Combustion characteristics of a single-cylinder spark ignition gasoline and ethanol dual-fuelled engine

Transcription

1 Combustion characteristics of a single-cylinder spark ignition gasoline and ethanol dual-fuelled engine G Zhu*, D Hung, and H Schock Department of Mechanical Enginering, Michigan State University, East Lansing, Michigan, USA The manuscript was received on 10 April 2009 and was accepted after revision for publication on 16 October DOI: / CASE STUDY 1 Abstract: The requirement of reduced emissions and improved fuel economy led to the introduction of direct-injection (DI) spark ignition (SI) engines. A dual-fuel injection system (DI and port fuel injection (PFI)) was also used to improve engine performance at high-speed highload conditions. Ethanol is one of the several alternative transportation fuels considered for replacing fossil fuels such as gasoline and diesel. Ethanol offers high octane quality but with lower energy density than fossil fuels. This paper presents the combustion characteristics of a single-cylinder dual-fuel injection SI engine with the following fuelling cases: case A, gasoline for both PFI and DI; case B, gasoline PFI and ethanol DI; case C, ethanol PFI and gasoline DI. For this study, the DI fuelling portion varied from 0 per cent to 100 per cent of the total fuelling over different engine operational conditions while the engine air-to-fuel ratio remained at a constant level. It was shown in all cases that the indicated mean effective pressure (IMEP) decreases by as much as 11 per cent as the DI fuelling percentage increases, except in case B where the IMEP increases by 2 per cent at a light load. The combustion burn duration increases significantly at a light load as the DI fuelling percentage increases, but only moderately at wideopen throttle (WOT). In addition, the percentage of the ethanol in the total fuelling plays a dominant role in affecting the combustion characteristics at a light load but, at a heavy load (WOT), the DI fuelling percentage becomes an important parameter, regardless of the percentage of ethanol content in the fuel. Keywords: automotive engines, ethanol fuel, duel-fuel systems, combustion characteristics 1 INTRODUCTION Increasing concerns about global climate change and ever-increasing demands on fossil fuel capacity call for reduced emissions and improved fuel economy. Vehicles equipped with a direct-injection (DI) fuel system have been introduced to markets globally. In order to improve the DI engine full-load performance at high speeds, Toyota introduced an engine with a stoichiometric DI system with two fuel injectors for each cylinder [1]. One is a direct injector generating a dual-fan-shaped spray with wide dispersion, while the other is a port fuel injector. The dual-fuel system introduces one additional degree of *Corresponding author: Department of Mechanical Enginering, Michigan State University, 148 ERC South, East Lansing, MI 48824, USA. zhug@egr.msu.edu freedom for engine optimization to reduce emissions with improved fuel economy. Ethanol has been used widely as a fuel additive or an alternative fuel because of its high-octane and clean combustion. Early research [2] provided the physical and chemical fuel properties of ethanol that affect spark ignition (SI) and compression ignition engine performance. Owing to the wide ranges of fluid and chemical properties of ethanol fuels, it is widely recognized that alternative fuels such as 15 vol % gasoline 85 vol % ethanol (E85) would impose additional challenges for a fuel injector to deliver acceptable spray quality to meet the requirements of combustion chemistry and emissions in DI engines. For example, the high-viscosity low-evaporation characteristics of E85 fuels probably result in degraded spray quality, particularly in cold-start conditions [3]. Therefore, for DI SI engines capable of operating with ethanol fuels, the injection control

