100 HP / 200 Nm Diesel Motorcycle with 6 Speed Automated Manual Transmission

Transcription

1 SAE JSAE HP / 200 Nm Diesel Motorcycle with 6 Speed Automated Manual Transmission Bernhard J. Graf, Wolfgang J. Schoeffmann, Peter W. Bartsch, Hermann Pecnik, Andreas Mair and Manfred Madler All of AVL List GmbH, Graz 2004 Small Engine Technology Conference Graz, Austria September 27-30, Commonwealth Drive, Warrendale, PA U.S.A. Tel: (724) Fax: (724) Web:

2 All rights reserved. No part of this publication may be reproduced, stored in a retrieval system, or transmitted, in any form or by any means, electronic, mechanical, photocopying, recording, or otherwise, without the prior written permission of SAE. For permission and licensing requests contact: SAE Permissions 400 Commonwealth Drive Warrendale, PA USA permissions@sae.org Fax: Tel: For multiple print copies contact: SAE Customer Service Tel: (inside USA and Canada) Tel: (outside USA) Fax: CustomerService@sae.org Copyright 2004 SAE International Copyright 2004 Society of Automotive Engineers of Japan, Inc. Positions and opinions advanced in this paper are those of the author(s) and not necessarily those of SAE. The author is solely responsible for the content of the paper. Printed in USA

3 / HP / 200 Nm Diesel Motorcycle with 6 Speed Automated Manual Transmission Bernhard J. Graf, Wolfgang J. Schoeffmann, Peter W. Bartsch, Hermann Pecnik, Andreas Mair and Manfred Madler AVL List GmbH, Graz Copyright 2004 SAE International and Copyright 2004 Society of Automotive Engineers of Japan, Inc. ABSTRACT Diesel engines, especially CR (Common Rail) DI (Direct Injection) TCI (Turbo Charged Inter-cooled), share a wide acceptance in the passenger car market due to the enormous torque and flexibility at low engine speed. A pre - condition for the use of a diesel engine in a motorcycle is that the disadvantages like combustion noise and visible smoke are reduced or eliminated. Moreover the fuel economy and performance characteristics of a diesel engine are dedicated to be used in a touring or large displacement motorcycle. The AVL engine concept is the first high performance diesel engine to be specially designed for motorcycles in terms of packaging and styling. To compensate for the limited engine speed range a gearbox with a wide ratio spread is required. This leads to a manual transmission with at least 6 gears or an automatic transmission. For the AVL concept an AMT (Automated Manual Transmission) was selected. CONTENT Besides the design work a 1D thermodynamic engine simulation with AVL's software code BOOST was carried out in order to optimise the gas exchange of the engine including the VTG (Variable Turbine Geometry) turbocharger characteristics to predict the engine performance. To optimise the engine in terms of cyclic speed irregularity a torsional vibration equivalent system was set up using AVL's software code BRICKS. This simulation included the variation of the flywheel mass and the analysis of different cylinder pressure characteristics. Additional simulations of the entire motorcycle in order to determine the driving performance were carried out using AVL's Vehicle simulation software CRUISE. The engine design is based on AVL know-how from development results and measurement data of state of the art HSDI (High Speed Direct Injection) diesel engines. The packaging of the engine in the vehicle is shown using rapid prototype parts. INTRODUCTION Diesel engines in the area of two wheeled vehicles or compact engines are not very common. There are only small niches where diesel engines find place in this field of applications. The range starts with single cylinder engines for generators and goes up to one or two cylinder engines for vehicle applications such as small capacity tractors or three wheelers. Only a small number of diesel powered motorcycles have been introduced over the last years in commercial production. Usually diesel engines for these vehicles were originally designed for stationary applications and modified to be fitted in the vehicles. The layout of these engines is simple and the power output is too low for the use in large displacement motorcycles. To be competitive with this type of motorcycle an engine power of at least 100 HP is required. This fact combined with a reasonable motorcycle displacement of 1200 to 1400cc demands a turbocharged engine and the special know-how from current passenger car diesel engine development to comply with the stringent diesel emission regulations. In state of the art diesel engines for passenger cars, new technologies to increase fuel efficiency and specific engine power, to improve sound quality, reduce combustion noise, along with the improved methods for the exhaust gas after-treatment opened up a new perspective for the use of diesel engines in motorcycles. Due to the high low-end torque of TCI diesel engines the bad perception on the image of diesel engines was no longer justified. Furthermore the tendency to a higher number of gears in automatic transmissions, and also in manual or automated manual transmissions allows for more efficiency in terms of driveability and fuel consumption. All these characteristics of diesel engines are only benefits if the engines fulfill the stringent future emission regulations.

4 The challenge to package a new engine, designed especially for motorcycles with all the required state of the art technologies into a big displacement Naked Bike needed a lot of considerations. The design of the engine has to be similar to typical motorcycle engines because the engine in this type of vehicle is not covered and therefore is very much part of the vehicle design. Additional aims were to comply with future emission regulations and to be competitive with big displacement Naked Bikes on the market in terms of acceleration, maximum speed, weight and additional performance. The Technical targets for the diesel motorcycle are: Meet future Pollutant and Noise Emission limits Performance (100 HP/200Nm) Comparable vehicle weight (Benchmarks) Motorcycle engine appearance Automated gearbox The roadmap of how to meet all these targets is shown with results from calculations and simulations and is covered in the following chapters. Fig. 2 ECE R40 Driving cycle for Motorcycles (present) [1] Compression ignition engines must also comply with the smoke emission limits to be tested under full load (as required by ECE R24 shown in Fig. 3) and free acceleration. FUTURE EMISSION SCENARIO FOR MOTORCYCLES AND DIESEL PASSENGER CARS The motorcycle manufacturers have to deal with stringent emission regulations (see Fig. 1) but also with new planned emission test procedures starting at EU III which is valid for TA (Type Approval) starting on January 1 st 2006 and ANR (All New Registrations) starting on 1 st January Present and future emission regulations for motorcycles > 150cc in the EU Valid from Measuring mode EU II TA: ANR: ECE EU III TA: ANR: x UDC + 1x EUDC CO [g/km] 5,5 2,0 Fig. 3 ECE R24 Smoke test cycle (present) [1] The future emission limits will be tested along using a proposed driving cycle (see Fig. 4 ). This driving cycle is similar to the NEDC but includes 6 instead of 4 Elementary Urban Cycles (see Fig. 6). The proposed test cycle includes cold start, the two warm-up cycles are cancelled. HC [g/km] 1,0 0,3 NOx [g/km] 0,3 0,15 Fig. 1 Emission Regulations Motorcycles [1] The present motorcycle test procedure to determine the pollutant emissions is the ECR R40 test cycle (see Fig. 2). This test cycle starts with hot conditions after two warm-up cycles for preparation. Fig. 4 Proposed future Driving cycle for motorcycles As a comparison the future emission limits and the driving cycle for diesel passenger cars are shown below. The present stage of emission regulations for diesel passenger cars is EU III. All development activities are

