Elements of a Mechatronic Vehicle Corner

Copyright @ IF AC Mechatronic Systems, Darmstadt, Germany, 2000

ELEMENTS OF A MECHA TRONIC VEHICLE CORNER

Steffen Gruber, Martin Semsch, Thomas Strothjohann, Bert Breuer

Darmstadt Uni versity of Technology, Automoti ve Engin eering Department, Petersenstrasse 30, D-642B7 Darmstadt, Germ any

As partne r in So nderforschungs bere ich 24 1 .. New Integrated Mec ha nic-Electro nic So luti ons fo r Mechani cal Engineerin g Sys te ms" of Deut sc he Forsc hungsgeme in sc ha ft at Darmstadt Uni ve rsity of Technology th e A uto mo ti ve E ngin ee rin g Departme nt (fzd ) is do in g research in its 3 SFB-projects : Tire with an Integrated Se nsor (A3 ), Adapti ve Whee l S uspen sion (A 7) and Electri call y Actuated Whee l Brake (B6). In thi s paper researc h targets, workin g princ ipl es. prototype perfo rma nces a nd co mpone nts' contri buti o ns to a smart interlinked so called mec hatro ni c ve hicle corner will be desc ri bed . Copyrigh t © 2000 IFAC Keyword s: acti ve brake contro l, ac ti ve ve hicle suspe ns io n, brakes . e lectroni ca ll ycontro ll ed , networks, subsystems. tires. wheels.

ginee rin g De partme nt has bee n do in g researc h s in ce 1988 w ith th e "Darmstadt Tire Sensor". now in the 4th gene rat io n ( figure] ).

I. INTRO DUCTI ON New mec ha troni c sys tems will co ntribute greatl y to safety, conse rvation of reso urces. comfort and e nvironme ntal impac t of future road vehi cles by sharin g informati o ns de li ve red by smart mechatroni c co mpone nts and matc hin g the ir behav iour in common control arrange me nts to wards an e nhanced overall system perfo rma nce .

With the" Darmstadt Tire Sensor" the deformati o n o f a tread e le me nt in the tire contac t area in x. v a nd [directi on can be measured in th e ro tating tire with o ut affec ting its performance c harac teri stics. Thi s a ll ows not o nl y to ca lc ul ate the local indi vidual forces. but also the g lo bal forces transmitted by the tire as a n Integra l of the local fo rces. The tread e le me nt de fo rmatio n is measured by the se nsor as a pos iti o n change of a mag net re lati ve to four cross wise ly arranged Ha ll sensors. The moveme nts in x, y a nd zdirecti o n can be resolved by the interconnecti on of the fo ur e le ments.

2. SU BSYSTEMS

2. 1 Th e Darm stadt Tire Sensor

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The evaluati o n and on-line assessme nt of the characteri sti c sig nal patterns of the tread ele me nt deformati on ac hie ved durin g tire rotati on yie ld info rmati on on:

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coe fficient o f fri cti o n / use of fri ctio n coeffic ie nt, tire pressure , wheel load , braking and accele ration forces, lateral forces, risk of hydroplaning , tire temperature (temperature sensing eleme nt at sensor).

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The adhesio n be twee n tire and road must transmit all forces required for acceleration, deceleration , change of directi o n and course holding o f a vehicle. The fri cti on-coeffi cient and the d ynamic wheel load are the main factors for the achie vabl e frictional force and a dec isive fac tor for drivin g c harac teristics and safety of a vehicl e. In thi s area, the Automoti ve En-

For e xa mple, the de vi ati on between the curves under free-ro lling and brakin g co nditi o ns a ll o ws to cal c ulate an index for the curre nt brakin g fo rce (figure 2).

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Fig. 2: Tire sensor x-signal under variation of the braking force [Bachmann 1998] .

Fig. 4: Tire sensor x-signal of a free-running wheel, variation of road surface [Bachmann 1999].

Figure 3 shows a characteristic sensor signal under hydroplaning conditions: The penetration of water into the contact area of the tire with increasin v driving speed is seen from the very beginning . Thus, a descriptor for hydroplaning danger has been found enabling host systems to actively oppose the danger and/or warn the driver. Figure 4 depicts the evaluation of the friction-coefficient: the gradient chan!!e of the s ignal between the curves of high and low cc;efficient of fricti o n can be clearly detected and identifies the beginning of local slip inside the contact patch. The available coefficient of friction can be calculated o n-line from the curves.

