP at Rs=3.5 caused CA50 to advance from 10.2° in case to 7° ATDC, which nearly matches the CA50 at Rs=2.2 (6.9° ATDC) Although CA50 is matched, combustion duration (CA10-75) is longer in the higher swirl case (9.3 CAD vs 8.1 CAD, or a 14.8% increase), perhaps indicating greater temperature heterogeneity This larger temperature distribution may create a staged combustion event as ignition begins in the hottest region and proceeds to the next hottest, and so on [9] At 2000 RPM, adjustments in IMAP to match Φ cause CA50 for both Rs=1.5 and 3.5 to advance, as shown in Figure 16, indicating pressure enhancement of autoignition kinetics Figure 17 First law energy accounting compares fuel energy pathways for Rs=1.5 and 3.5 at bar net IMEP, 2000 RPM Although CA50 is within 0.5 CAD, turbulent enhancement of heat transfer at Rs=3.5 primarily contributes to a 4-point reduction in gross ITE Increasing load at 2000 RPM from to bar net IMEP reveals similar trends to those at bar net IMEP, as shown in Figure 18 For this case, EGR was utilized (13.65% inlet O2 concentration) for both NOx emission reductions (specifically, to keep NOx below 0.5 g/kg-FI) and phasing control (to keep PRR