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Figure 15. Comparisons of NO (a), CO (b) and HC (c): Unbalanced A/F ratio vs. Balanced Stoichiometric A/F ratio at varying the engine operating condition.

Figure 15 Comparisons of NO (a), CO (b) and HC (c): Unbalanced A/F ratio vs. Balanced Stoichiometric A/F ratio at varying the engine operating condition.

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Table 1. Engine characteristics.
Table 1. Engine characteristics.
Figure 1. Engine layout with LP EGR circuit. The engine is designed following the so-called “downsizing” concept. At high loads the engine works under knock-limited conditions and this requires to delay the combustion phasing (50% of mass fraction burned—MFB 50%), especially at lower speeds where a higher knock tendency usually occurs. On the other hand, at high speeds the A/F mixture has to be particularly enriched to maintain the Temperature at Turbine Inlet below a certain maximum allowable level. These control strategies, although mandatory for the examined engine, greatly penalize the fuel consumption at high loads. For this reason, the base engine configuration is properly modified for research purpose through the installation of an external low-pressure (LP) exhaust gas recirculation (EGR) circuit as depicted in Figure 1
Figure 1. Engine layout with LP EGR circuit. The engine is designed following the so-called “downsizing” concept. At high loads the engine works under knock-limited conditions and this requires to delay the combustion phasing (50% of mass fraction burned—MFB 50%), especially at lower speeds where a higher knock tendency usually occurs. On the other hand, at high speeds the A/F mixture has to be particularly enriched to maintain the Temperature at Turbine Inlet below a certain maximum allowable level. These control strategies, although mandatory for the examined engine, greatly penalize the fuel consumption at high loads. For this reason, the base engine configuration is properly modified for research purpose through the installation of an external low-pressure (LP) exhaust gas recirculation (EGR) circuit as depicted in Figure 1
Table 2. Overall set of experimental engine points. The measured net IMEP (Table 2) is properly derived from the acquired in-cylinder pressure traces over the entire engine cycle and represents the difference between the gross IMEP (in-cylinder pressure over compression and expansion strokes) and the pumping mean effective pressure (PMEP), evaluated over the intake and exhaust strokes.
Table 2. Overall set of experimental engine points. The measured net IMEP (Table 2) is properly derived from the acquired in-cylinder pressure traces over the entire engine cycle and represents the difference between the gross IMEP (in-cylinder pressure over compression and expansion strokes) and the pumping mean effective pressure (PMEP), evaluated over the intake and exhaust strokes.
ee As discussed in the introduction section, experiments carried out on the examined engine have highlighted significant differences in cylinders, mainly ascribed to a non-uniform effective Cylinder-by-cylinder variations are apparent in Figure cylinder pressure trace and the rate of heat release (RO he combustion evolution between in-cylinder air/fuel (A/F) ratio [15]. 2a which shows the experimental in- HR) for two representative operating points (3000@9 and 3000@13, 0% EGR). As aforementioned, the pressure data refer to the ensemble average over 270 consecutive cycles. For both operating conditions, the pressure curves of two cylinders are overlapped in the compression stage, denoting the same air volumetric efficiency, while Cyl #2 provides higher than Cyl #1. The same behavior is found at each inves pressure peak and combustion rate igated operating point, suggesting a systematic difference in pressure cycles. Further experimental tests on the adopted fuel injection system have variation in port injec collected through injec highlighted a different fuel supply to cylinders, resulting from a ors fuel rates (Figure 2b). The results plotted in Figure 2b were tion tests realized at am-bent air pressure with fuel injection system removed from the engine. A drive signal to injectors was used to simulate the injection timing; a certain num mass per stroke delivered by injec pressure traces shown in Figure 2a, fuel mass flow rate. In of two injectors was switched, ind ber of injection timings were considered to measure the gasoline ors. In particular, consistently with the in-cylinder he injector corresponding to Cyl #2 provides a higher erestingly, no difference was found also when the relative position icating the fuel rail geometry as responsible of the cylinder-by-cylinder variation. Indeed, the gasoline rail mounted on the examined engine at test bench represen rail, mounted on the commercial S ts anon-optimized prototype geometry. The optimized version of engine, does not exhibit differences in the gasoline mass flow rates between port injectors. All the previous considerations underline that rail geometry is main responsible for the experimental cylinder-by-cylinder variation analyzed in this research activity. Figure 2. Comparison of experimental in-cylinder pressure traces and burn rates profiles at 3000 rpm, two IMEP levels (9 and 13 bar) and EGR = 0% (a). Differences of the injected gasoline mass per stroke between port injectors (b).
