You will find the figures mentioned in this article in the German issue of MTZ 12/2005 beginning on page 978.
Neuer Ottomotor mit Direkteinspritzung und Doppelaufladung von Volkswagen Teil 2: Thermodynamik
The New Dual-Charged FSI Petrol Engine by Volkswagen Part 2: Thermodynamics The new 125 kW engine from Volkswagen in the Golf GT represents a milestone in the development of direct-injection spark-ignition engines. Its performance far surpasses that of conventional engines with the same capacity while at the same time offering substantially improved fuel economy. This result has been achieved by a combination of downsizing, direct fuel injection and double supercharging. While part 1 described the engine design, part 2 deals with the thermodynamics.
1 Introduction
Authors: Rudolf Krebs, Rüdiger Szengel, Hermann Middendorf, Helmut Sperling, Werner Siebert, Jörg Theobald and Karsten Michels
By launching direct-injection onto the market in the form of its FSI engines, Volkswagen has taken an important step forwards towards reducing consumption in spark-ignition engines. In consistently further developing this technology, additional, significant consumption can be achieved via downsizing. At a pre-specified, rated output, the displacement is lowered, resulting in a shift in operating points from low engine load ranges to performance map areas with higher load. On one hand, this operating point shift leads to a reduction in gas cycle losses due to increased dethrottling; on the other hand, friction power losses are lowered as displacement and the size of the en-
gine are reduced. Both effects lead to a significant improvement in overall efficiency. However, reducing displacement also leads to less torque, particularly at low engine speeds, something which is unacceptable in operation in customers' hands. This effect can be compensated with the aid of charging. Besides turbocharging, mechanical charging has also been implemented in series production in spark-ignition engines. In addition to minimising consumption, customers' increased vehicle dynamics and ride comfort requirements are vital considerations in designing the supercharger system. To attain its objectives in the optimal manner, the new TSI engine concept therefore combines mechanical charging with turbocharging in such a way that the specifMTZ 12/2005 Volume 66
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ic advantages of both systems ideally supplement each other.
bocharged engine without compressor support would achieve a significantly lower charge pressure ratio in this case. Thanks to compressor charging, air throughput is significantly increased in the lower engine speed range, which also benefits the turbine and therefore the turbocharger compressor due to the greater availability of exhaust gas energy. The compressor can therefore be relieved by opening the charger bypass at an early stage. As a result of this, the mechanical charger's operating range is limited to a small performance map area with primarily low power consumption. The turbocharger is designed in such a way that its full-load operating curve in the lower engine speed range runs close to the compressor's surge limit, Figure 4. The compressor already reveals good efficiency here, contributing towards speedy acceleration behaviour on the part of the turbocharger. Taken together, the overall charge pressure ratio of the turbocharger and mechanical charger lies far beyond the compressor surge limit in the engine speed range up to 2000 rpm. Single-stage turbocharging would therefore be unsuitable for the desired full-load target due to the compressor's limited operating range. Based on compressor support, the turbocharger can be designed for good efficiency at rated output. This results in a low charge pressure and exhaust back pressure level and generous turbocharger altitude reserve. Figure 5 shows the transient torque buildup measured for an engine speed ramp simulated on the engine test rig, corresponding roughly to full-load acceleration in 3rd gear. In a pure turbocharger engine without compressor support, the naturally-aspirated engine full-load point is reached after approx. 0.5 s. The target torque value (100 %) is only achieved after approx. 4.8 s. Besides slow charge pressure build-up, the unsteady torque build-up, which is attributable to the turbocharger's dynamics, is perceived as unpleasant by the driver. These characteristics change fundamentally on operation with the compressor. As soon as the compressor is engaged, it contributes towards an increase in intake manifold pressure. The torque increase gradient is therefore steeper than in the case of pure turbocharger charging. This gradient is also maintained above full intake load. Torque build-up remains constant until the target torque is achieved. In subjective terms, an engine which is supercharged in this manner also feels like a naturally-aspirated engine with significantly higher displacement.
