You will find the figures mentioned in this article in the German issue of MTZ 07-08I2007 beginning on page 538.
Der TSI-Motor mit 90 kW – Erweiterung der verbrauchsgünstigen Ottomotoren-Baureihe von Volkswagen
The TSI Engine with 90 kW Extension of the Economical Petrol Engine Series by Volkswagen
Authors: Rüdiger Szengel, Hermann Middendorf, Stefan Voeltz, Alfons Laumann, Lutz Tilchner, Jörg Theobald, Thomas Etzrodt and Rudolf Krebs
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The 90 kW TSI is based on the 125 kW and 103 kW 1.4 l TSI engines with twin charging, which were successfully introduced by Volkswagen and have received numerous awards including the “Paul Pietsch Prize”, “Best New Engine, 2006” in the 1.0 l to 1.4 l class and “Best of What‘s New from Popular Science, 2005“ . Thus the 1.4 l 90 kW engine with single-stage charging stands for the rigorous further development of turbocharged downsizing designs.
1 Introduction With this, Volkswagen continues among the volume models as well to pursue the Group goal of reducing fuel consumption while improving the driving comfort and sets a further milestone in the implementation of its CO2 strategy. The new 1.4 l 90 kW TSI engine, Table, has been developed for world-wide use in all vehicles with substantially increased power output compared to its predecessor, the 1.6 l 85 kW engine. It possesses substantially better driving dynamics and further reduces fuel consumption.
2 Design and Developmental Goal Employing modular construction, the design of the new turbocharged engine is extensively based on the proven 1.4 l 125 kW TSI with twin charging and the engine replaces the previously used, naturally aspirated 1.6 l FSI with 85 kW. The main focus during the development was on the further reduction of friction, the economical realisation as a volume model for all Group vehicles and economical running costs for the customer while simultaneously increasing performance and improving robustness. The further development of the combustion process as well as the new design of the inlet ports enables the elimination of the tumble flaps. At the same time, fuel consumption, exhaust emissions, performance and smoothness of running are comparable to engines with tumble flaps. In spite of the high compression of F = 10.0 for a charged engine, it was possible to achieve compatibility with the fuel quality 95 RON. Because the target performance of 90 kW with a specific torque of 144 Nm/l for the new 1.4 l engine is moderate, the booster group and the timing could be designed without compromise for good response at low engine speeds. Even with single-stage boosting, an outstanding response behaviour could be achieved. Figure 1 shows a comparison of torque development of the new 1.4 l TSI engine and the 1.6 l naturally aspirated FSI engine. An increase in torque of up to 66 % was realised in the lower engine speed range. The maximum torque of 200 Nm is available at just 1500 rpm. The design with extreme low-end torque offers the ideal conditions for the combination of this new engine with long gear ratios. The nominal output of 90 kW is already attained at 5,000 rpm
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and is available to just below the speed limit of 6400 rpm.
sation of the entire engine. Because the exhaust temperature is limited to 950°, sodium-filled exhaust valves are not necessary. The weight of the cylinder head cover could be reduced by about 150 g compared to the 125 kW TSI valve cover by adapting the structure to the reduced diameter of the camshaft bearings. The 10 mm diameter, single-piston highpressure pump is operated by a four-lobed cam with 3 mm lift. This enables extremely fast pressure development during a cold start. Thus, during a cold start, just 0.5 seconds after the beginning of the starter phase, the pressure in the fuel rail exceeds 60 bar. With this design, it is possible for the first time to employ retarded high-pressure start throughout the entire cold-start temperature range.
