DE102014114751A1 - Method for operating a motor vehicle - Google Patents

Abstract

The invention relates to a method for operating a motor vehicle (2), which has a steering system (10) with a rack (12), with the wheels (6, 8) on an axle (4) on which tires are mounted, including associated Wheel suspensions of the motor vehicle (2) are directed, wherein the motor vehicle (2) is moved on a roadway, wherein a coefficient of friction or a derived value between a respective tire and the road in consideration of a driving condition, a kinematics of the steering system (10), a Kinematik of the axis (4), a behavior of the tires on the wheels (6, 8) and / or the trailing distance is determined as a characteristic size, wherein in the course of application of the method changing operating parameters of the motor vehicle (2) and the tire by a Learning phase that is performed during at least one trip of the motor vehicle (2), recognized and taken into account.

Description

The invention relates to a method and a system for operating a motor vehicle.

A motor vehicle has contact with its wheels to a traveling surface. Among other things, a so-called coefficient of friction is to be considered, which may depend on a tire tires, for example. Of a tire tread depth and the tire type, but also on a condition of the ground, usually a road, as well as weather conditions. If the motor vehicle drives through a curve, it is also exposed to a lateral force and thus to a lateral acceleration. Spurs safety when driving through the curve depends on a speed of the motor vehicle, a radius of the curve and on the coefficient of friction.

A method for determining the coefficient of friction between tires of a motor vehicle and a roadway is from the publication DE 10 2009 002 245 A1 known. Here, an actual, acting on the steerable wheels of the motor vehicle actual tire restoring torque is determined and compared with a reference tire return torque. To determine the coefficient of friction, the ratio of the actual tire return torque to the reference tire return torque is used. Furthermore, the actual tire return torque is calculated from a current rack force acting in a rack of a steering system of the motor vehicle.

In one from the publication DE 10 2006 036 751 A1 known methods for controlling or controlling at least one driving state variable of a motor vehicle, the cornering force is determined on a tire of the motor vehicle for adjusting an actuator in the motor vehicle according to a predetermined kinematic relationship as a function of the rack force acting in the rack of the steering system of the motor vehicle. In addition, the rack-and-pinion force is determined as a function of the steering moments acting in the steering system.

A method for determining the coefficient of friction in a motor vehicle with at least one steered wheel is in the document DE 10 2010 036 638 A1 described. In this case, the cornering force is determined on the at least one steered wheel, wherein a racking force detected and determined a restoring moment on the steered wheel in dependence of the rack force and the cornering force as a function of caster and determined on the basis of the return torque of the coefficient of friction.

The estimation of the coefficient of friction can be carried out by various already known approaches. For example, a wheel slip can be determined during a braking or acceleration process and the coefficient of friction determined therefrom. Another possibility is to perform an exclusion method using a currently measured lateral acceleration. In this case, low friction coefficients can be excluded at high body accelerations of a motor vehicle, since a high lateral acceleration on a slippery road surface can not be achieved.

Another possibility provides for the use of a motor vehicle model with the aid of which a theoretical restoring torque or a theoretical rack force of the steering can be determined. For this purpose, various input signals are required when carrying out previous measures, usually a lateral acceleration, a yaw rate, a motor vehicle longitudinal speed and a curve radius. Such reference values which apply to high friction coefficient roadways are compared with the values of a currently measured and / or estimated restoring torque or the current rack force. The coefficient of friction can then be estimated from the moment or force ratio via another model or a characteristic diagram.

These measures are, for example, in the publications DE 103 19 662 A1 and DE 10 2010 036 638 A1 mentioned.

Against this background, a method having the features of claim 1, a method having the features of claim 12 and a system having the features of claim 18 are presented. Further embodiments of the invention will become apparent from the dependent claims and the description.

The method according to the invention is intended for operating a motor vehicle, which by way of example here has a steering system with a rack, with the wheels on an axle on which tires mounted, including associated suspension of the motor vehicle are steered. The motor vehicle is moved on a roadway or a subsoil. In the method, a coefficient of friction or a quantity derived therefrom between a respective tire and the roadway taking into account a driving condition, a kinematics of the steering system, a kinematics of the axle, a behavior of the tires on the wheels and / or their trailing distance as a characteristic size or Operating condition of the motor vehicle determined. In the course of an application of the method, changing operating parameters of the motor vehicle and of the tires will be recognized and taken into account by a learning phase, which is carried out during at least one drive of the motor vehicle, and the method will be adapted accordingly in design.

The coefficient of friction or the quantity derived therefrom as an operating parameter is determined between measured forces and / or torques in the axle, the steering gear and / or the tires as well as from driving dynamic operating parameters, for example the speed and / or lateral acceleration, of a respective tire and the road Motor vehicle determined.

The friction coefficient is calculated as a function that depends on at least one other operating parameter. As the at least one further, usually driving dynamic operating parameters is a steering angle, a lateral acceleration acting on the motor vehicle, a yaw rate of the motor vehicle, a speed of the motor vehicle, acting on the rack rack force, a restoring torque of at least one tire and / or a steering angle at least a wheel used.

In an embodiment, a current, measured rack force F _{Zst_akt is subdivided} into a first portion, which comprises the trailing _{path of} the tires on the wheels, and into a second, constructive portion, which comprises a kinematics of the axle or axle kinematics. Here, the second, constructive portion is referred to here as a constructive rack force F _{Zst_akt_k} = k _Sv · F _Sv , which is proportional to a side force acting on the motor vehicle.

Furthermore, the trailing distance is determined on a route with a high coefficient of friction as a function of the lateral force F _Sv acting on the motor vehicle. Based thereon, the reference value for the rack-and-pinion force to be determined for a roadway having a high coefficient of friction is compared in design with a value of the currently measured rack-and-pinion force, the coefficient of friction being estimated taking into account a current driving state of the motor vehicle.

In addition, a dynamic wheel load is determined for at least one front wheel as a function of the lateral acceleration or an equivalent quantity, for example the lateral force, which acts as a driving dynamic operating parameter on the motor vehicle, in particular during cornering. The lateral acceleration results in a dynamic displacement or change in the wheel load, wherein a change in the wheel load is directly proportional to the lateral acceleration.

In another embodiment, the axle kinematics of the axle, in particular the front axle of the motor vehicle, and their influence on a behavior of a force and torque transmission from a resultant point of force application in a footprint or wheel contact surface at least one wheel to a place of measurement of the force or the torque on the axle.

A behavior of the tires and / or the trailing distance of the at least one wheel in the resulting point in the footprint of the at least one wheel and / or tire on the road is determined by variable driving conditions, by forces acting on the at least one wheel and by different ones Loads that act on the at least one wheel influenced.

The determination of the coefficient of friction or the quantity derived therefrom is usually carried out only if the following conditions are met. These conditions are met when the speed of the motor vehicle is greater than a minimum value provided for this purpose, if the lateral acceleration of the motor vehicle is greater than a value provided for this, if influences of driving and braking torques on the axle are negligible, if a second derivative the steering angle after the time, d. H. a steering angular acceleration, and a third derivative of the steering angle with time, d. H. a steering angle pressure, falling below each intended maximum values and / or the steering angle is outside of a range of the zero crossing, and when the ratio of the speed of the steering angle to the speed of the motor vehicle and the brake pressure are each smaller than a threshold provided for this purpose.

In a further refinement, an effective follow-up distance and a theoretical follow-up distance to be expected in the case of a high coefficient of friction are calculated from recorded operating parameters, an exclusion criterion being checked for the calculated effective follow-up distance. If a test of this exclusion criterion is positive, it is determined that no new reliable estimation of the coefficient of friction between the tires and the driven lane is possible. If the check of the exclusion criterion is negative, a value of the lateral acceleration is checked. If the value of the lateral acceleration is greater than a lower limit value and less than an upper limit value, the operating parameters for evaluating the rack force are readjusted or adapted, in the event that the lateral acceleration is greater than the upper limit value or equal to the upper limit value , a comparison of values of a theoretical trailing path with values of the effective trailing path is carried out. If values of the theoretical trailing distance are smaller than values of the effective trailing distance, the operating parameters for evaluating the rack force are adjusted again. If values for the theoretical caster distance and the effective caster distance differ by a maximum of one tolerance value, it is determined that the lateral acceleration is currently too low, and that no new reliable estimate of the coefficient of friction between the tires and the traveled lane is possible. If values of the theoretical trailing distance are greater than those of the effective trailing distance by the tolerance value, a maximum usable frictional connection and thus the coefficient of friction is estimated.

In addition, in the context of the method, at least one parameter of the motor vehicle, for example the steering system, at least one wheel and / or tire, or for a motor vehicle-specific operating parameter, a stored value for the at least one characteristic is taken into account. In this case, a current value for the at least one parameter of the motor vehicle is determined at regular times and / or in the presence of a characteristic operating range or operating state of the motor vehicle and adapted the previously stored value for the at least one parameter to the current value, wherein the at least one parameter is updated. Such a characteristic is z. B. a load state of the motor vehicle or a tread depth and an air pressure of the at least one tire. Taking into account the characteristic operating range of the motor vehicle, modified parameters of the motor vehicle and / or of the at least one tire are detected when the method is carried out and the method is adapted or adapted accordingly. Thus, a long-term quality of the method is to be improved. Usually, the operating range to be used in this case can be chosen such that there is no measurement of the coefficient of friction due to the process. Regular times at which a monitoring of the at least one parameter and possibly an adaptation of the method is carried out are determined and / or repeated on the basis of an optionally variably adjustable time interval, for example one minute, one hour or one day. Such a time interval can be adjusted taking into account a respective running time of the motor vehicle independently of a standstill of the motor vehicle.

The further method according to the invention is intended for operating a motor vehicle which has a steering system with a toothed rack, with which wheels are guided on an axle on which tires are mounted, including associated wheel suspensions of the motor vehicle, wherein the motor vehicle is moved on a roadway. In this case, in a learning phase which is carried out during at least one drive of the motor vehicle, value pairs are respectively calculated from values for a current rack force F _{Zst_akt of} the rack and from values for a current side force F _{Sv_akt} acting on the motor vehicle, which are calculated from driving dynamic operating parameters. determined. Here, a plurality of pairs of values a slope m _act for at least one line that describes a proportionality between the current rack force F _{Zst_akt} and the current side force F _{Sv_akt} determined, wherein m _akt applies to the slope:

Thus, in an embodiment for determining the friction coefficient, inter alia, a relationship between the rack force F _Zst and the lateral acceleration or the lateral force F _Sv , which acts on the motor vehicle, used. For this purpose, the currently acting lateral force is calculated from operating parameters of the motor vehicle, for example driving dynamics or kinematic variables, for describing a movement of the motor vehicle.

After the learning phase, a possible deviation of the actual rack force F _{Zst_akt} from the reference rack force F _{Zst_ref} is determined via a danger _{indicator GI} taking into account a constructive rack force F _{Zst_akt_k} , whereby the following applies to the danger _indicator GI:

The learning phase is usually carried out after wheels of the motor vehicle have been repurposed with tires.

In addition, the method can be operated with the lowest possible signal volume, so that only a small number of values of operating parameters, which are usually provided to a control unit via signals, are required. This relates for example to values for the steering angle of at least one wheel, the speed of the motor vehicle in the forward or longitudinal direction and / or the lateral acceleration. Only a small number of measured and / or determined values of operating parameters are used by the control unit of the motor vehicle in order to avoid having to use any further measuring technology and high control unit resources.

