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Improving Performance Parameters of Combustion Engine for Racing Purposes

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Improving Performance Parameters of Combustion Engine for Racing Purposes Res. Agr. Eng. Vol. 60, 2014, No. 3: 83–91 Improving performance parameters of combustion engine for racing purposes T. Polonec, I. Janoško Department of Transport and Handling, Technical Faculty, Slovak University of Agriculture in Nitra, Nitra, Slovak Republic Abstract Polonec T., Janoško I., 2014. Improving performance parameters of combustion engine for racing purposes. Res. Agr. Eng., 60: 83–91. Mechanical parts of stock engine have a performance reserve which could be utilized when the engine is used under the race conditions. Especially normal turbocharged engines have their performance parameters designed to drive in traffic, where a good flexibility, reliability, fuel consumption and a long service life is required. It is possible to utilize the whole power of the engine, when changing or modifying some of its external parts and achieve better performance parameters without modifying or changing internal engine components. Performed changes must be realized thought- fully and on the admittable level, so the engine and other drive train components would not be damaged. In our study we design several changes of external parts of engine which have a significant impact on the improvement of engine performance parameters. Their contribution has been verified in practice by an engine dynamometer. Keywords: engine; performance parameters; turbocharger; roller dynamometer It is generally known that the engine perfor- – decrease of intake air temperature (behind tur- mance is substantially dependent on the amount of bocharger), air (oxygen), which enters to the combustion cham- – reduction of mechanical and airflow losses, ber. Turbocharged engines are using turbochargers – optimization of intake and exhaust manifolds, or compressors to increase the amount of induced – optimization of combustion processes by sophis- air. The most commonly used charging system tur- ticated motor-management. bocharger powered by kinetic energy of exhaust The purpose is to design the best solutions to gases (Ferenc 2004; Sloboda et al. 2008; Čupera, improve the performance of normal supercharged Šmerda 2010; Hromádko et al. 2010;). engine, to improve acceleration of the vehicle as There are several ways to increase power of tur- much as possible, to verify these modifications by bocharged engine: measurements on roller dynamometer, to assess – increase of engine displacement, their contribution and propose other solutions to – increase of turbocharger’s boost pressure and airflow, achieve even better results. Supported by the Scientific Grant Agency VEGA of the Ministry of Education of the Slovak Republic and Slovak Academy of Sciences, Grant No. 1/0857/12 and by the Scientific Grant Agency KEGA of the Ministry of Education of the Slovak Republic and Slovak Academy of Sciences, Grant No. 044SPU-4/2014. 83 Vol. 60, 2014, No. 3: 83–91 Res. Agr. Eng. (a) (b) Fig. 1. Measured vehicle (a) Fiat 127A and engine (b) Lancia 2,0 16V Turbo MATERIAL AND METHODS compressor maps we have chosen the best turbo- charger Turbo Tech 103 (Honeywell International Measured vehicle. Performance measurement Inc., Morris Township, USA). was done on a special prototype vehicle designed for a drag race (Janoško, Polonec 2011). The base Airflow needed to achieve the performance of the vehicle was bodywork of Fiat 127A (Fiat Auto target: = × λ × S.p.A., Torino, Italy). As the power unit Lancia 2.0 Qv Pm SpSB (1) 16V Turbo engine was used, which was placed in the where: vehicle across, front of rear axle (Fig. 1). Q – airflow (kg/min) The vehicle was two-door hatchback, with frame- v P – performance target (kW) less steel body. Total weight of vehicle without driver m λ – air/fuel ratio (–) was 830 kg. Sp – brake specific fuel consumption (kg/kW·min) A powerful engine from Lancia Thema, made by SB Fiat Auto S.p.A., Italy was used. It was petrol engine Required absolute manifold pressure to achieve with charging by turbocharger. Displacement of en- performance target: 3 gine was 1,995 cm (bore: 84 mm, stroke: 90 mm). Q × R × (255.6 +T ) p = v P Engine had 4 cylinders in-line block with 16 valve ABS n (2) ηVOL × ×Vm DOHC head. Compression ratio is 8:1. Max. pow- 2 –1 er of stock engine was 147 kW at 5,500 min and where: –1 torque 298 Nm at 3,750 min . pABS – required absolute manifold pressure (kPa) Fuel delivery was provided by simultaneously Qv – airflow (kg/min) multi point port fuel injection, controlled by elec- R – gas constant tronic control unit Bosch LE2 – Jetronic (Robert Tp – intake manifold temperature (°C) Bosch GmbH, Gerlingen, Germany). Ignition was ηVOL – volumetric efficiency (–) fully electronic, “wasted-spark” type, controlled by n – engine speed (min–1) 3 electronic control unit Magneti Marelli MED 601E Vm – engine displacement (cm ) (Magneti Marelli S.p.A., Corbetta, Italy). Calculation of suitable