Application of CAE Methods to Increase Efficiency in Powertrain Development

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Application of CAE Methods to Increase Efficiency in Powertrain Development

Sean Biggs Technical Specialist, Jaguar Land Rover Ltd

Dr Antonio Turturro MBS Engineer, Jaguar Land Rover Ltd

Salim Chaker Engineering, ITI GmbH, Dresden

KurzfassungDie Konzipierung eines raffinierten Fahrzeugs erfordert detaillierte Kenntnisse der Antriebsstrangeigenfrequenzen und des vom Verbrennungsmotor geleisteten Erre-gungsgrads. Um einen Kompromiss bei der Beseitigung von Geräusch hörbaren bzw. Vibration spürbaren Schwingungen (NVH: Noise, Vibration and Harshness) im Kraft-fahrzeug und bei der Verbesserung von Leistungseigenschaften vermeiden zu kön-nen, soll in früherer Designphase eine tiefgründige Planung der Modenplazierung und der Steuerung der Erregerkräfte erfolgen.

Da der Antriebsstrang eine Reihe von Wechselwirkungen mit den restlichen Fahr-zeugaggregaten aufweist, ist eine fundierte Systemanalyse unabdingbar. Der vorlie-gende Beitrag beschreibt den durch Jaguar Land Rover (JLR), ZF und ITI GmbH ent-wickelten Prozess zur Vermeidung von Iterationsschleifen in der Entwicklungsphase.

AbstractThe delivery of a refined vehicle requires a detailed knowledge of the powertrain na-tural frequencies and the level of forcing applied by the engine. Modal placement and the management of the forcing levels have to be planned and executed early in the design phase, if compromises in the NVH (Noise, Vibration and Harshness) and other attribute performance are to be avoided.

The Powertrain as installed has a number of interactions with the vehicle which re-quires a system level analysis to be carried out. This paper will describe the process jointly being developed by Jaguar Land Rover (JLR), ZF and ITI GmbH to avoid late changes in the development process.

Introduction

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CAE MEthods to InCrEAsE EffICIEnCy In PowErtrAIn dEvEloPMEnt

Engineering a luxury vehicle to have acceptable levels of refinement, delivered in an efficient manner is a technical challenge. To do this requires a knowledge of the na-tural frequency distribution in a mechanical system and how these frequencies may change during operation of the vehicle. The potential for modal interaction, where one mode magnifies the forcing levels for a second mode (with an insufficient fre-quency separation) must be avoided and thus requires the development of a modal placement strategy.

The forced response of the system and the levels of forcing change during the ope-ration of the vehicle are important factors. The torsional forcing from an internal combustion engine is comprised of two significant components in opposition to each other. The first is the torque component generated by the cylinder pressure acting through the crank slider mechanism. This can be very non-linear as shown in Figure 1 (l) below.

Figure 1: (l): Full Load Cylinder Pressure Map vs. Crank Speed; (r): Single Cylinder Inertia Torque vs. Crank Angle

The second torsional forcing component is the torque required to overcome the in-ertia of the reciprocating pistons. This is known as the inertia torque and it changes magnitude and direction as the crankshaft rotates. Figure 1 (r) shows the torque for a single cylinder engine being motored at various speeds. Note from Figure 1 (r) that the instantaneous torque values are comparable to the mean torque values of a typical automotive engine and that the total inertia torque for the engine would require the contribution from each cylinder to be added with the correct phase re-lationship. Taking the instantaneous peak values of Inertia torque from Figure 1 (r) for each speed shows that the peak torques increase at the square of the crankshaft speed. Piston mass is also a very important parameter especially for high speed diesel engines which require a lightweight design piston in order to generate the minimum inertia force but is robust enough to withstand the resultant forces during the engine work-cycle. The combustion and inertia torques continually interact with the relative influence of the components dependant on the speed and load on the engine. The torsional forcing during a full load acceleration of the vehicle will initially be domi-nated by the combustion torque but as the engine speed builds the inertia torque increases until it becomes dominant as the combustion torque reduces. If the driver lets the vehicle decelerate using the engine to brake the vehicle then the inertia tor-

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que will be dominate as the engine fuelling will only ensure that it continues to run, significantly reducing the combustion torque.

The torsional output from the engine acts on the torque converter and damper system before acting on the automatic transmission. The torque converter (TC) is a fluid cou-pling used to connect an internal combustion engine to an automatic transmission. One of these features is the introduction of damping into the torsional system, but with this comes a significant loss of efficiency. Running the TC in unlocked mode is minimised and an additional damper system using springs is added to attenuate the engine cyclic input to the transmission. The attenuation of the torque converter damper under varying speed and load conditions is very non-linear and when com-bined with the combustion engine forcing the system is complex. A typical single stage damper system with the drive plate (see Figure 2 (l)) consists of a number of springs installed at a fixed radius of the damper which are compressed by the cyclic torque generated by the engine. The spring coils are also subjected to centrifugal forces generated from the speed of rotation and significant friction forces results at the contact points with the housing. If these friction forces are higher than the cyclic forces generated by the engine then the individual coils will stop moving and change the attenuation performance of the damper. TC Damper designs can also consist of more than one set of springs which act in series to give an overall rate. The stiffness characteristics for a two stage damper are shown in Figure 2 (r). The stiffness and tra-vel characteristics are different for the drive and over run condition further increasing the complexity of the system.