The system according to the invention or a corresponding arrangement is designed for operating a motor vehicle which has a steering system with a toothed rack, with which wheels are steered on an axle on which tires are mounted, including associated wheel suspensions of the motor vehicle, the motor vehicle being mounted on a roadway is moved. The system includes a control unit configured to calculate a coefficient of friction or a quantity derived therebetween between a respective tire and the road in consideration of a running condition, a steering mechanism kinematics, a kinematics of the axle, a behavior of the tires on the wheels, and / or determine their caster distance as a characteristic size. The control unit is further configured to recognize and take account of changing operating parameters of the motor vehicle and of the tires by a learning phase that is to be carried out during at least one drive of the motor vehicle.

In the context of the presented method, a linear approach for determining the lateral force or lateral acceleration is used with the at least one straight line with the gradient m _act . The coefficient of friction is alternatively or additionally determined only taking into account the speed of the motor vehicle and the steering angle of at least one wheel. In another embodiment, the axle kinematics of the axle and / or the steering kinematics of the steering system are taken into account for determining the coefficient of friction for more precise determination of the coefficient of friction.

The method presented here is based on the derivation of force or torque ratios in the area of wheel-guiding components of the motor vehicle. This can be done, for example, by determining a rack force in the steering system of the motor vehicle. A resulting signal can be derived from internal operating parameters of the steering system, as z. B. in implementing a measure for a so-called. Electric Power Steering (EPS) is possible. Alternatively, measures for measuring equivalent forces or moments, z. B. by means of an additionally mounted sensors in the field of a wheel, for example. By strain gauges on tie rods or torque sensors used.

If only for two of the mentioned operating parameters, here for example for the speed and the steering angle, incoming signals with measured values are available, a simulation model is used to determine the friction coefficient for which at least one assumption is made and / or in the at least one Query is integrated.

In an embodiment, the friction coefficient of the tires of the motor vehicle on the roadway is estimated via a change in a restoring torque of the at least one tire.

If a significant deviation from currently measured value pairs occurs from a straight line which was determined at high frictional value ratios and thus at a high coefficient of friction, a degree of the deviation can be used to quantify a decrease in the coefficient of friction.

As long as less than about 50% of the available coefficient of friction μ is utilized by a tire, no noticeable sliding movements occur in the tire contact patch. As a result, the trailing distance of the tire is approximately constant, ie the lateral force always attacks at the same distance behind a center of the tire or the wheel. As a result, the restoring torque of the tire increases in proportion to the transferred lateral force. At least for small steering angles, the rack-and-pinion force is also proportional to the restoring torque of the tire, so the approximate proportionality between the actual rack-and-pinion force F _{Zst_akt} and the current side force F _{Sv_akt is} approximated. If significantly more than 50% of the Friction coefficient be exploited, the tire begins to slide in the rear of the tire contact patch, the trailing distance decreases. The measured rack force is thereby reduced. From the deviation between the occurring rack force and the rack force, which is calculated at this side force from the previously learned proportionality between the rack force and the lateral force, can be concluded that the currently available adhesion between the tire and the road.

In a further embodiment, it is considered that the trailing distance is dependent on a lateral force which acts on the motor vehicle. Usually, the trailing distance is determined on a route with a high coefficient of friction μ as a function of the acting lateral force. The trailing distance increases as a function of the wheel load by a larger footprint of a wheel, wherein a maximum lateral force and the footprint increase almost in proportion to an increase of a respective wheel load.

As a rule, values for a selected steering angle of at least one wheel and for a longitudinal speed of the motor vehicle are used by a control device of the motor vehicle as measured operating parameters.

Usually, a lateral force is calculated which acts on an axle formed as a front axle and / or steerable axle of the motor vehicle.

The effective overrun distance is calculated in an embodiment. This may mean that a high coefficient of friction μ or a high friction coefficient, which theoretically affects the effective tracking distance, is calculated.

The theoretical caster distance here is the caster distance which results on the tire when no sliding movements occur and optimal conditions for the friction coefficient prevail. Assuming that the theoretical caster distance is constant independently of the lateral force acting or lateral acceleration, the proportionality between the rack force and the lateral force, which can be represented by the straight line provided for this purpose, can be derived from the theoretical caster distance. The effective overrun distance, on the other hand, is actually the overrun track that actually acts on the tire while driving. With low utilization of the coefficient of friction, the two overrun sections are to a good approximation identical. At a utilization of significantly more than 50% of the coefficient of friction, the effective overrun distance is noticeably less than the theoretical overrun distance. The occurring rack power is thus also lower.

In one possible embodiment, the exclusion criterion is checked for the calculated effective overrun distance. In this case, in the event that the value of the lateral acceleration is greater than a lower limit, z. B. 0.5 m / s ² , and / or less than an upper limit, z. B. 1 m / s ² , is the operating parameters for the evaluation of the rack force to be adjusted again. If the lateral acceleration is greater than the upper limit value or equal to the upper limit value, a comparison of values of a theoretical trailing _distance n _{R_theo} with values of an effective trailing _distance n _{R_wirk is} performed.

In this case, an estimate of the coefficient of friction is only trusts if the exclusion criterion is not satisfied, for example evaluating the exclusion criterion considered values of the drive torque on the axle, an acceleration d ² δ / dt ^{2 of} the steering angle δ and a derivative of the acceleration of the steering angle is avoided, thereby avoiding that at a strongly fluctuating acceleration d ² δ / dt ^{2 of} the steering angle zero crossings of the acceleration d ² δ / dt ^{2 of} the steering angle are considered valid. As a further operating parameter, a ratio of the speed dδ / dt of the steering angle to the speed v _{x of} the motor vehicle and the brake pressure are considered. The exclusion criterion is when a value of one of the above-mentioned operating parameters is greater than a respective limit value to be defined and / or defined for this purpose.

Alternatively or additionally, an estimate of the coefficient of friction is only carried out if the speed of the motor vehicle is greater than a minimum value provided for this, so that inter alia a resolution of the yaw rate or yaw rate is sufficiently accurate, and if there is no control intervention by a chassis control system.

In a further embodiment, the case is taken into account that an effective trailing distance is greater than the theoretical trailing distance. Accordingly, values of operating parameters of the motor vehicle do not correspond to values provided for this, which have been used to determine the theoretical tracking distance. In this case, the determined overrun distance for adjusting the operating parameters for evaluation the rack power used. However, if the value of the effective trailing distance becomes smaller than the theoretical value provided therefor, this value is used to estimate the maximum usable coefficient of friction.

In general, a dependency of the coefficient of friction is represented either by a mathematical function or by a characteristic map, wherein in the event that a determined, effective trailing distance is below a limit value which depends on the lateral force and no exclusion criterion exists, the coefficient of friction is estimated becomes.

The system according to the invention or the corresponding arrangement comprises at least one sensor for detecting at least one, for example, driving dynamics and / or motor vehicle-specific operating parameter of the motor vehicle and is designed to carry out the method presented above.

The control unit comprises a module which is designed to estimate the friction coefficient μ taking into account a simulation model.

The control unit is also designed to determine a reference value for the rack-and-pinion force exclusively from a steering angle of at least one wheel and a speed of the motor vehicle.

The presented method is usually adaptive feasible, wherein various, usually motor vehicle-specific operating parameters or characteristics of the motor vehicle over its lifetime, for example. By in-service updates are adjusted.

The method achieves high stability by considering the tire slip as a characteristic amount for determining the friction coefficient. Here, a wheel load distribution (left - right) on the front axle due to a lateral acceleration when cornering must be considered. In addition, a kinematics of the front axle and their influence of a behavior of the power transmission to the tie rods to be considered. In the method, a function of the tires, an influence of a tire wake and a resultant force introduction point in the tire footprint by tire longitudinal forces and transverse tire forces and different wheel loads are considered, wherein the front axle can also be driven by an engine of the motor vehicle.

Thus, a reduction of necessary operating parameters on the steering angle and the speed to simplify the process and for quick implementation on an EPS controller with simple algorithms is possible.

In the method, in a further refinement, during at least one drive of the motor vehicle, a variable characterizing the tire return torque is set as the operating parameter in relation to the effective lateral force acting on this wheel between the tire and the roadway. This consideration can also be carried out axle by axle or based on the entire motor vehicle.

In one embodiment of the method, for example, it is assumed that there is a linear relationship between the current rack force and the current lateral force that can be represented by the straight line. In an approximation to be considered in this case, all nonlinear effects from a suspension of the motor vehicle, the z. B. of steering angles and roll angles are neglected. In this case, from a plurality of pairs of values, a gradient _{m.sub.actu is determined} for at least one straight line or equalizing straight line which describes a proportionality between the rack-and-pinion force and the lateral force.

Thus, the slope m _akt to be determined in this embodiment is calculated via a compensation calculation, wherein the product of the value pairs F _{Zst_akt} (t) for the current rack force and F _{Sv_akt} (t) for the current side force, which are determined at the same time, is formed and then added up. A sum determined in this case is divided by the sum of all squares of the lateral forces F _{Sv_akt} (t) respectively determined at the times.

Alternatively, an equation for a reference rack force F _{Zst_ref} = m _akt · F _Sv + b and thus for a desired rack force, which should prevail during operation of the motor vehicle, is determined. In this case, an offset and / or a value for a constructive rack force F _{Zst_akt_k is} described with a constant value b. The slope m _akt depends inter alia on a pressure and a temperature of a tire of a wheel.

Values for the current rack force F _{Zst_akt} are measured for their determination or alternatively calculated, whereas values for a side force F _Sv acting on the motor vehicle are calculated for their determination or alternatively measured.

In order to determine the slope _m.sub.akt , a direct relationship between the measured, actual rack _{force F.sub.Zst_akt} and the current side force _{F.sub.S_akt is} evaluated. Alternatively, value pairs can also be evaluated from values of the rack force F _{Zst_akt} and associated values of the lateral acceleration.

For determining the current rack force F _{Zst_akt} at a non-constant steering angle δ, a value for the, for example, measured rack _force F _{Zst_gem} can be corrected with at least one correction _value F _{Zst_korr} . For the current rack power applies:

Here, F _{Zst_gem is} the measured rack force, J _{Rad is} an inertia of those components, including the wheels of the motor vehicle, which are moved by the steering angle δ in a steering. d ² δ / dt ² or δ  is a steering acceleration, Θ _Rad a rotational inertia of the wheels, d δ / dt or δ  a steering speed, n _{wheel is} a speed of a moving wheel and h _{w is} an effective steering lever. c _γ is the ratio of the camber change with respect to a change in steering angle when the wheels are turned. A first correction _value F _{Zst_korr1} describes a mass inertia, for which the following applies: F _{Zst_korr1} = -J _Rad · δ 

A second correction _value F _{Zst_korr2} describes centrifugal _forces , where:

As a rule, only the first correction _value F _{Zst_korr1 is} used here. An additional use of the second correction _value F _{Zst_korr2} is only optional.

On the basis of a currently determined characteristic size of the tire return torque and an additional variable which is equivalent to the currently acting lateral force, a currently available adhesion between the road and the tires mounted on the wheels and thus the coefficient of friction μ is concluded.

The tire return torque or restoring torque of the tire arises from the fact that the lateral force engages offset by the caster distance behind the center of the wheel. The restoring moment is a product of the lateral force and the trailing distance. With known lateral force can thus be calculated from the restoring torque, the trailing distance and vice versa from the trailing distance, the restoring moment. The trailing distance is thus a characteristic variable of the tire return torque.

The learning phase, which can be carried out during the method, is carried out, inter alia, when changing an operating range of the motor vehicle and thereby updates the gradient _m.sub.new after wheels of the motor vehicle have been re-frosted. Alternatively or additionally, a learning phase is carried out after a mode or an operating range of a chassis of the motor vehicle has been changed. In this case, values for different gradients can be stored for different suspension modes.