2 2 G Zhu, D Hung, and H Schock strategy and fuel system must be modified to accommodate for the higher fuel pump and injector flow capacities, fuel break-up, and spray characteristics of E85 fuels. Recently, renewed research in ethanol is mainly due to the concerns of global warming and transportation energy shortage [4 6]. Discussions on the application of ethanol to DI and turbocharged engines can be found in references [7] and [8]; finally, reference [9] presented the combustion and emission characteristics of a port fuel injection (PFI) ethanol homogeneous charge compression ignition (HCCI) engine. The use of a gasoline PFI ethanol DI dual-fuel system to increase the gasoline engine efficiency substantially has been described in reference [10]. The main idea is to use a highly boosted small turbocharged engine to match the performance of a much larger engine. DI of ethanol is used to suppress engine knock owing to its substantial air charge cooling which results from its high heat of vaporization. This paper investigates the combustion characteristics, and in particular the fuel economy properties, of a single-cylinder engine equipped with a dual-fuel system when different combinations of fuels (gasoline and E85) are used for PFI and DI fuel systems. For a given relative air-to-fuel ratio (AFR) l, the engine indicated mean effective pressure (IMEP), mass fraction burned (MFB), and burn duration (BD) were studied by varying the DI fuelling percentage from 0 per cent to 100 per cent (the fuelling percentages throughout are mass percentages) while either maintaining a fixed spark timing (ST) or conducting an ST sweep around the timing for the minimal advance for the best torque (MBT). Emissions-related characteristics are not discussed in this paper. The following fuelling combinations were studied: case A, gasoline for PFI and DI; case B, PFI gasoline and DI ethanol; case C, PFI ethanol and DI gasoline. The test results show that the percentage of ethanol in the total fuelling plays a dominant role in affecting the combustion characteristics at a light load but, at a heavy load (i.e. wide-open throttle (WOT)), the DI fuelling percentage becomes an important parameter, regardless of the percentage of ethanol content in the fuel. The paper is organized as follows. Section 2 describes the test set-up, and the test results of the three cases are presented in section 3. The combustion characteristics of the three cases are compared and discussed in section 4. Finally conclusions are drawn in section 5. Table 1 Engine parameter 2 TEST SET-UP The test parameters for the single-cylinder engine Value Bore 90.2 mm Stroke mm Connecting rod mm Compression ratio 9.8 Rated power per cylinder 37.5 hp at 5000 r/min Intake valve opening timing 330u BTDC* Intake valve closing timing 102u BTDC* Exhaust valve opening timing 86u ATDC { Exhaust valve closing timing 404u ATDC { *BTDC, before top dead centre. { ATDC, after top dead centre. The test data shown in this paper were obtained using a three-valve l single-cylinder engine (the engine parameters are given in Table 1). The singlecylinder engine was built based upon a PFI 5.4 l V8 engine with a rated power of 260 hp at 4500 r/min and a compression ratio of 9. The single-cylinder engine was equipped with a conventional PFI and Visteon s low-pressure direct-injection (LPDI) fuel systems [11, 12] (Fig. 1). The rail pressures of the PFI and LPDI fuel systems were operated at 3.5 bar and 20 bar respectively for all the dynamometer tests described in this paper. The end-of-injection (EOI) timings of the PFI and DI fuel systems are fixed at 360u and 300u respectively before top dead centre (BTDC). The EOI timings for both the PFI and the DI fuel systems were selected to have the best fuel economy based upon the EOI timing sweep. A fixed EOI was used for all tests because the IMEP sensitivity of the EOI at the selected values is relatively low for both E85 and gasoline. The fuel injection system incorporates two injectors, i.e. one on the intake manifold and one on the cylinder head. The port fuel injector, mounted on the intake port, was a production injector with a Fig. 1 Test set-up

3 Spark ignition gasoline and ethanol dual-fuelled engine 3 two-hole director plate and it formed a dual-plume spray pattern. The nominal fuel injection pressure was set to 350 kpa and the corresponding static flowrate was about 3 g/s. On the other hand, the direct injector was side mounted on to the cylinder head at an angle of 35u from the horizontal axis. The spray had a 60u spray angle with a 5u bent axis. The injector was designed to reduce direct fuel wetting on the spark plug and cylinder wall, and to propagate the spray towards the centre of the cylinder. The side-mounting location of 35u was chosen owing to the packaging constraint in the cylinder head without interference between the fuel injector and the existing coolant passage surrounding the cylinder head. It was a multi-hole production-intent injector configured with a nine-hole orifice plate. The injector was pressurized to about 2 MPa, giving a static flowrate of approximately 12.0 g/s. The internal nozzle geometry and geometrical parameters were designed to offer different spray characteristics. Using n-heptane as the standard test fuel according to SAE J2715 [13] recommended practice, the Sauter mean diameters (SMDs) measured across the spray at 50 mm below the injector tip were between 20 mm and 42 mm. These radial scan pointwise SMD measurements were then converted into a single line-of-sight SMD value by weighing and normalizing the measurement at each location with its corresponding flux density. The calculated SMD value was 33.4 mm at 2 MPa. Details of other spray characterizations performed on the direct fuel injectors and the LPDI fuel systems can be found in reference [14]. The SMDs were measured using phase Doppler interferometry. The test set-up and procedure were carried out according to SAE J2715 [13]. SMDs were only measured for the direct injector sprays, and not for the port fuel injector. The SMD value was given to represent a statistical drop diameter which was typical for an LPDI system with an injection pressure of 2 MPa. A laboratory-instrument-grade pressure sensor was flush mounted on the cylinder head and a relative AFR l meter was installed for relative AFR measurement and control. The universal exhaust gas oxygen sensor used for the l meter was installed on an exhaust pipe connected directly to the exhaust port of the cylinder head, where there was no exhaust manifold used for the single-cylinder engine. For each test point the engine speed is maintained at a desired constant speed and the engine throttle was kept constant to maintain a constant airflow while the DI-to-PFI fuelling ratio varies. The single-cylinder engine was controlled by a prototype engine controller for ST and dual-fuel injections, and the engine throttle and speed were regulated by the engine dynamometer controller. The single-cylinder engine has a single camshaft for both intake and exhaust valve timing regulations. The camshaft timing was manually advanced to optimize engine combustion. No exhaust gas recirculation was used for all the tests conducted. The in-cylinder pressure and relative AFR l signals were collected using a dynamometer data-sampling system with 1u crank angle resolution. For each test point, 300 cycles of test data were collected with 1u crank angle resolution. In the rest of this paper, the pressure signals used are averaged over 300 cycles, or otherwise specified. The engine IMEP, MFB, and BD are calculated on the basis of the pressure signals averaged over 300 cycles. The gasoline test fuel used for all consequent dynamometer tests is indolene, and the ethanol used for these tests is laboratory-grade E85, which contains 15 vol % gasoline and 85 vol % ethanol. 3 EXPERIMENTAL RESULTS This section presents the experimental results for three cases. They are as follows: case A, gasoline PFI and DI; case B, gasoline PFI and ethanol DI; case C, ethanol PFI and gasoline DI. The following sections describe the experimental results of the three cases. 3.1 Case A: gasoline PFI and DI In this section, the combustion characteristics of the dual-fuel SI single-cylinder engine when gasoline was used for both fuel injection systems are discussed. For each test point, the engine throttle and speed were held constant. The engine test started at 100 per cent PFI fuelling with a constant relative AFR l, and then the PFI fuelling was reduced to the desired level (e.g. 70 per cent of the original fuelling quantity) while the corresponding DI fuelling was increased to maintain the same relative AFR l. The test continued until 100 per cent DI fuelling was reached. Five PFI fuelling percentages were selected and they are 100 per cent, 70 per cent, 50 per cent, 30 per cent, and 0 per cent. For the ST sweep at each given PFI fuelling percentage, MBT timing was selected by adjusting the ST such that the 50 per cent MFB location remains at around 8 10u after top dead centre (ATDC). This method was discussed in reference [15].

