2008-01-0611
A New 3 Cylinder 1.2l Advanced Downsizing Technology Demonstrator Engine Dave Hancock, Neil Fraser, Mike Jeremy, Richard Sykes and Hugh Blaxill MAHLE Powertrain Ltd Copyright © 2008 SAE International
ABSTRACT This paper introduces the new 3 cylinder 1.2l downsizing technology demonstrator engine from MAHLE. The purpose of the paper is to describe the design approach and technologies applied. Emphasis is given to the low speed torque and transient response issues associated with advanced downsized engines. An overview of the design of all engine systems is provided, including the predictive analysis results used to validate, guide and optimize the design process. The design targets outstanding levels of performance, fuel consumption & drivability.
This has led to a significant increase in the number of downsized engines on the market today [3, 4, 5, 6] with more engines to be introduced in the near future. Further improvements in fuel economy have been shown to be possible through increased levels of engine downsizing [7,8]. However, as specific output increases so too do the technical challenges; the foremost of these being: •
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INTRODUCTION Since the agreement to reduce average new car CO 2 emissions to 140g/km by 2008, fuel consumption improvement has been one of the main drivers for engine development within the automotive industry. Current status of the EU fleet indicates that significant work is still required to achieve this target, particularly in the field of gasoline spark ignition engines [1]. Furthermore recent proposed legislation in California indicates the US market will follow the EU’s lead. In recent times various technologies have been applied to gasoline engines to improve operating efficiency. The most significant of these being: Direct injection (homogenous and stratified lean) Variable valvetrains (profile switching, cylinder deactivation, fully variable lift and duration) Controlled auto-ignition or homogenous charge compression ignition
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A robust combustion system that allows a high compression ratio to maintain part load efficiency Good low speed torque and transient performance Real world fuel consumption benefits through a reduction in full load fuel enrichment Base engine robustness and durability
These challenges require unique engine design solutions to enable the significant efficiency improvements of an aggressively downsized engine to be realized. This paper summarizes the design of a new 1.2l, 3cylinder high specific output gasoline downsize engine. The purpose of this engine is to demonstrate the potential of different downsizing strategies and to provide a research tool for the development of new technologies by the business group.
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Engine downsizing
Gasoline engine downsizing is the process whereby the speed / load operating point is shifted to a more efficient region through the reduction of engine capacity whilst maintaining the full load performance via pressure charging. It has long been known as one of the most effective technologies for immediate implementation [2].
ENGINE DESIGN TARGETS The design targets for this engine were initially to produce an engine capable of replacing a typical 2.4l V6 PFi engine used in a C or D class European vehicle platform (circa 1600kg), delivering comparable performance with a fuel consumption reduction of circa 30% on the NEDC driving cycle. The engine would be compliant with EU5 / ULEV2 emissions from conception. The concept approach was based on the use of technology, which would be ultimately available for mass production techniques.
The engine was initially conceived as a high technology demonstrator with a twin turbo charger approach. However, a variant was also designed and produced at a lower performance level. This derivative is considered an intermediate step to the twin turbo solution. The I3 engine was designed with a high level of protection for a V6 version. As a downsizing demonstrator the engine was designed to deliver the best possible levels of friction & weight but also a high level of robustness to support the future requirements as an R & D platform. The design was independent of any carry over requirements from manufacturing facilities or existing engine architecture, however, it was intended to demonstrate a maximum content of group technology. Table 1 shows the basic performance and fuel economy targets for the twin turbo & single turbo engine derivatives. Target Performance at Full Load Single turbo: Torque Maximum at n = 2,500 - 3,000 rpm Tmax = 210 Nm (BMEP = 22 bar) Power Output Pmax = 90 kW Maximum at n = 6,000 rpm (Specific Power = 75 kW/l) Twin turbo: Torque Md = 153 Nm at n = 1,000 rpm (BMEP = 16 bar) Md = 286 Nm Maximum at n = 2,500 - 3,000 rpm (BMEP = 30 bar) Power Output Pmax = 144 kW (Specific Power = 120 kW/l) Maximum at n = 6,500 rpm Fuel Consumption at Part Load Optimum bsfc = 235 g/kWh at n = 2000 rpm / BMEP = 4 bar bsfc < 295 g/kWh Emissions EU5 / ULEV2 compliant Hydrocarbon emissions at n = 2000 rpm / BMEP = 2 bar <5.4g/kWh at n = 2000 rpm / BMEP = 4 bar <4.4g/kWh