It is also possible to carry out the learning phase for the adaptation of operating parameters at regular time intervals and / or after a certain distance covered, and thus to update the gradient m _akt , whereby wear of the tires used can be taken into account.

In a further embodiment, a first straight line with a first pitch m _akt for a direction of travel or steering to the left and a second straight line with a second pitch m _akt for a direction of travel or steering to the right are determined.

The system according to the invention comprises the at least one sensor for continuously detecting at least one operating parameter, for example the steering angle of at least one wheel, a steering wheel angle about which the steering wheel is rotated, the lateral force F _Sv and the lateral acceleration a _y and of force or Torque ratios in Radführenden elements of the suspension or the steering system or a steering system of the motor vehicle. The rack force F _{Zst_akt} is derived in design from other sensory operating parameters.

In an embodiment of the method, among other things, the rack force of a rack and pinion steering system is determined.

Further advantages and embodiments of the invention will become apparent from the description and the accompanying drawings.

It is understood that the features mentioned above and those yet to be explained below can be used not only in the particular combination indicated, but also in other combinations or in isolation, without departing from the scope of the present invention.

1 shows a schematic representation of a detail of a motor vehicle and an embodiment of a system according to the invention.

2 shows a first diagram with operating parameters, which are taken into account in a first embodiment of the method according to the invention.

3 shows a second diagram with operating parameters, which are taken into account in the first embodiment of the method according to the invention.

4 shows a third diagram with operating parameters, which are taken into account in the first embodiment of the method according to the invention.

5 shows a fourth diagram with operating parameters, which are taken into account in the first embodiment of the method according to the invention.

6 shows a schematic representation of an example of a control device as a component of the system according to the invention.

7 shows a first and a second diagram with operating parameters, which are taken into account in a second embodiment of the method according to the invention.

8th shows a third and a fourth diagram with operating parameters, which are taken into account in the second embodiment of the method according to the invention.

9 shows a first flowchart for carrying out a third embodiment of the method according to the invention.

10 shows a detail 1 for determining operating parameters, which are taken into account in the third embodiment of the method according to the invention.

11 shows a diagram for determining operating parameters, which are taken into account in the third embodiment of the method according to the invention.

12 shows a second flowchart for determining operating parameters, which are taken into account in the context of the third embodiment of the method according to the invention.

13 shows another detail 1 for determining operating parameters, which are taken into account in the third embodiment of the method according to the invention.

14 shows a fifth diagram with operating parameters, which are taken into account in the third embodiment of the method according to the invention.

15 shows a third flowchart for determining operating parameters, which are taken into account in the context of the third embodiment of the method according to the invention.

The figures are described coherently and comprehensively, like reference numerals designate like components.

In 1 is a schematic representation of a front portion of a motor vehicle 2 shown, the here designed as a front axle axis 4 with two steerable wheels 6 . 8th , which are designed here as front wheels, has, on which tires are mounted. In addition, the motor vehicle comprises a steering system 10 for steering the two wheels 6 . 8th , this steering system 10 a housing 11 which includes a steering pinion and a rack 12 having. The enclosure encloses 11 the rack 12 , That inside the case 11 arranged steering pinion engages teeth of the rack 12 one. The steering system 10 also includes a manually operated steering wheel 13 and a handlebar 15 , wherein the steering pinion and the steering wheel 13 over the handlebar 15 connected to each other.

The presented here embodiment of the system according to the invention 14 includes a controller 16 and at least one sensor 18 , where the in 1 shown, at least one sensor 18 the rack 12 and the housing 11 the steering system 10 assigned and is designed to operate as an operating parameter during operation of the motor vehicle 2 resulting rack force, ie one on the rack 12 and / or between the housing 11 and / or the rack 12 acting rack force to measure directly or at least indirectly. Here are the sensor 18 in an embodiment control variables, ie control and / or controlled variables in an actuator as operating parameters of the steering system 10 to detect, from which the rack power is derived.

The in the 2 to 5 Diagrams shown show gradients 34 . 50 . 84 . 86 . 104 . 106 . 108 of operating parameters, which are taken into account in the context of the method according to the invention and will be described in more detail below.

The diagrams off 2 and 3 each comprise an abscissa 30 along the transverse acceleration a _{y of} the motor vehicle 2 in m / s ^{2 is} applied. In addition, the two diagrams include the 2 and 3 one ordinate each 32 , along which a rack-and-pinion force F _zst , _resting on the rack 12 acts, is applied.

The diagram 2 includes a curved course 34 of measured values of the rack-and-pinion force F _{Zst_gem} at a coefficient of friction μ, which here is approximately 1 when the road is dry. In addition, the diagram includes 2 a first straight line 36 for a reference rack force F _{Zst_ref} with a positive slope m _akt , a second straight line 38 for a limit value of a hazard indicator GI and a third straight line 40 for representing a constructive portion of the rack-and-pinion force F _{Zst_akt_k} , this third straight line 40 has a slope k _Sv .

The current rack force F _{Zst_akt} is in an embodiment in a first portion, the trailing _distance or a wake of the tires on the wheels 6 . 8th or a tire overrun, and in a second structural portion, the axis kinematics of the axis 4 includes, subdivide. The second, constructive portion is referred to here as a constructive rack force F _{Zst_akt_k} = k _Sv · F _Sv . In the described embodiment, it is assumed on the basis of a simplifying approach that there is a linear relationship between the lateral force and the rack-and-pinion force. In this case, all non-linear effects of a suspension and the tires for the wheels remain 6 . 8th unconsidered.

By a double arrow 42 is off in the diagram 2 Also, a defined by a limit value of the lateral acceleration range of validity for the hazard indicator and for the reference rack force, by the line 36 . 38 are shown. Within the scope are slopes of the line 36 . 38 positive, wherein the reference rack force F _{Zst_ref} reaches a maximum at the limit value of the lateral acceleration a _y or side force F _Sv . Above the limit value of the lateral acceleration for defining the range of validity for a lateral acceleration, wherein in the example shown here the limit value is approximately 7.5 m / s ² , the diagram shows 2 a fourth straight line 44 , with a negative slope a _Sv for representing a dependence of the reference rack force F _{Zst_ref} on the lateral acceleration and the lateral force F _Sv outside the validity range.

The diagram 3 also shows a curved course 50 of measured values of the rack-and-pinion force F _{Zst_gem} in the case of a wet track, if the coefficient of friction μ is less than 1. These Considering the condition, the diagram comprises 3 also a first straight line 52 for a reference rack force F _{Zst_ref} with positive slope m _akt , a second straight line 54 for a limit value of a hazard indicator GI and a third straight line 56 for a constructive portion of the rack force F _{Zst_akt_k} with a slope k _Sv . Here, the reference rack force F _{Zst_ref} also has a maximum here, which is achieved for a limit value of the lateral acceleration or the lateral force F _Sv .

A fourth, indicated here by dashed lines 58 with a slope a _Sv illustrates a course of the measured rack force F _{Zst_gem} or reference rack force F _{Zst_ref} for values of the lateral acceleration which are greater than the limit value, this fourth straight line 58 has a negative slope a _Sv . One in the diagram 3 additionally indicated fifth straight line 60 is parallel to the fourth straight line 58 Accordingly, it also has the slope a _Sv for describing the proportionality between values of the measured rack force F _{Zst_gem} and the lateral acceleration or lateral force F _Sv . To determine a coefficient of friction μ, this fourth straight line is obtained 58 with which the dependence of the measured rack force F _{Zst_gem} or reference rack force F _{Zst_ref} on the lateral acceleration or side force F _Sv above a limit value provided for the lateral acceleration or lateral force F _Sv is described, shifted in parallel and with the third straight line 56 for the constructive part of the rack-and-pinion force F _{Zst_akt_k} cut.

In the present case, the fifth line intersects 60 with the slope a _Sv the third straight line 56 for the structural part of the rack-and-pinion force F _{Zst_akt_k} , which here also has the slope k _Sv , in one operating point 62 for which the coefficient of friction is here, for example, μ = 0.45. At a projection of this operating point 62 on the abscissa 30 as shown here by an arrow 64 indicated here is a value for a maximum lateral acceleration of about 4.5 m / s ² in this operating point 62 to investigate. A straight 58 . 60 here describes a proportionality between the lateral acceleration a _y and the side force F _Sv and the reference rack force F _{Zst_ref} , with the example, the measured rack force F _{Zst_gem} can be adjusted, where F _{Zst_ref} = a _Sv · F _Sv + c with a corresponding Offset c applies. By such a shift of this line 58 . 60 is the above equation of the line 58 . 60 for the reference rack force F _{Zst_ref} for values of the side force F _Sv that are greater than the limit, with the constructive rack force F _{Zst_akt_k} = k _Sv · F _Sv , which is hereinafter referred to as equation (G3), equated and thus the operating point 62 determined with the friction coefficient μ to be determined.

The diagram 4 includes an abscissa 80 , along which the time is plotted in seconds, as well as an ordinate 82 along which a force is applied in Newton. In the diagram is a first history 84 for a reference rack force F _{Zst_ref} to be calculated from the lateral force F _SV when carrying out the method. This reference rack force F _{Zst_ref} is proportional to a lateral force F _{Sv generated} while driving on the motor vehicle 2 acts. Possible courses of this reference rack force F _{Zst_ref} are based on the diagrams from 2 and 3 been presented. In addition, the diagram shows 4 a course 86 for the current rack force F _{Zst_akt} determined during the measurement.

The diagram 4 shows the courses 84 . 86 the calculated and measured current rack force F _{Zst_akt} when driving on a country road as a lane. The courses 84 . 86 cover themselves very well here.

The diagram 5 includes an abscissa 100 , along the one to the axis 4 of the motor vehicle 2 acting lateral force F _{Sv is} applied. Along an ordinate 102 of the diagram 5 is a rack force F _Zst applied. Furthermore, the diagram comprises 5 three different courses 104 . 106 . 108 for a measured rack force F _{Zst_gem} with different wheels tires 6 . 8th and different configuration of a chassis of the motor vehicle 2 , This results in the first course 104 for winter tires (WR) with a comfort-oriented modification of the chassis. Regarding this first course 104 shows 5 in addition a first straight line 110 for representing a slope of the reference rack force F _{Zst_ref} as a function of the lateral force F _Sv . The second course 106 results in summer tires (SR) and comfort-oriented modification of the chassis, in this respect in the diagram 5 also one from this second course 106 derived slope of a second straight line 112 is shown. The third course 108 results from the use of summer tires and vehicle dynamics-oriented modification of the chassis. From a respective value of the measured rack force F _{Zst_gem} , a current rack force F _{Zst_akt is} determined taking into account at least one correction factor F _{Zst_korr} .

Additionally is in 5 through an ellipse 114 an area for training slopes of the two straight lines 110 . 112 indicated. In the first embodiment of the method, a pitch _m.sub.actu becomes representative of the proportionality between the reference rack force _{Fzst_ref} and the side force F _Sv low side forces F _Sv recorded and calculated on the compensation method. An thereby determined value of a respective slope m _akt is extrapolated by conducting the first embodiment of the process for higher lateral forces F _Sv.

With the present first embodiment of the method, inter alia, an estimate of a coefficient of friction μ of the tires of the motor vehicle 2 about a change of a restoring torque of the tires possible, with a relationship between the rack force F _Zst and thus the force of the rack 12 the steering system 10 of the motor vehicle 2 and a lateral force F _Sv or lateral acceleration of the motor vehicle 2 is being used.