4 4 G Zhu, D Hung, and H Schock Table 2 Test matrix for gasoline PFI and DI (case A) Engine speed (r/min) IMEP (bar) l ST (deg BTDC) WOT WOT Sweep The test matrix for this case is shown in Table 2, where l (the inverse of the equivalence ratio) is defined as the engine AFR divided by the stoichiometric AFR for the fuel mixture used in the test. For the engine operated at 1500 r/min with an IMEP load of 3.3 bar, the STs were selected as 34u, 37u, 40u, 43u, and 46u BTDC, where 40u BTDC was the MBT timing with 100 per cent PFI fuelling. Again, for each test point, the PFI fuelling percentages were selected as 100 per cent, 70 per cent, 50 per cent, 30 per cent, and 0 per cent. Figure 2 shows the engine IMEP as a function of the DI fuelling percentage when the engine was operated at 1500 r/min with an IMEP of 3.3 bar. It can be observed that in both cases the engine IMEP is reduced as the DI fuelling percentage increases, and the IMEP is reduced even more when the ST remains unchanged (and is not optimized for the MBT). Therefore, the solid curve with open circles shows the IMEP reduction due to the increased DI charge cooling; and the differences between the solid and dashed curves are due to fixed (and not optimized combustion phasing) ST as the DI fuelling percentage increases. Therefore, it is important to optimize the ST for combustion efficiency as the DI fuelling percentage increases. In summary, the reduction in the IMEP with increased DI fuelling percentage is mainly due to the DI charge cooling effect, which reduces the thermal efficiency of the combustion process. Figure 3 shows similar information to Fig. 2 for the engine operated at 2500 r/min and 3000 r/min with WOT. It is clear that the IMEP is reduced as the DI fuelling percentage increases at both speeds but, as the engine speed increases at WOT, the reduction in the IMEP becomes moderate. It is believed that at high engine speeds with WOT the ability of the charge cooling to affect the combustion rate is reduced significantly owing to the relatively high in-cylinder air fuel mixture temperature at high engine speeds. This is why at 3000 r/min the IMEP is almost unchanged with low DI fuelling percentages. Figure 4(a) presents the averaged engine in-cylinder pressure signals, and Fig. 4(b) the MFB signals calculated from the averaged in-cylinder pressure signals, when the engine was operated at 1500 r/min with an IMEP of 3.3 bar. The ST was adjusted to its MBT timing at each given DI fuelling percentage. Fig. 2 IMEP versus DI fuelling percentage for case A when the engine was operated at 1500 r/ min with an IMEP of 3.3 bar

5 Spark ignition gasoline and ethanol dual-fuelled engine 5 Fig. 3 IMEP versus DI fuelling percentage for case A when the engine was operated at 2500 r/ min or 3000 r/min with WOT Note that these operational conditions are associated with the solid curve with open circles in Fig. 2. The 70 per cent DI fuelling line was omitted in Fig. 4 since its MFB and in-cylinder pressure are very close to those for 100 per cent DI fuelling. It can be observed in Fig. 4 that, as the DI fuelling percentage Fig. 4 (a) Pressure and (b) MFB for case A when the engine was operated at 1500 r/min with an IMEP of 3.3 bar