ENGINE SPECIFICATION Table 2 shows the most important mechanical design features and dimensions of the base engine and Figure 2 shows a general view of the complete twin turbocharged engine. Main Data, Cylinder Block and Cranktrain Cylinderblock (material) A 356 (COSCAST) Bore 83,0 mm Stroke 73,9 mm Stroke-Bore ratio 0,89 Cylinder displacement 0,400 l Cylinder number and arrangement 3 in line Displacement 1.200 l Bore spacing 91 mm Conrod length 123 mm Block height 189,5 mm Compression ratio 9,75 Firing order 1-3-2 Crankshaft Main Journal Diameter 48 mm Crankshaft Pin Journal Diameter 48 mm Cylinder Head and Valvetrain Cylinderhead (material) A 356 (COSCAST) 4 valves/cylinder, Pentroof DOHC architecture with roller finger followers and dual independant cam phasing Central Gdi with Piezo Injector Valve head diameter Intake / Exhaust 31.4 / 25.5 mm Valve stem diameter Intake / Exhaust 6/6 mm Maximum valve lift Intake / Exhaust 11 / 11 mm º Valve angle Intake / Exhaust 21.5 / 20.0 Overal Dimensions Length (TVD to RfoB) x Width x Height 438x675x706 mm Dry Weight - fully dressed 145 kg
Table 2 – Base Engine Data
Table 1 - Performance targets 350 300 250
] m N [ e 200 u q r o T
150 100 MAHLE 1.2l Downsize Engine Gasoline Turbocharged Production Engines 1.0 to 1.6l
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Figure 1 - Target Torque Curve – Twin Turbo Figure 2 - Twin Turbocharged Engine Figure 1 shows the target torque curve for the twin turbo derivative compared to current commercially available turbocharged gasoline engines of between 1.0l and 1.6l swept volume.
ENGINE SUBSYSTEMS & COMPONENTS CRANKCASE, BEDPLATE & NUTPLATE During the concept phase of design, a variety of crank case architectures were considered, including: •
Monoblock crankcase and cylinder head
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Cast-in aluminium liners
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Cast-in iron liners
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Bedplate verses deep skirt construction
Consideration was given to the best overall solution for function, weight and cost. The eventual choice of an all aluminium construction with Nikasil coating for the cylinder bore was driven by the combination of excellent heat transfer properties and minimum bore spacing. The crankcase and all aluminium bedplate are cast in A356 alloy using a low pressure sand casting technique (COSCAST)[9] to give the best possible strength to weight ratio, taking advantage of the excellent tolerance capabilities of the process. The block has a closed deck construction for maximum stability. Through bolts run from the cylinder head through the crankcase and bedplate assembly into a separate nut plate to allow all of the aluminium components to be held in compression. This allows the use of large bay to bay breather passages to reduce pumping losses and also results in excellent bore distortion characteristics. The nut plate provides good load distribution for the main bearings giving low main bearing distortion. The bolts are inclined at 2º to run the bolts close to the main bearing bore without compromising the ports in the cylinder head. Figure 3 shows the cross section through the cylinder head bolts. Modern CNC machining centers enable the use of the angled cylinder head bolt for little cost effect.
Figure 3 – Cross section through crankcase and cylinder head Both bedplate and nut plate also form an integral part of the lubrication system using a combination of cast-in and machined passages. The bedplate also includes a cast oil scraper / baffle system. The crankcase incorporates a split cooling system, supplying the cylinder head with coolant from a separate channel to the main crankcase coolant. Verification of the base engine structure was carried out using FEA to analyse both bearing panel conditions (Figure 4) & the thermo mechanical loading in the whole system. CFD of the cooling system was also carried out to optimise pressure drop, local velocities and distribution of the coolant flow within the cylinder block. Safety Factor
10000
LC7 0 0 0 0 0 1 -
LC6
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Figure 4 – Bearing Panel Finite Element Analysis – Crankcase Half Finishing of the bearing and cylinder bores was completed with the addition of ‘dummy’ cylinder heads and nut plates to replicate the final assembled condition of the assembly to obtain the best possible final roundness of the features. Figure 5 shows typical measured bore distortion data.
The cooling system in the cylinder head is in a cross flow configuration fed from the outboard channels in the crankcase. CFD was used to confirm that the flow across the cylinder head was consistent for all cylinders (Figure 7).
