RESE ARCH G A S OLINE ENGINES IC AR TI LO IN MTZ. | STR DU IN ED B Y Y. RECEIVED 2014-04-30 REVIEWED 2014-05-02 ACCEPTED 2014-09-04 IEW 46 FOR SCIENTIF PEER REVIEW REV PROF. DR.-ING. STEFAN PISCHINGER is Director of the Institute for Internal Combustion Engines (VKA) at the RWTH Aachen University (Germany). VA L ES | DIPL.-ING. MARCO GÜNTHER is Chief Engineer at the Institute for Internal Combustion Engines (VKA) at the RWTH Aachen University (Germany). RO F P AP CL DIPL.-ING. PATRICK HOPPE is Research Associate at the Institute for Internal Combustion Engines (VKA) at the RWTH Aachen University (Germany). To reduce CO2 emissions, a reduction in swept volume while at the same time fulfilling the same performance requirements, also referred to as downsizing, has long been considered to be an established solution in the automotive industry. The limitations of the reduction in displacement and its potential, in particular using biogenic, ethanolcontaining fuels, were emphasised as part of a FVV research project. For this purpose, a direct injection 0.8-l three-cylinder gasoline engine with a specific power output of 120 kW/l was developed, set up, and its combustion system was subjected to thermodynamic investigations at the Institute for Internal Combustion Engines of the RWTH Aachen University. SEA DIPL.-ING. BASTIAN LEHRHEUER is Research Associate at the Institute for Internal Combustion Engines (VKA) at the RWTH Aachen University (Germany). CO 2 REDUCTION THROUGH SMALL DISPLACEMENT IN COMBINATION WITH BIOFUELS THE AUTHORS EX PE RTS R FROM RESEA CH AN D 1 MOTIVATION 2 TEST OB JECTS AND FUELS 3 RESULTS OF THE TEST BENCH INVESTIG ATIONS 4 SUMMARY AND OUTLO OK 1 MOTIVATION In order to reach the same torque, it is necessary to increase the mean effective pressure when the displacement is reduced. This initially leads to less gas-exchange losses. Additionally, the relative wall heat losses and friction losses also decrease due to the higher load [1]. Consequently, the operating point of the downsized engine lies in an engine performance range with a higher efficiency, which leads to lower specific fuel consumption. Achieving comparable maximum performance and torque characteristics requires a high specific power output and therefore an increase in the maximum mean effective pressure by using suitable charging units. However, higher temperatures and pressures at the end of the compression stroke in the end gas require higher knock limits and thus lead to losses in efficiency [2] due to retarded spark timing at high loads or an adjusted lower compression. Fuels with a higher knock-resistance counteract this disadvantage so that using recoverable, biogenic blended ethanol fuels which feature a high octane number is a good idea. In addition to the above mentioned direct CO2 reduction potential thanks to engine efficiency, the savings during the production of fuel must be rated positively, if the energy required for the production process is not disproportionally high [3, 4]. 2 TEST OBJECTS AND FUELS The engine block and the crank assembly of a 0.8 l three-cylinder diesel engine are used as a basis for the test object. The high stability of the diesel engine permits maximum mean cylinder pressures of up to pcyl, max = 142 bar, which takes the high peak pressures to be expected into account with reasonable certainty, in particular when using ethanol-containing fuels. The strength benefits are accompanied by a frictional torque that is unusually high for gasoline engines. This must be taken into account when evaluating the effective efficiency. A new cylinder head was developed using 1-D gas cycle simulations as well as 3-D CFD computations with the emphasis on FUEL E10MB Density at 20 °C 746.2 kg/m Research octane number fatigue, cooling, and flow. The 3-valve concept that is preferable because of the flow and cooling benefits features an integrated, cooled exhaust manifold as well as a tumble port that is increased by masking the combustion chamber and thanks to the special design of the intake ports. In contrast to the basic engine, the valve assembly is driven by a dry-running timing belt. The valve assembly has cam phaser at the intake and exhaust side, which permit an adjusting range of 50 °CA. This results in a maximum valve overlap of ∆(EVC-IVO) = 73 °CA with a valve lift of hv = 1 mm. As a result, a very wide adjusting range is available for dethrottling by means of internal exhaust gas recirculation at partial load or scavenging to increase the torque at full load. The central location of the spark plug in combination with the small bore diameter of d = 65.5 mm has a positive effect on the knock resistance. This spark plug positioning