DE19704954A1 - Wheel rotation signal detection for motor vehicle - Google Patents

Abstract

The method involves using a sensor signal to determine a wheel rotation signal, with which the rotational speed of the wheel is determined, relatively to a reference point of the wheel carrier. The wheel rotation signal is produced under consideration of a relative movement between the wheel mounting point and the vehicle, and preferably also suspension, during the evaluation of the sensor signal.

Description

The invention relates to a method for determining a wheel speed signal Use of a sensor signal with which the rotational speed of the wheel is determined relative to a reference point of the wheel carrier.

Such a method is widely used in motor vehicles in which the Wheel rotation speed or the wheel slip is used as an input variable by means of a regulation to affect the vehicle. Such systems include, for example Anti-lock braking systems with which the brake pressure is reduced at least for individual wheels, if the respective wheel tends to lock up during braking. They also belong to it for example, traction control systems with which one Accelerating the spinning of the drive wheels should be prevented. It is in In this context, it is known to reduce the engine torque when the drive wheels both sides of the vehicle tend to spin or the corresponding drive wheel to brake by an active brake intervention, if only one drive wheel one Vehicle side tends to spin. Such systems have recently also been added systems in which the vehicle is engaged, for example, by active brake intervention is stabilized on individual wheels when the vehicle tends to skid. It is also known, a switchable differential lock and / or a switchable all-wheel drive in Switch automatically depending on the wheel speeds. It is expected that the Significance of these signals in systems of future development is still increasing.

According to the prior art, the wheel speed signal is obtained from a sensor, wherein a part of the sensor is attached to the wheel carrier and, for example, a to the wheel Sprocket. When the ring gear rotates relative to that on the wheel carrier attached part of the sensor an inductive evaluation takes place by the time interval it is determined between two successive pulses of the sensor rotating sprocket.

To determine the wheel slip, the wheel speed of the wheel is related brought to the wheel speed, which at the current vehicle speed would set a freely rotating wheel.  

So it is usually the difference between the (fictitiously determined) wheel speed of the free rotating wheel and the determined wheel speed and by the (fictitiously determined) Divided wheel speed. The wheel speed of the freely rotating wheel is, for example won by defining a wheel as the reference wheel. By this reference wheel then rotates at least temporarily without braking, its wheel speed is determined as fictitious Wheel speed of the freely rotating wheel used.

It is an object of the present invention to determine the wheel speed more precisely as well as the wheel slip.

According to the invention, this object is achieved in that the wheel speed signal is obtained, by making a relative movement between the Suspension point and the vehicle is taken into account.

It has been shown that the accuracy, in particular with a smaller wheel slip in determining the wheel slip is improved when determining the Wheel slip an appropriately corrected wheel speed signal is used.

In a particularly advantageous embodiment (see claim 7), another Improvement can be achieved if in addition the relative movement between tire belts or the wheel contact point and wheel center point is taken into account.

In the embodiment of the method according to claim 2, the wheel speed signal won by taking into account the deflection of the vehicle wheel.

With regard to accuracy, this has a particularly advantageous effect if the Wheel slip is in an area where the maximum braking force from the wheel to the Road is transmitted. This is because of the vehicle deceleration Nodding movement of the vehicle, which leads to shifts in the wheel load distributions. As a result this load shift and the changing load on the individual wheels change the radii of the vehicle wheels, d. H. the distance of the wheel contact point from Wheel center. Because of the dynamic change in this wheel radius, the Peripheral speed of this wheel.

In the embodiment of the method according to claim 3, the wheel speed signal obtained by adding to the rotational speed of the wheel determined by the sensor a term is added relative to the wheel carrier which is proportional to the speed of the  Wheel center point relative to the vehicle in the horizontal direction and inversely proportional to the strut length.

This shows that the consideration of the appropriate sizes in a simple manner can be carried out.

In the method according to claim 4, the wheel speed signal is obtained by the Evaluation of the rotational speed of the wheel determined by the sensor relative to the wheel carrier as further variables the brake pressure and the spring deflection of the wheel or other sizes corresponding to these further sizes are taken into account.

It shows that the correspondingly more complex measurable quantities by means of these simple measurable sizes can be determined. So it can be made from these measurable sizes improved wheel speed signal can be determined.