6 6 G Zhu, D Hung, and H Schock increases, the combustion process slows down and the corresponding peak cylinder pressure is reduced. This confirms the fact that the engine IMEP is reduced as the DI fuelling percentage increases (see Fig. 2). The MFB signals were calculated by normalizing the net pressure signal obtained from the in-cylinder pressure signal described in equations (1) and (2) [15, 16]. The net pressure change DP(i) between two crank angles is defined by ( ) V (i) n~1:3 V (i) DP(i)~ P(iz1){P(i) ð1þ V (iz1) V Ig and therefore the net pressure at each crank angle is P NET (i)~p NET (i{1)zdp(i) ð2þ where n is the constant polytrophic index, P is the in-cylinder pressure, V is the chamber volume, and V Ig is the chamber volume at the ignition point. Note that the net pressure represents the pressure increase due to the combustion only since the pressure variation due to piston movement is excluded during the calculation. After the net pressure is obtained, the MFB can be obtained by the normalization calculation MFB(i)~ P NET(i) P NET MAX where ð3þ P NET MAX ~ maxfp NET (1) P NET (2) P NET (m) g with m an index corresponding to the end-ofcombustion crank angle. The calculation of the MFB from the net pressure signal assumes that the net pressure has the same shape as the MFB. Figure 5 shows the calculated BD from 10 per cent to 90 per cent MFB for the engine operated at 1500 r/ min with an IMEP of 3.3 bar, while the data at 2500 r/ min and 3000 r/min were with WOT. As mentioned before, the ST at 1500 r/min with an IMEP of 3.3 bar was adjusted to the engine MBT timing, while at 2500 r/min and 3000 r/min the ST was fixed at 20u BTDC which is very close to its MBT timing. As discussed before, the per cent BD increases in all three cases as the DI fuelling percentage increases, which indicates that, for all three operational conditions, the BD increases because of increased charge cooling. As a summary of this section, it was concluded that for case A (gasoline for both PFI and DI fuel systems), both the engine IMEP and the peak Fig. 5 BD for case A when the engine was operated at various speeds and under different loading conditions

7 Spark ignition gasoline and ethanol dual-fuelled engine 7 cylinder pressure decrease as the DI fuelling percentage increases. In addition, the IMEP reduction becomes severe at WOT compared with part-load conditions because of the heavier charge cooling effect at WOT than at a light load. Similarly, the per cent BD increases as the DI fuelling percentage increases. 3.2 Case B: gasoline PFI and ethanol DI In this section, the combustion characteristics of the dual-fuel single-cylinder engine when gasoline was used for the PFI fuel system and E85 was used for the DI fuel system are discussed. Similar to all gasoline case described in the previous section, for each test point, the engine throttle and speed were held constant. Engine tests started at 100 per cent PFI gasoline fuelling with a given relative AFR l, and then the PFI fuelling was reduced to the desired level while the corresponding DI E85 fuelling percentage was increased to maintain the same relative AFR l or the same equivalence ratio. The tests continued until 100 per cent DI E85 fuelling was reached. Similar to case A, five PFI fuelling percentages were selected at 100 per cent, 70 per cent, 50 per cent, 30 per cent, and 0 per cent. An ST sweep was also conducted at 1500 r/min, where, for a given PFI fuelling percentage, MBT timing was selected by adjusting the ST such that the 50 per cent MFB location remained at about 8 10u ATDC. The test matrix for this case is shown in Table 3. For the operational condition at 1500 r/min with an IMEP load of 3.3 bar, the swept STs were selected as 34u, 37u, 40u, 43u, and 46u BTDC, where 40u BTDC is the MBT timing for 100 per cent PFI fuelling. At 1500 r/min with WOT, the STs were selected as 14u, 17u,20u,23u, and 27u BTDC, where the MBT timing is 20u BTDC for 100 per cent PFI fuelling. Again, for each test point, the PFI fuelling percentages were selected as 100 per cent, 70 per cent, 50 per cent, 30 per cent, and 0 per cent. Tests were also conducted with fixed ST at 1500 r/ min with WOT, with an IMEP of 3.3 bar, and with an IMEP of 5.5 bar and at 2500 r/min with WOT (see Table 3 Test matrix for gasoline PFI and E85 DI (case B) Engine speed (r/min) IMEP (bar) l ST (deg BTDC) WOT WOT Sweep 1500 WOT Sweep Table 3 for the ST), where for each test point, the gasoline PFI fuelling was selected to be at 100 per cent, 70 per cent, 50 per cent, 30 per cent, and 0 per cent. Figure 6 shows the engine IMEP as a function of the DI E85 fuelling percentage when the engine was operated at 1500 r/min with an IMEP of 3.3 bar and with WOT. Note that the engine ST was optimized at its MBT. It is interesting to see that, for the engine operated with an IMEP of 3.3 bar, the actual engine IMEP increases as the DI E85 fuelling percentage increases while, when the engine is operated at WOT, its IMEP decreases as the DI E85 fuelling percentage increases. It can be seen that the IMEP with 50 per cent DI fuelling is slightly lower than that with 0 per cent and the IMEPs with 70 per cent and 100 per cent DI fuelling are higher than that with 0 per cent when the engine operates with an IMEP of 3.3 bar. When the engine operates at WOT, the IMEP with 70 per cent DI fuelling is higher than those with 50 per cent and 100 per cent DI fuelling. To investigate why under different conditions the engine IMEP varies in the opposite directions as the DI ethanol fuelling percentage increases, the total fuel energy injected was calculated on the basis of the fuel injection quantity of both DI and PFI. Specifically, the DI fuelling percentage was adjusted according to the change in the PFI fuelling dynamic flow (injected quantity per pulse) on a mass basis. Then, the injection mass for each injector was determined from the injector mass duration calibration curves. Therefore, the total energy injected TE Inj is the sum of the energy injected from each injector based on the lower heating value (LHV) of the specific fuel and the injected quantity according to TE Inj ~LHV gasoline m gasoline zlhv E85 m E85 ð4þ where m gasoline and m E85 are the injected masses of gasoline and ethanol (E85) respectively. Figure 7 shows the calculated injected energies, normalized to the fuel energy with 100 per cent PFI fuelling, in terms of the DI fuelling percentage. It can be seen from this figure that, in both cases, the total energy injected increases as the DI fuelling percentage increases. However, at a light load (IMEP, 3.3 bar), the increase is more significant (9 per cent) than that at WOT (2.6 per cent). Consider the two main factors that affect the engine combustion efficiency: DI charge cooling and injected fuel energy. It becomes clear that, in the light-load condition (IMEP, 3.3 bar), the increase in the injected energy (9 per cent) due to increasing DI