Here, for small lateral accelerations, as already explained above, a simplifying, linear approach or relationship between the lateral force F _Sv or F _{Sv_akt} and the rack force F _Zst or F _{Zst_akt is} used, where an ascertainable function is given by the straight line with the gradient m _act can be described. This straight line is usually learned when carrying out the method, for example newly trained, and therefore updated. With the function to be described about the straight line operating parameters of the motor vehicle 2 that is characterized by new tires, a change in the track of the motor vehicle 2 or other influences can be detected.

When carrying out the method, according to equation (G1), the side force F _{Sv of} the motor vehicle 2 calculated. Alternatively, instead of the lateral force F _Sv, a transverse acceleration a _{y of} the motor vehicle can be measured and used directly, but the lateral acceleration a _{y is} in dynamic maneuvers of the motor vehicle 2 inaccurate.

Equation (G1) comprises the mass m of the motor vehicle and a distance l _{h of} a center of gravity of the motor vehicle to its rear axle in the longitudinal direction of the motor vehicle, the lateral acceleration a _{y of} the motor vehicle 2 in the y-direction and thus transversely to a direction of travel in the x-direction of the motor vehicle 2 , an inertia Θ _{z of} the motor vehicle 2 in the z-direction and thus parallel to a vertical axis of the motor vehicle 2 as well as perpendicular to the x-direction and the y-direction, a yaw rate d ² ψ / dt ² or ψ , a steering angle δ _VA on the axis of the motor vehicle 2 and a wheelbase l of the motor vehicle.

In order to determine a current rack force F _{Zst_akt} , a currently measured rack force F _{Zst_gem can be designed} at least via the mass inertia Θ _{wheel of} the wheel and an acceleration d ² δ / dt ^{2 of} the steering angle or a steering angle of a wheel 6 . 8th Getting corrected. In a further embodiment, the measured rack force F _{Zst_gem} can be corrected if necessary by shares of centrifugal _forces , with which a camber change based on the steering angle change c _γ , the inertia Θ _wheel , a speed of dδ / dt steering angle, a speed n _wheel and the effective Tracking lever h _{w is} taken into account. Taking into account all the correction _factors, the actual rack force F _{results Zst_akt} equation (G2):

The constructive rack force F _{Zst_akt_k} , as equation (G3) shows, calculated from an axle kinematics and is proportional to the side force F _Sv , where k _{Sv describes} a slope and thus a ratio of the constructive rack force F _{Zst_akt_k} to the side force F _Sv . This is done on the straights 40 . 56 from the 2 and 3 referenced, wherein instead of the lateral force F _Sv a proportional lateral acceleration is taken into account. The constructive rack-and-pinion force F _{Zst_akt_k} is also present when tires are in full sliding slip and thus the dynamic rack-and-pinion force is zero: F _{Zst_akt_k} = k _Sv · F _Sv (G3)

From the lateral force F _Sv or lateral acceleration a _y , a reference rack force F _{Zst_ref is} calculated via equation (G4), which describes a straight line. The slope m _akt the straight line varies over the tire pressure and the temperature, which correction factors can be specified. A parameter b, which can also be taught in accordance with the equation (G4), describes an offset or minimum value that the reference rack force F _{Zst_ref} also has when the side force F _{Sv is} zero: F _{Zst_ref} = m _act · F _Sv + b (G4)

Furthermore, in accordance with equations (G5.1) and (G5.1), first the mean value of all determined rack forces and of all determined lateral forces is formed:

In a subsequent step, according to equation (G6.1), the deviations of all determined, time-dependent values of the rack-and-pinion force are formed by the mean value of the rack-and-pinion force: ΔF _{Zst_akt} (t) = F _{Zst_akt} (t) - F _{Zst_akt_mittel} (G6.1)

Correspondingly, according to the equation (G6.2), the deviations of all determined time-dependent values of the lateral force are formed by the mean value of the lateral force: ΔF _{Sv_akt} (t) = F _{Sv_akt} (t) - F _{Sv_akt_mittel} (G6.2)

These values are used to determine the slope m _{akt of} a regression line according to equation (G7):

Thereafter, in a further step, a value for the offset b is determined for which equation (G8) holds: b = F _{Zst_akt_mittel} - m _akt · F _{Sv_akt_mittel} (G8)

Examples of the reference rack force F _{Zst_ref} are based on the straight line 110 . 112 represented with a respective slope m _act . With this reference rack force F _{Zst_ref} , a measured rack force F _{Zst_gem} or a current rack force F _{Zst_akt} derived _{therefrom is} adjusted.

The maximum possible lateral force F _{Sv_max} is calculated via equation (G9). It is assumed that this is achieved when the actual rack force F _{Zst_akt has} the value of the constructive rack force F _{Zst_akt_k} . Under this assumption, equation (G9) yields:

Here is another slope a _Sv to take into account, such as the straight lines 44 . 58 . 60 from the 2 and 3 exhibit. As with the preceding description of the 2 and 3 cited, these are straight lines 44 . 58 . 60 in an embodiment by F _{Zst_ref} and F _{Zst_akt} = a _Sv * F _Sv + c, where c corresponds to another offset. In this case, the slope a _{Sv describes} a proportionality between the reference rack force F _{Zst_ref} and the side force F _Sv for values of the side force F _Sv that are greater than a limit value provided for this, at which the reference rack force F _{Zst_ref reaches} a maximum value.

Equation (G10) is used to calculate a hazard indicator GI which is used to generate no false warning with slight deviations of the measured, actual rack force F _{Sv_akt from} the reference rack force F _{Zst_ref} .

If the hazard indicator GI falls below a limit specified for this, which is approximately 0.8 depending on the measurement accuracy, an estimate is made for a coefficient of friction μ, which is calculated using equation (G11):

Here F _{Nv describes} one on the wheels 6 . 8th the axis 4 acting normal force.

If the motor vehicle is changed over from summer tires (SR) to winter tires (WR) without new learning of the gradient m _akt , a false warning is permanently provided if the danger indicator GI is less than the limit value. This can be avoided if for the inside of the ellipse 114 arranged area in 5 a new slope m _akt is trained. Here, equation (G10) is used.

With equation (G10) can z. B. in the range of a lateral acceleration of 0.5 to 2 m / s ² or a corresponding lateral force F _Sv a new slope m _{akt be} learned for the line.

Learning and / or adaptation is carried out within the scope of the method, wherein the gradient m _{akt is} averaged over several cornering movements without the driver of the motor vehicle 2 of something noticed. Here, for the sake of simplification, it is taken into account that the slopes m _{akt of} the straight lines barely differ in a range of the friction coefficient μ of 0.5 to 1.2. Thus, knowledge of a precise condition of the roadway is not required.

In 5 you can see that the straights 110 . 112 for different tire configurations of the chassis have different slopes m _akt .

For left and right curves, the slope m _akt and the offset b can be determined separately. If this results in different values for left and right turns, it can be different settings of Radführungsgrößen, such. B. lane and camber, to different tire conditions, such. B. type, tread depth and pressure of a tire on a left and a right side of the motor vehicle and unilateral loading conditions are closed as operating parameters of the motor vehicle.

In this way, new procedures for determining these operating parameters or variables of the motor vehicle are generated in an embodiment.

In the above presented first embodiment of the method is based on simplifications, taking into account that for all operating parameters signals, via which these the control unit 16 be provided available.

However, it is possible that the control unit 16 not all described operating parameters are available. Thus, it is conceivable that in addition to the usually estimated rack force only one each struck steering angle of at least one wheel 6 . 8th and a longitudinal speed of the motor vehicle are available, whereas signals for the lateral acceleration and the yaw rate are not present.

That already on the basis of 1 presented control unit 16 as a component of the system according to the invention is for carrying out the above-described first embodiment of the method but also for carrying out the second embodiment of the inventive method described below provided. A possible example of such a control unit 16a , which is designed here as an EPS control unit is in 6 shown schematically. In this case, this controller includes 16a a module 120 , which is designed to estimate a coefficient of friction μ between tires and a traveled lane taking into account a simulation model and in 6 is shown as a block diagram. In the schematic diagram 6 includes this module 120 several submodules 122 . 124 . 126 . 128 . 130 which are designed to determine and / or process individual operating parameters with the simulation model, which are taken into account in at least one embodiment of the method according to the invention.

Here is a first sub-module 122 for calculating and / or determining a yaw rate dψ / dt, wherein this yaw rate dψ / dt is to be calculated as a function f, which is dependent on a speed v, a steering angle δ and a factor K ₁ . In this case, values for the speed v and the steering angle δ the control unit 16a provided by signals from at least one sensor. A map for the factor K ₁ is in the first sub-module 122 stored. A second submodule 124 is for determining a lateral acceleration a _{y of} the motor vehicle 2 formed, wherein this lateral acceleration a _y as a function of the speed v, whose values the control unit 16a from which at least one sensor is provided, as well as calculated and / or determined values of the yaw rate dψ / dt, that of the first sub-module 122 be provided, is representable.

A third submodule 126 is for determining set values of a side force F _Sv and values of a target side force F _{Sv_soll} which are on the motor vehicle 2 acts, educated. The values for the target lateral force F _{Sv_soll} as a function of the steering angle δ and the lateral acceleration a _{y are} to be represented. Values for a reference rack force F _{Zst_ref} are with a fourth submodule 128 to calculate. Here, the reference rack force F _{Zst_ref} as a function of the target lateral force F _{Sv_soll} and a second factor K ₂ can be _displayed . Values for the target side force F _{Sv_soll} become the fourth submodule 128 from the third submodule 126 provided. Values for the factor K ₂ are in another characteristic field, the fourth submodule 128 is assigned, filed.

Values for the friction coefficient μ are calculated using the third submodule 130 determined. Here, the friction coefficient μ as a function of the reference rack force F _{Zst_ref} , the measured rack force F _{Zst_gem} and a query A can be _displayed . To calculate the coefficient of friction μ, the fifth submodule 130 Values for the reference rack force F _{Zst_ref} from the fourth submodule 128 provided. Values for the measured rack force F _{Zst_gem} are detected by the at least one sensor and the control unit 16a provided. The query A can be represented as a function of the speed v, usually the longitudinal speed, the steering wheel angle and / or the steering angle δ, the yaw rate dψ / dt and the lateral acceleration a _y .

In the second embodiment of the method, a set of operating parameters is used to calculate the coefficient of friction μ. The following applies for a function to describe the friction coefficient μ either μ = f _(y a, dψ / dt, v, F _Zst) or μ = f _(y a, dψ / dt, v, M _re). Here, in addition to the operating parameters M already described _above, a restoring moment must be taken into account. The proposed set and thus a corresponding function for calculating the coefficient of friction μ is further reduced to μ = f (δ, v, F _Zst ).

The two each in 7 as in 8th Illustrated diagrams each include an abscissa 132 along which the time t is plotted in seconds. Along an ordinate 134 A first of the two diagrams plots values for a force. Along an ordinate 136 A second of the two diagrams plots values for a coefficient of friction μ.

In the first diagram 7 shows a first curve 138 a time course of values of a measured rack force F _{Zst_gem} and a second curve 140 a course of a reference rack force F _{Zst_ref} . A difference between the measured values of the rack force and the values of the rack force ΔF = F _{Zst_ref} -F _{Zst_gem} provided as reference values is shown by a third curve 142 in the first diagram 7 shown. A second diagram 7 shows a curve 144 for a constant course of a high, maximum coefficient of friction μ = 1, which is also referred to here as a high friction coefficient.