Figure 7 – Cylinder Head Cooling Jacket Heat Transfer Coefficient Distribution Across the Cylinder Head Figure 5 – Typical Measured Cylinder Bore Distortion
CYLINDER HEAD The layout of the cylinder head and, within, the combustion chamber, ports and cooling jacket are key to achieving the project targets for performance and robustness.
In order to obtain stable lambda 1 combustion the inlet port was designed for a high level of tumble. Performance of the inlet port and combustion chamber design was initially verified using moving geometry CFD (Figure 8 shows in-cylinder flow vectors). The design was further verified using a rapid prototyped flow box. Both the inlet port and the combustion chamber are machined into the cylinder head casting to allow for minor design changes and to ensure consistent performance across all 3 cylinders.
The cylinder bore is, at 83mm, slightly larger than ideal from the perspective of surface area / volume ratio. This compromise was necessary in order to ensure excellent cooling provision for the spark plug and injector, whilst maintaining valve sizes necessary to achieve the performance objectives of the engine The fuel injector is situated on the inlet valve side of the combustion chamber at a slight angle, with the long reach M10 spark plug further offset towards the exhaust side. Further use of the low pressure sand casting process allows a complex water jacket to be wrapped around the plug, injector and ports with a complex form of fins directing coolant flow around the injector tip. The flow in this area was carefully optimized through the application of CFD. A section of water jacket core is shown in Figure 6.
Figure 6 – Cylinder Head Water Jacket
Figure 8 – Air Motion In Cylinder The combustion chamber was designed to give a compression ratio of 9.75, with a variety of piston crown profiles. A cross section of the complete combustion chamber is shown in Figure 9.
The cylinder head was designed to suit both 3 cylinder and V6 configurations. With minor changes in machining to enable degas function, the V6 cylinder head can be produced from a common casting, which is produced with a chain case cast on both ends. Fuel pump & breather system features are also duplicated in the design with no compromise to the function of either derivative. The cylinder head also includes the base feature for the 3 cylinder engine mount for a transverse vehicle installation. CRANK TRAIN & BALANCER SYSTEM Figure 9 – Combustion Chamber Cross Section A cam ladder was chosen for the upper part of the cylinder head to give a flat cam cover surface, with a raised front section to form the rear half of the upper chain case thus allowing a single front cover for the whole engine. The ladder also includes features for the breather part load and make up air circuits. Oil is fed through the ladder via the front cam bearing to feed the two VCT units. A notch is included in the ladder to maintain a constant oil feed to the VCT.
GT Power was used for the initial gas exchange cycle simulations and suggested that the peak combustion pressure could be expected to be as high as 140bar (Figure 11). 160
140 Cylinder Pressure @ 6500 rpm bar Cylinder Pressure @ 6000 rpm bar Cylinder Pressure @ 5000 rpm bar Cylinder Pressure @ 4000 rpm bar Cylinder Pressure @ 3000 rpm bar Cylinder Pressure @ 2250 rpm bar Cylinder Pressure @ 2000 rpm bar Cylinder Pressure @ 1750 rpm bar Cylinder Pressure @ 1500 rpm bar Cylinder Pressure @ 1250 rpm bar Cylinder Pressure @ 1000 rpm bar
120
100
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60
The cylinder head gasket is a 3 layer multi layer steel design with an integral metal temperature sensor [10,11,12], which is able to pick up rapid coolant temperature changes, thus enabling the implementation of a precision cooling strategy. The gasket is designed to be robust with in-cylinder pressures of up to 140bar.
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Figure 11 – Combustion Pressure Prediction from GT Power With this in mind the Cranktrain was made to be robust but still relatively lightweight with good frictional properties.
Figure 10 – Cylinder head
The crankshaft was designed in conjunction with the balancer system to use a four counterweighted crank to give the best compromise of weight without excessive main bearing loading. The 48mm main bearing and pin journal diameters offer a good overlap for stiffness and strength, whilst as a fully machined part, the back of the webs and pins was also removed to reduce weight. The high combustion pressures however, still yielded areas of high stress in the fillet between the rear crank pin and web which required careful geometry optimization (Figure 12).