5 focusing on EU IV and beyond. The current limits together with the future emission limits are shown in Fig. 5. Present and future emission regulation for diesel passenger cars in the EU EU V EU III EU IV UBA till 10/ 2003 Valid from Present TA: ANR: ~2010 Measuring mode NEDC 98/69/EC NEDC 98/69/EC NEDC 98/69/EC CO [g/km] HC [g/km] NOx [g/km] PM [g/km] Fig. 5 Emission Regulations diesel passenger cars [1] EU IV limits are already fixed and the date of implementation for Type Approval is the 1 st of April All new registrations follow at the 1 st of July EU V limits are a still discussed proposal of the German UBA (Deutsches Umwelt Bundesamt) to make common the diesel and gasoline emission limits starting in The Test cycle for passenger, NEDC (New European Driving Cycle) is shown in Fig. 6 and combines UDC (Urban Driving Cycle) and EUDC (Extra Urban Driving Cycle). The test cycle has to be started in cold starting conditions. [1] Fig. 6 NEDC 98/69 EC Driving cycle for Passenger cars [1] PREDICTED FUTURE EMISSION SCENARIO FOR DIESEL MOTORCYCLES There is no future emission scenario available for diesel powered motorcycles, therefore a combination of future emission limits for motorcycles and future emission limits for diesel passenger cars was used to create a predicted future emission scenario for diesel powered two wheelers. These future emission limits consist of the limits for HC, CO and NOx taken from future emission limits for motorcycles and a PM (Particulate Matters) limit from the diesel passenger car limits equivalent with EURO IV (see Fig. 7). Predicted future emission regulation for diesel powered motorcycles in the EU Diesel Diesel Motorcycle Motorcycle Passenger Car EU III EU IV Valid from Measuring mode TA: ANR: x UDC + 1x EUDC 6 x UDC + 1x EUDC CO [g/km] HC [g/km] NOx [g/km] TA: ANR: NEDC 98/69/EC PM [g/km] Fig. 7 Predicted future emission regulations for diesel motorcycles The diesel motorcycle was designed to fulfill these predicted limits using the proposed future driving cycle for motorcycles as shown in Fig. 4. REQUIRED TECHNICAL FEATURES TO MEET THE PREDICTED EMISSION SCENARIO The minimal equipment to meet the predicted future emission limits with the diesel motorcycle is strongly directed by the relatively low power requirement in the future test cycle for motorcycles. Because of the low power requirement during the new motorcycle test cycle (cycle average 1.55 kw) as a result of the low vehicle weight compared to passenger cars (cycle average approximately 5 kw) the engine load is quite low and thus offering excellent prerequisites for exhaust gas and noise emission optimization. The predicted minimum requirements for the exhaust after-treatment system to comply with the future emission limits for diesel motorcycles are: Modulated, non-cooled EGR (Exhaust Gas Recirculation) Diesel Oxidation Catalyst (DOC) system as for diesel passenger car EURO III applications To further suppress visible smoke emission during full load periods a flow-through particulate filter could be applied. This filter is mentioned as an add-on solution to oxidize particulates but it is not necessarily required to meet the future emission targets for diesel motorcycles. The benefit of this flow-through type particulate filter is a PM reduction of 24% achieved in a passenger car equipped with a 6 cylinder 3.0 litre diesel engine in the

6 European stage III driving cycle. The filter was applied together with an upstream oxidation catalyst. [2] Many passenger cars have to use a Wall Flow particulate trap with much higher efficiencies to comply with EURO IV PM emission limits. A predicted roadmap which gives an overview of the EGR system specifications and of the required exhaust after-treatment technologies planned for the diesel passenger car future emission scenario is shown in Fig. 8. For the stationary test procedure according to ISO 5130 (see Fig. 10) there are no limits fixed. The highest recorded value is entered in the type approval document and defines the in-use noise limit. Fig. 10 Test procedure for stationary noise test (ISO 5130) BENCHMARK OF LARGE DISPLACEMENT NAKED BIKES UF... Under Floor DOC... Diesel Oxidation Catalyst CC... Closed Coupled DPF... Diesel Particulate Filter Fig. 8 Technology Roadmap for diesel vehicles The mentioned technologies for passenger cars can vary if they are applied in different vehicle types with different motorizations and different weights. EUROPEAN NOISE EMISSION SCENARIO FOR MOTORCYCLES The current noise emission regulation for motorcycles in Europe consists of two measurement procedures (passby and stationary) and is defined as ECE R The pass-by noise test has to be performed according to ISO 362 and the stationary test according to ISO The limits of the pass-by noise test for motorcycles with a displacement > 175cc is 80 dba. The test procedure is defined in ISO 362 and shown in Fig. 9. The gear which has to be selected for motorcycles with a displacement of > 175cc is the 2 nd or 3 rd gear. If the engine speed exceeds 110% of max. power speed in 2 nd gear only the 3 rd gear result is valid. Fig. 9 Test procedure for pass-by noise test (ISO 362) It has been decided to place the AVL diesel Motorcycle Study in the Naked Bike class since the development of a vehicle of this type is the most demanding task in terms of styling and visual appearance of the engine because it can not be hidden behind a fairing. Furthermore the absence of any covering requires a lot of work to reduce the engine noise by using internal engine measures such as a special combustion process in part load conditions to achieve an acceptable noise level without encapsulation. At the same time it would not make sense to place a diesel motorcycle in the Supersport class in order to compete with high revving high performance engines. Nevertheless most motorcyclists are performance oriented so it is absolutely necessary that the diesel motorcycle offers the vehicle performance which is competitive in the chosen vehicle category. Additionally it had to be taken into consideration not to exceed the usual weight and size of the Naked Bikes on the market since these parameters have a big influence on the driving performance and handling of a motorcycle. As a benchmark value in terms of vehicle size the wheel base was chosen. Further reference values are the fuel consumption and tank capacity which result in a certain travelling range. The travelling range may not be of high importance in the Naked Bike class but for the application in a touring motorcycle it is of high interest. In order to compare the potential in terms of emissions, the emission values which have been measured during type approval for Europe are shown in the benchmark matrix (Fig. 11). Of course it has to be said that these motorcycles are capable of reaching lower emission values than shown in the matrix. For example significantly lower emissions will be required to comply with the European regulations which will come into force in 2006 (see Fig. 1). Besides all the aforementioned reference values, also shown are the main engine specifications of the chosen benchmark vehicles in the matrix (Fig. 11).