The measured rack force and the reference rack force are approximately equal. No reduction coefficient is detected.

The first diagram 8th includes a first curve 146 for a time course of values of a measured rack force F _{Zst_gem} and a second curve 148 for a course of a reference rack force F _{Zst_ref} . A difference between the measured values of the rack force and the values of the rack force ΔF = F _{Zst_ref} -F _{Zst_gem} provided as reference values is shown in the first diagram 8th on the basis of a third curve 150 shown. A second diagram 8th includes a curve 152 for a course of a friction coefficient μ. In regions with peaks where the reference rack force is significantly greater than the measured rack force value, a reduced coefficient of friction μ = 0 is detected. In the areas between the peaks no Niederreibungsbeiwert, but a Hochreibungsbeiwert or maximum coefficient of friction μ = 1 is detected because the queries regarding a minimum steering angle prohibit the output. An associated misdiagnosis to an unrecognized, low coefficient of friction is of no importance, since such a low coefficient of friction is required neither in situations with very low speeds, nor at very low steering angles.

The curves 138 . 140 . 142 . 144 in the diagrams 7 arise based on operating parameters of a first measurement, whereas the curves 146 . 148 . 150 . 152 of the diagrams 8th result from a second measurement. Based on the diagrams, an operation of the method is explained.

In the second embodiment of the method is within the controller 16a , which is designed as an EPS steering controller, a reference value for the rack force determined solely from the steering wheel angle and the speed. With an algorithm used here, a reference value for the rack-and-pinion force to be determined for a roadway having a high coefficient of friction is compared with a value of the currently measured rack-and-pinion force, the coefficient of friction being estimated taking into account a current driving state of the motor vehicle.

It is taken into account that both the rack-and-pinion force as well as the lateral force on the wheels increase with increasing lateral acceleration in uncritical driving conditions. In addition, on a lane with high or maximum coefficient of friction ( 7 ) Amounts of the forces or accelerations to be taken into account, ie at least the rack force and the lateral acceleration, greater than on a roadway with a low or minimal coefficient of friction ( 8th ). To determine the friction coefficient, a reference value for the rack force is first formed from a theoretically determined side force with the aid of a correction function. The correction function is determined by means of reference measurements on high-friction road surfaces and adjusts a lateral force target value to the measured value of the rack-and-pinion force. The determined reference value of the rack-and-pinion force is compared with the currently measured rack-and-pinion force. When an amount of the value of the actual rack force is substantially lower than an amount of a theoretically determined reference value of the rack force, a reduction in the coefficient of friction is inferred. An associated detection of a reduction coefficient is thus attributed to the difference between the current value and the reference value of the rack force. Another possibility is to consider a ratio of the current value and the reference value to one another and / or a percentage deviation of the current value from the reference value.

The determination of the theoretical reference value for the rack-and-pinion force is carried out with the aid of a simulation model, for example a single-track model, exclusively on the basis of steering-internal signals for operating parameters, ie for the speed, the steering angle and the geometry of the motor vehicle. In this case, no incoming signals are required for values of the lateral acceleration and the yaw rate to be taken into account as further operating parameters, since values for these operating parameters are determined within the scope of FIG 6 be approximated to the presented simulation model. Consequently, the simulation model is about the submodules 122 . 124 . 126 . 128 . 130 out 6 is to implement, for estimating the coefficient of friction easily, since this only small computing capacity within the controller 16a required are.

Due to the fact that incoming signals are only available for two operating parameters, here for the speed v and the steering angle δ, some simplifying assumptions can be made for the simulation model. Thus, an approximate calculation of the yaw rate is first made on the assumption that a skew angle difference between the front axle and the rear axle is zero. An influence of the skew angle difference is then taken into account via a further correction function, which is to be used analogously to that reference rack force. In addition, an influence of the axial shift due to a longitudinal acceleration of the motor vehicle and an influence of inertia and / or yaw acceleration about the vertical axis of the motor vehicle to the reference rack force are neglected, especially since the signals required for this purpose are also not available. The influences are approximately contained in the correction function. On the basis of this, in an embodiment of the method, an approximated and / or reduced coefficient of friction, which depends only on a few further operating parameters, is identified.

In critical driving conditions, such as oversteer or understeer, a progression of rack force over lateral acceleration also decreases as the vehicle begins to slip. In order to avoid a misdiagnosis of the simulation model to a reduced coefficient of friction for this case, various queries are integrated into the simulation model. Thus, it is considered that an output of a reduced coefficient of friction is only permissible if the motor vehicle in a certain driving condition is located. The coefficient of friction is, for example, dependent on the speed, the steering angle, the estimated lateral acceleration and the estimated yaw rate. By a suitable choice of said operating parameters, the avoidance of misdiagnosis can be ensured.

This in 9 illustrated flowchart for carrying out the third embodiment of the method according to the invention comprises a plurality of steps 601 . 602 . 603 . 604 . 605 . 606 . 607 . 608 . 609 . 610 . 611 . 612 . 613 . 614 . 615 ,

It is in a first step 601 a start of this embodiment of the method feasible. Starting from this, in a second step 602 Values for operating parameters, namely for a speed of the motor vehicle in the x-direction v _x , a lateral acceleration of the motor vehicle a _y , a yaw rate of the motor vehicle dψ / dt, a steering wheel angle δ _H and a rack force F _Zst detected. It is possible in a third step 603 Values for operating parameters of the motor vehicle from a memory of the control unit 16 . 16a read. Based on the second step 602 recorded values of the operating parameters are in a fourth step 604 as further operating parameters so-called primary variables for spring travel f, pitch angle θ, roll angle φ, drive or braking forces F _L , the steering angle δ _{VA of} the front axle and wheel loads F _{N are} calculated.

Based on this, in a fifth step 605 Calculated values of an effective, frictional-dependent restoring torque of the tires of the motor vehicle. In a sixth step 606 an effective lateral force is calculated at the front axle of the motor vehicle. In a seventh step 607 Values or parameters, which are determined during an evaluation of the rack force, are adjusted. In an eighth step 608 an effective caster distance of the tires is calculated. In addition, in a ninth step 609 calculated a tracking distance, which would theoretically result on a track or road with a high coefficient of friction μ or a Hochreibungsbeiwert.

In a tenth step 610 an exclusion criterion must be checked. If an examination of this exclusion criterion is positive (yes), this results in a fourteenth step 614 in that no new reliable estimation of the frictional connection and thus of the coefficient of friction between the tires and the driven lane is possible. However, if this exclusion criterion in the tenth step 610 denied (no) becomes in a subsequent eleventh step 611 a value of the lateral acceleration a _y compared with at least one limit value. Alternatively, instead of the lateral acceleration, the lateral force acting on this lateral acceleration during stationary cornering can also be considered as a criterion. If the value of the lateral acceleration a _y z. B. is greater than 0.5 m / s ² or less than 1 m / s ² , the method at the seventh step 607 continued, the operating parameters for the evaluation of the rack force are adjusted again. However, if the lateral acceleration a _y z. B. is greater than or equal to 1 m / s ² , is in the twelfth step 612 a comparison of a theoretical _{tracking distance} n _{R_theo performed} with values of an effective _{tracking distance} n _{R_wirk} . If in this twelfth step 612 shows that values of the theoretical trailing _distance n _{R_theo are} smaller than values of the effective trailing _distance n _{R_wirk} , the method also goes to the seventh step 607 reset. However, if the values for the theoretical caster _distance n _{R_theo} and the effective caster _distance n _{R_wirk are} approximately equal, the result is a thirteenth step 613 in that the lateral acceleration a _{y is} currently too low, whereupon the fourteenth step 614 is carried out. If, during the comparison, the twelfth step 612 however, results in that values of the theoretical caster _distance n _{R_theo are} greater than those of the effective caster _distance n _{R_wirk} becomes in a final fifteenth step 615 a maximum usable adhesion and thus the coefficient of friction estimated.

The representation in 10 based on 1 and shows how a manual steering angle δ _H about which the steering wheel 13 is rotated manually, with a steering angle δ _{l of} the first, left wheel 6 and a steering angle δ _{r of} the second, right wheel. In the context of the method, a steering angle δ is taken into account as the operating parameter, which is usually of at least one steering angle δ ₁ and / or δ _{r of} one of the two wheels 6 . 8th is dependent. Accordingly, it is possible for the steering angle δ, for example, an average value δ _{VA, to be} the steering angle δ ₁ , δ _{r of} both wheels 6 . 8th corresponds, since these wheel-individual steering angle δ _l , δ _r may differ from each other. However, it is also possible within the scope of the method to consider the larger or the smaller of the two wheel-specific steering angles δ ₁ , δ _r as the steering angle.

The diagram 11 shows a vectorial and / or trigonometric relationship of individual operating parameters, the wheels when driving the motor vehicle 7a . 7b influence, with in 11 In this regard, without limitation of generality, a one-track model of the motor vehicle is considered, wherein the wheels each have an axis on a wheel arranged in the middle of the motor vehicle 7a . 7b be reduced. The steering angle δ _{VA of} the front wheel 7a corresponds to the mean of the steering angle of the two bikes 6 . 8th on the front axle 4 out 1 , A slip angle β gives the angle between the speed of a center of gravity 9 of the motor vehicle and a median plane of the motor vehicle. It acts on the front wheel 7a that the wheels 6 . 8th replaced in the model, a front side force F _Sv . On the focus 9 of the motor vehicle acts a lateral acceleration a _x in the longitudinal direction and a transverse acceleration a _y in the transverse direction. The rear wheel 7b , which replaces the wheels on a rear axle in the model, points from the center of gravity 9 a distance l _h . It also acts on the rear wheel 7b a rear side force F _Sh .

Details for calculating the effective frictional-dependent restoring torque according to the fifth step 605 out 9 are out in the flow chart 12 shown. This includes a first step 200 in which a division of the restoring moments in a frictional- _dependent portion and other shares, for example. A left restoring moment M _konl and a right restoring moment M _konr , made. In a second step 202 is a to the wheels on the left and right track levers effective _wl h, h _wr as a function of steering wheel angle δ _H is determined. In a third step 204 a ratio of the friction- _{dependent restoring moments of} the tires M _rμl / M _{rμr is determined} as a function of the lateral acceleration. In a fourth step 206 are determined proportions of the rack force F _{Zst_kon} , which are caused by shares of the restoring _moments M _konl , M _konr . In a fifth step 208 Values of the effective rack force F _Zst are provided. In a sixth step, values for the lateral acceleration a _y and in a seventh step 212 Values for the steering wheel angle δ _H provided.

Based on the fourth step 206 certain proportion of the rack force F _{Zst_kon} and the fifth step 208 provided, measured values of the rack force F _Zst is in an eighth step 214 a proportion of the rack force F _ZSRμ , which is caused by frictional- _{dependent restoring moments} calculated. On the basis of this calculated portion of rack power as well as in the steps 210 and 212 provided values for the lateral acceleration a _y and the steering wheel angle δ _H is in a ninth step 216 an effective track lever length h _w calculated as a function of the steering wheel angle δ _H and the lateral acceleration a _y . A final calculation of the effective traction-dependent tire return torque M _Rμges is made in a final tenth step 218 calculated on the basis of the calculated portion of the rack force and the calculated effective track lever length.

The diagram 13 is based on the diagram 10 and additionally shows how a rack-and-pinion force F _Zst acting in the left-hand direction acts on a track lever length h _{wl of} the left-hand wheel 6 and a track lever length h _{wr of} the right wheel 8th effect.