2 Procedure 2.1 Layout of the Supercharger Units Turbocharging is highly suitable for achieving high, specific engine output with moderate exhaust back pressure if low-end torque is consciously dispensed with. In turn, mechanical charging is used at low engine speeds, primarily because of its good response behaviour. The advantages of both concepts are combined in the case of turbocharging with additional, mechanical charging. The basic principle is already familiar [1, 2]. Figure 1 shows the principle layout of the system. A magnetic coupling for activating the compressor is integrated into the water pump module. The mechanical charger is driven by the crankshaft via a belt. In the lower engine speed range, the compressor is activated as required to support the turbocharger. If the turbocharger alone supplies sufficient charge pressure at higher engine speeds, the compressor control valve can be opened and the compressor can be switched off. During stationary operation, the compressor is only required in the upper load range up to engine speeds of max. 2400 rpm, Figure 2. As the turbocharger reaches its nominal charge pressure following a time lag in dynamic operation, depending on which gear is selected, the compressor remains dynamically switched on for longer. At the latest, the compressor is switched off when the engine reaches a speed of 3500 rpm. In this case, the turbocharger is always – i.e. also dynamically on transition from overrun to fullload operation – able to supply the desired charge pressure by itself. At this engine speed, the compressor reaches its maximum speed of 18,000 rpm with a ratio of 5:1. Thanks to the combination of charging procedures, the exhaust turbine can be dimensioned more generously, thereby reducing the exhaust back pressure level. In this case, however, consideration must be given to minimising the mechanical charger's cutin frequency and operating time, so that the power required to drive the charger does not impermissibly increase consumption. Finally, a continuous torque curve must be ensured during both stationary and dynamic operation. The maximum charge pressure ratio during stationary full-load is approx. 2.5 at 1500 rpm, whereby the turbocharger and mechanical charger are operated at roughly the same pressure ratio, Figure 3. A tur24
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2.2 Layout of the Charge Air Cooling System Fitting an effective charge air cooler, which makes use of the entire frontal area in a package together with the radiator and air conditioner condenser, enables the charge air to be cooled to 5 °C above ambient temperature over the route from the turbocharger to the throttle valve. An intake temperature of 25 °C was therefore assumed for the basic design, enabling engine operation with ignition timing which is optimal as regards consumption over wide performance map ranges. Very good performance and outstanding consumption can be achieved by designing the torque curve for a maximum value of 240 Nm as of 1750 rpm and the gear ratio which has been selected, Table. The NEDC cycle values over elasticity (80-120 km/h) in 5th and 6th gear are shown for different engines fitted in the VW Golf in Figure 6. This clearly shows that the 1.4-litre TSI engine achieves the lowest consumption, at 7.2 l/100 km, together with the best acceleration values in comparison with higher-displacement engines. The consumption attained by the 125 kW TSI in comparison with competitors also represents a new milestone in spark-ignition engine development, Figure 7.
3 Combustion Process With stroke of 75.6 mm and a bore diameter of 76.5 mm, the 1.4-litre TSI engine has a very compact combustion chamber, which is designed as a roof-shaped combustion chamber with central spark plug system on the cylinder head side. The pistons have a flat, wide recess. This basic geometry, which has been optimised in terms of knock resistance, enables a compression ratio of 10.0, despite the high charge pressure of 2.5 bar, giving rise to outstanding, specific functional values: – effective mean pressure 21.6 bar – specific torque 172.6 Nm/l – specific output 90 kW/l . By using the tumble flaps already fitted in the naturally-aspirated FSI engine [3], which close the intake ports by 50 % in the lower range when activated, the intensity of in-cylinder flow is adapted to achieve the most optimal combustion speed possible throughout the entire engine performance map. As of an engine speed of approx. 2800 rpm, the tumble flaps release the entire intake port cross-section. The intake ports can therefore be consistently designed for optimised flow, enabling high cylinder filling to attain the output target of 125 kW.
In order to heat the catalytic converter, the 1.4-litre TSI engine employs the strategy of twin injection with early fuel injection during the intake cycle and subsequent, secondary fuel injection prior to ignition TDC (approx. 50°). Precise co-ordination of the parameters of in-cylinder flow, piston recess geometry and high-pressure injector spray pattern is essential to achieve the extremely retarded ignition angle with good engine smoothness required to enable the most extensive possible flow of exhaust gas heat, Figure 8. The contradictory piston recess design requirements needed to achieve this during homogeneous engine operation – small piston surface to reduce thermal entrainment at full-load and avoid HC nests – lead to a multitude of possible configurations. Based on CFD simulation results, the most promising variants were selected and compared against each other on the engine test rig. The scope of engine testing during combustion process development was therefore significantly reduced. A multi-hole, high-pressure injector with 6 fuel outlet bores, Figure 9, is fitted for the first time in the 1.4-litre TSI engine. As in the naturally-aspirated FSI engines, the injector is located on the intake side between the intake port and the cylinder head gasket level. In contrast to a conventional swirl injector, the layout of the 6-hole, high-pressure injector's individual jets, which can be chosen virtually at will, enables formation of the fuel injection spray. In particular, coating the open injectors in the case of early injection during the intake cycle is avoided. Better air-fuel mixture homogenisation and therefore lower HC emissions and lower cyclical fluctuations are the positive result. The images from VW's internal pressure chamber, Figure 10, reveal the emerging, individual sprays in both the side and frontal view. Thanks to the special fuel outlet bore geometry, success has been achieved in reducing individual spray penetration to such an extent that coating the combustion chamber surfaces with fuel can be reliably avoided. This helps to reduce raw exhaust emissions, particularly when the engine is cold. In the laser section, Figure 9, additionally shows the distribution of liquid and vaporised fuel in the fuel spray (30 mm away from the tip of the injector) under catalytic converter heating conditions, using laser-induced fluorescence (LIF) measuring technology. The higher concentration of fuel in the 6 individual sprays is clearly visible. Homogenisation of the fuel-air mixture is sig-