1.29 Assembly Stress 0.79 Unrealistic Pin Bending Condition 1.31
I ni ti al C on di ti on
F in al C on di ti on
Figure 13 – Connecting Rod Small End Analysis Figure 12 – Fatigue Factor of Safety of Crankshaft Rear Pin Fillet High piston strength and robustness was maintained by the use of a forged piston [13] with a 3 piston ring configuration and a 22mm diameter gudgeon pin. The piston top ring groove has a hard anodised coating for additional resilience. The gudgeon pin has a DLC coating to reduce friction and give additional resistance to pin bore scuffing. Alternative piston blanks are also available for the lower power derivatives of the engine. The piston ring pack is made up of a barrel faced compression ring, a Napier second ring and a 3 piece oil control ring. Initial targets for the ring pack are to produce a good blow by characteristic with low friction. Further tangential load and ring style optimization with be carried out during engine development.
A 3 cylinder in-line engine has free rotating couples of 1st and 2nd order. The primary couple imbalance, being the more significant, is counteracted by means of a pair of masses housed within two crankshaft-driven plastic gears. This design affords a robust, low weight solution, free of the package or oil aeration issues associated with more conventional balancer shaft layouts . The balancer gears are housed at the rear of the engine within the rear seal housing, and at the front of the engine, behind the chain drive, thus avoiding the usual issues of increased package width, or the running of balancer gear in the oil pan, associated with normal balancer shafts. The complete Cranktrain layout is shown in Figure 14.
The connecting rod is designed as a forged, fracture split part using the latest high strength steel 36MnVS4By [14]. The rod has webs running down into the big end for high strength and a profiled small end for reduced reciprocating mass and optimized bearing length in both rod and piston (see Figure 13 for optimization – acceptable Fatigue Factor of Safety is 1.2 in this case). The fracture split big end cap is held in place with M9 fasteners. Use of an aluminium bedplate without an iron insert results in potential clearance problems at elevated temperatures. This has been overcome by the use of thick wall bearing shells for the main journals. The big end shells are manufactured using the latest HVOF process to give a high load capacity at low cost. 180º thrust bearings are located in the upper half of bearing panel 2 and include ramps to ensure effective oil entrainment. All bearings used in the engine are lead free to comply with legislative requirements.
Figure 14 – Cranktrain and Camshaft Drive
CHAIN DRIVE A new 8mm pitch chain system from IWIS (Polygon Effect Reduced Chain – PER Chain) provides a high strength, simple drive for the camshafts with a reduction in noise when compared with normal bushed chain systems, but with equivalent resistance to harsh combustion products in the lubrication system (common in diesel and gasoline direct injection systems). All guides and blades have replaceable plastic facings mounted on aluminium carriers (in mass production these would be reinforced plastic parts [15]). The crankshaft sprocket is unique, requiring a slightly larger diameter to lead in the specially shaped side plates (Figure 15) of the low noise chain to reduced the polygon effect of the chain on the small diameter sprocket. A hydraulic tensioner is used to retain tension. A separate 8mm chain is used to drive the oil pump with no tensioner or guide blades.
chamber. Figure 16 shows a cross section through the exhaust camshaft
Figure 16 – exhaust camshaft cross section. The front plugs of the camshaft house control valves integrated into the cam phaser fasteners and also provide oil feed for the cam phasers from the front camshaft bearing. Verification of the valve train was carried out through analysis with GT Valvetrain and with dynamic measurements on a cylinder head rig (Figure 17). The valve train remains in control to over 7000 RPM.
Figure 15 – Polygon Effect Reduced Chain from IWIS
VALVE TRAIN Double overhead camshafts with dual independent cam phasers actuate the valves via roller finger followers for reduced friction. Hydraulic lash adjusters provide maintenance free valve clearances. Lightweight valves [13,16] run in powder metal guides and seat inserts enabling the use of reduced spring rates. This gives rise to reduced seating force and hence lower seat wear and friction. Exhaust valves are sodium filled for enhanced heat transfer performance. The camshafts are positioned outboard of the valves to allow optimum packaging of spark plug and injector and to protect the package for the use of profile switching or deactivating roller finger followers [17]. The camshafts themselves consist of an assembled tube with steel lobes and end plugs. The exhaust camshaft also forms part of the oil separation system [18] incorporating a small turbine device, which acts as a pre separator. Breather gas enters the camshaft through annular holes behind a shield before running through the centrifugal separator, the gas and separated oil then flowing from the end of the camshaft into a further separation
Figure 17 – Correlation of Valve Velocity from GT Valvetrain and Cylinder Head Rig Test
LUBRICATION SYSTEM Key design objectives for the lubrication circuit were low parasitic loss, weight and a compact package. Separate galleries are used to supply the bottom end of the engine, the cam phasing system, piston cooling jets and cylinder head lubrication. The oil feed to the cylinder head for the cam bearings and lash adjusters runs through a restrictor reducing oil consumption in the cylinder head. The cam bearings are also fed through the top of the bearing into the high pressure region, further reducing the oil flow to the bearings. The piston cooling gallery is supplied with unfiltered oil from a channel in the oil pump pressure control valve, removing the requirement for separate valves and their inherent pressure loss and cost. The pump body is cast in aluminium and uses a Gerotor type pumping element, which mounts directly to the underside of the bed plate.