results in a slightly off-centre injector position. The design of the asymmetric spray pattern of the multi-hole injectors represents a compromise between homogenisation and wetting of the cylinder wall or the valves. Furthermore, there should be the option for an ignitionlinked injection to optimise the cold start and warm-up behaviour. As part of the investigations, three different injection nozzles were compared with each other. In order to achieve high brake mean effective pressures of up to BMEP = 27.5 bar, the engine is equipped with a two-stage charging system. It consists of a turbocharger and a mechanical compressor that can be coupled via an electromagnetic dry clutch. In addition to high boost pressure, this permits very good instationary behaviour, even in the low engine speed range, whereby fuel consumption and the effort required for the regulation are acceptable. In view of the systematic design of the research engine that is geared towards operation with ethanol-containing fuels, compression ratios of CR = 11 that are relatively high for a supercharged engine, or CR = 13 in a further expansion stage, were analysed. The design of the piston crown represents a compromise between high compression ratio, valve clearance, and the longest possible unobstructed path length of the individual spray jets of the injectors. ❶ provides an overview of the properties of the three analysed fuels E10MB, E20SB, and E85SB, as well as that of pure ethanol in comparison. The stoichiometric air-to-fuel ratio (Lst) and the lower heating value (HU) of E20SB are reduced compared to E10MB because of the presence of the OH group in the ethanol molecule. In E85SB, this effect becomes more pronounced due to the increased amount of ethanol. In spite of the higher combus- E20SB 3 767.2 kg/m 3 E85SB ETHANOL 781 kg/m³ 808.8 kg/m 3 95.7 102.2 106.1 108.6 Lower calorific value 41.68 MJ/kg 39.71 MJ/kg 29.16 MJ/kg 27.06 MJ/kg Carbon mass fraction 83.38 % 79.27 % 56.9 % 49.80 % Hydrogen mass fraction 13.36 % 13.49 % 13.2 % 13.00 % Oxygen mass fraction 3.26 % 7.24 % 29.9 % 37.20 % 94.9 g/mol 87.4 g/mol 55 g/mol 42.9 g/mol Ethanol mass fraction 9.2 % 20.8 % 86.1 % 99.9 % Stoichiometric air requirement 14.6 13.3 9.8 9 428.5 kJ/kg 495.5 kJ/kg 778.2 kJ/kg 1014.0 kJ/kg Mean molar mass Vaporisation enthalpy ❶ Chemical and physical fuel properties 11I2014 Volume 75 47 RESE ARCH G A S OLINE ENGINES ❷ Variation of valve timing at n = 2000 rpm and BMEP = 2 bar tion efficiency, the lower calorific value nevertheless results in higher volumetric fuel consumption. However, the carbon content that decreases with increasing ethanol content compensates for this effect when considering at CO2 emissions. tage of better homogenisation due to early opening of the intake valves outweighs a later exhaust event, which is preferable when it comes to expansion losses. Taking a safe distance of 10 °CA 3 RESULTS OF THE TEST BENCH INVESTIGATIONS The goal of the experimental investigations was to highlight the potential of this extreme downsizing concept aimed at reducing fuel consumption, as well as its limitations with regard to the increase in the specific power output. For this purpose, the engine was tested several hundred hours on the engine test bench. In addition to performing basic investigations during partial load and wide open throttle, the oil dilution tendency, the heat-up characteristics, and the transient properties of the engine were analysed. A few of the results will be presented by way of example in the following. 3.1 PARTIAL LOAD The part load investigations started by varying the valve timing at an engine speed of n = 2000 rpm and a brake mean effective pressure of BMEP = 2 bar with the goal of achieving an optimum gas cycle as a compromise between low specific fuel consumption and adequate combustion stability. ❷ shows the selected variables for this variation in each case above the valve timing exhaust valve closes (EVC) and intake valve opens (IVO) at a lift of 1 mm. It is clear to see that the fuel consumption drops with an increasing valve overlap and the resulting dethrottling. On the other hand, the increased residual gas portion leads to an unstable combustion all the way down to misfires. The location of the maximum fuel conversion leads to the conclusion that the advan- 48 ❸ Variation of start of injection and rail pressure to timing with first combustion misfires into account, a moderate overlap of EVC = -5 °CA after TDC and IVO = -18 °CA after TDC was selected. As a compromise between friction, combustion stability, and soot emissions, injection timing and pressure were derived for partial load operation in further investigations. An early start