8 8 G Zhu, D Hung, and H Schock Fig. 6 IMEP versus DI fuelling percentage for case B when the engine was operated at 1500 r/ min with an IMEP of 3.3 bar or with WOT fuelling percentage dominates the combustion process and also, in the light-load condition, the charge cooling is relatively less dominated than at WOT; in contrast, in the heavy-load condition (WOT), the charge cooling effect becomes a key factor since the fuelling energy increment is moderate at 2.6 per Fig. 7 Fuel energy injected versus DI fuelling percentage for case B when the engine was operated at 1500 r/min with an IMEP of 3.3 bar or with WOT

9 Spark ignition gasoline and ethanol dual-fuelled engine 9 cent. This indicates that, for a dual-fuel (gasoline PFI and ethanol DI) system engine to improve the combustion efficiency at high DI fuelling percentage, it is necessary to reduce the charge cooling effect in high-load operation conditions. One approach is to have an adjustable engine compression ratio (e.g. to operate the engine in an Atkinson cycle) for the best combustion efficiency. Figure 8 shows similar information to Fig. 6 for fixed-st tests at 1500 r/min and 2500 r/min under different load conditions. It can be observed that, at 1500 r/min with a light load (IMEP, 3.3 bar) the normalized IMEP is reduced and then increases as the DI fuelling percentage increases. As the load increases at 1500 r/min, the IMEP is reduced while the DI E85 fuelling percentage increases; see the curves in Fig. 8 for 1500 r/min with an IMEP of 5.5 bar and with WOT. This effect is repeated at 2500 r/ min with WOT. It is believed that this mainly arises because the engine ST was not optimized at its MBT. Figure 9 shows the calculated BD from 10 per cent to 90 per cent MFB for the engine operated at 1500 r/ min with an IMEP of 3.3 bar and with WOT. As discussed before, the ST was adjusted to the engine MBT timing. These test points are associated with the IMEP curves shown in Fig. 6. It can be observed that the per cent BD increases in both conditions as the DI E85 fuelling percentage increases. This is also true for the cases of fixed timing shown in Table 3 and, therefore, the plot is not presented. It can be concluded that, for the case of gasoline PFI and E85 DI at WOT, the engine IMEP and peak cylinder pressure decrease as the DI E85 fuelling percentage increases. However, at a light load, the engine IMEP and peak cylinder pressure decrease first and then increase as the DI E85 fuelling percentage increases. This is mainly because, at a light load, the combustion process is dominated by the increase in the total fuel energy injected while, at WOT, the charge cooling effect is more important. In general, the per cent BD increases as the DI E85 fuelling percentage increases. 3.3 Case C: ethanol PFI and gasoline DI In this section, the combustion characteristics of the dual-fuel single-cylinder engine when ethanol (E85) was used for the PFI fuel system and gasoline was used for the DI fuel system are discussed. Similar to the two previous cases, for each test point, the engine throttle and speed were held constant, and all tests started at 100 per cent PFI E85 fuelling with a Fig. 8 IMEP versus DI E85 fuelling percentage with fixed ST for case B when the engine was operated at various speeds and under different loading conditions