For extreme cold starts the pump also includes a dump valve. Figure 18 shows the oil pump.
Figure 18 – Oil Pump The oil filter module is manufactured in plastic [19] for low weight and cost and has an integrated aluminium oil cooler. The assembly mounts directly to the side of the bedplate. The oil cooler specification can easily be modified depending on engine requirements. The filter module includes a fully ashable filter element and nonreturn, bypass and drain valves. The drain valve is automatically operated on removal of the cap thus avoiding oil spillage during maintenance. Figure 19 shows the oil filter module.
Figure 20 – Oil Pump Characteristic COOLING SYSTEM In order to optimize combustion, friction and wear within the engine a split cooling circuit was used with separate feeds and returns from the cylinder head and crank case circuits. Water is fed via an electric pump and distribution valve into the circuits, controlled by the engine management system. Feedback is provided by the integrated cylinder head gasket temperature sensor. CFD and Flowmaster modeling was carried out to optimise the circuits. The oil cooler, charge air cooler, and final EGR cooler, are fed by a separate low temperature circuit to maintain required temperature levels. The main circuit engine coolant is used to cool the exhaust manifold on the single turbo derivative and the EGR pre cooler on both derivatives.
Figure 19 – Oil Filter Module The oil pan is cast in A356 alloy and includes oil drainage channels, which return oil from the cylinder head and breather system to below the running oil level. The oil pan also includes an oil level and quality sensor and a central pick up point. Oil pump sizing (Figure 20) and circuit simulation was carried out with the aid of Flowmaster, to reduce overall system requirements and hence parasitic losses.
EXHAUST SYSTEM, TURBOCHARGING & EGR The twin turbocharger configuration is seen as the most cost effective route to satisfying the performance, drivability and fuel consumption targets for the project. The single turbocharger variant offers lower specific outputs, but at a lower cost, through the application of a diesel type VGT turbocharger. The EGR system is a key part of both installations for the control of exhaust gas temperature and improved knock sensitivity [20]. Twin Turbo system The twin turbochargers are configured in a sequential layout (Figure 21), and are positioned as close as is feasible to give reduced gas paths. The exhaust gas side of the system is designed to run with gas temperatures of 1050ºC, which in conjunction with full load cooled EGR should enable running at lambda 1 under most conditions. The high pressure turbocharger turbine housing is integrated into the cast stainless steel manifold directly
opposite one of the runners to benefit from pulsation effects. The manifold also has a direct connection for EGR, and a bypass valve block leading to the low pressure turbocharger, (for peak power delivery) and turbine housing (also both cast steel). The low pressure turbine housing also contains a traditional wastegate. Exhaust gas from the high pressure turbine flows through a separate passage in the valve block and through the low pressure turbine housing to improve response.
low speed most work is provided by the high pressure turbo, whilst the low pressure turbo is driven purely by the exhaust from the high pressure device. As the speed and load increases, the valve to the low pressure turbine from the manifold is opened, and the valve on the high pressure compressor is opened, transferring the load onto the low pressure turbine. The high pressure turbine is not isolated in this case and the continuous flow of exhaust gas through it ensures rapid response on demand.
EGR valve
Charge air cooler
EGR cooler
Compressor bypass
C
C
HP stage Turbine Bypass Valve
T
T
LP stage Turbine Wastegate
Figure 21 – Twin Turbo Schematic
Figure 23 - Compressor Maps with Proposed Run Up Curves
On the compressor side of the assembly, all air passes through the low pressure compressor housing, and either through the high pressure compressor or via an integrated bypass valve direct into the intake system. The twin turbo assembly is shown in Figure 22.