In the method according to claim 5 is taken into account from the wheel speed signal a variable distance of the wheel contact point from the wheel center Circumferential speed of the wheel signal obtained or from the Circumferential speed of the wheel taking into account a variable distance a signal representing the wheel speed of the wheel contact point from the wheel center, with the variable distance a vertical movement of the wheel and / or Bumps in the road are taken into account.

The vertical movement of the wheel can be done in a comparatively simple form by means of the Travel and various characteristic curves determined by a differential equation will. Bumps in the roadway can be caused, for example, by a suitably mounted Infrared distance sensor can be detected.

In the method according to claim 6, the variable distance is taken into account by an average value of the distance of the wheel contact point from the center of the wheel vertical deflection of the wheel center added and / or vertical components of Uneven road surfaces are subtracted.

As a result, the parameters can be taken into account in a comparatively simple manner. It there is an improvement in signal accuracy if the subtraction of the vertical Components of road bumps in addition to the addition of the vertical deflection the center of the wheel.  

In the method according to claim 7, the peripheral speed is determined of the wheel the change in time of the shift of the wheel contact point to Wheel center taken into account.

It has been shown that more realistic wheel slip values can be determined, if the wheel slip is greater than the slip at which the maximum braking force from the wheel is transmitted to the road and when the wheel slip decreases (a negative Gradient). This can be used especially in the transition to the unstable area thus improve the accuracy of the wheel slip signal.

In the method according to claim 8, at least one of the corrected variables is used Determination of wheel slip used.

In the method according to claim 9, the wheel slip is determined by the Signal representing vehicle speed corresponding to the relative movement between the suspension point and the vehicle.

This shows an improvement in the wheel slip signal with a wheel slip that is smaller than the wheel slip at which the maximum braking force from the wheel onto the road is transmitted.

In the method according to claim 10, the relative movement between the Suspension point and the vehicle added to the vehicle speed.

This size can be easily taken into account.

In the method according to claim 11, the sizes are obtained by in addition to the Sensor signal with which the speed of rotation of the wheel relative to the wheel carrier still the brake pressure and the spring deflection or others corresponding to these measured variables Measured variables are taken into account.

This allows the corresponding corrections to be made with a comparatively simple Measurement effort can be carried out. Especially in a vehicle with an electrical Brake system, which is known under the keyword "brake by wire", is the signal of the Brake pressure or the clamping force before. This also applies, for example, to the Travel in a vehicle with a regulated suspension or one load-dependent adjustment of the headlights.

An embodiment of the invention is shown in more detail in the drawing. Show it included:

Fig. 1 is a representation of the horizontal movement of the wheel due to the longitudinal compliance of the suspension,

Fig. 2, the axial displacement of the tire belt _x relative to the rigid rim as a result of the braking force _{F b,}

Fig. 3, the vertical movement of the wheel center as a result of wheel load,

Fig. 4 is a simplified representation of radial (RKS), lateral (LKS) and tangential force (TKS),

Fig. 5 is an illustration of a spring-damper strut type suspension under the influence of a braking force,

Fig. 6 is an illustration of a spring-damper strut type suspension with spring deflection of the front wheels,

Fig. 7 shows a simplified model of the spring-damper strut type suspension, and the tire,

Fig. 8 is a vertical free-cut model for determining the cutting forces with measured characteristics,

Fig. 9 is a horizontal cut-free model for determining the cutting force,

Fig. 10 shows the time profile of the braking force and the slip with and without consideration of the influencing variables,

Fig. 11 shows the μ-slip curve with and without taking into account influencing variables, and

Fig. 12 is a schematic representation of the model-based compensation.

The following are the influences on the wheel speed signal and the slip during driving described. Derived from the presented interference, a method for Determination of an improved slip value and a correction of the wheel speed signal presented. Modeling continues to generate the required influencing variables suspension and tire using a quarter vehicle model. The at Test drives or static measurements on the test vehicle determined, non-linear Process coefficients and the signals of the model are explained. In addition, on Example of a drive with braking maneuver of the slip and the µ-slip characteristic with and without Correction represented and discussed by the presented method.

When determining the slip between the brake slip and the Differentiated traction slip. Since the interfering influences on these sizes are similar, the only the brake slip should be considered below for better clarity.  