10 10 G Zhu, D Hung, and H Schock Fig. 9 BD versus DI E85 fuelling percentage for case B when the engine was operated at 1500 r/ min with an IMEP of 3.3 bar or with WOT given relative AFR l. Then, PFI fuelling was reduced to the desired level while the corresponding DI gasoline fuelling was increased to maintain the same relative AFR l or the same equivalence ratio. The test continued until 100 per cent DI gasoline fuelling was reached. Five PFI fuelling percentages were selected and they were 100 per cent, 70 per cent, 50 per cent, 30 per cent, and 0 per cent. An ST sweep was also conducted at two test points (1500 r/min with an IMEP of 3.3 bar and 1500 r/min with WOT) where, for a given PFI fuelling percentage, the MBT timing was selected by adjusting the ST such that the 50 per cent MFB location remains at around 8 10u ATDC. The test matrix for this test case is shown in Table 4. Owing to the significant MBT timing variations for different DI gasoline fuelling percentages at 1500 r/min with an IMEP load of 3.3 bar, the ST sweep windows were quite different. For each Table 4 Test matrix for E85 PFI and gasoline DI (case C) Engine speed (r/min) IMEP (bar) l ST (deg BTDC) WOT WOT WOT Sweep 1500 WOT Sweep fuel ratio test point, a nominal ST, called the centre timing (CT), was selected. For this test, the CTs were selected as 35u, 42u, 44u, and 47u BTDC corresponding to 100 per cent, 50 per cent, 30 per cent, and 0 per cent PFI fuelling respectively. The STs used for each test point were CT, CT 3, and CT 6. At 1500 r/min with WOT, the ST sweeps were fixed as 14u, 17u, 20u, 23u, and 27u ATDC, where the engine MBT timing is around 20u BTDC. Again, for each test point, the PFI fuelling percentages were selected as 100 per cent, 70 per cent, 50 per cent, 30 per cent, and 0 per cent. Tests were also conducted with a fixed ST at 1500 r/min with WOT, an IMEP of 3.3 bar, and an IMEP of 5.5 bar and at 2500 r/min and 3000 r/min with WOT where, for each test point, the ethanol PFI fuelling percentages were selected as 100 per cent, 70 per cent, 50 per cent, 30 per cent, and 0 percent. Figure 10 shows the normalized engine IMEP as a function of the DI gasoline fuelling percentage when the engine was operated at 1500 r/min with an IMEP of 3.3 bar and with WOT. It can be observed that, when the engine was operated with WOT, the engine IMEP is reduced by 3 per cent at 100 per cent DI gasoline compared with that at 100 per cent PFI E85; at an IMEP of 3.3 bar, the engine IMEP decreases by as much as 11 per cent at 100 per cent DI fuelling

11 Spark ignition gasoline and ethanol dual-fuelled engine 11 Fig. 10 IMEP versus DI fuelling percentage for case C when the engine was operated at 1500 r/ min with an IMEP of 3.3 bar or with WOT compared with that at 100 per cent PFI fuelling. It is believed that the IMEP reduction is due to both the increased charge cooling and the reduced injected energy. Figure 11 shows similar information to Fig. 10 for the fixed-st tests at 1500 r/min, 2500 r/min, and 3000 r/min under different loads. It can be observed that in general the normalized percentage IMEP is reduced as the DI gasoline fuelling percentage increases, mainly because of the DI charge cooling. For the case of PFI E85 fuelling and DI gasoline fuelling, Fig. 12(a) presents the averaged engine incylinder pressure signals, and Fig. 12(b) the MFB signals calculated from the averaged in-cylinder pressure signals, when the engine was operated at 1500 r/min with an IMEP load of 3.3 bar. The ST was adjusted to its MBT timing at each given DI fuelling percentage. Note that this corresponds to the IMEP signal indicated by the solid line in Fig. 10. Only the signals associated with 0 per cent, 50 per cent, and 100 per cent DI gasoline fuelling are plotted to improve the readability. It is clear that, as the DI percentage of gasoline fuelling increases, the combustion slows down and the peak cylinder pressure is reduced. Figure 13(a) shows the averaged engine in-cylinder pressure signals and Fig. 13(b) the MFB signals calculated from the averaged in-cylinder pressure signals, when the engine was operated at 1500 r/min with WOT. The engine was operated at its MBT timimg at each given DI fuelling percentage. Note that these operational conditions are associated with the IMEP indicated by the dashed curve in Fig. 10. For clarity, only the signals associated with 0 per cent, 50 per cent, and 100 percent DI E85 fuelling are plotted. Figure 13 illustrates that the peak in-cylinder pressure signal decreases as the DI gasoline fuelling percentage increases, and the MFB signals show that the combustion slows down when the DI fuelling percentage increases. This correlates to the IMEP curve presented in Fig. 10; i.e. the IMEP decreases as the DI gasoline fuelling percentage increases. Figure 14 shows the calculated per cent BD for the engine operated at 1500 r/min with an IMEP load of 3.3 bar and with WOT. As discussed before, the ST was adjusted to engine s MBT timing, which corresponds to the IMEP curves shown in Fig. 10. It can be observed that the per cent BD increases significantly at 1500 r/min with an IMEP load of 3.3 bar but, with WOT, the BD remains relatively unchanged. This correlates well with the IMEP signal shown in Fig. 10 since, with WOT, the IMEP is reduced moderately as the DI gasoline percentage increases. On the other hand, the IMEP reduction is also correlated with the reduced injected energy, similar to case B shown in Fig. 7.

12 12 G Zhu, D Hung, and H Schock Fig. 11 IMEP versus DI fuelling with fixed timing for case C when the engine was operated at various speeds and under different loading conditions Fig. 12 (a) Pressure and (a) MFB signals with for case C when the engine was operated at 1500 r/ min with an IMEP of 3.3 bar For the case of PFI E85 fuelling and DI gasoline fuelling, in general, the engine IMEP and peak cylinder pressure signal decrease as the DI gasoline fuelling percentage increases. In addition, the reduction in the IMEP is more significant at a light load than at WOT. The per cent BD increases

13 Spark ignition gasoline and ethanol dual-fuelled engine 13 Fig. 13 (a) Pressure and (b) MFB signals for case C when the engine was operated at 1500 r/min with WOT Fig. 14 BD versus gasoline DI percentage for case C when the engine was operated at 1500 r/min with an IMEP of 3.3 bar or with WOT