Figure 24 – Wastegate and Bypass Valve Control Strategy, and Turbocharger Speed from GT Power Control of all exhaust gas and air path valves uses vacuum actuators due to the extremes of temperature experienced around the turbocharger package. Single Turbo system
Figure 22 – Twin Turbocharger Assembly GT-Power has been used initially to both verify the turbocharger characteristics and to develop the switching strategy for the system (Figures 23 and 24). At
The single turbo derivative (Figure 25) uses a combination of water cooled aluminium exhaust manifold, and full load cooled EGR to limit the inlet temperature to a variable geometry turbine diesel turbocharger (temperature limit circa 850ºC) thus minimising the requirement for over fuelling
Figure 25 – Single turbocharger installation with water cooled manifold.
Figure 26 – Twin Turbo EGR System
Control of the variable geometry mechanism is via a stepper motor. EGR system
INTAKE & FUEL SYSTEM
Cooled EGR is one of the most promising technologies to enable stoichiometric operation at full load and increased compression ratio through enhanced knock performance [20]. Both engine specifications run with similar systems, however the single turbo will use a rotating air pulse valve to improve the level of EGR available over a broader range. Target EGR level is circa 15% in the full load map, with an inlet temperature to the inlet manifold of 120ºC.
The fuel system is designed around a piezo fuel injector, however, solenoid injectors may also be accommodated with minor machining changes. The piezo injection system offers sufficient dynamic range for all gasoline and alcohol applications, and with its outward opening nozzle should be less prone to soot build up. The fuel pump is driven via a drive dog from the end of the inlet camshaft.
The EGR system will include a tube in tube pre-cooler, running on the main coolant circuit to reduce the gas temperature to 700ºC. A plate type main cooler with a bypass valve will be used to control final gas temperature to 120ºC under all speed and load conditions. EGR will be introduced through a rotating barrel valve into each intake manifold tract, close to the port entry. The valve will offer a high level of control and extremely low volumes of residual EGR during transient operating conditions. The valve will also act to keep the individual intake tracts isolated. The total system for the single turbo is shown in Figure 26.
Primary design targets for the intake system were reduced system volume to aid transient performance and a minimum of flow loss. Plastic components are used where possible to reduce weight. The design had to be robust with post turbocharger pressures in excess of 2.8bar (absolute). Positioning of the throttle became a major design issue due to predicted compressor outlet temperatures of over 200ºC. This required that the throttle should be sited post charge air cooler, however, with commercially available throttle systems, an unacceptable level of pressure loss would occur due to multiple changes of direction and cross section in the package space available. A triple blade throttle was therefore used actuated via a single shaft and motor. The throttle bore profile was designed to replicate the opening characteristic of a single 50mm diameter throttle to improve initial response (Figure 27)
Figure 29 – GT Power Prediction of Manifold Volumetric Efficiency Figure 27 - Throttle Bore Profile The throttle was designed for production in plastic, however it was manufactured in aluminium as a prototype part due to the lack of availability of reinforced prototype material to give the required dimensional stability. An aluminium charge air cooler assembly is mounted on to the engine with intake and exit cast ducts welded onto the charge air cooler. Extensive CFD analysis was used to give the best possible air distribution onto the charge air cooler face (Figure 28), through the use of internal ribbing in the pre cooler duct, and hence into the throttle set and on into a small post throttle plenum.
Figure 28 – Comparison of CFD on Pre Charge Air Cooler Duct Before and After Optimisation
The low volume air cleaner assembly mounts directly onto the cam cover, also providing closure for the breather system for a serviceable fibre filter. The air cleaner element uses the latest micro fibre technology for improved dirt removal performance with reduced pressure drop. NVH analysis (Figure 30) was used to predict the requirement for counter NVH measures. Helmholz and gap resonators are included in the clean air duct to the turbocharger inlet to remove high frequency content. In vehicle installations, a 2 – 3 litre water trap would also be required in the dirty air duct. The manifold air flow (MAF) sensor is integrated into the air cleaner lid, with an internal bellmouth to give consistent flow across the sensor. The complete air intake system is shown in Figure 31.
Figure 30 – Initial Induction System NVH Analysis
The intake manifold was designed as a two piece plastic moulding, the two halves being friction welded together. The manifold includes entry passages for EGR and breather gas into each runner to give the best possible distribution. The runner length was tuned using GT Power (Figure 29) to help fill in a small torque dip in the transition between high pressure and low pressure turbochargers and still retain high levels of volumetric efficiency elsewhere in the engine running range.