ee As discussed in the introduction section, experiments carried out on the examined engine have highlighted significant differences in cylinders, mainly ascribed to a non-uniform effective Cylinder-by-cylinder variations are apparent in Figure cylinder pressure trace and the rate of heat release (RO he combustion evolution between in-cylinder air/fuel (A/F) ratio [15]. 2a which shows the experimental in- HR) for two representative operating points (3000@9 and 3000@13, 0% EGR). As aforementioned, the pressure data refer to the ensemble average over 270 consecutive cycles. For both operating conditions, the pressure curves of two cylinders are overlapped in the compression stage, denoting the same air volumetric efficiency, while Cyl #2 provides higher than Cyl #1. The same behavior is found at each inves pressure peak and combustion rate igated operating point, suggesting a systematic difference in pressure cycles. Further experimental tests on the adopted fuel injection system have variation in port injec collected through injec highlighted a different fuel supply to cylinders, resulting from a ors fuel rates (Figure 2b). The results plotted in Figure 2b were tion tests realized at am-bent air pressure with fuel injection system removed from the engine. A drive signal to injectors was used to simulate the injection timing; a certain num mass per stroke delivered by injec pressure traces shown in Figure 2a, fuel mass flow rate. In of two injectors was switched, ind ber of injection timings were considered to measure the gasoline ors. In particular, consistently with the in-cylinder he injector corresponding to Cyl #2 provides a higher erestingly, no difference was found also when the relative position icating the fuel rail geometry as responsible of the cylinder-by-cylinder variation. Indeed, the gasoline rail mounted on the examined engine at test bench represen rail, mounted on the commercial S ts anon-optimized prototype geometry. The optimized version of engine, does not exhibit differences in the gasoline mass flow rates between port injectors. All the previous considerations underline that rail geometry is main responsible for the experimental cylinder-by-cylinder variation analyzed in this research activity. Figure 2. Comparison of experimental in-cylinder pressure traces and burn rates profiles at 3000 rpm, two IMEP levels (9 and 13 bar) and EGR = 0% (a). Differences of the injected gasoline mass per stroke between port injectors (b).
Figure 3. Relative air-fuel ratio (A) in Cyl #1, Cyl #2 and overall engine at different operating points as a function of EGR rate. Figure 3 shows the relative air—fuel ratio (A) in Cyl #1, Cyl #2 and the engine exhaust at different engine o perating points and EGR rates. A of Cyl #1 was estimated from the composition of the exhaust gases of that cylinder through the carbon balance of species. Similarly, the overal This allows to indirectly evaluate the A value of Cyl #2 under the hypothesis o volumetric efficiency in the compression s age. Even though the engine conditions are close to stoic lean (A between 1.03 and 1.08) and rich (A between 0.96 and 0.98) mixtures are o Cyl #1 and Cyl #2, respectively. Considering that the engine does not provide an individual set of injection and combustion phasing for each cylinder, the same spark advance and injection parameters are chosen for both cylinders, optimizing the engine load. As a consequence, a difference in the single cylinder load is found at each operating condition. engine A was verified from the overall exhaust gas composition. f an equal for both cylinders, as suggested by the comparison of pressure traces hiometric, btained in
Figure 3. Relative air-fuel ratio (A) in Cyl #1, Cyl #2 and overall engine at different operating points as a function of EGR rate. Figure 3 shows the relative air—fuel ratio (A) in Cyl #1, Cyl #2 and the engine exhaust at different engine o perating points and EGR rates. A of Cyl #1 was estimated from the composition of the exhaust gases of that cylinder through the carbon balance of species. Similarly, the overal This allows to indirectly evaluate the A value of Cyl #2 under the hypothesis o volumetric efficiency in the compression s age. Even though the engine conditions are close to stoic lean (A between 1.03 and 1.08) and rich (A between 0.96 and 0.98) mixtures are o Cyl #1 and Cyl #2, respectively. Considering that the engine does not provide an individual set of injection and combustion phasing for each cylinder, the same spark advance and injection parameters are chosen for both cylinders, optimizing the engine load. As a consequence, a difference in the single cylinder load is found at each operating condition. engine A was verified from the overall exhaust gas composition. f an equal for both cylinders, as suggested by the comparison of pressure traces hiometric, btained in
Figure 4. Indicated load in Cyl #1 and Cyl #2 at different engine operating points as a function of EGR rate.