7 SUZUKI GSX 1400 KAWASAKI ZRX 1200 R YAMAHA XJR 1300 HONDA CB 1300 Engine Type 4 - Stroke 4 - Cylinder in-line 4 - Stroke 4 - Cylinder in-line 4 - Stroke 4 - Cylinder in-line 4 - Stroke 4 - Cylinder in-line Cooling Air / oil Liquid Air Liquid Valve train 4 Valve DOHC 4 Valve DOHC 4 Valve DOHC 4 Valve DOHC Bore x Stroke 81.0 mm x 68.0 mm 79.0 mm x 59.4 mm 79.0 mm x 63.8 mm 78.0 mm x 67.2 mm Displacement 1402 cc 1164 cc 1251 cc 1284 cc Power 78 kw / rpm 90 kw / rpm 72 kw / rpm 85 kw / rpm Torque 126 Nm / rpm 110 Nm / rpm 104 Nm / rpm 117 Nm / rpm Carburation Fuel injection 4 x CV Carburettor 4 x CV Carburettor Fuel injection (36 mm) (37 mm) Gearbox 6 - speed manual 5 - speed manual 5 - speed manual 5 - speed manual Top Speed 225 km/h 230 km/h 213 km/h 230 km/h Dry weight 229 kg 223 kg 232 kg 252 kg Wheelbase 1520 mm 1465 mm 1510 mm 1515 mm Tank capacity incl. reserve 22 l 19 l 21 l 21 l Travelling range 328 km 322 km 375 km 323 km Fuel consumption 1) 6.7 l / 100 km 5.9 l / 100 km 5.6 l / 100 km 6.5 l / 100 km Acceleration km/h 3.0 s 3.0 s 3.4 s 3.1 s km/h 5.4 s 5.1 s 5.9 s 5.4 s km/h 4.1 s 3.9 s 4.4 s 3.7 s km/h 4.8 s 3.1 s 4.6 s 4.2 s Emissions 2) CO 6.35 g/km 7.10 g/km 1.81 g/km 3.51 g/km HC 1.21 g/km 1.88 g/km 0.74 g/km 0.83 g/km NO x 0.18 g/km 0.09 g/km 0.17 g/km 0.16 g/km Exhaust - emission control Remarks: 2 open - loop catalysts secondary air 1) Country road, 2) Homologation EU Fig. 11 Benchmark Large Displacement Naked Bikes [3], [4] open - loop catalyst secondary air open - loop catalyst secondary air open - loop catalyst secondary air DEVELOPMENT OF SPECIFIC POWER IN DIESEL ENGINES FOR PASSENGER CAR APPLICATIONS The tremendous increase in specific power and torque (shown in Fig. 12) of engines used for diesel passenger cars and SUV s (Sport Utility Vehicles) over recent years was rendered possible by the enormous improvements regarding fuel injection and charging system capabilities. The specific power values for passenger cars are currently around 60kW/l and for SUV`s around 55kW/l. The trend is steadily increasing. Today s max. rated speed of DI, CR diesel engines is at 4200rpm. There is an aim to have the rated speed higher to get a wider engine speed range which results in an improved driveability. In consequence this opened up the possibility to use state of the art diesel technology for a high performance motorcycle. In Fig. 12, the indicated value for the diesel motorcycle shows the compromise between the specific power and peak firing pressure. The peak firing pressure was limited to 140bar to keep the noise emission and the mechanical friction at an acceptable value for this nonencapsulated engine application. Furthermore the engine structure can be more lightweight with this compromise. PREDICTED LOAD DISTRIBUTION OVER VEHICLE LIFECYCLE Fig. 12 Specific power and torque development in diesel engines for passenger cars and SUV`s For the development of an engine in a new engine category, it was necessary to estimate and fix a standard vehicle lifecycle. This assumption includes the total mileage over the vehicle lifetime, average speed and

8 load distribution in all gears and was a result of the combination of experiences from passenger car and motorcycle development. As a total mileage km was fixed for the diesel motorcycle. Under consideration of the estimated vehicle load distribution this total mileage corresponds with a calculated lifetime of 1390 hours and an average speed of 86.4 km/h over the vehicle lifecycle. All calculations of engine and gearbox relevant parts were performed based on this estimation. Fig. 13 shows the distribution of the total vehicle lifetime over all gears and the assumed average speed of the vehicle in all gears is shown. Fig. 15 Frequency of gears for a diesel powered passenger car with 6 speed gearbox [5] MAIN ENGINE SPECIFICATIONS AND PACKAGING The engine specifications were fixed under consideration of the required performance, emission and noise targets. Additionally the engine weight and width (with focus on the competitors in the vehicle benchmark) and the styling and design due to the engine appearance in the motorcycle gave further inputs to decide the main dimensions. The predicted load distribution and AVL know-how supported the component dimensioning and design process. Fig. 17 gives an overview of the main engine specifications for the engine. Fig. 13 Assumed load distribution over vehicle lifecycle of a diesel powered motorcycle Fig. 14 shows the frequency of the diesel powered motorcycle in all gears, based on the assumption over the lifecycle in overall conditions (a mix of City traffic, interstate and Express way). Fig. 14 Assumed frequency of gears for the diesel powered motorcycle with 6 speed gearbox As a comparison Fig. 15 shows the frequency of a diesel powered passenger car with a 6 speed gearbox in the different driving conditions of City traffic, Interstate and Express way [5]. Main Engine Specifications Engine type 4 - stroke diesel Number of cylinders 3 Cylinder arrangement in - line Cooling water - cooled Displacement 1331 cc Bore mm Stroke mm Valves per cylinder 4 Valve train DOHC Aspiration turbocharged intercooled Turbocharger Variable Turbine Geometry Intercooler air - air Injection system common rail EGR system non - cooled, modulated Piston cooling oil - jet - cooled piston Oil cooler oil - air Fuel diesel Catalyst Diesel oxidation catalyst Engine target weight incl. gearbox 80 kg Fig. 16 Main engine specifications The decision to go for a 3 cylinder in-line engine was driven by the overall engine width, the engine weight and the better thermodynamic behavior. The cylinder displacement of 443.6cc gave a nearly square bore/stroke ratio which improves packaging situation of the engine (overall engine height).