3 Design and Form 3.1 Engine Components The following components from the 1.4 l 125 kW TSI were taken over for the design. – cast iron crankcase with open deck design – timing chain housing with integrated oil filler, crankcase breathing and oil filter module – steel crankshaft – connecting rods. The piston is designed as a light-weight cast construction and the pouring takes place via the middle feeder. The mechanical reworking of the recess in the piston crown which is consequently required has the additional effect of increasing the hardness because the oxides which are washed out during the casting on the combustionchamber side and which could act as local stress raisers, are removed. The valve recesses are cast and the geometry of the asymmetrical shaft and the wall thicknesses are optimised appropriately for weight and load. The weight of the cylinder head, the design of which is based on the 1.4 l 125 kW TSI, could be reduced by 600 g through a structurally optimised design. Based on the 125 kW TSI, the inlet port of the 1.4 l 90 kW TSI was further developed with the goal of attaining a degree of tumble which would allow the omission of the tumble flap. To achieve this, the new inlet port approaches the inlet valve seat ring at a flatter angle and more tangentially, and for the first time in the EA111 engine family, port masking to interrupt flow was introduced on the lower side of the inlet port directly in the line of flow ahead of the seat ring. As can be seen in Figure 2, the flow breaks off along edge of the casting on the lower side of the port. As a result, the greater portion of the flow is directed over the upper side of the inlet valve head so that a stable tumble effect develops in spite of the absence of the tumble flap. The valves are operated by two optimised, assembled camshafts with inlet camshaft adjusters. The weight of each camshaft was reduced by 304 g through the reduction of the width of the cams, reduction of the diameter of the camshaft bearings by 3 mm and the structural optimisation of the sintered bearing races, whereby the reduction of the bearing diameter contributes to the friction optimi-
3.2 Fuel Mixture Components 3.2.1 Turbocharger Dynamics and fuel economy were the main focus during the design of the turbocharger, Figure 3. The very good dynamics could be attained through the low inertia of the rotor group: the outer diameter of the turbine is 37 mm and that of the compressor wheel, 40 mm. The stated goal of 90 kW shows that only a low degree of boost is required at full load. Therefore the waste-gate channel is generously proportioned at 26 mm. With this turbocharger, 80 % of the maximum torque is already attained at 1250 rpm, and the maximum torque of 200 Nm is available from 1500 rpm. The turbocharger is designed as an integrated module and is made of cast Ni-Resist D5S. The optimal design of hardness and thermodynamics of the combined manifold and turbine housing was achieved using FEM and CFD calculations. The turbine, cast from INCO 713, is equipped with three-dimensionally formed vanes with the goal of an optimal degree of efficiency, whereby the flow is prevented from separating from the vanes. The shaft is borne on floating sleeve bearings so that the differential speed is reduced by about one half at the bearing points. Thus the maximum speed of the bearing group could be set at 220,000 rpm. The relief valve is integrated directly on the compressor housing with a cast flange. The use of an electric valve ensures fast and flexible actuation as well as acoustically unnoticeable operation. The large diaphragm diameter of the vacuum unit requires only a slight actuaMTZ 07-08I2007 Volume 68
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tion pressure to open the waste-gate flap. This makes it possible to use the basic boost pressure even in the lower part-load range, which helps lower fuel consumption.
The time required to attain 1700 mbar pressure in the intake manifold could be reduced by 100 milliseconds following a jump from overrun to load at 2000 rpm and, at 1500 rpm even by 250 milliseconds compared to a conventional charge-air cooling system with an air-air intercooler located in the front, Figure 5. The absolute times for attaining maximum charge pressure document that this turbocharged engine exhibits the best values regarding transient behaviour. A differential temperature at the same level as an air-air system over a broad range was attained for typical customer use, while simultaneously an advantage in packaging was gained. Under extreme driving conditions, for example in hot climates (maximum speed at ambient temperature of 40 °C), the difference in temperatures between air after the charge air temperature and ambient temperatures is 25 K. The intake manifold is a shell construction composed of PA6-GF30. The development of a uniform flow of the charge air cooler in the intake manifold to maximise the cooling effect and minimise the temperature differential of the charge air in the four runners proved to be especially exacting. To this end, ribs below the charge air cooler to guide the air were designed with the aid of CFX interior flow calculations, Figure 6. The ends of the charge air pipe are pushed onto the compressor outlet and the throttle valve and clipped onto the throttle valve end. A retaining bar assures secure seating on the turbocharger, whereby the chain of tolerances between the turbocharger and the throttle valve can be managed to facilitate assembly while maintaining the quality requirements.