The absolute brake slip λ _{x, b} is defined as the ratio of the differential speed between the speed of the wheel contact point v _a and the peripheral speed of the tire v _r to the speed at the wheel contact point. When braking straight ahead, the following applies for absolute brake slip:

the circumferential speed of the tire can be replaced by multiplying the dynamic tire diameter r _dyn by the angular speed of the wheel ω _r . Simplifications in the slip calculation are, for example, the assumption of a constant tire diameter r _dyn = r ₀ or the equation of the speed at the wheel contact point v _a with the vehicle speed v _{motor vehicle} the suspension and tire dynamics are taken into account.

The influences of the horizontal movement of the wheel will be explained with reference to FIG. 1.

Horizontal or longitudinal movements or vibrations of the wheel can cause fluctuations in the frequency of rotation, which affect the wheel speed signal ω. The degree of freedom of the wheel center in the longitudinal direction is created by an elastic mounting (rubber bearing) of the axle suspension on the vehicle body. This absorbs hard impacts such as z. B. occur when braking or driving over obstacles, and thus contributes to driving comfort. The longitudinal movement of the wheel center point relative to the body changes the speed of the wheel contact area v _a , which is required to calculate the slip. In order to correctly calculate the slip, the speed of the wheel center dx _r / dt relative to the body must be taken into account. The speed v _{a of} the wheel contact point is thus obtained by adding the vehicle speed v _Kfz and that of the wheel center dx _r / dt.

(2) v _a = v _vehicle + dx _r / dt.

The influence of dx _r / dt on the speed of the wheel contact area v _a is essentially determined by the course of the braking force and the horizontal wheel suspension parameters.

Another influence that has to be taken into account in the slip calculation is the tangential contact or belt speed of the tire relative to the rim. These relationships can be illustrated with the aid of FIG. 2. This speed arises from the rotation of the very stiff tire steel belt connected to the rim via the elastic carcass as a result of forces acting on the wheel contact surface. The resulting tangential speed component at the wheel contact point dx _g / dt is therefore not detected by the wheel speed sensor and can therefore advantageously be taken into account when calculating the slip in addition to the peripheral speed v _{r of} the tire.

Taking the tire deformation into account, the following relationship for the wheel speed results with the measured wheel speed ω:

(3) v _r = r _dyn . ω - (dx _g / dt - dx _r / dt)

The influences of the vertical movement will be explained below with reference to FIG. 3.

The dynamic tire radius r _{dyn is} required to calculate the wheel rotation speed v _r from the speed sensor signal ω. This radius changes constantly while driving due to wheel load fluctuations, which will be explained with reference to FIG. 3. Wheel load fluctuations can result on the one hand from excitation on the roadway and on the other hand from excitation on the vehicle side. Roadside suggestions are caused by bumps in the roadway. Suggestions from the vehicle arise during the respective driving maneuvers such as braking, accelerating, steering or through the influence of vehicle components such as wheel imbalances and tire irregularities.

Driving maneuvers initially result in forces at the wheel contact points in the longitudinal or lateral direction. These cause the vehicle to pitch or roll due to the inertia of the body. Such body movements and the resulting deflection cause a change in the dynamic wheel load and thus the dynamic tire diameter r _dyn . Wheel imbalances and tire irregularities can occur on the one hand while driving (brake plates, lost weights, curb crossing etc.) or even during tire production. When assembling the many individual components of a tire in tire production, deviations from the ideal shape easily occur, which results in fluctuations in thickness, fluctuations in mass, and fluctuations in stiffness and damping. In addition to imbalance, these also cause longitudinal, transverse and radial force fluctuations due to different mass and stiffness distributions on the tire circumference, which are explained in FIG. 4. Fig. 4 shows a simplified representation of radial (RKS), lateral (LKS) and tangential force fluctuations (TKS).

The disturbing forces resulting from tire irregularities lead to vehicle-side vibration excitation. Fig. 4 schematically shows a circulation of a tire, the formation and the courses of the radial, lateral and tangential force variations.

Due to the elasticity of the tire, the mentioned fluctuations in force on the roadway and vehicle side lead to a change in the dynamic wheel contact force and thus to a change in the dynamic tire diameter r _dyn .