Figure 4. Indicated load in Cyl #1 and Cyl #2 at different engine operating points as a function of EGR rate.
Figure 5. Engine Indicated Specific Fuel Consumption at different operating points as a function of EGR rate.
Figure 5. Engine Indicated Specific Fuel Consumption at different operating points as a function of EGR rate.
Figure 6. MFB 10% in Cyl #1 and Cyl #2 at different operating conditions as a function of EGR rate. leaving a shorter period available to complete the combustion process. Further increase in combustion duration against the EGR is found at lower engine load, because of the greater impact of internal EGR which contributes to increase the overall in-cylinder residual content at decreasing the load.
Figure 6. MFB 10% in Cyl #1 and Cyl #2 at different operating conditions as a function of EGR rate. leaving a shorter period available to complete the combustion process. Further increase in combustion duration against the EGR is found at lower engine load, because of the greater impact of internal EGR which contributes to increase the overall in-cylinder residual content at decreasing the load.
Figure 7. MFB 10-50% in Cyl #1 and Cyl #2 at different operating conditions as a function of EGR rate.
Figure 7. MFB 10-50% in Cyl #1 and Cyl #2 at different operating conditions as a function of EGR rate.
Figure 9. Numerical/experimental comparisons of combustion core duration MFB 10-50% for Cyl #1 (a) and Cyl #2 (b) at various operating conditions.
Figure 9. Numerical/experimental comparisons of combustion core duration MFB 10-50% for Cyl #1 (a) and Cyl #2 (b) at various operating conditions.
Figure 10. Numerical/experimental comparisons of ISFC (a) and TIT (b) at various operating conditions
Figure 10. Numerical/experimental comparisons of ISFC (a) and TIT (b) at various operating conditions
Figure 11. Numerical/experimental comparisons of nitrogen oxide (NO) for engine (a) and Cyl #1 (b) at various operating conditions. As for pollutants, Figures 11-13 show the comparison between numerical and experi- mental levels of regulated gaseous emissions: NO, CO and HC, respectively. Consistently with the presentation of the experimental results, the variations of considered pollutant emissions are reported as a function of the EGR rate.
Figure 11. Numerical/experimental comparisons of nitrogen oxide (NO) for engine (a) and Cyl #1 (b) at various operating conditions. As for pollutants, Figures 11-13 show the comparison between numerical and experi- mental levels of regulated gaseous emissions: NO, CO and HC, respectively. Consistently with the presentation of the experimental results, the variations of considered pollutant emissions are reported as a function of the EGR rate.
Figure 12. Numerical/experimental comparisons of percent carbon monoxide (CO%) for both engine and Cyl #1 at various operating conditions.
Figure 12. Numerical/experimental comparisons of percent carbon monoxide (CO%) for both engine and Cyl #1 at various operating conditions.
Figure 13. Numerical/experimental comparisons of engine unburned HC at various operating conditions.
Figure 13. Numerical/experimental comparisons of engine unburned HC at various operating conditions.
Figure 14 shows the ISFC comparisons in the analyzed engine operating points. Interestingly, this figure highlights that not negligible ISFC improvements are realized and the lower ISFC levels resulting from the balanced A/F ratio (blue bars in Figure 14) mostly follow the EGR related trend. A maximum ISFC percent benefit equal to 5% (AISFC = 12.4 g/kWh at 3000@7, EGR% = 9%) is attained, while the average percent gain is around 2%. Figure 14. ISFC comparisons: Unbalanced A/F ratio vs. Balanced Stoichiometric A/F ratio at varying the engine operating condition.
Figure 14 shows the ISFC comparisons in the analyzed engine operating points. Interestingly, this figure highlights that not negligible ISFC improvements are realized and the lower ISFC levels resulting from the balanced A/F ratio (blue bars in Figure 14) mostly follow the EGR related trend. A maximum ISFC percent benefit equal to 5% (AISFC = 12.4 g/kWh at 3000@7, EGR% = 9%) is attained, while the average percent gain is around 2%. Figure 14. ISFC comparisons: Unbalanced A/F ratio vs. Balanced Stoichiometric A/F ratio at varying the engine operating condition.
Related topics:
Materials ScienceAutomotive EngineeringExhaust Gas RecirculationHomogeneous Charge Compression Ignition (hcci)Ignition System
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