14 14 G Zhu, D Hung, and H Schock Fig. 15 Comparison of IMEPs for (a) case B when the engine was operated at 1500 r/min with an IMEP of 3.3 bar and (b) case C when the engine was operated at 1500 r/min with WOT as the DI gasoline fuelling percentage increases sharply at a light load, but only moderately with WOT. 4 DISCUSSION OF THE CASES In this section, the test results for different test cases are compared. Particular attention is devoted to comparing the results between cases B and C. Figure 15 shows how the engine IMEP changes as a function of the DI fuelling percentage. Figure 15(a) is related to the engine operated at 1500 r/min with an IMEP load of 3.3 bar, and Fig. 15(b) is for the engine operated at 1500 r/min with WOT. It is interesting to see that, with the WOT condition, the IMEPs in the two cases are reduced in a similar way but, in the light-load condition (IMEP, 3.3 bar), the situation is quite different. For case B, the engine IMEP increases slightly as the DI ethanol fuelling percentage increases but, for case C, the IMEP is reduced sharply, as shown in Fig. 15(a). As illustrated in Fig. 7, the percentage of ethanol fuelling increases while the total fuel energy injected increases for a given relative AFR l. For case B, as the DI gasoline fuelling percentage increases, the sharp decrease in the engine IMEP could be due to both the reduction in the total fuelling energy and the increase in the DI charge cooling effect. In contrast, for case C, as the DI ethanol (E85) fuelling percentage increases, the fact that the engine IMEP remains almost unchanged might be due to the combined effect of the increase in the total fuelling energy (the increase in the IMEP) and the increase in the DI charge cooling effect (the decrease in the IMEP). It can be concluded that, at a light load, the percentage of ethanol (E85) plays a dominant role. However, at a heavy load (WOT), the percentage of DI fuelling, regardless of the percentage of ethanol fuelling, is the controlling factor. Table 5 summarizes the variations in the IMEP and BD when the DI fuelling percentage increases Table 5 Case Summary table for an engine speed of 1500 r/min Variation in the following from 0 per cent to 100 per cent DI fuelling IMEP (3.3 bar) (per cent) IMEP with WOT (per cent) BD with an IMEP of 3.3 bar (deg) A 22 N/A* +2 N/A* B C *N/A, not applicable. BD with WOT (deg)

15 Spark ignition gasoline and ethanol dual-fuelled engine 15 from 0 per cent to 100 per cent. In almost all cases, the IMEP is reduced except in case B with a light load, when the IMEP increases. The BD increases more in the light-load condition than at WOT. What was learned from the results shown in Table 5 is that, with WOT, there is no significant combustion characteristic deviation for both cases B and C. In general, the engine IMEP is reduced slightly as the DI fuelling percentage increases, while the per cent BD remains almost unchanged. Therefore, both dual-fuel system configurations (gasoline PFI ethanol DI and ethanol PFI gasoline DI) provide similar combustion characteristics at WOT. On the contrary, in the lightload condition, case B (gasoline PFI ethanol DI) results in an improved engine IMEP as the DI fuelling percentage increases and, in case C (ethanol PFI gasoline DI), the engine IMEP decreases significantly. Therefore, special attention needs to be paid to the light-load operating conditions when designing a mixed dual-fuel system to optimize the engine performance. Figure 16 shows the coefficient of variation (COV) of the IMEP as a function of the DI fuelling percentage for cases A, B, and C when the engine is operated at 1500 r/min with WOT. It can be observed that at 100 per cent PFI fuelling the engine has the lowest COV for all three cases (below 2.5 per cent); this is primarily because the PFI wall-wetting dynamics reduce the fuel injection shot-to-shot variation. As the DI fuelling percentage increases, it can be seen that the COV increases; this could arise because of the increased shot-to-shot variations in DI and decreased wall-wetting dynamics effect due to reduced PFI fuelling. Also, when both injections split the fuel injection quantity, the relative shot-toshot variation increases since the injectors may be operated close to their non-linear operational conditions owing to low injection masses. At 100 per cent DI rate the COV reduces to a level higher than the 100 per cent PFI case since the shot-to-shot variation in the DI fuel system is normally higher than the PFI shot-to-shot variation but, on the other hand, the PFI injection shot-to-shot variation is eliminated. For future work, further experiments will be performed to study the in-cylinder pressure and combustion duration together with the emissions data to provide a more comprehensive investigation of the dual-fuel injection combustion process. Fig. 16 COV of the IMEP as a function of the DI fuelling percentage when the engine was operated at 1500 r/min with WOT