Figure 2: TC Damper: (l) Typical single stage TC Damper; (r) Dual Stage TC Damper Characteristics

The TC damper is connected to the input shaft of an 8 speed automatic transmission supplied by ZF as detailed in reference [1]. The output from the transmission is cou-pled to a conventional rear wheel drive automotive driveline.

Model DescriptionIn an attempt to accelerate the development process, JLR, ZF and ITI are collaborating in the development of a complete Powertrain torsional system modelled using Simu-lationX. JLR provide the engine and driveline elements of the model with ZF providing the damper and transmission elements. For the purpose of profound system analysis

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and for determining the system vibration behavior all these components were mode-led by ITI in the modeling and simulation environment SimulationX. The engine model (Figure 3) consists of a rigid crankshaft coupled to number of crank mechanisms. Each crank mechanism acts on a mass element representing the reciprocating mass of each cylinder. The inertia torque contribution is adequately captured using this method. The combustion torque is captured using maps of cylinder pressure for the load cases of interest.

Figure 3: Four cylinder engine SimulationX-model generated by JLR

The cylinder pressure maps are generated using a combustion simulation code and then as prototypes are developed and the engine calibration is refined. These are re-placed with test cylinder pressure maps. The SimulationX model has been compared with a similar model generated in the MBS code MSC/ADAMS with no significant differences. To perform a whole system analysis this component will be integrated into complete drive line model.

The next element in the system is the torque converter damper. JLR uses various torsional damper models. One of the torsional damper versions is given as FMU (“Black-Box”), which was provided by the manufacturer ZF and adjusted according to the provided torsional damper component. The modelling principle is to capture the response of each active coil in the damper stage, under the influence of the cyclic torque input and the friction forces acting on it. The modelling method is similar to that detailed in Ref [2].

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Figure 4: Torque Converter Damper FMU Black Box generated by ZF

In order to carry out fundamental investigations and parametric studies regarding attenuation power, several models (with different spring mass configurations) of the torsional damper have been created by ITI as open structure. The Figure 5 shows the hierarchical structure of these models. The internal structure consists primarily of BowSpring elements. In accordance with Ref [3], each set of bowed springs is mo-delled as a purely rotary multi-mass system with a two-stages elastic characteristic and discretized in 8 spring segments and 7 masses. Their friction interaction with the bearing shells is modelled in discrete points considering the dependency on the radial component of the springs’ elastic reaction, and the centrifugal force and including constructional end stops and gaps. The BowSpring element represents a set of bowed springs acting in a torsional damper and disposed at the same mean radius.

Figure 5: BowSpring element – icon (connectors) and internal structure

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Analogous to the 9HP Gearbox-model presented in Ref [4] an 8-speed automatic transmission model was created by ITI (see Figure 6). This model captures the change in power flow in each of the gears. The desired engaged gear or sequence of gears can be assigned to a controller model, which implements all the information needed for the 8HP gearbox shift logic and sends proper actuation signals to all clutches, brakes and claws of the gearbox structure. The element type discClutch models Clutch and brakes. It has a signal input to get the actuation signal generated by the controller and two 1D rotational connectors for exchanging kinetic and kinematic information with the gearbox model structure.

Figure 6: 8 Speed Automatic Transmission generated by ZF

The driveline system (Figure 7) consists of the propeller shaft connected to a final drive unit, which drives through two half shafts to a simple dynamometer to control the system. The flexible shafts are discretised to ensure that the modal content of the system is captured and potentially excited by the cyclic torques from the engine model. The complete system can be run at a fixed speed or in a transient condition for either run up or over run conditions.

Figure 7: Driveline (RWD) and Dynamometer System generated by JLR

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Analysis ResultsUsing the presented models both transitory and steady state simulations can be performed. Figure 8 shows the speeds reached during a complete gearshift sequence at different components along a driveline model.

Figure 8: engine speed (no markers), gearbox input shaft speed (squared markers) and gearbox output shaft speed (triangular markers) over a run up

The consideration of the rotary backlash in the gear meshings allow to highlight possi-ble undesired vibrational phenomena in the transmission like the rattling or the clonk. Running a transient simulation the forces generated at the tooth contact represent the main physical quantities to be observed: The model makes this information available at both sides of the engaged teeth of a meshing. The clonk can be highlighted by nor-mal forces arising alternatively at opposite flanks of the teeth when the transmitted torque changes its sign.