9 The performance targets (shown in Fig. 17) were fixed under consideration of the aforementioned boundary conditions. Maximum Power max. Power Specific power Torque / engine speed max. Torque max. Engine speed max. cylinder pressure (engine operation) max. cylinder pressure (dimensioning, design) Performance Target Values Fig. 17 Performance target values 74 kw / rpm 14.8 bar 55.6 kw/l rpm 18.9 bar 5000 rpm 140 bar 150 bar The overall vehicle arrangement shows the new diesel engine packaged in the naked bike (Fig. 18). The challenge was to arrange all the unusual (for a motorcycle) engine components which are required by a diesel engine in a way that still looks like a motorcycle engine and part of the vehicle design. engine together with the intercooler in a sandwich type position. Additionally, an oil cooler is foreseen to keep the engine oil temperature at a certain level (piston cooling). The air supply while the vehicle is stationary, or running at low velocities, is ensured by a large fan behind the radiator. The turbocharger is placed directly behind the steering head. This reduces the distance between the exhaust valves and the turbine intake to use the pulse charging in addition to the continuous exhaust gas flow. The turbocharger receives direct airflow from the frontal direction but has to be shielded against the cylinder head, the cooler package, and the frame. The VTG (Variable Turbine Geometry) control is realized with an electromechanical actuator instead of a vacuum type actuator. The oil supply of the turbocharger is realized in combination with the cylinder head oil gallery. The airbox with the air filter is placed in the area above the cylinder head (shown in Fig. 19). The pressure pipe from the intercooler is running from the bottom end of the cooler up to the cylinder head into a plenum. Out of this plenum the intake runners lead the air to the intake manifold. The pipe length from the intercooler to the plenum and the length of the intake runners was calculated with BOOST to use a resonance charge effect together with the turbocharger. Fig. 19 Vehicle arrangement view top The exhaust pipe is directly connected to the turbocharger and is running downwards underneath the engine and towards the engine center. The DOC (Diesel Oxidation Catalyst) is placed under the engine and requires approximately the volume of the engine displacement (shown in Fig. 20). Fig. 18 Vehicle arrangement view left/right side The engine is water cooled and uses a standard motorcycle radiator, which is located in front of the Fig. 20 Vehicle arrangement view bottom

10 ENGINE COMPONENTS The cylinder head is designed as a typical 4V state of the art cylinder head as used in high performance HSDI diesel engines. The valves are actuated with roller cam followers to reduce friction and to increase the center distance of the camshafts because of the small valve angles (intake 2, exhaust 3 ). This allows sprockets with a sufficient diameter for the conventional chain drive valve train (shown in Fig. 21). Fig. 23 Exhaust port arrangement The camshaft (shown in Fig. 24) uses a decompression device to ensure the startability of the engine with a motorcycle electric starter. The decompression device is actuated with oil pressure and acts on all cylinders. Fig. 21 Cylinder head inner parts The intake port layout shows one tangential and one swirl port (shown in Fig. 22). Fig. 24 Intake and Exhaust Camshaft Fig. 22 Intake port arrangement The exhaust ports are standard and are directed towards the turbocharger to keep the distance between exhaust valve and turbine as short as possible (shown in Fig. 23). The common rail injection with a working pressure of 1600 bar uses solenoid type injectors and a common rail fuel injection pump. With these injectors it is possible to have two pre-injections, one main injection and 2 post injections during one injection period. The injectors are positioned facing into the center of the cylinder bore to realize the best conditions for the injection in the

11 direction of the piston recess typically used for direct injected engines (shown in Fig. 25) This allows for adaptation of the injection to each engine operating point: BOOST. Therefore valve pockets were introduced to ensure enough clearance between piston and valve in top dead center. The piston is cooled with an oil jet out of the crankcase. This jet sprays oil into the direction of a hole on the bottom face of the piston. This hole is connected to an oil passage inside the piston which surrounds the piston recess. This cooling feature is often used in HSDI engines with high power outputs. Fig. 25 Injector position The timing drive is designed using a tooth type chain (also known as silent chain). This tooth type chain is driven by the crank train via right hand side balancer shaft to keep the engine width in this area as small as possible and to allow for a standard chain guide and tensioner arrangement (shown in Fig. 26). Tooth type chains are not very common in passenger car diesel engines because of the high chain wear due to the relatively high amount of abrasive particles in the engine oil from the combustion process. The diesel motorcycle with an assumed lifetime of kilometers (low compared to passenger cars) allows to use this chain type as another measure to reduce the engine noise. Fig. 27 Piston assembly The crankshaft (shown in Fig. 27) of the diesel engine was designed in order to achieve the required torsional stiffness with a minimum weight. Since a three cylinder engine is free of 1 st and 2 nd order forces of inertia only the rotating component was balanced on the crankshaft. Additionally the usual balancing of 1 st order moments of inertia by a balancer shaft (split in two parts) and counter weights on the crank webs was introduced to improve the vibration behavior of the engine. The 2 nd order mass moments have not to be balanced in this engine category. The connecting rod ratio was set to to reduce friction by decreasing the pressure on the thrust face of the piston. Fig. 26 Timing drive The piston shows a typical shape as used for DI (Direct Injection) diesel engines (Fig. 27). The top shape of the piston is influenced by the valve overlap in top dead center. The valve overlap was optimized with AVL Fig. 28 Crank train layout The balancer shaft is split into a right hand side and a left hand side component (shown in Fig. 29). This was necessary to achieve the typical motorcycle design

12 requirements. The balance weights on two separate shafts are driven independently from the crankshaft and are aligned with the cylinder in the view from the engine side direction. To reduce the size of the balancer gears heavy metal inserts are pressed in to increase the effectiveness of the balance weights. in 3 and 4 cylinder inline motorcycle engines to decrease the overall engine width. Outside the main bearing the right hand side balance shaft with the timing drive are located. Fig. 31 Primary drive Fig. 29 Balancer shaft arrangement Primary damper between driven primary gear and outer clutch cage (using 6 springs) and clutch shows a typical motorcycle arrangement. The clutch was designed to transmit the high engine torque and is a wet type multiplate clutch using 10 plates (shown in Fig. 32). Alternator and electric starter are arranged on the left hand side of the crankcase in addition to the left hand side balance shaft (shown in Fig. 29). The starter is placed in front of the oil pan underneath the radiator and fan. This arrangement is a common solution in motorcycle engines. The alternator arrangement is prepared to add a metal ring for additional flywheel mass. Fig. 32 Primary damper and clutch Fig. 30 Alternator and starter arrangement The hub of the outer clutch cage is driving the oil pumps for the engine oil circuit and the hydraulic pump for the automated shifting and clutch actuation underneath the gearbox input shaft (shown in Fig. 33). These pumps are placed on the same shaft. The primary drive is located on the right hand side of the crankshaft (shown in Fig. 31). The toothed outer crank web is used as the driving gear. This solution is common