of injection has a positive effect on the mixture formation here and thus on the stability, expressed by the standard deviation of the indicated mean effective pressure, ❸ (middle). In contrast to this trend, a start of injection before 340 °CA after TDC leads to increased piston wetting and consequently to increased soot emissions. The rail pressure also affects the development of soot emissions through droplet diameter and penetration depth or wall wetting. ③ below shows that rail pressures of below 100 bar are not an effective solution here in spite of a low frictional mean pressure. While no significant efficiency advantages can be noted by using high-octane ethanol fuels in the lower partial-load range, their increased knock resistance becomes beneficial in the upper load ranges and at wide open throttle, ❹. In contrast to conventional E10MB fuel with an octane number of 95.7, the engine with E20SB (RON = 102.2) and E85SB (RON = 106.1) can also be operated in the upper load range for a longer period of time or with complete optimal efficiency with a centre of combustion of approximately 8 °CA after TDC. As shown at the bottom of ③, the presence of oxygen in the ethanol molecule significantly reduces the tendency of soot formation. This advantage is especially pronounced in the higher load range. As a result of the low calorific value of ethanol a larger volumetric fuel flow rate must be supplied with increasing ethanol content at the same load. In order to avoid long injection times and associated poor mixture formation, the injection pressure has been increased to the maximum of 250 bar, when using the E85SB. 3.2 WIDE OPEN THROT TLE Taking a look at the maximum achievable indicated mean effective pressure at an engine speed of 1500 rpm, ④ (right), the advantage of the lower stoichiometric air-to-fuel ratio in addition to the improved efficiency becomes obvious. An up to 16 % higher power output is achieved with high-octane, ethanol-containing fuels than with conventional fuel. The rated power output of Pe = 96 kW is reached at an engine speed of n = 5500 rpm using E20SB. This corresponds to a specific power output of 120 kW/l. The earlier mentioned investigations for finding optimum valve timing, start of injection, as well as injection pressure were performed at this operating point as well. ❺ shows an extract from the results of the valve timing variation at the rated output. We can see that the targeted mean effective pressure of BMEP = 26.2 bar is achieved over a wide adjusting range. Only at early intake valve timing is it not possible to build up sufficient boost pressure. The advantage of the high evaporation enthalpy of E20SB fuel and the associated early centre of combustion as well as the lower need for enrichment is not as pronounced as expected in the experiment. This implies that the mixture formation is inadequate. This assumption is confirmed by an above average oil dilution level in the range of the rated power output. The cause and a possible countermeasure are illustrated in ❻. It shows the droplet distributions at 510 °CA after TDC as the result of a 3-D CFD simulation of the rated power output point on the left for the current, slightly decentralised position of the injector and on the right in the centre at the location of the spark plug. ⑥ (left) shows a clear drift of the spray in the direction of the combustion chamber wall. High flow velocities via the intake valve of over 150 m/s already reach the fuel jet at a very early stage and thus prevent an optimum mixture formation while severely wetting the wall. A central position of the injector would remedy the problem, ⑤ (right). However, the risk of an extended flame propagation and the associated increased tendency of pre-ignition with a decentralised spark plug position have to be taken into consideration in this case. ❹ Comparison of different operation points 11I2014 Volume 75 3.3 NEW EUROPE AN DRIVING CYCLE To evaluate the steps taken to reduce CO2 emissions, a conventional 1.4 l turbocharged engine with a comparable power output of 94 kW can be used for comparison. Based on the measurements, vehicle longitudinal dynamics simulations were carried out and the fuel consumption in the New European Driving Cycle (NEDC) was determined using the same general conditions. A vehi- 49 RESE ARCH G A S OLINE ENGINES ❺ Variation of valve timing at n = 5500 rpm and full load cle from the B-segment with a mass of 1360 kg, a seven-speed dual clutch transmission, as well as a start/stop system was used as platform. The transmissions of both engines were designed in each case paying attention to comparable driving performances in particular with regard to