However, the measurement of this dynamic tire radius is very difficult and costly when driving (e.g. infrared distance sensor). Known systems therefore assume a constant tire radius r _dyn = r ₀ for simplification. However, this leads to a falsification of the determined wheel rotation speed. To calculate the slip, the change in radius of the tire due to wheel load fluctuations can advantageously be taken into account. The dynamic tire radius is then composed of a static radius r ₀ and a dynamic component due to the vertical movement of the wheel center z _r and the bumps in the road h _f , as shown in FIG. 3.

(4) r _dyn = r ₀ + z _r - h _f

It has been shown that with speed changes up to 150 km / h, a change in the static radius r ₀ due to the change in stiffness of the belt tire is negligible.

The influences of the rotation of the wheel carrier will be explained below with reference to FIG. 5.

The is of crucial importance for the usability of the wheel speed signal Location of the sensor and the wheel suspension kinematics. The speed sensor measures that Difference between the absolute rotation of the wheel and that of the wheel carrier. Since the sensor in generally attached to the wheel carrier, its position in space must be suitable as a reference variable. This is not sufficient for all axle designs. Depending on the axis kinematics the wheel carrier rotates under deflection or under the influence of the braking torque. This Twisting causes the sensor to measure a speed that is not actually on the wheel occurs. The speed signal and thus the slip determined from it is thereby adulterated. An axle construction is therefore for the installation of the speed sensor on the wheel carrier only suitable if there is little or none at all in all driving conditions Self-rotating movement, i.e. change in the caster angle Δτ occurs.  

Fig. 5 shows the behavior of a spring damper strut wheel suspension with an inclined strut under the influence of the braking force F _{x, b} . The wheel-guiding spring-damper leg is articulated in the village and firmly connected to the steering knuckle ( 1 ). As a result of the braking torque, the steering knuckle and thus the wheel center point are shifted from ( 1 ) to ( 2 ) by the amount x _r . The wheel carrier rotates by the angle Δτ _e , which reduces the original caster angle τ. The angular movement Δτ _{e is} superimposed on the wheel rotation during the measurement and falsifies its value.

The following formulaic relationship is used to correct the wheel speed signal considered:

Assuming small deflections Δτ _e (permissible for standard cars, x _{r, max} << l _om ):

(5) x _r ≅ l _om . Δτ _e , with l _om = l ₀ + Δz

where l _{om represents} the effective length of the spring damper when the angular position of the spring damper is neglected. Differentiating and exempting Δτ _e results in:

(6) d (Δτ _e ) / dt = (dx _r / dt) / l _om

For the corrected wheel speed signal ω _corr one obtains with the measured wheel speed ω:

(7) ω _corr = ω - (dx _r / dt) / l _om

The spring-damper strut type suspension with obliquely standing strut according to Fig. 6, the point (1) moves during compression parallel to the rotation axis (3) and pushes the fastened in the dome strut together. The distance from the dome to the point ( 1 ) is shortened when compressing or extended when rebounding. Here, there is a rotation by the angle Δτ _k , which causes a falsification of the wheel speed signal.

Since the vehicle compresses at the front during the braking process due to the dynamic axle load shift, both self-rotating movements are superimposed, with Δτ _k acting opposite to Δτ _e . Δτ _k is low and can be neglected compared to Δτ _e . Axle constructions in which the wheel carrier carries out only slight rotary movements are e.g. B. the double wishbone axle with a correspondingly rigid construction and spring-shock absorber axles. The trailing link axle and the multi-link axle, however, are problematic.

Taking into account all influences and the resulting relationships according to the aforementioned Eq. (2) to Eq. (7) the expanded slip formula according to Eq. ( 1 ) ultimately set up as follows:

The determination of the kinematic influencing variables required for the calculation of the slip according to Eq. ( 8 ) can be done on the one hand by measurements, on the other hand by simulation.