turbocharger. In the Compressor discharge pressure: = + Δ calculations of suitable turbocharger we took ac- p2C pABS pSTR (3) count of future application of the vehicle in races where: and we defined a max. engine power to 300 kW at p2C – compressor discharge pressure (kPa) –1 6,000 min . The best turbocharger for the intended pABS – absolute manifold pressure (kPa) use of vehicle was calculated using the following re- ΔpSTR – pressure loss between the compressor and the lations (Estill 2008). Substituting the results into manifold (determined to 14 kPa) 84 Res. Agr. Eng. Vol. 60, 2014, No. 3: 83–91 4.5 Fig. 2. Compressor map of Garrett GT3076R turbocharger p2C/p1C – pressure ratio of inlet and outlet of turbocharger 4.0 3.5 3.0 1C p / 2C p 2.5 2.0 1.5 1.0 0 4.5 9.1 11.3 15.9 20.4 27.2 Corrected airflow (kg/min) Compressor inlet pressure: quirements and the expected use of the vehicle for = − Δ racing purpose. p1C pATM pSTR.S (4) where: Calculation of theoretical injectors fuel flow: p1C – compressor inlet pressure (kPa) = QV × ρ QP P (6) pATM – ambient air pressure (at sea level) (kPa) λ N ΔpSTR.S – pressure loss in air filter and piping (determine where: to 7 kPa) QP – flow of fuel (kg/min) Q – flow of air (kg/min) Pressure ratio: v λ – numerical value of lambda (–) ∏ = p2C N TD ρ – fuel density (kg/m3) p1C (5) P where: Because of lower heat stress of injectors we cal- ΠTD – pressure ratio culated with approximately 80% duty cycle. Con- p1C – compressor inlet pressure (kPa) sidering this duty cycle RC Racing injectors with p2C – compressor discharge pressure (kPa) fuel flow 750 cm3/min were chosen (at 300 kPa fuel Based on calculations of operating parameters pressure). and substituting them into various compressor Calculation of heat ratios in the intercooler maps we chose a Garrett GT3076R turbocharger (Estill 2008). (ΔT − ΔT )÷ F (Honeywell International Inc., Morris Township, W = U × S × 1 2 ⎛ ΔT ⎞ (7) USA). As seen on the compressor map (Fig. 2), this ln 1 ⎝⎜ Δ ⎠⎟ turbocharger is the most suited to performance re- T2 85 Vol. 60, 2014, No. 3: 83–91 Res. Agr. Eng. (a) F (b) T1 T in Tex T P 2 Fig. 3. Heat diagram of correction factor F (a) and intercooler temperature scheme of inlet/outlet flow (b) F – correction factor (–); P – temperature ratio (–); T1 – outside (cooling) air temperature on inlet (°C); T2 – outside (cooling) air temperature on outlet (°C); Tin – compressed air temperature on inlet (°C); Tex – compressed air temperature on outlet (°C) where: − = Tex Tin W – total transfer of heat energy (J) P (8) T1 −Tin U – heat transfer coefficient (W/m2·K) 2 T −T S – heat transfer surface (m ) R = 1 2 (9) Tex −Tin ΔT1 – difference between intercooler input air tempera- ture and temperature of cooling air behind inter- where: P, R – temperature ratios cooler (Tin – T2) (°C) T – compressed air temperature on outlet (°C) ΔT2 – difference between output air temperature from ex intercooler and temperature of cooling air in Tin – compressed air temperature on inlet (°C) T – outside (cooling) air temperature on inlet (°C) front of intercooler (Tex – T1) (°C) 1 F – correction factor T2 – outside (cooling) air temperature on outlet (°C) Determination of the correction factor F. Cor- Calculating the amount of lost or received heat rection factor F, taking into account the unequal on one side of exchanger: distribution of heat at exchanger area, could be W = Qm × CP × ΔT (10) read from the diagram according to the calculated values of temperature ratios of P and R (Fig. 4). For where: calculation of temperature ratios P and R we need W – heat energy transfer (J) to know temperature of compressed air (Tin, Tex) Qm – mass airflow (kg/min) and cooling air (T1, T2) on the inlets and outlets of Cp – heat capacity of air (J/K·mol) the intercooler (Fig. 3). ΔT – difference of input and output temperatures (K) T2 Tex Fig. 4. Airflow through the intercooler T1 – outside (cooling) air temperature on inlet; T in T2 – outside (cooling) air temperature on out- let; Tin – compressed air temperature on inlet; T1 Tex – compressed air temperature on outlet 86 Res. Agr. Eng. Vol. 60, 2014, No. 3: 83–91 RESULTS AND DISCUSSION – intake manifold replaced by shorter type from Lancia Kappa, Performed engine modifications – throttle body replaced by bigger one with inter- nal diameter 73 mm, The engine power can be measured by the dy- – stock exhaust manifold replaced by custom steel namometer directly or through the power take-off manifold with pipes with diameter 42 mm, shaft, or possibly on a roller bench or by the road- – boost pressure controlled by electronic control board test (Semetko, Janoško 2005). After per- unit with solenoid valve. formance measurement of stock engine the follow- ing modifications were made: Dynamometer Stage 1: – turbocharger replaced by more powerful type The measurements were performed on the roll- Garrett GT3076R, with rotor on ball bearings, er dynamometer MAHA LPS 3000 PKW 4 × 4 – injectors replaced by more powerful (RC Racing; (MAHA Maschinenbau Haldenwang GmbH & Co.
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