JLR have tested a 2WD Jaguar powertrain with a ZF 8-speed gearbox (see Figure 6) with a number of TC dampers of various constructions. A sample of the correlation being achieved can be seen in Figure 9 (l) which details the crankshaft velocity fluc-tuations for an engine running at a steady state condition. Every mechanical system has a number of natural frequencies which magnify any applied forcing at resonance. One important mode of interest is that of the propshaft torsional mode. When this system is at resonance it results in an increase in the cyclic torque fluctuations reacted by the final drive assembly, mounted to a very sensitive part of the vehicle body via rubber bushings. If these torque fluctuations are excessive this can cause a booming noise in the vehicle cabin. The vehicle under test showed a very significant difference in the propshaft boom response between full load and no load conditions, at the speed at which the engine forcing was close to the propshaft resonance frequency. The interior noise increased when the load on the engine was suddenly reduced. The change from full load to no load condition resulted in a reduction in the peak cylinder pressure by 75% together with a change in shape as can be seen in Figure 9 (r).

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Figure 9: (l): Engine Primary Side Correlation (Steady State Condition); (r): Full and No Load Cylinder Pressures

Another factor that influenced the difference in the boom response was that the en-gine speed at propshaft resonance also corresponded to the minimum cyclic torque at full load. As stated previously this is the speed at which the inertia torque starts to dominate the combustion torque as seen in Figure 10 (l) and labelled as the torque transition point.

Figure 10: TC Damper Full Load Torque vs Crankshaft RPM; (r): TC Damper Torque for Full and No Load

The forcing for the no load condition was investigated by changing the cylinder pres-sure maps to the no load condition, spinning the powertrain system at high speed and allowing the vehicle dyno to decelerate against the engine resistance. The resulting torque was recorded at the TC Damper and can be seen in Figure 10 (r). Note that without the influence of the combustion torque the inertia torque is significantly gre-ater than the full load torque seen at propshaft resonance/full load torque transition. When analysing the influence of the different forcing mechanisms on the crankshaft cyclic velocity (at the propshaft resonance speed) we can see that the amplitude for the no load (over run) condition is higher together with a change in the shape and hence the frequency content. It should be noted that some powertrain systems exhi-bit a change in propshaft resonance frequency with selected gear and with modern transmissions offering 8 to 9 gears with potentially the addition of a selectable 2/4 wheel drive system. This increases the opportunity to miss an error states during te-

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sting. There are other torsional modes of the system influenced by engine and TC damper parameters which also need to be considered during development. The use of SimulationX allows JLR and ZF engineers to test the robustness of the system design and to identify potential issues early in the development phase.

Summary and outlookWe presented a new vibration model for the dynamic simulation of complete drive line including a combustion engine, torsional damper and an 8-speed automatic transmission. In addition to the detailed nonlinear damper model consisting of multi-ple masses and nonlinear elastic and friction properties, the transmission model takes into account the gear contact stiffnesses depending on rotational angles, the profile shift and the contact ratio. The model is valid for all gears and can simulate gearshifts. It is suitable for the analysis and prevention of combined forced and parametrically excited vibrations, which are typical for gears. The physical modeling enables a si-mulation which takes into account the interactions with the neighboring oscillatory components of the drive system. Thus the used modeling approaches promote the essential understanding of the inner relationships of the system and differ significantly from traditional models with regard to adaptability, flexibility and adjustability of the properties to be considered.

The presented simulation model can be used for both, dimensioning tasks as well as diagnosis. It is applicable to a vast variety of tasks and applications in the transmission development process ranging from concept design to virtual testing, such as NVH assessment of the powertrain. This will be increasingly important as the automotive industry strive for more efficiency. The trend to downsize the capacity of the engi-ne is leading to boosted cylinder pressures in fewer cylinders, this is also leading to stratagies which de-activate cylinders when the engine load levels allow. All of the previously mentioned factors will significantly impact the system performance. The application of electric motors will also continue with the challenge of combining the two power sources in a refined manner. The opportunity to examine the system perfor-mance virtually using SimulationX will be vital to deliver a refined and robust system.

References[1] ZF 8 Speed Transmission Product information. http://www.zf.com/corpo-

rate/en/products/innovations/8hp_automatic_transmissions/8hp_automa-tic_transmission.html

[2] Converter Damper Attenuation Performance using Multi-Body Systems Ana-lysis. Sean Biggs, NAFEMS World Congress. Salzburg 2013

[3] Meingaßner, G. J.: Entwicklung von Simulationsmodellen zur Abbildung der Hyste-reseeffekte in der Dämpfereinheit und des Drehzahlschlupfes in der Wandlerüber-brückungskupplung von Pkw-Drehmomentwandlern. Diplomarbeit. Landshut, 2010

[4] V. Langella, U. Schreiber, H. Bertels: A New Physical Simulation Model for the 9HP Gearbox Incorporating a Nonlinear Twin Torsional Damper and Backlashes. In: VDI-Berichte Nr.2158: Getriebe in Fahrzeugen 2012, VDI-Verlag. Düsseldorf, 2012