nificantly more advanced in the centre of the spray. The combination of good partialhomogenisation and co-ordinated penetration of the individual sprays is the prerequisite of stable engine operation during the above described catalytic converter heating process with piston wall-controlled, secondary fuel injection and low raw exhaust emissions. The spread between the minimum and maximum quantity of fuel to be injected from idle speed up to a volumetric efficiency of 90 kW/l is very wide. To prevent negative deviation from the injector's minimal injection period at idle speed, and to guarantee adequate atomisation of the injection sprays and therefore good mixture formation, injection takes place at a pressure of 60 bar. To adhere to low emission and consumption values under full-load conditions, injection must not commence too early, in order to minimise the quantity of fuel sprayed onto the pistons; on the other hand, sufficient time must remain for mixture preparation between the end of injection and the start of combustion. Particularly at high engine speeds, this temporally limited injection period is achieved by raising the injection pressure to 150 bar. The above described measures lead to a consumption map, Figure 11, which reveals wide areas with very low, specific consumption values. At 235 g/kWh, the optimum consumption point is outstanding. The onroad partial-load curves shown in the Figure reveal that the engine functions at operating points with extremely low, specific consumption even at high speeds.
4 Engine Management System The TSI engine is managed by a further-developed version of the control unit fitted as standard in Volkswagen's naturally-aspirated FSI engines [4]. The engine control unit is a sensor-guided system, in which the engine charge is registered with the aid of a pressure sensor. Initial use in a dual-charged spark-ignition engine with adjustable inlet camshaft and wide air requirement spread necessitates precise modelling and co-ordination of the charging model.
4.1 Load Requirement Control In the MED 9.5.10 engine management system's torque structure, the driver's torque command is converted into nominal intake manifold pressure via the relative air mass. Whilst this nominal value is regulated solely by the throttle valve and camshaft adjuster in intake mode, interaction between
the magnetic coupling, compressor, compressor control valve and turbocharger wastegate is additionally required in supercharging mode. As turbocharging is more favourable than mechanical charging in energy terms, a check is generally carried out to determine whether the turbocharger can generate the required charge pressure on its own when charge pressure is demanded. If the required air mass cannot be provided by the turbocharger, the compressor is engaged on a model basis. In the engine control unit's charge pressure modules, the cold and hot side of the turbocharger are calculated with the aid of mass flows and are corrected as regards efficiency; leaked flows, e.g. via the wastegate, are taken into consideration. The compressor's percentage of supercharging is determined with the aid of compressor efficiency, taking the position of the compressor control valve into account. Precise knowledge of the two units' supercharging levels is used to define the parameters for the participating actuators and the basic variables for engine operation.
4.1.1 Actuation of the Compressor's Magnetic Coupling The compressor is engaged and disengaged via a magnetic coupling, which is integrated into a module together with the water pump. To meet customers' quality standards in this regard, great importance was attached to ensuring that engagement and disengagement take place with neutral torque and without jolting. Besides compressor torque loss, the coupling's air gap, which changes as a result of wear, also has to be taken into consideration in this case. These stringent requirements necessitate precise temporal co-ordination of solenoid actuation within the coupling, which is achieved by means of pulse-width modulation. Differences in the magnetic coupling's air gap are detected via additional current measurement, which determines the precise time of positive connection. Adaptation routines are implemented to give consideration to these data as correction variables for subsequent activation and deactivation processes, these finally being overlaid with classic handling intervention measures such as load impact damping and anti-jerk functions. As it has to be ensured that the compressor is not operated in excess of its permissible speed of 18,000 rpm, cut-off and diagnostic strategies are integrated into the corresponding software modules for magnetic coupling actuation. MTZ 12/2005 Volume 66
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4.1.2 Compressor Output Control via the Compressor Control Valve
alytic converter coating developed specifically for this purpose, the TSI can be continually operated at an exhaust temperature of 1050°C upstream of the turbocharger. With the chosen ratio, the Golf can therefore be operated at a speed of 200 km/h with Lambda = 1 in 6th gear. Thanks to the robust combustion process, the engine can also be operated stoichiometrically at high loads, with the result that the TSI engine can be fitted with a simple two-point transient probe as the starter catalytic converter probe. As in the case of Volkswagen's naturally-aspirated FSI engines, the 1.4-litre TSI engine is also started at high pressure [4, 5]. Initial fuel injection only takes place once a pressure of approx. 25 bar is present in the fuel system. Raw emissions are reduced even when the engine is cold thanks to lower fuel requirements as a result of improved mixture processing. The dual injection possible in the case of direct-injection leads to rapid catalytic converter heating. Air gap insulation of the pipe connecting the turbocharger and catalytic converter partly compensates the temperature losses caused by the turbocharger. Thanks to the totality of these emission measures, reliable and also inexpensive achievement of the Euro 4 emission limits has been accomplished.