3.2.2 High-pressure Pump and Lower Part of Intake Manifold A new generation of pumps, Figure 4, is employed to develop high pressure for the fuel. The special features of the new model include full delivery in the non-energised state (the opposite actuation concept compared to previous high-pressure pumps in the EA111 engine family) and the integrated pressure-limitation valve. The greater piston diameter enables dependable highpressure starting of the engine even under extremely marginal conditions due to faster pressure development. Maintaining constant pressure in the high-pressure system and the refining the acoustics presented special challenges during the development. To this end, the cross-section and the design of the spring and solenoid in the metering valve, the diameter of the high-pressure line as well as the non-return valve on the high-pressure side were optimised. The further development of the combustion process made it possible to eliminate the tumble flap without disadvantage to fuel consumption, exhaust emissions or performance. However, the lower part of the intake manifold still has the function of high-pressure fuel distribution. The pressure-limitation valve is integrated in the new high-pressure pump, so the return could be eliminated from the high-pressure rail.
3.2.3 Intake Manifold, Charge-air Cooling and Charge Air Route The 1.4 l 90 kW TSI marks the first time that water cooling for charge air is being used in the EA111 family, with a water cooler directly in the intake manifold. It is integrated in a low-temperature circuit (LT circuit) independent of the engine cooling system. This design exhibits the significant advantage of a low volume for the charge air system in comparison to a design with a charge-air cooler situated in the front. As a result, there is a dynamic gain in power as a consequence of minimal delay in attaining maximum combustion chamber filling, which clearly can be felt by the driver. The volume of the system between the compressor of the turbocharger and the intake valve could be reduced by more than half, from about 11 to 4.8 l. 10
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3.2.4 Low-temperature Coolant Circuit The heat dissipation of the water-air cooling system is managed through a separate lowtemperature coolant circuit. The water cooling circuit for the bearing housing of the turbocharger runs parallel. The distribution of the coolant flow is managed in accordance with the requirements through restrictions in the plastic connecting pieces. An electric coolant pump ensures the circulation in the low-temperature coolant circuit. The low-temperature circuit, Figure 7, is joined to the main circuit in the expansion tank to equalise pressure in the cooling system and for filling. A restriction limits the interchange of water between the two cooling systems to a minimum. This is necessary because the temperature differential between the two systems may exceed 100 K.
4 The Combustion Process 4.1 High-pressure Injectors A new type of multi-hole, high-pressure injector with 6 fuel jets, Figure 8, is being used; the spray cone does not have the typical round or oval shape, but rather, the possibilities for spatial distribution of the spray are optimally exploited. This reduces the moistening of the piston during the very early start of injection at full load as well during as late points of injection shortly before compression stroke TDC in the catalytic converter heating mode. Consequently, the injection can take place earlier, lengthening the time for forming the fuel mixture and improving the homogenisation. The results are very low HC emissions and a reduction in fuel entering the engine oil during cold operation. An undesirable interaction between the individual streams could be avoided by optimising the penetration of each spray cone individually in regard to the ratio of length to diameter of the exit hole. The figures on the right in Figure 9 show in laser cross section (30 mm below the tip of the injector and perpendicular to axis of injector) the distribution of liquid and vaporised fuel in the fuel spray under catalytic converter heating conditions, employing the laser-induced f luorescence (LIF) technique. It can clearly be seen that, in comparison to the production TSI injector, the spray pattern is wider and asymmetric, with one stream in the centre. The combination of good partial homogenisation and finely tuned penetration of the individual streams of spray form the basis for stable engine running during the catalytic converter heating mode. In the process, during the split homogenous operating mode with the second fuel injection guided by the recess in the piston deck, the catalytic converter can be heated extremely quickly with low raw exhaust emissions.