Figure 31 Application
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Induction
System
For
Twin
Turbo
SEALING, COVERS & BREATHER SYSTEM Air intake
The front cover is an aluminium cast part which carries the front crankshaft seal, cam phaser control solenoids, and a sealed pad for engine mounting. The rear cast aluminium cover carries the rear crankshaft seal and contains the rear balancer gear. A plastic semi isolated cam cover with a T section seal is also used to mount the fuel rail, high energy coils and air cleaner (Figure 32). The non-return valves and passages for the breather system are also integrated into the cam cover to avoid breather icing. An impactor chamber with microfibre filter, for final breather system separation, is also integrated into the cover with bypass valve and drain system. The cam cover shares fasteners with the cam bearings in 6 places around the periphery of the cover to reduce the number of fasteners required. Since the fasteners pass through an area which includes the cam bearing oil feed, an extra sealing element is also included on the fastener. All cam cover fasteners are retained in the cover for ease of assembly and supply, and include integrated compression limiters for the cover sealing system.
Air filter AFM
Wide Open throttle
Ps -5...0 kPa
EGR
Throttle
Positive ventilation
EGR Cooler Ps -85...+70 kPa
Circa Ø5mm
OilSeparator
Additional check valves Blowby-Gas
Exhaust gas
Intercooler
Part. Open throttle
Positive ventilation Blow-byGas
Turbo Charger
PCV
PC + pSO
Drain line under oil level in oil pan or back into crankcase with drain back valve
) (
PC -5...0 kPa
Figure 33 – Breather System Schematic The return of part load breather system gas is routed to a distribution rail integrated into the intake manifold. Full load breather gas is routed to the entry of the low pressure compressor housing. This passage also serves to supply filtered make up air back into the engine during part load operation.
ANCILLARY SYSTEM An increasing use of electric PAS is occurring in vehicle platforms as pressure for fuel consumption improvement increases [21]. The use of the electric water pump system for engine and charge / EGR cooling is also key to improved cooling system control, and reduced parasitic loss.
Figure 32 – Components
Cam
Cover
and
Breather
System
The breather system brings together a number of technologies not previously integrated into a single system. The use of camshaft separator, impactor device, and filter should result in the class leading levels of oil separation required for a compact engine with high combustion pressures and use of multiple turbochargers. The breather system schematic is shown in Figure 33.
The simple ancillary drive (Figure 34) remains to drive both the air conditioning compressor, (a pre-requisite on most vehicles), and a starter generator unit. The base fuel consumption target for the project does not include the use of stop-start technology, and as such this system should result in a further benefit. Further development may also result in the unit being used as a torque boosting system to further enhance drivability. The air conditioning compressor is mounted direct onto the bedplate and oil pan to reduce weight, however the requirement for a robust pivot point for the starter generator has driven the requirement for a separate bracket.
paper 2003-01-0081 SAE World Congress Detroit March 2003. [3] Welter, A. et al. “The new turbocharged BMW six cylinder inline petrol engine” 15. Aachen Colloquium, Aachen, 2006. [4] Groff, E. et al. “The New 2.0l High Performance Turbo Engine with Gasoline Direct Injection from GM th Powertrain” 27 International Vienna Motor Symposium, 2006. [5] Nishida, M. et al. “The new 2.3l Direct Injection Turbo th Gasoline Engine from Mazda” 14 Aachen Colloquim Aachen 2005
Figure 34 – Ancillary Drive
CONCLUSION Maximizing the fuel economy and minimizing CO2 emissions are the most important objectives for the development of future vehicles and engines. Downsizing is an important option to achieve these objectives, especially for gasoline engines, together with the achievement of good drivability. For this scenario a downsizing demonstrator engine with a displacement of 1.2 l was designed. This engine has targets to achieve full load performance, which will enable a downsizing of 50 %, resulting in a potential of 30 % fuel consumption and CO 2 emission reduction in the NEDC. The focus of the concept design was the pressure charging, (single and sequential twin turbo), gasoline direct injection, and variable valve timing systems. The thermodynamic and mechanical properties (friction and weight) of the engine were optimized carefully to achieve high levels of efficiency. The engine has been designed not only as a technology demonstrator, but also as a production viable solution, to deliver a robust research and development platform to satisfy future requirements.