The diagram 14 includes an ordinate 250 along which values for lateral force are plotted on the axle formed as the front axle 4 of the motor vehicle 2 acts. Along an ordinate 252 of the diagram 14 are values for an entire overrun of tires that are on wheels 6 . 8th of the motor vehicle 2 are applied, applied. Inside the diagram 14 is a family of elliptical curves 254 . 256 . 258 . 260 . 262 . 264 . 266 . 268 . 270 . 272 . 274 . 276 for the side force dependent overrun distance for different values of a coefficient of friction. This results in a first curve 254 for a minimum value of the coefficient of friction and a last, twelfth curve 276 for a maximum value of the coefficient of friction. This results in the first curve 254 with a coefficient of friction μ = 0.1. A second turn 256 results in a coefficient of friction μ = 0.2, a third curve 258 results in a coefficient of friction μ = 0.3, a fourth curve 260 results in a friction coefficient μ = 0.4, a fifth curve 262 results in a coefficient of friction μ = 0.5, a sixth curve 264 results in a coefficient of friction μ = 0.6, a seventh curve 266 results at a coefficient of friction μ = 0.7, an eighth curve 268 results in a friction coefficient μ = 0.8, a ninth curve 270 results in a coefficient of friction μ = 0.9, a tenth curve 272 results in a coefficient of friction μ = 1.0 and an eleventh curve 274 results at a coefficient of friction μ = 1.1. The last, twelfth curve 276 results at a coefficient of friction μ = 1.2.

A straight 281 describes the theoretically acting on a high coefficient of friction overrun distance. Also, in the diagram is off 14 a first point 278 for an estimated maximum lateral force and a second point 280 entered for a measured value. The diagram 14 further includes a dashed line 282 parallel to the line 281 is.

As in 14 can be seen, the straight line approaches 281 a first section of all elliptical curves 254 . 256 . 258 . 260 . 262 . 264 . 266 . 268 . 270 . 272 . 274 . 276 very good. This first section corresponds in each case to a low utilization of the available friction coefficient with low lateral force. The slope of the line 281 , which alternatively can also be replaced by a curve, is dependent on the motor vehicle and can assume positive values or, as shown here, negative values.

In the first embodiment of the method, the slope of the straight line 281 set to zero. In this case, the restoring torque and the side force are proportional to each other, which can usually represent only an approximation, at least for large values for the lateral force and the lateral acceleration. In this first embodiment of the method also effects of the steering kinematics are neglected and assumed a linear relationship between the rack force and the restoring moment. As a result, a linear relationship between lateral force and rack force is then also obtained at a constant follow-up distance, on the basis of which, according to equation (G7), the learnable gradient m _{akt is} determined.

However, since a second linearization leads to negligible errors in most motor vehicles only in the case of small steering angles, which are common in overland driving, the first embodiment of the method represents a special case of the third embodiment. The first embodiment is easier to implement, but limits the scope and the accuracy of the procedure.

The diagram 14 Thus, for example, shows the relationship between the lateral force and the effective trailing distance as a function of the coefficient of friction. With this diagram, the maximum possible side force and the currently available maximum friction coefficient are determined from a value pair from the lateral force at the front axle F _Sv and the effective trailing distance n _RVA . Since, for example, a road as a road surface has unevenness, the resulting fluctuations in the wheel load lead to fluctuations in the restoring torque and thus also to fluctuations in the determined overrun distance. In order not to interpret fluctuations of the trailing distance in the range of low values as a lack of adhesion, a tolerance threshold is provided, which is shown in the diagram 14 through the dashed line 282 given is. This line 282 is a preselected value below the parallel lines 281 , which describes the theoretically expected at a high coefficient of friction lag.

The flowchart off 15 shows further steps 302 . 303 . 304 . 308 . 310 . 312 . 316 . 318 for calculating an effective return torque. This will be done in a first step 302 a to the wheels left and right effective toe lever h _wl , h _{wr determined} as a function of the steering wheel angle δ _H. In a second step 303 dynamic wheel loads of the front wheels are determined as a function of the lateral acceleration. In a third step 304 a ratio of the restoring moments of the tires M _l / M _{r is determined} as a function of the lateral acceleration. In a fourth step 308 Values of a rack force F _Zst are provided. In a fifth step 310 become values for the lateral acceleration a _y and in a sixth step 312 Values for the steering wheel angle δ _H provided. Based on in the steps 310 and 312 Provided values for the lateral acceleration and the steering wheel angle is in a seventh step 316 an effective track lever length h _{w is} calculated as a function of the steering wheel angle and the lateral acceleration. A calculation of the effective restoring torque M _Rμges is in a final eighth step 318 calculated on the basis of the calculated portion of the rack force and the calculated effective track lever length.

When carrying out the third embodiment of the method, a higher detection quality can be achieved for the operating parameters, wherein u. a. a kinematics and a follow-up behavior of the tires are taken into account.

Taking advantage of detailed knowledge of a kinematics of the steering system 10 and thus a kinematics of the rack 12 , the axis 4 , the wheels 6 . 8th and the tire, it is possible to dispense with the simplistic linearization for determining the rack force provided in the first embodiment and replace it with more accurate modeling. This also makes a more accurate estimate of the coefficient of friction possible.

The operating parameters or measured variables which are standard available in a motor vehicle with vehicle dynamics control, d. H. the longitudinal acceleration, the lateral acceleration and possibly a vertical acceleration, the steering angle, the yaw rate, the speed and the rack force, which is determined as already described, are sufficient for the third embodiment of the method.

In the second step 602 of the method, as shown in the flowchart 9 represented, operating parameters or parameters for the longitudinal, lateral and vertical acceleration, for the steering angle, for the yaw rate and for the driving speed, if any, detected. In the fourth step 604 are determined as operating parameters or primary variables spring travel, pitch angle, roll angle, drive or braking forces and wheel loads. In this case, only a part of the above-mentioned measured quantities is sufficient to approximately determine the primary quantities.

Due to the lateral acceleration a _y results in a dynamic Radlastverlagerung. The dynamic wheel load change ΔF _Nv during stationary cornering is thus a function of the lateral acceleration according to equation (G12): .DELTA.F _Nv ≈ f (a _y) (G12)

Simplified, it can also be assumed that the ratio of a rolling support of both axles of the motor vehicle, which is caused by axle kinematics, spring and stabilizer rates, corresponds to the ratio of the static axle load. If the effects of a shift of the center of gravity within a curve are also neglected, then from a height h _{s of} the center of gravity and a track width b _v, a change of the wheel load on an axis can be approximated using equation (G13):

Thus, a change in the wheel load is directly proportional to the lateral acceleration a _y . If a ratio of differences in the wheel loads of the front and rear axles during cornering, for example, from the simulation model or a simulation calculations known in the above equation (G13), the term l _h / l by the ratio of roll support of the axis formed as a front axle entire roll support of both axles of the motor vehicle to be replaced.

The total axle load of the front axle, neglecting aerodynamic forces as a function of the longitudinal acceleration a _x and gravitational acceleration g, is approximately determined by the following equation (G14):

und für die Radlast am rechten Rad 8 Gleichung (G16):

This results for the wheel load on the left wheel 6 Equation (G15):

and for the wheel load on the right wheel 8th Equation (G16):

Are other sensors or algorithms in the vehicle 2 available, which allow a more accurate estimate of the front axle load, alternatively, the thus determined axle load on the front axle can be used. In this case, as described above, the term l _h / l can also be replaced by the ratio of the roll support of the axle formed as a front axle to the total roll support of both axles.

Next, in a further step, taking advantage of the knowledge of the steering kinematics, the steering wheel angle δ _H and the two steering angle δ _l , δ _{r of} the wheels 6 . 8th the axis 4 is calculated via equation (G17) an average value δ _VA for the steering angle (cf. 10 ): δ _VA (δ _H ) = 0.5 · (δ _l (δ _H ) + δ _r (δ _H )) (G17) Here, too, the indices l and r stand for left and right.

From the lateral acceleration a _y , the yaw rate or yaw rate dψ / dt measured with the vehicle-internal sensor, and the average steering angle δ _{VA of} the wheels 6 . 8th will be the next step on the axis 4 acting lateral force neglecting vertical accelerations with equation (G18) (cf. 11 ) in the sixth step 606 of the method (flowchart from 9 ) certainly:

Here, m is the mass of the motor vehicle, l _{h is} the distance of the center of gravity of the motor vehicle 2 measured from its rear axle in the longitudinal direction of the motor vehicle and l the wheelbase of the motor vehicle 2 , A yaw acceleration d ² ψ / dt ² is determined from the yaw rate or yaw rate dψ / dt by deriving. Θ _z is the moment of inertia of the motor vehicle 2 around its vertical axis.

In the parallel fifth step 605 is determined from the rack force, the effective adhesion-dependent restoring torque of the tires. Details on this are in the diagram 12 shown.

In a steering knuckle steering, the inside wheel is hit stronger than the outside wheel (see. 13 ). This means that a ratio between a path of the rack s _Zst and a steering angle is dependent on a steering angle δ ₁ , δ _{r of} the respective wheel. This relationship can be described by the effective tracking levers h _Wl and h _Wr via equations (G19) and (G20): h _Wl = ds _Zst / dδ _l = f (δ _H ) (G19) h _Wr = ds _Zst / dδ _r = f (δ _H ) (G20)

For the wheel on the outside, an increase in the wheel load that occurs causes an approximately proportional increase in the wake of the tire and a larger proportion of the total transmitted lateral force. On the inside wheel, an approximately proportional reduction in the wake of the tire and a smaller proportion of the total transmitted side force is caused by an increase in the wheel load, ie, the two wheels contribute to the total restoring torque depending on the lateral acceleration to varying degrees. Since the effective steering levers vary in size depending on the steering angle, the restoring torques on both wheels have different effects on the rack-and-pinion force. For rack force, equation (G21) applies: F _Zst = M _l / h _Wl + M _r / h _Wr (G21)

The at the motor vehicle 2 Steering restoring torques M ₁ , M _r occurring in each case consist of the restoring torque and at least one of the following operating parameters, namely a lateral force in connection with the trailing distance of the tires or a tire trailing, a restoring moment of the tire due to fall, a lateral force in connection with a geometric trailing resp a geometric trailing distance, a wheel load associated with an effective Radlasthebelarm, a driving force associated with an effective Störkrafthebelarm, a braking force associated with an effective steering roller radius, an inertial moment, a Kreiselmoment and a friction torque.

In this case, only the first portion of the restoring moment is dependent on the _frictional connection and is hereinafter referred to as the left wheel with M _Rμl and for the right wheel with M _Rμr . Therefore, all other parts of the restoring torque while driving taking into account the known kinematics of the chassis and the steering system 10 from at least one in the motor vehicle 2 arranged sensor calculated. These portions of the return torque are with M _konl for a left side of the motor vehicle 2 and with M _konr for a right side of the motor vehicle 2 designated. This applies to the proportion of the restoring torque M _konl the left side of the motor vehicle 2 with external brake and conventional drive via a drive shaft Equation (G22): M = _KONl .DELTA.M (γ) + F · _S n + _kl F _Nl * h + F _{wFNl Al} · r + F _{wsl Bl} · r + Θ _{wrl zRadl} · d ² δ _l / dt ² + M + M _{kreil reil} ( G22)

The restoring moment M _konr for the right side of the motor vehicle 2 is calculated from corresponding operating parameters on the right side of the motor vehicle from equation (G23), where index I is to be swapped by index r: M _konr = ΔM (γ) + F _Sr ·n _kr + F _Nr ·h _wFNr + F _Ar · r _wsr + F _Br ·r _wrr + Θ _zRadr ·d ² δ _r / dt ² + M _circle + M _reir ( G23)

Here, ΔM (γ) describes the fall by the left or right wheel 6 . 8th caused tire return torque. The further variables from the equations (G22) and (G23) correspond to the lateral force F _S1 or F _Sr , the driving force F _Al and F _Ar and the braking force F _Bl and F _Br on the left and right wheels, respectively 6 . 8th , n _kl or n _kr is the geometric caster, h _wFNl or h _wFNr the effective with respect to Radrückstellmoments lever arm of the wheel load and r _wsl or r _wsr the effective Störkrafthebelarm and r _wrl or r _wrr the effective steering roller radius of the respective left or right wheel 6 . 8th , These operating parameters n _kl and n _kr , h _wFNl and h _wFNr , r _wsl and r _wsr , r _wrl and r _wrr are respectively from the spring travel f _l and f _r at the left and right wheel 6 . 8th , the roll angle φ, the pitch angle Θ and the steering angle δ _l , δ _r calculated. In a possible embodiment, only a part of the last-mentioned variables or operating parameters is used for the approximate calculation of the respective restoring torque M _konl or M _konrr .