13 Fig. 33 Oil pump arrangement Furthermore the fuel injection pump, (sitting above the gearbox input shaft), is driven by this hub (shown in Fig. 34). The total transmission ratio from crankshaft to the fuel injection pump was kept at the value of 2. This is a typical ratio for 3 Cylinder engines and uses the pressure pulses of the radial type CR (Common Rail) pump with three radial pistons (piston offset 120 ). Fig. 35 Gearbox arrangement Cylinder block and upper gearbox housing are combined in the same part called upper crankcase concept (shown in Fig. 36). The upper crankcase is made of aluminum with cast in gray iron cylinder liners. All main shafts such as crankshaft and gearbox input and output shafts are located in the horizontal split plane between the upper crankcase and the lower ladder type bearing frame. This arrangement improves the manufacturing and assembly process of the engine. Fig. 36 Upper crankcase concept Fig. 34 Fuel injection pump drive The gearbox layout with gearbox input and output shaft is shown in Fig th and 6 th gear are shifted with one movable gear on the gearbox input shaft. 1 st and 3 rd gear together with 2 nd and 4 th gear are shifted at gearbox output shaft using two movable gears. A ladder type frame supports all bearings positioned in the above mentioned split plane. To increase the stiffness in the area of the plain bearings the ladder type frame includes cast iron inserts in the area of the crankcase bearings. This bearing beam is the functional part in the bottom end of the engine and carries the oil pumps, the shifting mechanism, part of the clutch actuation, the starter drive and mounting. It also carries

14 the lower half of the mounting flanges for the side covers (clutch cover and alternator cover). This frame has the same function as so called bed plates in passenger cars but has no outer covering surface. It is a rigid aluminum structure with cast iron inserts where all functional parts are mounted. The covering part at the bottom end of the engine is a special de-coupled oil pan. The connection to the upper crankcase is de-coupled with a special gasket. This bottom end cover is planned to be made of magnesium and has no rigid connection with the ladder type frame. This improves the noise emission of the engine noise towards the ground and is another engine integrated noise reduction measure. The overall engine arrangement with all engine components is shown in Fig. 37 and Fig. 38 Fig. 37 View of entire engine right and front Fig. 38 View of entire engine left and rear

15 GEARBOX LAYOUT AND CALCULATION As mentioned in the abstract a wide spread in the gearbox ratios is necessary to compensate for the limited engine speed range of the diesel engine and to use its torque capacity over the whole speed range. Another requirement was to keep the primary transmission ratio as small as possible to decrease the gearbox input torque. The gearbox setup was performed using AVL`s vehicle simulation software tool CRUISE. Fig. 39 shows a comparison of the transmission ratios for the diesel motorcycle with 6 speed gearboxes of 4 motorcycles on the market. Transmission ratio and spread AVL diesel MC Honda VTR1000 Suzuki TL1000 Aprilia RSV 1000 Kawasaki ZXR900 Primary st spread nd spread rd spread th spread 1, th spread th total spread secondary Overall gear rato in 6th gear Fig. 39 AVL 6 speed gearbox in comparison to benchmark vehicles The challenge in the gearbox design was to develop a motorcycle gearbox for a gearbox input torque of 240Nm. This problem could be solved by an increased center distance (75mm) between the main and counter shafts. The benchmarked vehicles showed a maximum axle base distance of 68mm and a maximum input torque of 200Nm. The increased axle base helped to keep the gearbox almost at the same width as the conventional 6 speed gearboxes of the benchmark. Together with the calculation for the load distribution of the vehicle a calculation of the maximum transmittable torque in all gears was performed. The torque in the gearbox was calculated from the engine side and from the rear wheel side under consideration of max. wheel load and a maximum friction value between rear tire and road (see Fig. 40). The load distribution for the gearbox layout is shown in Fig. 13 and the lifetime was considered as described in the foregoing chapters. Because of the high torque provided by the diesel engine a module of 2.75 was selected for all gears except for the fifth with a module of 2.5, to be able to keep the desired center distance. COMPARISON OF DIFFERENT TOOTHING TYPES SHOWN IN 1 ST AND 2 ND GEAR Furthermore the emitted noise from the gearbox was a factor which had to be taken into consideration. For a low noise level a helical gear set is the optimal solution, but considering the boundary conditions in a motorcycle gearbox with limited width and center distance it was only possible for first and second gear. The reason for this was that helical gears require selector sleeves, which need more room in axial direction than sliding gears. Therefore helical gears for all gear sets would have had a negative effect on the width of the gearbox. For the movable gears it was decided to use increased addendum and dedendum reference profile to get a transverse contact ratio of at least 2 to reach a sufficient noise behavior. Nevertheless for first and second gear which have fixed mating gears, helical gearing can be used. In order to check if helical gearing of the 1 st and 2 nd gear will lead to the expected benefits a study comparing three versions in terms of noise emission was carried out. The different versions were calculated using the software package KISSsoft. The KISSsoft calculation considers the influence of the gear geometry on contact stiffness and noise behavior. The basis for the calculation by KISSsoft is single tooth contact. Additional effects due to contact ratios higher than 2.0 were considered using AVL s knowledge and experience. Calculated versions of 1 st and 2 nd gear: 1. Spur Gearing 2. Gear set with increased tooth addendum 3. Helical Gearing For the comparison of the three versions the default values of the weighting for the different parameters implemented in KISSsoft were used [7]. Fig. 40 Calculated torque value at clutch In addition to the three versions mentioned above another calculation was carried out to show the influence of tooth quality on noise emission. Because of the manufacturing process no better quality (tooth quality according DIN 3916) than 7 to 8 should be used for