flexibility. ❼ shows the fuel consumption corrected to a lower heating value of Hu = 42.5 MJ/kg as well as the CO2 emissions of the 1.4 l turbocharged engine in comparison to the illustrated 0.8 l downsizing engine while using different fuels. The volumetric fuel consumption shows the actual advantage in efficiency of the concept or alternatively the fuel. In the case of CO2 emissions, the chemical composition of the ethanol-containing fuel, to be exact the lower ratio between carbon (C) and hydrogen (H), additionally has a positive effect. Due to the downsizing, consumption and emission reductions of 11.1 % are already achieved using conventional E10MB. As shown, conventional fuel has high knock tendency in the area of higher loads, while ignition is already possible with optimal efficiency with as little as 20 vol.-% of ethanol (RON102.2) in the operating points that are relevant for the cycle. This is why no further consumption advantage can be noted with E85 (RON106.1) in terms of volumetric consumption. However, when we look at CO2 emissions, the aforementioned advantage of the lower C/H ratio becomes apparent and a reduction in CO2 emissions of 14.9 % can be achieved. During the interpretation of the results, we must always include the frictional level of the diesel engine that is relatively high in comparison to other gasoline engines, which is approximate 20 % above the level of current 1.0 l turbocharged engines. Despite the necessary high peak pressure resistance as well as the smaller displacement, there still is a considerable potential for reductions here. 4 SUMMARY AND OUTLOOK ❻ 3-D-CFD simulation at rated power for different injector positions 50 By systematically gearing the system towards high-octane ethanol fuels, it was possible to reach a specific power output of 120 kW/l ❼ Fuel consumption and CO2 emissions in NEDC vehicle simulation with a displacement of only 0.8 l. The compression ratio that is relatively high for this degree of charging leads to substantial efficiency advantages during partial load, which also have a particularly positive effect in the NEDC in addition to the shift in the load point resulting from the displacement reduction. It is therefore possible to forecast a reduction in CO2 emissions for the developed engine in the cycle of up to 15 % in comparison to a 1.4 l turbocharged engine with comparable output. The challenges of this type of extreme displacement reduction become apparent in the area of the rated power output. High compression pressures promote unwanted combustion phenomena such as pre-ignition and knocking. Due to their high evaporation enthalpy and knock resistance, high-octane, biogenic fuels offer a high potential for counteracting these problems. A compromise between a central spark plug position in favour of short flame propagation and a central injector position to prevent wetting the walls has been proven to be essential. In particular the long fuelrelated injection periods require a precise design of the spray pattern, the injector position, and the injection pressure. Thanks to the development of this extreme concept, a new research engine is available to the Institute for further investigations. The engine has already been extended by a cooled, external exhaust gas recirculation system for example. Moreover, the system is supposed to be analysed with camshafts designed for the Miller cycle in order to determine its potential, in particular with regard to a further reduction in the tendency to knock and therefore a reduction in CO2 emissions in the rated power output range. 11I2014 Volume 75 REFERENCES [1] Golloch, R.: Downsizing bei Verbrennungsmotoren. Berlin/Heidelberg: Springer-Verlag, 2005 [2] Weinowski, R. et al.: CO 2-Potenzial eines zweistufigen VCR-Systems in Kombination mit zukünftigen ottomotorischen Antriebskonzepten. 33. Internationales Wiener Motorensymposium, 2012 [3] N. N.: Bioenergy – Chances and Limits. Leopoldina: Nationale Akademie der Wissenschaften, 2012 [4] Picard, K. (Hrsg.): Zukunft der Kraftstoffe: Bewertung fossiler und erneuerbarer Kraftstoffe unter Berücksichtigung der Nachhaltigkeitskriterien. 14. Aachener Kolloquium Fahrzeug- und Motorentechnik, 2005 THANKS This report is the scientific result of the research project design of a downsized DI gasoline engine for use with biofuels to reduce CO2 emissions (Project number: 1041), which was provided by the Research Association for Internal Combustion Engines e. V. (FVV, Frankfurt, Germany). Support for the project was provided by a project management committee from FVV, headed by Dr.-Ing. Ulrich Kramer (Ford Werke GmbH, Cologne). The authors want to thank this work group for their great support. 51