2.2 New Mechatronic Parr Disc Brake The increasing function potential of current braking systems and the lack of beneficial auxiliary energy (vacuum) due to new types of engine concepts require new methods in brake technology. Future vehicle concepts will require brake systems which operate without complex auxiliary energy components and which ha ve a su itabl e interface to hivher leve l vehicle systems. However, the economically reasonable reali zation of these brake sys tems will be depending on the system's benefit s, such as its dynamic behaviour, energy and power de mand , space required, mass and reliability. Especially the connection between actuator and friction brake throuvh a suitable tran sm ission system appears as a problem that has not yet been solved adequately. For this reason , a mechatronic wheel brake is bein v investigated at the Automotive Engineering D;partment

In the same way , indices for further parameters are calculated and transmitted to host systems. These systems, in turn , are able to recognize the driving situation and intervene respectively to increase driving safety and economic viability. They may also be used to immediately warn the driver.

Fig. 3: Tire sensor x-signal of a slick-tire under hydroplaning conditions and descriptor for hydroplaning danger [Stocker 1995]. 1068

motor follows an initial rotary translation stage, which then works on a spindle drive . The spindle introduces a translation movement into the friction component of the self-energizing part disc brake . Figure 8 shows the realized prototype brake and how it is fitted into the test stand in a series produced 17" rim .

in Darmstadt. Current research topics are both the improvement of the efficiency of several components and the use of the vehicle's kinetic energy to actuate the brakes .

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Fig. 5: Electromechanical wheel brake. Figure 5 gives the struc ture of an e lectrically actuated wheel brake. Input variables are the e lec trical energy and the data fl ow to control the sys tem; the output vari able are the required cl amping force and information for hi gher systems like the vehicle controller.

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Fig . 6: Sketch of the reali zed wheel brake. In addition to calculating a nd s imulatin g the brake. tests must also be carried o ut to valid ate the theoreticall y deri ved results. Currently th e first tests are be ing performed on a roller dynamometer.

The schemati c des ig n of the rea li zed wheel brake shows figure 6 . The main units of motor, transmission, and friction brake are designed modularly. The

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1.300 Nm braking moment. The standard spindle force is just less than 3 kN. The resultant clamping force was around 15 kN. A brake factor of C* = 5 was derived. The electrical output during this test was 25 Watt, the peak output for overcoming the brake pad gap was around 150 Watt.

market, that actively operate in horizontal direction. Other and had shown, however, that there these mentioned conflicts.

Therefore, the Automotive Engineering Department of Darmstadt Technical University has developed an actuator to substitute a common rubber-metal element to link the wheel suspension to the body. The properties of the actuator with regard to stiffness, damping capacity and kinematics are variable in a wide range , i.e. they may be adjusted to the current driving situation. However, safety has priority over riding comfort. Experimental investigations were made on a HydropuIs test stand. The actuator itself consists of a modified rubber-metal element housing with two separate hydraulic chambers that can be individually loaded with a pressure medium via hydraulic connection. The adaptive bushing also contains connections to measure the pressure inside the chambers that are transmitted as control parameters to individual pressure controllers. The force resulting from deformation of the rubber is superposed by additi o nal fo rce from pressure s inside the chambers if the se were loaded by external pressure. The force produced by the pressure medium is calculated from chamber pressures Pkl and Pu and the effective cross sectional area. The vector of force s created by pres surization of the chambers in addition to rubber-elastic force resulting from external sinusoidal excitation (e .g. unbalance of wheel. brake judder) superimpose one another. Amplitude and phase relation of the resulting total force and thus dynamic stiffness and loss angle or damping charac teristics of the bearing can be modified in wide ranges in case of a sinusoidal differential pressure curve with equal frequency. Uneven filling of the chambers may be used to create kinematic effects in the form of wheel position changes for direct. comfortable and adequate modification of driving characteristics.

Fig. 8: Brake prototype in the 17"' rim.