16 mass production. Nevertheless higher qualities show a certain improvement in noise emission. Therefore in low volume production or for high quality products this possibility should be taken into account. Higher qualities also mean that the tolerance limits get smaller, and unfavorable combinations of tolerances, profile and tooth trace error do not lead to much higher noise. The calculation results for the first and second gear can be summarized as follows: 1. Spur Gearing is the conventional method for gearing in motorcycle gearboxes. Noise emission is high compared to deep tooth gearing and helical gearing, but tooth strength and scoring is better than with deep tooth form. Additionally standard gearing tools can be used for production. 2. Deep Tooth Gearing: The aim when using Deep Tooth Gearing is similar to the helical gear; a significant reduction in noise emission due to the softer mesh of the tooth. To achieve that goal a transverse contact ratio of two or higher is required. Good results were achieved with an addendum of 1.27 x module (originally 1 x module) and a dedendum of 1.6 x module (originally 1.25 x module) of the gear. Both values had to be adjusted in such a way that the bottom clearance is only reduced down to the permitted limit. The key for a good contact ratio is a high value for the dedendum, to which the addendum has to be adjusted properly. To obtain a good mesh the bottom clearance of the basic rack profile had to be slightly reduced and profile modifications like changing the fillet radius of the basic rack had to be done. Those adjustments have to be made for each step and gear separately and therefore it is not described any further here. As shown in Fig. 41 and Fig. 42 deep tooth gearing leads to a smaller gradient of tooth-stiffness due to reduced tooth thickness compared to spur gearing which results in a noise reduction. Another point necessary to mention is that when using deep tooth gearing a certain amount of tooth profile modification has to be done. That means tip and foot relief have to be conducted for each gear set individually [11]. The limiting factor for the deep tooth gearing is the tooth thickness on the tip cylinder, strength against scoring and the root strength. In order to achieve a positive effect on the noise matter and to reach a contact ratio higher than 2.0 the number of teeth on the pinion gear has to be well above the allowed minimum number of teeth (20 according to KISSsoft). Fig. 41 Spur gearing (standard addendum reference profile) Fig. 42 Deep tooth gearing 3. Helical Gearing: Helical gears require a minimum overlap ratio of at least one to have a significant improvement in noise emission. The calculation showed that values of the overlap ratio close to one can not be achieved due to the chosen face width of the gears, the limited distance between the two gear shafts and due to the helix angle which was set to 15 degrees to limit the axial forces. Nevertheless even with lower values good noise results were reached (see Fig. 43 and Fig. 44) and the values for the safety of the root stress are satisfying. Fig. 43 Noise first gear (0...bad, 6...very good)

17 control valves is planned. Alternatively a fine sieve could be used. Further components which need to be installed for this concept are a pressure reservoir, a pressure sensor, the control valves (these can be combined in one unit) and a potentiometer on the selector barrel which indicates its momentary position to the electronic control unit. When using semi-automatic shifting an automated clutch is also required too. Fig. 44 Noise second gear (0...bad, 6...very good) Additionally a combination of helical gearing and deep tooth gearing was calculated, but the safety against scoring was too low. All three versions have a chamfered tip edge due to manufacturing reasons (gears are bulk material). To avoid extra wear and noise caused by errors in the pitch straight transmission ratios were avoided. Based on the gearbox architecture, the optimization regarding torque capacity and noise behavior as well as production requirements and production cost (and the findings of the comparison of the three versions), the version with helical gearing was chosen for the first and second gear because of the good compromise between noise emission and at the same time acceptable safety in terms of root strength. For 3 rd to 6 th gear spur gearing with increased addendum was used. SEMI - AUTOMATIC GEAR SHIFTING In different fields of vehicle applications a strong tendency towards automatic and automated gearboxes can be detected. Especially in the upper class passenger car segment automatic gearboxes win recognition Meanwhile for more sportive cars semi-automatic gearboxes based on sequential gearboxes become more and more fashionable. In the field of 2 wheelers CVT drives are common for 50 cc scooters. During recent years a shift of opinion towards using automatic gearboxes for bigger displacements could be observed. Evidence for this is the introduction of scooters with a displacement up to 600 cc which use CVT drives. For motorcycles there are only some prototype concepts which have not been put into series production yet. A motorcycle gearbox works sequentially, therefore no new gearbox has to be constructed, only the placement of the new components has to be done. To keep both gear shifting possibilities (manual and semi automatic), the selector barrel, the gear index and the whole shifting mechanism of a conventional motorcycle gearbox were simply adapted to the new system. One of these modifications is the gearshift pattern. In a standard six speed gearbox, neutral is between first and second gear which leads to a different shift travel for shifting to neutral gear compared to shifting the remaining gears. For the automated gearbox this was modified. Neutral is now a gear of its own and is situated below the first gear since it is easier for the actuation system to deal with same shift travels for all gears. For manual shifting the rider of the bike has to be aware of the changed gearshift pattern. Following the operating principle of the hydraulic system is described: A hydraulic piston can have pressure applied on either side of the piston depending whether an up-shift or a down-shift is intended (shown in Fig. 45). The selector shaft, which engages in the middle of the piston translates the motion in a rotation of the shifting shaft, which acts as in a conventional gearbox upon the selector barrel. A spring resets the whole system to the initial position. The system is ready for the next gear shift. Other possibilities like a gear rack that acts upon a pinion mounted on the selector barrel were also taken into consideration but would require a rather long gear rack for a six speed gearbox and the combination of manual and automated shifting could not be easily preserved. The objective for the gearbox of the diesel motorcycle was to provide comfortable shifting without pulling the clutch lever, just by pressing a button on the handle bar. There are different systems for automation offered on the market such as electrical, electro-mechanical and electro-hydraulic systems. The electro-hydraulic system fits best in a motorcycle concept, because the extra oil pump can easily be positioned on the same shaft as the oil pump for engine lubrication. Additionally quite clean oil is provided since the oil circulates through an oil filter continuously. To obtain the quality needed for the actuator valves the use of a second filter just before the Fig. 45 Automated gear selector mechanism

18 As mentioned above a hydraulic clutch is required for this concept. The hydraulic clutch also shows the dual actuating feature (It can either be actuated by pulling the lever on the handle bar or it works automatically when a gear shift activated by one of the two buttons on the handle bar is in progress). The management of the actuating valves is done by the electronic control unit. Additionally an anti-hopping mechanism to ensure the driving safety could be realised mechanically (as state of the art in high performance motorcycles) or hydraulically with the clutch actuation mechanism. In addition to the added comfort the automated transmission also has a good effect on the wear of all transmission parts while fast and sportive shifting can be obtained. COMBUSTION AND THERMODYNAMIC SIMULATION Thermodynamic simulations were performed in order to optimise all parameters relevant to the engine s gas exchange. The code used was AVL BOOST, a state of the art gas exchange and cycle simulation program. AVL BOOST simulates the complete engine cycle including the gas exchange considering the 1D-gas dynamics in the intake and exhaust system. It can be applied to the investigation of a wide variety of problems including those of intake and exhaust orifice acoustics or exhaust gas after-treatment. In addition to the gas exchange layout of street legal engines, BOOST is also applied to the optimisation of high performance racing engines. [8], [12] The figure Fig. 46 shows the simulation model for the 3- cylinder DI TCI motorcycle engine. The complete intake and exhaust system was considered. As the main interest of the investigation was focused on the gas exchange the exhaust muffler was modelled by a simple volume. The gas exchange simulations aimed at the layout of the charging system, the valve timing and the manifold geometry. These tasks were performed at full load, when no EGR is added to the aspirated air. Therefore the EGR system was not considered and is not included in the simulation model. CL Air cleaner TC (Turbocharger) CO1 Intercooler C1, C2, C3 Cylinders CAT1 Catalyst PL1 Muffler 1 18 Pipes and Manifolds MP1 MP9 Measuring points J1 J4 Junctions R1 Restriction SB1, SB2, SB3 System Boundary Fig. 46 Gas exchange simulation model (BOOST) Among the parameters investigated were the valve timing and the intake and exhaust system geometry. In addition a turbocharger was matched to the engine. Especially for the turbocharger, the increased speed range (1000 rpm to 5000 rpm, 4500 rpm rated speed) made the balancing between high and low speed requirements more difficult. In the following some aspects in the layout of the engine s gas exchange system are explained in more detail. COMBUSTION SYSTEM The engine will be equipped with a common rail fuel injection system with multiple injection capabilities. This is mandatory when using a low compression ratio to ensure startability at low ambient temperatures without additional measures. In addition pilot injection has to be used at part load and low speed full load operation to reduce combustion induced noise. Minimising the latter at its source is important as the installation of the engine in a motorcycle (especially without fairing) does not allow the use of acoustic shielding for aesthetic reasons. Current emission targets do not necessitate the application of a variable swirl system or cooled EGR. This simplifies the overall arrangement and facilitates packaging.