In the following, a method is to be described with which the missing sizes can be calculated, based on the model, from signals that are easily measured on the vehicle and are already available. The measured wheel speed ω, the spring travel Δz and the brake pressure p _hyd are therefore selected as the model _input (the first two are already available in vehicles with ABS or chassis control or automatic headlight range adjustment, the brake pressure is easy to measure, or is available with brake-by- wire systems directly or as clamping force for fully electric brake-by-wire systems). In order to achieve a high level of accuracy in modeling, it is necessary to map the non-linear properties of the individual parameters of the wheel suspension and tires as precisely as possible. An improvement in the calculation model can also be achieved by taking into account the geometry data of the wheel suspension used and by modeling friction effects of the bearings and mass coupling effects. This geometry data is known from the prior art.

The modeling of the wheel suspension below is the spring-shock absorber front suspension of a BMW 318i Type E36, whereby the model is formed in the event of straight braking on a level road. Excitation on the road side due to bumps, potholes or the like is initially not taken into account in the modeling (h _f = 0). Fig. 7 shows the overall model of the suspension and the tire having two translational and one rotational degree of freedom.

The vertical dynamic behavior is decisively determined by the shock absorber, the spring and the tire. The stiffnesses and damping c _az , d _az , c _rz , d _rz and the proportional body mass m _a are time-variant quantities. The tire sprung mass m _r can be assumed to be constant. Since different partial masses are accelerated with horizontal or vertical vibrations of the wheel, a distinction is made between horizontal m _rx and vertical tire-sprung masses m _rz . The dynamic behavior of the dual mass transducer can be derived from the balance equations of the forces acting on the body and the wheel. With regard to the coordinate system shown in Fig. 8 and Fig. 9, the following applies to the balance equations of the structure and the tire:
Construction:

( 9 ) m _a . (d ² z _a / dt ² + g) = F _{caz, stat} + F _{caz, dyn} + F _daz - F _{a, dyn} , F _{caz, stat} + F _{caz, dyn} = F _caz

Wheel:

(10) _mar . (d ² z _r / dt ² + g) = -F _caz - F _daz + F _{N, dyn} + F _{N, stat} , F _{N, dyn} + F _{N, stat} = F _N

(11) F _{caz, stat} = m _a . g, F _{N, stat} = (m _a + m _rz ). G

F _caz and F _daz represent the spring or damper force of the wheel-guiding spring damper _strut . The dynamic wheel contact force F _{N, dyn} acts in the normal direction and characterizes the force applied in the tire contact patch. Here, the vertical portion of the belt or carcass _mass m _{rg is} neglected, since, in contrast to the horizontal direction, only the tire-sprung masses m _{rz are} accelerated. F _{a, dyn} is caused by driving maneuvers such as B. Brakes, steering or similar causes and takes into account the resulting change in mass of m _a . This force is due in particular to the vehicle's nod during braking. The static point of the spring force F _{caz, stat} and the wheel contact force F _{N, stat} determines the working point by which the dynamic variables are deflected.

The fact that these forces do not act in a vertical axis but rather offset by certain distances and angles can be taken into account in the equations in the form of scale factors. This procedure is known from the prior art. For reasons of clarity, these scale factors should not be included here. The differential equation for mapping the vertical wheel movement can be specified for the working point using the non-linear characteristic curves measured on the test vehicle and the spring travel as follows.

(12) m _rz .d ² z _r / dt ² + F _crz (z _r ) + F _drz (dz _r / dt) = - F _caz (Δz) - F _daz (d (Δz) / dt) = F _fd

Fig. 8 shows characteristics that describe the spring and the tire. The forces are also shown in detail in a model.

Starting from FIG. 9, a coupled dual mass oscillation system is used to describe the longitudinal movements of the wheel center and the longitudinal flexibility of the tire.

The longitudinal rigidity of the wheel suspension as well as the deformation property of the tire, e.g. B. under the influence of a braking force, in the longitudinal direction are described by non-linear spring-damper systems. To determine the stiffness, static measurements can be carried out on a test vehicle. The horizontal tire deflection can be determined by the relative path between the tire belt and a point on the rigid wheel body, which results from the tangential deformation of the carcass. The damping can be determined using spectral estimation based on an autoregressive model with measured signals from test drives. This is known from the general state of the art.

• - The largest part of the wheel mass (rim and parts of the carcass) is combined in the tire-sprung mass m _rx and
• - A small mass fraction m _rg (steel belt and carcass portion) acts in the area of the wheel contact patch and thus in the lower area of the belt.