In a further embodiment of the method, an effect of an elastokinematics can additionally be taken into account, with the above-mentioned operating parameters additionally being considered as a function of forces acting on the left and / or right wheel. Furthermore, a deformation of a respective tire, which is also dependent on the forces and moments acting on the left or right wheel and on the camber angle, can additionally be taken into account in the calculation of the above-mentioned operating parameters.

In addition, Θ _zRadl or Θ _{zRadr describes} the inertia or inertia of those components, including the left and right wheel 6 . 8th , the motor vehicle 2 when steering by the steering angle of the left or right wheel 6 . 8th to be moved. M _kreil , M _kreir is a _{gyroscopic moment} when steering on the left or right wheel 6 . 8th occurs when the kinematics causes a steering angle-dependent camber change. M _reir , M _reir is a friction torque on the left or right wheel 6 . 8th which is dependent on the speed of the steering angle left and right and the forces acting on the joints of the left and right wheel guide forces.

Depending on the accuracy requirement, only individual parts of the equations (G22) or (G23) can be taken into account.

To determine the at the individual wheels 6 . 8th acting lateral force is the lateral force F _Sv acting on the axle simplified to the two wheels 6 . 8th divided according to the distribution of the wheel load according to the equation (G24) for the left lateral force and the equation (G25) for the right lateral force: F _Sl ≈ F _Sv F F _Nl / F _Nv (G24) F _Sr ≈ F _Sv F F _No / F _Nv (G25)

Alternatively, a wheel-dependent division of the lateral force can also be considered as a function of the steering angle, the acting lateral force or other driving dynamic operating parameters or variables, for example the lateral acceleration.

With the restoring _moments M _konl and M _konr determined from the equations (G22) and (G23), the portion of the rack-and-pinion force which is not generated by _{restraint-dependent} restoring _{moments of} the tires and is independent of this is designated as F _{Zst_kon} and via equation (G26) certainly:

This constructive portion of the rack-and-pinion _{force is} deducted from the acting, current and / or measured rack-and-pinion force F _{Zst_gem} , from which equation (G27) the rack-and-pinion force is calculated, which is determined by the _{traction-dependent} restoring torque of the tires 6 . 8th causing:

From this, equation (G28) is used to calculate the mean track lever h _W :

Under the simplified assumption that the lateral forces are distributed according to the dynamic wheel loads and the influence of the wheel load on the caster distance is neglected, with an operating parameter F _NVA for an axle load or a normal force acting on the front axle for the effective toe lever equation (G29):

To increase the accuracy, a function (G30) expressing the ratio of the moments as a function of a change in the wheel load and thus the lateral acceleration can also be used: M _l / M _r = f (F _Nl / F _Nr) = f (a _y) (G30)

In the last step 218 in the diagram 12 the total return torque is determined from the rack force or from the difference between the two tie rod forces via equation (G31): M _Rμges = F _{Zst_Rμ} * h _W (G31)

The restoring moment thus obtained is in sufficient accuracy only dependent on the side force transmitted to the axle and the frictional connection. Such a two-dimensional dependence can be described either by a mathematical function or by a characteristic map and derived from a fundamental behavior of the restoring torque of a tire. The restoring moment of the respective tire arises from the fact that the lateral force acts eccentrically offset by a trailing distance of the tire to the rear. Therefore, the effective trailing distance is calculated from this restoring moment and the effective lateral force, which is the eighth step 608 out 9 is indicated. For the effective follow-up distance n _RVA , equation (G32) applies:

The reference here is in the ninth step 609 out 9 determines the trailing distance on a route with a high coefficient of friction μ as a function of the acting lateral force. This function is in the control unit 16 . 16a of the motor vehicle 2 deposited and is continuously tested. Will now the motor vehicle 2 loaded or reduced internal pressure of the tire, the restoring torque increases. Since the internal pressure may be known by a sensor for detecting the internal pressure of the tire, it is also possible to detect and / or determine a load of the axle by measuring the overrun distance with low lateral force. Theoretically, the recognition of a one-sided loading is also conceivable, since the trailing paths in left and right curves differ slightly from each other. An offset of the measured rack force or lateral acceleration applied to the controller 16 . 16a provided via a signal can also be detected and corrected to some extent. In this case, with a calculated zero crossing of the lateral force, the restoring torque measured via the rack-and-pinion force is not equal to zero, from which it follows that the trailing distance theoretically approaches infinity. Since a range of lateral acceleration with a value of +/- 1 m / s ^{2 is} often used, the offset of these two operating parameters can be recognized by observing this range and continue to correct if it is known by further estimates, which value of an operating parameter Offset caused in a respective zero crossing. This will be in the seventh step 607 for adjusting the parameters of the rack force evaluation in the diagram 9 , which will be described in more detail below.

The above equation (G32) for determining the caster distance may give erroneous results in certain operating conditions, if simultaneously over the steered wheels 6 . 8th a large longitudinal force is transmitted, since the wheel load and the fall when cornering relative to the roadway on the two wheels 6 . 8th the axis 4 are different in size, resulting in the longitudinal force for the wheels 6 . 8th different sized restoring moments occur. The differences that result from different effective Störkrafthebelarme are indeed taken into account, but are not responsible for this effect alone. Furthermore, the longitudinal force causes a shift of the load in the longitudinal direction and thus also changes the tire slip or the trailing distance.

In addition, erroneous results can be provided by equation (G32) when the motor vehicle 2 At the same time, it accelerates or decelerates in an operating situation and thus changes a dynamic axle load. The maximum lateral force and footprint of a wheel 6 . 8th increase almost proportionally to an increase in wheel load. The trailing distance as a function of lateral force also increases due to the larger footprint. Furthermore, it may be necessary to take into account that the effective geometric caster and thus the geometric caster distance are slightly changed by the pitch angle.

If the steering system 10 is operated very dynamically in an operating situation, caused by inertial forces of the wheels 6 . 8th of wheel carriers and due to centrifugal forces of the rotating wheels 6 . 8th additional moments, which theoretically can be determined relatively accurately. At a high compared to the driving speed steering speed results in a non-negligible, delayed build-up of lateral force, which can also lead to a change in the caster distance.

At very low lateral accelerations, the numerator and denominator of the equation (G32) result in very small values. Accordingly, small sensor errors and time-offset fluctuations can lead to large deviations in the trailing distance.

To account for the above errors, an estimate of the coefficient of friction will only be trusted if done in step tenth 610 there is no exclusion criterion (no). In this case, values of the drive torque on the axle are used, for example, to evaluate the exclusion criterion 4 , an acceleration d ² δ / dt ^{2 of} the steering angle δ and a derivative of the acceleration d ² δ / dt ^{2 of} the steering angle δ considered, which is to avoid that at a strongly fluctuating acceleration d ² δ / dt ^{2 of} the steering angle δ zero crossings than validly considered. As a further operating parameter for this purpose, a ratio of the speed dδ / dt of the steering angle δ to the speed v _{x of} the motor vehicle 2 and the brake pressure is considered.

If a value of one of the above-mentioned operating parameters is greater than a respective limit value to be defined and / or defined for this purpose, the exclusion criterion is met (yes). In addition, an estimate of the friction coefficient is only performed if the speed is greater than a minimum value provided for this purpose, so that inter alia the resolution of the yaw rate or yaw rate is sufficiently accurate, and if there is no control intervention by a chassis control system.

If there is no exclusion criterion, the lateral acceleration or the lateral force on the axis is considered. If a value of the lateral acceleration a _y , which corresponds to the side force during stationary cornering, is present in a range of, for example, 0.5 to 1 m / s ² , as in the flowchart 9 indicated, the determined trailing distance is used for learning or adjusting the operating parameters for the evaluation of the rack force.

If, for example, the value of the lateral acceleration is greater than 1 m / s ² , the twelfth step 612 compared the actual effective overrun distance with the theoretically effective overrun distance. If values of these trajectories deviate from one another by at most a predetermined minimum tolerance value, so that the values are approximately equal, a current frictional coefficient and thus a coefficient of friction are so much below the available frictional coefficient that no estimation is possible, which is the fourteenth step 614 is indicated. If the effective trailing distance is greater than the theoretical trailing distance at least by the tolerance value, then values of operating parameters of the motor vehicle, such as, for example, As its loading, not those intended for this purpose values that have been used to determine the theoretically effective follow-up distance. In this case, according to the seventh step 607 the thus determined trailing distance used to adjust the operating parameters for the evaluation of the rack force. If, on the other hand, the value of the effective follow-up distance is at least smaller than the theoretically effective value by the tolerance value, this value is used to estimate the maximum usable adhesion or friction coefficient in the fifteenth step 615 used.

In the diagram 14 are the elliptical curves 254 . 256 . 258 . 260 . 262 . 264 . 266 . 268 . 270 . 272 . 274 . 276 plotted for the caster distance over the effective lateral force, with different curves for different values of the coefficient of friction and thus of the frictional connection 254 . 256 . 258 . 260 . 262 . 264 . 266 . 268 . 270 . 272 . 274 . 276 result. This dependence of the friction coefficient and / or adhesion on the effective lateral force can be represented either by mathematical functions or by maps. Furthermore, in the diagram 14 theoretically to a Hochreibungsbeiwert, ie for a high coefficient of friction, effective tracking distance through the line 281 along the front portions of the elliptical curves 254 . 256 . 258 . 260 . 262 . 264 . 266 . 268 . 270 . 272 . 274 . 276 for small values of lateral force approximated. A dashed line 282 is offset by a tolerance threshold below this line 281 , this line 282 Also referred to as limit curve. Is the determined effective trailing distance below the dashed line 282 and if there is no exclusion criterion, a frictional connection and thus the coefficient of friction is estimated.

The point 280 results from the currently active trailing distance and the currently acting lateral force with a coefficient of friction μ of approx. 0.78. The point 278 applies to the same adhesion when a value of the trailing distance is zero and an available adhesion coefficient is fully utilized at the coefficient of friction of 0.78. The point 278 Readable lateral force corresponds to the maximum possible lateral force F _Svmax which can be easily calculated using equation (G33) for this estimated adhesion or coefficient of friction μ:

In a motor vehicle with a rear-wheel drive, an estimate of the coefficient of friction or frictional connection is also possible during acceleration processes in a curve, unless strong yawing moments are generated by a limited-slip differential. During the acceleration process with the acceleration a _x , this changes a dynamic axle load on the front axle.