19 VALVE TIMING Compared to other HSDI engines intake valve closing was slightly retarded in order to improve airflow at rated speed. Due to the three cylinder configuration exhaust valve opening could be chosen for best fuel consumption without compromising the air flow through the engine (by the impact of the blow down pulse on the gas exchange of the cylinder following in the firing order). Similarly valve overlap period can be made long and is only limited by piston to valve clearance. Therefore a calculation together with the piston design and the fixing of the deck height was performed. The results of the optimised valve lifts in relation to the piston movement is shown in Fig. 47 and Fig. 48. TURBOCHARGER MATCH: The engine will be equipped with a turbocharger with variable turbine geometry. The wide spread between maximum and minimum speed requires a wide compressor map. In order to lower boost pressure requirements at rated speed an intake system with tuned intake runners was considered. Fig. 47 Exhaust valve piston movement Fig. 50 Compressor Map with Engine Full Load Operating Points Fig. 48 Intake valve piston movement The optimized valve lift curves for intake and exhaust are shown in Fig. 49. Compared to a neutral system air delivery ratio can be increased significantly at the expense of increased pumping work (see Fig. 51). Fig. 49 Optimized Valve lift curves Fig. 51 Predicted Air-Delivery Ratio (related to intake manifold conditions) and Pumping Mean Effective Pressure

20 Keeping the engine output to the target value of 74 kw, the increased air delivery ratio cannot compensate for the increased pumping work. The increased pumping work even deteriorates fuel consumption and leads to higher required air flow rates thus making compressor match more difficult. Because of this and in view of packaging requirements the neutral intake system was chosen for the engine. The compressor map together with the engine s operating points at full load is shown in Fig. 50. FULL LOAD CHARACTERISTICS The target full load characteristics with a maximum torque of 200 Nm available between 2000 and 3500 rpm and a maximum power of 74 kw (100 HP) available between 4000 and 4500 rpm can be reached with a relatively low peak firing pressure of 140 bar. CALCULATION OF THE FLYWHEEL MASS The size of the Flywheel is always a compromise between large flywheel mass for good comfort and small flywheel mass for quick response dynamic engine behavior. Flywheel mass on diesel engines with a small number of cylinders is required for the damping of the cyclic speed irregularity caused by the enormous impacts of the high PFP (Peak Firing Pressure) combined with the long firing intervals especially at low engine speeds. The degree of cyclic speed irregularity is an indicator of the running smoothness of the engine. This parameter is defined as follows: Max. AngularVelocity Min. AngularVelocity CyclicSpeedIrregularity = MeanAngularVelocity Generally, for passenger car diesel engines, the recommended acceptance limit is 0.25 around the low end torque speed to ensure a certain comfort level. The AVL software code BRICKS was used to predict the influences caused by the reduced flywheel mass for the diesel motorcycle engine. The simulation code BRICKS can also be used for the calculation of crankshaft strength, the bearing calculation and the analysis of torsional vibrations. [10] To simulate the influence of the flywheel size a multibody system of the crankshaft was created (shown in Fig. 54). The cylinder pressure characteristics in full load conditions over the whole engine speed range were implemented using the BOOST results. Fig. 52 Predicted Engine Full Load Performance The above shown full load performance can be reached with very attractive fuel consumption (compare Fig. 53 below). Fig. 53 Predicted Brake Specific Fuel Consumption at Full Load Fig. 54 BRICKS model of the 3 cylinder crank train The simulation run was performed with different flywheel masses. For the base calculation the original setup without additional flywheel masses as shown in Fig. 55

21 was performed. The total crankshaft inertia for the related parts was calculated with kgm². to the requirement of full load and PFP (Peak Firing Pressure) starting at 1000 rpm. With the improved motorcycle engine layout it is possible to achieve the same speed irregularity (target value of passenger cars) at 2000 rpm. Fig. 55 Crankshaft configuration without flywheel mass To show the influence of the flywheel mass on the speed irregularity a steel ring was simulated which is pressed onto the alternator. This steel ring increased the total crankshaft inertia from kgm² to kgm². The position of the steel ring is shown in Fig. 56. Fig. 57 Comparison of cyclic speed irregularity of different vehicles and Flywheel masses To compare the cyclic speed irregularity of diesel and gasoline engines, the same model as shown in Fig. 55 was calculated using a gasoline cylinder pressure curve. The calculation was performed starting at an engine speed of 3000 rpm due to the different performance characteristics of gasoline engines compared to diesel engines. The result in Fig. 58 shows that the speed irregularity is strongly dependent of the PFP which is between 60 and 70 bar in gasoline engines over the shown speed range and 140 bar in turbocharged diesel engines. Fig. 56 Crankshaft configuration with additional flywheel mass For comparison, the motorcycle results are shown together with the speed irregularity of a 3 cylinder diesel passenger car engine. The value of the total crankshaft inertia including flywheel is kgm² in the passenger car engine and the maximum PFP (Peak Firing Pressure) is 150 bar. The comparison of the simulation results as shown in Fig. 57, indicates that the difference in flywheel mass has significant influences on the cyclic speed irregularity. Without any additional inertia in the crankshaft it is impossible to achieve acceptable passenger car target values at low engine speeds. These target values are based on experience and consider comfort and component durability of engine and gearbox. The passenger car configuration in Fig. 57 shows a maximum value of 0.27 at 1000 rpm and maximum engine load. The flywheel mass was designed according Fig. 58 Comparison of cyclic speed irregularity of gasoline and diesel There are two possibilities to solve the problem concerning speed irregularity for the diesel motorcycle. One possibility is to reduce the PFP below 2000 rpm to reduce the impact out of the combustion. The second possibility is to add more inertia mass on the crankshaft. The balance between these measures has to be found during the prototype development.