4.2 Intake Port Development The turbulent kinetic energy is an effective measure of the energy for creating the fuel mixture and is the determining factor in the engine‘s combustion behaviour. Figure 9 illustrates the turbulent kinetic energy (TKE at 2,000 rpm and full load) variations during the compression stroke of the production intake port with open or closed tumble flap in comparison to the two 90 kW port versions favoured in 3D simulation. When the tumble flap is closed, the tumble macro flow collapses shortly before the spark TDC into micro turbulence, and
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very high TKE values are attained. To a lesser degree, this can be identified for the two tumble intake ports as well. In contrast, the production TSI port with the tumble flap open exhibits no increase in TKE at the end of the compression stroke. At the moment of TDC on the compression stroke, about 55 % of the turbulent kinetic energy of the production port with closed tumble flap could be attained with port version 1, which was chosen for production. As an example of the combustion behaviour of the new tumble port, Figure 10 illustrates the elapsed time measured in degrees of crankshaft angle from ignition to the point of 50 % fuel burn as a function of engine speed at full load. Up to the point at which the tumble flap opens, fuel burn occurs about 4° crankshaft angle faster than the production port; on the other hand, the new tumble intake port is dramatically faster than the production TSI port with tumble flap open and, with a burn time of less than 27° crankshaft angle over the entire operating range, demonstrates a short ignition phase as well as high combustion speeds. Especially at high engine speeds, this is essential in order to hold fuel mixture enrichment to a moderate level at a maximum exhaust temperature of 950 °C. At the same time, the pressure increase gradient always remains below the acoustically critical limit of 5 bar/°crankshaft angle.
5 Fuel Consumption The rigorous continued development for the reduction of friction in the engine, the optimisation of the combustion process, the high compression ratio of 10:1 which could be attained as a result, the use of double-injection operating modes and the optimisation of functions in the engine control unit are the cornerstones for this economical engine. Because of the outstanding static and dynamic torque figures, the combination with a manual or dual clutch gearbox with long gear ratios for the reduction of engine speeds is also possible. The NEDC fuel consumption could be further reduced by about 6% in comparison to the 1.6 l 85 kW FSI with a manual gearbox in spite of significant improvement in performance, Figure 11.
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Numerous further developments in the engineering and a new design for charge-air cooling enable a substantial reduction in design costs while further increasing potential fuel economy. The great potential torque of the TSI engine permits its combination with a transmission having relatively long gear ratios, resulting in substantial reductions in fuel consumption while preserving superior driving performance. The result is a convincing drive design providing pleasurable driving without regrets. The combination of the new 90 kW TSI engine with the newly developed 7-speed dual clutch gearbox creates a drive design which is in every respect superior in comfort, fuel consumption and driving performance and which serves as a basis for further development.
References [1] Hagelstein, D. ; Pott, E.; Theobald, J.: Nachhaltigkeit und Mobilität, Globale Herausforderungen an die Entwicklung neuer Antriebstechniken. In: Sonderausgabe MTZ, Dezember 2006 [2] Szengel, R. ; Middendorf, H. ; Soehlke, G. ; Kuphal, A.: Der EA111 4V – Basis für die CO2-Strategie von Volkswagen. In: SAE Kongress Peking, Beijing, 2006 [3] Krebs, R.; Szengel, R. ; Middendorf, H.; Fleiß, M.; Laumann, A.; Voeltz, S.: Der weltweit erste Ottomotor mit Benzindirekteinspritzung und Doppelaufladung von Volkswagen. In: Sonderausgabe ATZ MTZ, November 2005 [4] Middendorf, H.; Voeltz, S.: Kurbelgehäuseentlüftung und Ölkreislauf des 1.6l-85kW FSI-Motors. In: Ölkreislauf von Verbrennungsmotoren, Essen 2005 [5] Krebs, R.; Szengel, R.; Middendorf, H.; Pott, E., Fleiß, M.; Hagelstein, D.: Der weltweit erste doppeltaufgeladene Otto-Direkt-Einspritzmotor von Volkswagen. In: 14. Aachener Kolloquium, Aachen, 2005 [6] Golloch, R.; Merker, P.: Downsizing bei Verbrennungsmotoren Grundlagen, Stand der Technik und zukünftige Konzepte. In: MTZ 2/2005 [7] Szengel, R.; Middendorf, H.; Wiedmann, M.; Wietholt, B.; Laumann, A.; Voeltz, S.; Stiebels, B.; Damminger, L.: Die Ottomotoren des neuen Volkswagen Golf. In: Sonderausgabe ATZ MTZ, Oktober 2003 [8] Krebs, R.; Spiegel, L.; Stiebels, B.: Ottomotoren mit Direkteinspritzung von Volkswagen. In: 8. Aachener Kolloquium, Aachen, 1999
6 Summary Volkswagen has rigorously pursued its TSI strategy with the 1.4 l 90 kW TSI engine. MTZ 07-08I2007 Volume 68
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