2.3 Adaptive Wheel Bushing When constructing a wheel suspension. one will sooner or later come to a trade-off between driving comfort, driving safety and viability. Even while including the so-called elastokinematics it is not possible to completely solve them. As an example, the compromise on stiffness and damping capacity of the rubber-metal-parts may be mentioned . To meet the high demands concerning vehicle dynamics on a wheel suspension like road holding , responsiveness etc. it would be necessary to use rather stiff bushings between suspension parts and chassis. At the same time this measure will worsen the driving comfort, because vibrations caused for example by road surface and wheel imbalances will then be much better transferred to the chassis and to the driver. Vibrations in the area of suspension are another problem in modern vehicles. Low levels accepted for vibration and noise brings this problem into the foreground on one hand but measures taken for weight reduction result in a higher susceptibility of the construction to vibrations on the other hand.

Rubber-elastic force is very low at low excitation amplitudes that are predominant during normal driving operation. The force that can be produced by different pressures between the two chambers is higher by many times over. Therefore , bearing properties at small excitation amplitudes are mainly defined by internal chamber pressures. Shifting of the phase relation between path and differential pressure has immediate effect on the phase of resulting overall force and thus to the phase angle defining the damping, that originates from the predominance of pressure. The phase angle can therefore be set within wide limits (figure 9).

A first step towards optimization of the transfer function of elastokinematic parts was the introduction of the so-called hydro-bushings, which allow the damping of certain frequency ranges. Currently no concrete attempts are known to solve the trade-off by using active or adjustable components. By now, sys-

The maximum dynamical stiffness achievable by the actor is reduced versus frequency as the amplitudes of differential pressure reachable in the chambers 1070

diminish at higher frequencies which is due to the applied hydraulic power pack. In case of frequencies up to approximately 20 Hz the differential pressure inside the chambers is high enough to enable setting of the phase angle over the entire range of -180° up to + 180 0 . The ratio of differential pressure force and rubber-elastic force decreases with rising frequency and thus the setting range of the loss angle is reduced accordingly. There IS, however. a certain interdependence of dynamic stiffness and set phase angle.

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All three described components were being developed as autonomous systems. that are supervised by their own controllers. Since a single system might benefit from information provided by other systems. it makes sense to bind each single system in to one compound. A first step is the exchange of data, which is used to operate the controllers more efficiently because of the increased number o f information. Plus , sensors that record redundantly the same signals in different s ubsystems can be omitted to save expenses.

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The second step not o nl y includes the communication , but also the interaction of the subsystems. that may be coordinated by a central controller. It is expected that sllch systems can take advantage with regard to critical driving situations opposite to conventional driver assistance. Thi s shall be shown on the basis of two examples:

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Example 1.- "Full braking on a straightline road" Currently elastokinematics are mostly realized by passive rubber-metal elements. The driver who recognizes a dangerous situation requiring full braking will move his foot from the accelerator to the brake pedal and press it down vigorously. Optical acquisition of the foot 's motion or measurement of the pedal path or acceleration initiates an approach of the brake lining to the brake disk and setting of reduced clearance so that a steeper rise of the brake torque is accomplished. If the rubber bearings are hydraulically pretensioned by external hydraulic pressure in addition to a reduction of the clearance , the response time of the brake can be further improved . This also helps to reduce undesired elasto-kinematic effects by increased bearing stiffness and improvement of wheel position stability. It is even possible to create desired elastokinematic effects such as toe-out on the lefthand and toe-in on the right-hand side in my-split conditions. This effect is therefore much higher than with passive elastokinematics which avoids a waste of tire-road-friction potential. Finally it is possible to

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Fig. 9: Comparison between a conventional and the adaptive bushing at an excitation amplitude of 0.2 mm. After successful tests on the Hydropuls test bench the Automotive Engineering Department presently equips a test vehicle with such an actor-prototype to make experimental tests on real roads. The actuator is placed between the wishbone of a McPherson front wheel suspensIOn and the corresponding cross member. A picture of the mounting position IS shown in figure 10.

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5. ACKNOWLEDGEMENTS

measure the current maximum admissible braking force by the brake system as the tire sensor supplies information about active tire-road-friction and available friction potential. Thus, an iteration loop used by today 's ABS systems to prevent wheel lock will not be necessary any more . Brake judder can be actively damped during braking. It is also possible to shift natural frequencies into currently unexcited ranges just by altering the stiffness.