22 VEHICLE SIMULATION The software AVL CRUISE was used to predict the vehicle performance. CRUISE is developed for vehicle simulation. The most typical tasks are the calculation of driving performance, fuel consumption, and emissions [9]. Fig. 59 shows the vehicle model of the diesel motorcycle. This model contains the components as in the real vehicle with all relevant inputs simulated. The engine component includes the engine performance curve from the BOOST calculation. Furthermore measured emission and fuel consumption maps from a comparable engine were adapted to simulate the emission and fuel consumption of the vehicle in the new motorcycle driving cycle. Fig. 59 CRUISE vehicle model The main vehicle specifications shown in Fig. 60 are a combination of values out of the vehicle benchmark, AVL experience, assumptions and measurements. Main Vehicle Specifications Curb weight 1) 240 kg Gross weight 1) 400 kg Riders weight 75 kg Tank volume 1) 21 litres Wheel base 1) 1510 mm Rear tire dimensions 1) 180/55-ZR17 Static rolling radius 314 mm Dynamic rolling radius 318 mm Rolling resistance Velocity dependent Frontal area 2) 0.72 m² Drag resistance 2) 0.6 1) Average value of benchmark vehicles 2) Motorrad Magazine [6] Fig. 60 Main vehicle specifications FULL LOAD ACCELERATION In this task the acceleration from vehicle stand still to the maximum vehicle speed (220 km/h) (shown in Fig. 61) and two other typical scenarios were calculated. The shifting time using the automated shifting mode was set to 0.15 seconds and the time of clutch release at the vehicle start was set to 0.15 seconds. Fig. 61 Full load acceleration from vehicle stand still One of the main criterion to compare the performance of vehicles is the acceleration time for 0 100km/h. The calculated acceleration time for the diesel motorcycle is below 3 seconds. A comparison with the benchmark vehicles is shown in Fig. 62. Vehicle comparison Acceleration time km/h Suzuki GSX Kawasaki ZRX 1200R 3.0 Yamaha XJR Honda CB AVL diesel motorcycle 2.96 Fig. 62 Full load Acceleration 0 100km/h The next task (Fig. 63) shows the acceleration from km/h Vehicle comparison Acceleration time km/h Suzuki GSX Kawasaki ZRX 1200R 5.1 Yamaha XJR Honda CB AVL diesel motorcycle 5.3 Fig. 63 Full load Acceleration 0 140km/h ROLL ON IN TOP GEAR Another way to compare the acceleration performance of vehicles is the roll on in top gear. These values are measured between 60 and 100 km/h (shown in Fig. 64) and 100 to 140 km/h using the highest gear. For the diesel motorcycle the 6 th gear was selected to show the excellent roll on behavior because of the high low end torque. The engine speed at 60 km/h was measured at 1290 rpm and at 100 km/h it was 2100 rpm. The torque at 1290 rpm is about 150 Nm and therefore the acceleration value even using the 6 th gear at 60 km/h is high enough to accelerate the vehicle to 100 km/h in 3.3 seconds.

23 Vehicle comparison Acceleration time km/h Suzuki GSX Kawasaki ZRX 1200R 3.9 Yamaha XJR Honda CB AVL diesel motorcycle 3.3 Fig. 64 Top gear roll on km/h The engine speed at 100 km/h was measured at 2100 rpm and at 140 km/h at 3020 rpm. This means that during this acceleration period the full engine torque is available. Vehicle comparison Acceleration time km/h Suzuki GSX Kawasaki ZRX 1200R 3.1 Yamaha XJR Honda CB AVL diesel motorcycle 3.2 Fig. 65 Top gear roll on km/h MAXIMUM TRACTION FORCE The maximum traction force of the vehicle in all gears together with the vehicle resistance force and the gear diagram is shown in Fig. 66. The maximum traction force in 1 st gear is 4200 N. Vehicle comparison Suzuki GSX 1400 Kawasaki ZRX 1200R Yamaha XJR 1300 Honda CB 1300 AVL diesel motorcycle Fig. 67 Maximum vehicle velocity Top speed 225 km/h 230 km/h 213 km/h 230 km/h 220 km/h Together with the excellent vehicle acceleration, the results of top gear roll on and the comparable vehicle top speed to the benchmarks shows, that this high performance diesel motorcycle would be competitive on the naked bike sector. The enormous traction force due to the high torque value of the engine means that the vehicle performance is hardly influenced by additional payload. EMISSIONS AND FUEL CONSUMPTION IN THE PROPOSED NEW MOTORCYCLE DRIVING CYCLE As already mentioned the Simulation software CRUISE allows to simulate the vehicle running in a driving cycle. To predict fuel consumption and the emissions of the diesel motorcycle along the new proposed driving cycle for motorcycles (as mentioned in the emission section of the paper and shown in Fig. 4), a computer simulation was created in CRUISE (shown in Fig. 68). Fig. 68 Proposed future motorcycle driving cycle created in CRUISE The following maps show measured emissions and the fuel consumption of an engine from a diesel passenger car. This maps were modified according to the changes in the engine layout and emission control features of the diesel motorcycle engine. The maps represent: Fig. 66 Maximum traction force in all gears The intersection of the line 6 th gear traction force and the vehicle resistance line shows the maximum velocity of the vehicle. The rider is in upright sitting position and the maximum velocity is 220 km/h. A comparison in Fig. 67 shows that this max. velocity value is in the area of the benchmark vehicles. BSFC Brake Specific fuel consumption in [g/kwh] shown in Fig. 69 CO Emissions in [g/h] shown in Fig. 70 HC Emissions in [g/h] shown in Fig. 71 NOx Emissions in [g/h] shown in Fig. 72 SOOT Emissions in [g/h] shown in Fig. 73

24 Fig. 69 BSFC in [g/kwh] (Brake Specific Fuel Consumption) Fig. 72 NOx Emission map in [g/h] Fig. 70 CO Emission map in [g/h] Fig. 73 SOOT Emission map in [g/h] The result of the diesel motorcycle in the proposed motorcycle driving cycle are shown in relation to the proposed future emission limits for diesel motorcycles. Comparison of calculated emission values with proposed future emission limits for diesel motorcycles Limits EU III CRUISE results TA: Valid from ANR: Measuring mode 6 x UDC + 1x EUDC 6 x UDC + 1x EUDC CO [g/km] HC [g/km] NOx [g/km] Fig. 71 HC Emission map in [g/h] PM [g/km] Fig. 74 Comparison of CRUISE simulation with proposed future emission limits