16 16 G Zhu, D Hung, and H Schock 5 CONCLUSIONS This paper presents the combustion characteristics of a single-cylinder dual-fuel injection SI engine with different configurations of its dual-fuel system that include gasoline PFI and DI (case A), gasoline PFI and E85 DI (case B), and E85 PFI and gasoline DI (case C). For each case, the DI fuel percentage was varied from 0 per cent to 100 per cent while the engine AFR remained constant. It has been shown that in all cases the IMEP decreases by as much as 11 per cent as the DI fuelling percentage increases, except for case B at a light engine load when the IMEP increases by 2 per cent. The combustion BD increases significantly as the DI fuelling percentage increases at a light load, but only moderately at WOT. Specifically, in light-load conditions, the DI fuelling increment leads to a sharp decrease (211 per cent) in the engine IMEP for case B, but only a slight increase (2 per cent) for case C. Also, test results shows that, as the percentage of ethanol fuelling increases, the total fuel energy injected increases for a given AFR. For case B, as the DI gasoline fuelling percentage increases, a significant decrease (11 per cent) in the engine IMEP was observed owing to both the reduction in the total fuelling energy injected and the increase in the DI charge cooling effect. In contrast, for case C, as the DI ethanol (E85) fuelling increases, a relative steady engine IMEP increase (2 per cent) might be caused by the combined effect of the increase in the total fuelling energy increment (the increase in the IMEP) and a more pronounced DI charge cooling effect (the decrease in the IMEP). As a result, the percentage of ethanol (E85) plays a dominant role in the combustion process at a light load. However, in heavy-load (WOT) conditions, the percentage of DI fuelling, regardless of the percentage of ethanol, is the controlling factor. Because of the wide ranges of fluid and chemical properties of ethanol fuels, it is widely recognized that alternative fuels such as E85 would impose additional challenges for a fuel injector to deliver acceptable spray quality to meet the requirements of combustion chemistry and emissions in GDI engines. For example, the high-viscosity low-evaporation characteristics of E85 fuels may probably result in degraded spray quality, particularly in cold-start conditions. Therefore, for SI DI engines capable of operating with ethanol fuels, the injection control strategy and fuel system must be modified to accommodate the higher fuel pump and injector flow capacities, fuel break-up, and spray characteristics of E85 fuels. F Authors 2010 REFERENCES 1 Ikoma, T., Abe, S., Sonoda, T., Suzuki, H., Suzuki, Y., and Basaki, Y. Development of V liter engine adopting new direct injection system. SAE paper , Sinor, J. E. and Bailey, B. K. Current and potential future performance of ethanol fuels. SAE paper , Tsunooka, T., Hosokawa, Y., Utsumi, S., Kawai, T., and Sonoda, Y. High concentration ethanol effect on SI engine cold startability. SAE paper , Li, L., Liu, Z., Wang, H., Deng, B., Xiao, Z., Wang, A., Gong, C., and Su, Y. Combustion and emissions of ethanol fuel (E100) in a small SI engine. SAE paper , Varde, K. S. and Clark, C. P. A comparison of burn characteristics and exhaust emissions from offhighway engines fueled by E0 and E85. SAE paper , Nakata, K. and Utsumi, S. The effect of ethanol fuel on a spark ignition engine. SAE paper , Taniguchi, S., Yoshida, K., and Tsukasaki, Y. Feasibility study of ethanol applications to a direct injection gasoline engine. SAE paper , Kapus, P. E., Fuerhapter, A., Fuchs, H., and Fraidl, G. K. Ethanol direct injection on turbocharged SI engines potential and challenges. SAE paper , Zhang, Y., He, B., Xie, H., and Zhao, H. The combustion and emission characteristics of ethanol on a port fuel injection HCCI engine. SAE paper , Bromberg, L., Cohn, D., and Heywood, J. Optimized fuel management system for direct injection ethanol enhancement of gasoline engines. US Pat , Xu, M., Porter, D. L., Daniel, C. F., Panagos, G., Winkelman, J., and Munir, K. Soft spray formation of a low-pressure high-turbulence fuel injector for direct injection gasoline engines. SAE paper , Hung, D. L. S., Mara, J. P., and Winkelman, J. R ; Tailoring the spray pattern of multi-hole fuel injectors for gasoline DI engine homogeneouscharge combustion. In Proceedings of the 18th Annual Conference on Liquid Atomization and Spray Systems (ILASS-Americas), Irvine, California, USA, May SAE J2715 Gasoline fuel injector spray measurements and characterizations, 2007 (SAE International, Warrendale, Pennsylvania). 14 Hung, D. L. S., Zhu, G. G., Winkelman, J., Stuecken, T., Schock, H., and Fedewa, A. A high speed flow visualization study of fuel spray pattern effect on mixture formation in a low pressure direct

17 Spark ignition gasoline and ethanol dual-fuelled engine 17 injection gasoline engine. SAE paper , Zhu, G. G., Daniels, C. F., and Winkelman, J. MBT timing detection and its closed-loop control using in-cylinder pressure signal. SAE paper , Mittal, M., Zhu, G., and Schock, H. Fast massfraction-burned calculation using the net pressure method for real-time applications. Proc. IMechE, Part D: J. Automobile Engineering, 2009, 223(3), DOI: / JAUTO1006. APPENDIX Notation AFR ATDC BD BTDC air-to-fuel ratio after top dead centre burn duration before top dead centre COV CT DI EOI E85 HCCI IMEP LHV LPDI MBT MFB PFI SI SMD ST WOT coefficient of variation centre timing direct injection end of injection fuel blended with 15 vol % gasoline and 85 vol % ethanol homogeneous charge compression ignition indicated mean effective pressure lower heating value low-pressure direct injection minimal advance for the best torque mass fraction burned port fuel injection spark ignition Sauter mean diameter spark timing wide-open throttle