The compressor's output and charge pressure downstream of the compressor are controlled by the compressor control valve. During compressor operation, the intake air is diverted to the compressor downstream of the air filter by closing the control valve. Spontaneous pressure build-up, even in the lower engine speed range, was one of the TSI engine's primary development objectives, Figure 12. The increase in pressure is influenced directly by the closing speed of the control valve, which can be closed in 0.2 s when maximum compressor output is required. The control valve closing process and compressor engagement have to be timed to co-ordinate precisely, in order to prevent a drop in charge pressure and therefore loss of torque which is perceptible to the driver at all times.
4.1.3 Turbocharger Control The turbocharger is controlled by altering the mass exhaust gas flow which is fed to it, in which case the turbine output is accordingly adapted to the operating point. The necessary turbine output is directly dependent on the compressor's output, which is determined in a calculation model within the control unit via the necessary compressor pressure ratio and efficiency map. If the turbocharger is to be operated at maximum output, the wastegate is closed via the spring in the charge air control pressure unit, and the entire mass exhaust gas flow supplied by the engine flows through the turbine. If reduced turbine output is required, a calculation model of the hot side of the turbine is used, with the aid of efficiency, to determine the mass flow of exhaust gas via the turbine, the nominal position of the wastegate valve and the actuation signal for the wastegate frequency valve. The frequency valve is used to set the corresponding pressure in the wastegate unit, which leads to the desired wastegate position. Evaluation of the pressure upstream and downstream of the turbocharger is necessary to actuate the electric deceleration air valve. In contrast to single-charged engines, in which virtually ambient pressure prevails on the compressor's intake side, the turbocharger's surge limit during compressor operation is reached at a later stage in the TSI engine. These new requirements have been taken into consideration by modifying the existing series software functions.
4.2 Exhaust Gas Measures and Emission Classification Thanks to its cast steel manifold and a cat26
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5 Summary With its new 1.4-litre 125 kW TSI engine, Volkswagen is consistently continuing to develop direct-injection in spark-ignition engines by combining it with dual charging for the first time ever in mass production. This engine's outstanding characteristics are achieved by a combination of downsizing, direct-injection and charging. The engine will be fitted first in the Golf GT. Particular focus was placed on the criteria of favourable consumption, driving enjoyment and ride comfort. At 7.2 l/100 km in the NEDC cycle, the Golf GT attains a value which is unique in the 125 kW output class. The engine's impressive torque characteristics, with maximum torque of 240 Nm in the engine speed range from 1750 to 4500 rpm, ensure outstanding vehicle dynamics. This plentiful torque curve is accomplished via the intelligent combination of a compressor with a turbocharger. The interaction of these components has been reliably solved by internally developing VW functional control and adaptation strategies. High dynamics, outstanding economy and simultaneous compliance with the most stringent of emission limits give this new Volkswagen engine the very best prospects for a rosy future.
References [1] Hiereth, H.; Prenninger, P., 2003: Aufladung der Verbrennungskraftmaschine, Springer Verlag, S. 125 ff [2] Lang, O.; Habermann, K.; Wolf, K.; und Pischinger, S.: Anwendung der Zusatzaufladung bei abgasturboaufgeladenen Ottomotoren, 9. Aufladetechnische Konferenz, Dresden, 2004 [3] Krebs, R.; Spiegel, L.; Stiebels, B.: Ottomotoren mit Direkteinspritzung von Volkswagen, 8. Aachener Kolloquium, 1999 [4] Szengel, R.; Middendorf, H.; Wiedmann, M.; Wietholt, B.; Laumann, A.; Voeltz, S.; Stiebels, B.; Damminger, L.: Die Ottomotoren des neuen Volkswagen Golf. In: Der neue VW Golf. Sonderausgabe ATZ/MTZ, Oktober 2003 [5] Szengel, R.; Kirsch, U.; Ebel, B.; Lieske, S.; Reschke, F.: Die neue V6-Motorengeneration mit Direkteinspritztechnik von Volkswagen. 26. Internationales Wiener Motorensymposium, Wien