REFERENCES [1] Commission of European Communities “Implementing the Community Strategy to Reduce CO2 Emissions from Cars. First annual report on the effectiveness of the strategy”, http://ec.europa.eu/environment/docum/00615_en.htm [2] Atkins, J M and Koch, C R “A Well-to-Wheel Comparison of Several Powertrain Technologies” SAE
[6] Krebs, R. et al “The New Audi 2,0l T-FSI Motor – The th First Direct Injection Turbo Gasoline Engine at Audi” 25 International Vienna Motor Symposium, 2004. [7] N Fraser, H Blaxill, “Engine Downsizing and the Application of Gasoline Direct Injection to a High Specific Output Turbocharged Engine”, Fuel Economy and Engine Downsizing Seminar Proceedings, May 2004. [8] Bandel, W. et al. “The Turbocharged GDI Engine: Boosted Synergies for High Fuel Economy Plus Ultralow Emission” SAE paper 2006-01-1266 SAE World Congress Detroit 2006.
[9] Smith, R.A and Wilkins, P.S.A.W. Low Pressure Sand Casting: Current Experience with a New Process. AFS Transactions 86-144 [10] M. Laske, J. Ludwig, G. Unseld, G. Weiss, Zylinderkopfdichtung mit integrierten Temperatursensoren - Aufspüren kritischer Bauteiltemperaturen im Herzen des Motors ; MTZ 12/2002 Jahrgang 63 [11] M. Laske, J. Ludwig, G. Unseld, A. Weiss, Thermal Examination of a MLS Gasket with integrated Sensor Technology for an effective Heat Management, SAE Technical paper series 2003-01-0473 [12] M. Laske, Intelligent abdichten – Zylinderkopfdichtungen mit integrierten Sensoren für Kraftstoffersparnis und Emissionsreduzierung, OEM & Lieferant · Messeausgabe zur IAA 2007 [13] Spangenberg S., Sander W., Geyer A. Hochbelastete Werkstoffe: Werkstoffe im Umfeld der Verbrennung; ATZ/MTZ extra, Wekstoffe im Automobilbau, November 2005 [14] Spangenberg S., Kemnitz P., Kopf E., Repgen B.: Massereduzierung an Bauteilen des Kurbeltriebs Pleuel im Fokus; MTZ 67 (2006) 4, S. 254
[15] S. Belmer, T. Fink, I. Lorenz, H. Neukirchner, Steuertriebe für Verbrennungsmotoren – Konzeption, Auslegung und Basiskonstruction ; MTZ 6/2005 Jahrgang 66 [16] Abele M.: Das Leichtbauventil - Ergebnisse bisher und Potenzial für die Zukunft; 2. Fachtagung "Ventiltrieb und Zylinderkopf", VDI Werkstofftechtik, Leonberg 25./26.10. 2006 [17] Korte V., Blaxill H., Schneider F., Hoffmann H.: Variable Valve Timing: General Options and Technologies; SIA Variable Valve Actuation Conference, IFP Rueil-Malmaison, November 2006 [18] Schellhase T., Sauter H., Ruppel S.: Assembled Camshaft with Integrated Oil Mist Separation; SAE Paper 2007-01-0260 [19] Klöpfer M., Kiemlen R., Siefert D.: Ölfiltermodule mit Vollkunststoffölfilter; MTZ 66 (2005) 10 [20] Ganser J., Blaxill H., Cairns A.: Hochlast-AGR am turboaufgeladenen Ottomotor; MTZ 68 (2007) 7-8, S. 564 [21] Korte V., Hemminger B.: Verbesserung der Wirtschaftlichkeit von Pkw-Motoren durch Reduzierung der Antriebsleistung von Hilfsaggregaten; Nebenaggregate im Fahrzeug, Haus der Technik Essen 1988
CONTACT Dave Hancock, MAHLE Powertrain Ltd., Costin House, St. James Mill Rd., Northampton, NN5 5TZ, U.K. Tel: +44(0)870 157 3014, Fax: +44(0)870 157 3082, Email:
[email protected] Neil Fraser, MAHLE Powertrain Ltd., Costin House, St. James Mill Rd., Northampton, NN5 5TZ, U.K. Tel: +44(0)870 157 3133, Fax: +44(0)870 157 3082, Email:
[email protected] Further information Powertrain website:
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