These masses can be weighed, for example, or taken from the information provided by the vehicle manufacturer or component manufacturer. The inertia of the belt and the rim, which were accelerated rotationally during the horizontal movement, were converted to the masses m _rx and m _rg . In order to model the dynamic behavior of the wheel suspension and the tire in the longitudinal direction, the balance equations of the vibration system are established on the basis of FIG. 8. By cutting the system free you get for the balance of forces in the direction x _r or x _g :

Suspension:

(13) m _rx . d ² x _r / dt ² = F _crx + F _drx - F _{x, f}

Carcass:

(14) m _rg . d ² x _g / dt ² = F _crk + F _drk - F _{x, b}

The following also applies:

(15) F _crx + F _drx = - c _rx . x _r - d _rx . dx _r / dt

such as

(16) F _k = F _crk + F _drk = c _rk . (x _r - x _g ) + d. (dx _r / dt - dx _g / dt)

where c _rx and d _{rx represent} non-linear variables of the wheel suspension and c _rk and d _{rk of} the carcass. The force F _{x, f} can be determined from the moment balance around the attachment point of the spring damper strut "0" from the force F _k acting on the tire patch (see FIG. 9):

The horizontal output variables can be calculated using the measured characteristic curves and input variables with the following coupled differential equation system:

Suspension:

Carcass:

(19) m _rg . d ² x _g / dt ² = F _k (x _r , dx _r / dt, x _g , dx _g / dt) - F _{x, b} (dω / dt, p _hyd ).

As a suggestion for the model, the braking force F _{x, b} applied in the tire flap is suitable. If this is known, all sizes of the horizontal model can be simulated. From the torque balance around the center of the wheel, however, is only a reconstruction of the Karkassenkraft F _k possible (cf. Figures 7 and 9...):

(20) Θ _r . dω / dt = r _dyn . F _k - M _b

where Θ _r corresponds to the combined moment of inertia of the rotating masses (tires and rim), dω / dt to the angular acceleration of the wheel and M _{b to} the braking torque. With the assumption permitted for the braking process:

(21) m _rg . d ² x _g / dt ² << F _k → F _{x, b} ≈ F _k

follows for the suggested excitation of the model F _{x, b} :

The angular acceleration dω / dt of the wheel can be obtained directly from the measured wheel speed signal by one-time differentiation. R _dyn = r ₀ + z _r and Θ _r are known quantities. With the measured brake pressure p _hyd , the braking torque M _b can finally be determined according to the following relationship

(23) M _b = r _b .F _{u, b} = r _b .2. µ _b .p _hyd .A _k ,

where F _{u, b represent} the circumferential force between the brake pad and brake disc and r _b the effective braking radius. The circumferential force F _{u, b} is obtained by multiplying the coefficient of friction µ _b between the two brake pads (factor 2) and the brake disc, the brake pressure p _hyd and the effective piston area A _{k of} the hydraulic brake cylinder ( 8 ). The coefficient of friction µ _b is generally time-variant. An average was chosen for the simulation. The braking force can thus be reconstructed using the input signals wheel speed and brake pressure. With Eq. 22 and Eq. 23 finally follows for horizontal excitation:

In experiments, results of this model with measurements of the corresponding sizes were obtained compared. This showed a good agreement between the calculated quantities and the  respective measured quantities. The following are some additional ones Explanations are given. For this purpose, the slip with and without Compensation of the described dynamic influences is shown.

The time profiles of the braking force F _{x, b} and the corrected and uncorrected slip λ _{x, b are} plotted in FIG. 10. The signal curves shown in FIG. 10 result from Eq. (8), Eq. (24) and Eq. (25).

In area 1 , the correction is mainly achieved by taking the horizontal wheel center speed dx _r / dt into account. Area 2 also shows a clear difference between the two types of slip calculation. Here the wheel center is already at rest. The correction in this area can be traced back to the change in the dynamic tire radius r _dyn caused by the vehicle nodding off. In area 3 , the uncorrected slip shows strong vibrations, which even deflect into the negative slip area, although braking force is still available. The correction, which is mainly achieved in this area by taking into account the wheel center speed dx _r / dt and the belt speed dx _g / dt, shows a clear smoothing and a shift in the slip curve into the positive and thus plausible range.