We would like to thank Deutsche ForschungsGemeinschaft for sponsoring the three sectional projects in Special Research Area 241 and many German companies which support our work.

6. REFERENCES

Example 2: "Braking durin g cornering" Load changes during cornering cause a transfer of weight especially to the outer wheel of the front axle related to occurring engine braking torque or forces which finall y results in a change of wheel loads and tread forces . The cornering force applied to the wheel must be kept which changes the slip angle and leads to oversteering. Loading of the vehicle adds to this effect. Friction potential may also be affected by cornering forces resulting from tensioning as experienced in the case of conventional elastokinematics. This effect can be remedied by active rubber bearings helping to optimize the self steering properties of the vehicle by improved kinematic compensation also with respect to best possible exploitation of tire-roadfriction on all wheel s.

Bachmann. V.; Fach, M.: Breuer, B.: Future CarTires as Provider of Info rmation for Vehicle Systems to enhance Primary Safety. SAE Paper 981944 , Future Transportati on Tec hn ology Conferen ce , Costa Mesa, USA , 1998 Bachmann. V.: Unte rsucllllngen ZlIm Eillsatz. van Reifensensoren im Pkw-Reifen. Dissertation TU Darmstadt , 1998. Fortschritt-Berichte VDI Reihe 12 Nr. 38 1. VDI-VerJag Diisse ldorf 1999 Balz. 1.; Bill , K.-H .; Ba hm, 1. ; Scheerer, P. ; Semsch, M .: Konzept fiir eine elektromechanische Fahrzellgbremse. ATZ Automobiltechnische Zeitschrift 98 (1996) Barz, M. : Experimentelle Untersllchungen an passiven IInd aktivell Fahr.verkslagern. fzd -B eri ch t Nr. 260/00 (unpubli shed ), Automotive Engineering Department TUD Bill. K.-H.: Grundsatzllnrersll chllngen zum Einsarz. elekrrischer Radb remsen in Personenkraftwagell. Fortse hritt-Beri chte 166. Reihe 12. VDI-Vcrl ag GmbH Diisse ldorf 1992 Bill, K.- H.; Semsc h, M. : Trallslarionsgerriebe jiir elekrrisch betiifigre Fahrzellgbremsen. A TZ Automobiltechnische Ze itsc hri ft 100 ( 1998) 1 Huinink , H. ; Schra der. c.: DYllamische Interaktioll Brel1lse - Reifell - SrrafJe. XVIII. Internati onales !l-Symposium , Bremsen-Fachtagung. Fortsc hrittBeri chte VDI Reihe 12 Nr. 373. VDI- Vcrl ag Diisse ldorf 1999 N.N.: DFG-Sonderforschungsbereich 241. Nelle integrierre m echanisch -elektrische Systeme fiir den Maschin enbau (IMES). Arbeits- und Ergebnisbericht 1uli 1996 - Marz 1999 Semsc h, M .: Neuartige mechatronische Teilb elag scheibenbremse. Fortschritt-Berichte 405. Reihe 12, VDI-Verlag GmbH Diisseldorf 1999 Semsch, M.; Bill , K.-H .; Dausend , U.; Breuer. B.: Kon zeptioll lInd Auslegung d es Getriebes fUr eine mechatronische Teilbelagscheibenbremse. Fortschritt-Beri chte 743, Reihe 8, VDI-VerJag GmbH Diisseldorf 1998 Stacker, 1.; Kirschbaum . A.; Aller, K. ; Breuer, B.; Glesner , M.; Hartnagel , H .-L. : Der " Intelligente Reifen " erste Ergebllisse einer illterdiszipiindren Forschungskooperation. Automobiltechnische Zeitschrift ATZ 97 (1995) 12 Wick, A.: Erminlung dynamischer Kennwerte an Gummi-Metall-Elementen. Automobiltechnische Zeitschrift ATZ 88 (1986) 4

4. OUTLOOK

To study the behaviour of such a "mechatronic vehicle corner", a test bench is currently being fitted with all three subsystems and its completi on is scheduled for the middle of year 200 I. A central controller will then coordinate their interaction und serve as interface and gateway to a personal computer, which will simulate different boundary conditions.

Fig. I I : The test bench for the "Darmstadt mechatronic vehicle corner".

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