By applying the adhesion coefficient µ _{x, b} over the slip courses from FIG. 10, the µ slip course plotted in FIG. 11 is obtained.

The coefficient of adhesion results from this

where F _N with the help of Eq. (10) to (12) can be calculated. The areas of clear correction from FIG. 10 are also identified here. In this form of representation, area 1 can be found in the shift of the ascending slip curve (λ _{x, b} ≦ 6%) towards larger slip values. The characteristic curves shown show a significant improvement of the course in the sense of an approximation to an optimal µ-slip curve course by taking into account the dynamic processes presented in this work when calculating the slip. The hysteresis is greatly reduced and a clear smoothing of the course can be seen.

Fig. 12 shows a block diagram of the model. The brake pressure p _hyd , the spring deflection Δz and the measured wheel speed ω are used as measured _variables . From this, the other quantities can be determined using equations 24, 12, 18 and 19, which are used to calculate the wheel slip according to Eq. 8 are needed. A corrected wheel speed signal and a corrected slip of the wheel can thus be obtained.

The braking force can be determined by measuring the brake pressure, taking into account other vehicle-dependent but constant variables according to equation 24. A coupled differential system of equations is then obtained from equations 12 and 18 (via the influence of z _r on r _dyn ). These equations can be solved, since the respective quantities can either be determined using characteristic curves (F _crz , F _caz , F _daz , F _crx ) or by spectral estimation of the parameters (F _drz , F _drx ). The spring deflection Δz is measured. This means that z _r , dx _r / dt and dx _g / dt can be determined.

Claims (11)

1. A method for determining a wheel speed signal using a sensor signal with which the rotational speed of the wheel is determined relative to a reference point of the wheel carrier, characterized in that the wheel speed signal is obtained by a relative movement between the wheel suspension point and the vehicle when evaluating the sensor signal is taken into account (dx _r / dt). 2. The method according to claim 1, characterized in that the wheel speed signal is obtained by deflection of the vehicle wheel is taken into account (Δz). 3. The method according to claim 1 and 2, characterized in that the wheel speed signal is obtained by adding a term to the rotational speed of the wheel determined by the sensor relative to the wheel carrier, which term is proportional to the speed of the wheel center relative to the vehicle in the horizontal direction and inversely proportional to the strut length ((dx _r / dt) / l _om ). 4. The method according to any one of claims 1 to 3, characterized in that the wheel speed signal is obtained by evaluating the rotational speed of the wheel determined by the sensor relative to the wheel carrier as further variables of the brake pressure and the spring deflection of the wheel (p _hyd , Δz ) or other sizes corresponding to these other sizes are taken into account. 5. The method in particular according to one of claims 1 to 4, characterized in that a signal representing the peripheral speed of the wheel is obtained from the wheel speed signal taking into account a variable distance of the wheel contact point from the wheel center or from the peripheral speed of the wheel taking into account a variable distance of the Wheel contact point from the wheel center a signal representing the wheel speed, with the variable distance taking into account vertical movement of the wheel and / or bumps in the road (z _r , h _f ). 6. The method according to claim 5, characterized in that the variable distance is taken into account by adding a vertical deflection of the wheel center point to a mean value of the distance of the wheel contact point from the wheel center point and / or subtracting vertical components from uneven road surfaces (z _r , h _f ) . 7. The method in particular according to claim 5 or 6, characterized in that the time change of the displacement of the wheel contact point to the wheel center is taken into account when determining the peripheral speed of the wheel (dx _g / dt). 8. The method according to any one of claims 1 to 7, characterized in that at least one of the corrected quantities for determination of wheel slip is used. 9. The method according to any one of claims 1 to 8, characterized in that when determining the wheel slip, the signal representing the vehicle speed is taken into account in accordance with the relative movement between the wheel suspension point and the vehicle (dx _r / dt). 10. The method according to claim 9, characterized in that the relative movement between the wheel suspension point and the vehicle is added to the vehicle speed (dx _r / dt). 11. The method according to any one of claims 1 to 10, characterized in that the quantities are obtained by taking into account, in addition to the sensor signal with which the rotational speed of the wheel relative to the wheel carrier, the brake pressure and the spring deflection or other measured quantities corresponding to these measured quantities ( p _hyd , Δz).