In order to obtain a satisfactory estimation quality also in this case, the dynamic change is also taken into account in the calculation of the effective return torque and the theoretically effective at high coefficient of friction μ tracking distance. Here, in the same way as in the division of the restoring moments on the two wheels 6 . 8th the axis 4 proceed.

By changing the ratio of the axle load on the front axle to a pressure of the tires on the wheels 6 . 8th , which are arranged on the front axle, the overtravel of the tires changes approximately proportionally. This shifts the line, which describes the behavior with large values of the coefficient of friction, and thus also the line 282 (see. 14 ).

If the motor vehicle 2 already has a static roll angle, the z. B. may be due to a one-sided load, so even when driving straight without lateral inclination a lateral acceleration is sensed. Values of up to approx. 0.1 m / s ^{2 are} conceivable for lateral acceleration at normal spring rates. This leads to the fact that the determined trailing distance is extremely falsified, especially at low values of the lateral acceleration. If, for example, the motor vehicle has a static roll angle which occurs in a left turn, slightly too low values are determined for the follow-up distance in the left turn, since the actually acting side force is less than the lateral force calculated via the lateral acceleration. In a right-hand turn, on the other hand, too large values are determined for the trailing edge, since the lateral acceleration and thus also the acting lateral force are greater in magnitude than those detected by the sensor for detecting the lateral acceleration, usually calculated. To avoid these errors, a training of these changes is provided, which has already been explained above with reference to the first embodiment presented first.

The learning becomes at occurring lateral accelerations for a left turn in a range of about 0.5 m / s ² to about 1 m / s ² and for a right turn from about -0.5 m / s ² to about -1 m / s ² performed.

For pairs of values for a ratio of the rack force F _{Zst_Rμ} and the side force F _Sv in at least one of the above-mentioned ranges for the lateral acceleration respectively measured at the same time points becomes continuous or after reaching a preselected number of pairs of values for these two Operating parameters according to the method of least square error according to equation (G34) a compensation curve and determined by the measured values: F _{Zst_Rμ} ≈ A _l * F _Sv + ... + A _l * F l / Sv + | F _Sv | * (A ₂ * F 2 / Sv + ... + A _m * F m / Sv) (G34) {l ≤ n, for odd values and m ≤ n for even values}

Here, n should be at least 1, but more accurate results will be obtained for values of n that are 2 or greater.

The value b is subtracted from the calculated rack force to correct a possible offset. From the value A ₁ , the theoretical caster distance at a lateral force F _Sv equal to zero is calculated according to equation (G35): n _Rtheor (F _Sv = 0) = A ₁ · h _W (G35)

This value is used to adapt the theoretical tracking distance to changed operating parameters of the motor vehicle 2 and to adjust the height of the line 282 as a limit curve in the diagram 14 used.

When determining the compensation curve, at least a part of the coefficients A _n can be specified for n ≥ 2, so that in an embodiment all the coefficients A _{n are also} specified. By this measure, at a high adhesion and / or coefficient of friction μ a to a restoring moment of the respective tire 6 . 8th to achieve a suitable course.

Shows itself when comparing the overrun sections in the twelfth step 612 in that values of the measured trailing distance are greater than values of the theoretical trailing distance, these values can be used to correct the limit curve or line 282 be used at higher lateral forces F _Sv .

QUOTES INCLUDE IN THE DESCRIPTION

This list of the documents listed by the applicant has been generated automatically and is included solely for the better information of the reader. The list is not part of the German patent or utility model application. The DPMA assumes no liability for any errors or omissions.

Cited patent literature

• DE 102009002245 A1 [0003]
• DE 102006036751 A1 [0004]
• DE 102010036638 A1 [0005, 0008]
• DE 10319662 A1 [0008]

Claims (20)

Method for operating a motor vehicle ( 2 ), which has a steering system ( 10 ) with a rack ( 12 ), with the wheels ( 6 . 8th ) on an axis ( 4 ) on which tires are mounted, including associated suspensions of the motor vehicle ( 2 ), wherein the motor vehicle ( 2 ) is moved on a roadway, wherein a coefficient of friction or a value derived therefrom between a respective tire and the roadway taking into account a driving state, a kinematics of the steering system ( 10 ), a kinematics of the axis ( 4 ), a behavior of the tires on the wheels ( 6 . 8th ) and / or whose trailing path is determined as a characteristic variable, wherein in the course of an application of the method changing operating parameters of the motor vehicle ( 2 ) and the tire through a learning phase during at least one journey of the motor vehicle ( 2 ) is performed, recognized and taken into account. Method according to Claim 1, in which the coefficient of friction or the quantity derived therefrom between a respective tire and the roadway is determined from measured forces and / or torques in the axle ( 4 ), the steering system ( 10 ) and / or the tire as well as from driving dynamic operating parameters of the motor vehicle ( 2 ) is determined. Method according to Claim 1 or 2, in which the coefficient of friction is calculated as a function which depends on at least one further operating parameter, wherein as the at least one further operating parameter a steering angle, a transverse acceleration which is applied to the motor vehicle ( 2 ), a yaw rate of the motor vehicle ( 2 ), a speed of the motor vehicle ( 2 ), one on the rack ( 12 ) acting rack force, a restoring moment of at least one tire and / or a steering angle of at least one wheel ( 6 . 8th ) is used. Method according to one of the preceding claims, in which a current, measured rack-and-pinion force in a first portion, which determines the trailing distance of the tires on the wheels ( 6 . 8th ), and in a second, constructive portion, the kinematics of the axis ( 4 ) is divided. Method according to one of the preceding claims, in which the trailing path on a line with a high coefficient of friction as a function of the vehicle ( 2 ) acting side force F _{Sv is} determined. Method according to one of the preceding claims, in which a dynamic wheel load for at least one front wheel ( 6 . 8th ) as a function of a lateral acceleration or an equivalent quantity, which is applied to the motor vehicle ( 2 ), in particular when cornering, acts, is determined, wherein a behavior of a curve inside and a curve outer wheel ( 6 . 8th ) is displayed differently as needed. Method according to one of the preceding claims, in which the axle kinematics of the axle ( 4 ), in particular the front axle ( 4 ) of the motor vehicle ( 2 ), and their influence on a behavior of a force and torque transmission from a resulting point of a force introduction in a footprint of at least one wheel ( 6 . 8th ) to a location of measurement of the force or torque on the axle ( 4 ) is taken into account. Method according to Claim 7, in which a behavior of the tires and / or the trailing path of the at least one wheel ( 6 . 8th ) in the resulting point in the footprint of the at least one wheel ( 6 . 8th ) on the road by variable driving conditions, by forces acting on the at least one wheel ( 6 . 8th ) and by different loads acting on the at least one wheel ( 6 . 8th ), is influenced. Method according to one of the preceding claims, in which the determination of the coefficient of friction or the quantity derived therefrom is carried out only if the following conditions are met: - the speed of the motor vehicle ( 2 ) is greater than a minimum value provided for this purpose, - the lateral acceleration of the motor vehicle is greater than a value provided for this purpose, - influences of driving and braking torques on the axle ( 4 ) and / or the wheels ( 6 . 8th ) are negligible since they are below critical thresholds, - a second derivative and a third derivative of the steering angle with respect to the time fall below this maximum values and / or the steering angle is outside of a range of zero crossing for steering angle, and - the ratio the speed of the steering angle to the speed of the motor vehicle ( 2 ) and the brake pressure are each smaller than a threshold provided for this purpose.  Method according to one of the preceding claims, in which an effective follow-up distance and a theoretical coefficient to be expected at high coefficient of friction are calculated from recorded operating parameters and an exclusion criterion is checked for the calculated effective follow-up distance, whereby in the event that a check of this exclusion criterion is positive, it is determined that no new reliable estimate of the coefficient of friction between the tires and the driven lane is possible, and in the event that the check of the exclusion criterion is negative, a value of the lateral acceleration is checked, in which case the value of the lateral acceleration greater than a lower limit and less than an upper limit, the operating parameters for rack force evaluation are readjusted, and in the event that the lateral acceleration is greater than the upper limit or equal to the upper limit In the case that values of the theoretical caster distance are smaller than values of the effective catenary distance, the operating parameters for evaluating the rack force are readjusted, and wherein for the If values for the theoretical caster distance and the effective caster distance deviate from each other by a maximum of one tolerance value, it is determined that the lateral acceleration is currently too low and no new reliable estimate of the coefficient of friction between the tires and the traveled roadway is possible in the event that values of the theoretical trailing distance are greater than those of the effective trailing distance by the tolerance value, a maximum usable frictional connection and thus the coefficient of friction is estimated. Method according to one of the preceding claims, in which for at least one parameter of the motor vehicle ( 2 ) is taken into account, wherein at regular times and / or in the presence of a characteristic operating range of the motor vehicle ( 2 ) a current value for the at least one parameter of the motor vehicle is determined, and wherein the previously stored value for the at least one parameter is adapted to the current value, so as to adapt the method by adapting to changing parameters of the motor vehicle ( 2 ) and / or make the tire long term. Method for operating a motor vehicle ( 2 ), which has a steering system ( 10 ) with a rack ( 12 ), with the wheels ( 6 . 8th ) on an axis ( 4 ) on which tires are mounted, including associated suspensions of the motor vehicle ( 2 ), wherein the motor vehicle ( 2 ) is moved on a carriageway in which, in a learning phase, during at least one drive of the motor vehicle ( 2 ) pairs of values each from values for a current rack force F _{Zst_akt} the rack ( 12 ) and values for one on the motor vehicle ( 2 ), the current side force F _{Sv_akt} , which are calculated from driving dynamic operating parameters, are determined, wherein from a plurality of value pairs a slope m _akt for at least one straight line, which describes a proportionality between the current rack force F _{Zst_akt} and the current side force F _{Sv_akt} , is determined where the slope is:

Method according to one of claims 1 to 11, wherein in a learning phase, during at least one journey of the motor vehicle ( 2 ) pairs of values are in each case made of values for the current rack force F _{Zst_akt of} the rack ( 12 ) and values for the motor vehicle ( 2 ), the current side force F _{Sv_akt} , which are calculated from driving dynamic operating parameters, are determined, wherein from a plurality of value pairs a slope m _akt for at least one straight line, which describes a proportionality between the current rack force F _{Zst_akt} and the current side force F _{Sv_akt} , is determined where the slope is:

Method according to one of the preceding claims, in which, for determining the coefficient of friction, a relationship between the rack-and-pinion force F _Zst and a side force F _{Sv acting} on the motor vehicle ( 2 ) is used, is used. Method according to one of claims 12 to 14, _wherein after the learning phase via a danger _indicator GI a possible deviation of the current rack force F _{Zst_akt} from a reference Rack force F _{Zst_ref} is determined taking into account a constructive rack force F _{Zst_akt_k} , whereby the following applies to the danger _indicator GI:

The method of claim 15, _wherein a maximum possible lateral force F _{Sv_max is} calculated, which is achieved when the current rack force F _{Zst_akt} the constructive rack force F _{Zst_akt_k} corresponds, wherein for the maximum possible lateral force taking into account a proportionality k _Sv between a constructive rack force F _{Zst_akt_k} and the side force F _Sv and a slope a _Sv for describing a proportionality between the reference rack force F _{Zst_ref} and the side force F _Sv . applies:

Method according to claim 15 or 16, in which, in the event that a value of the hazard indicator GI becomes greater than a limit value, a warning signal is provided, and in the event that a value of the hazard indicator GI becomes smaller than a limit value Considering a normal force F _Nv a coefficient of friction μ is calculated, for which applies:

Method according to one of the preceding claims, in which the learning phase is carried out after wheels ( 6 . 8th ) of the motor vehicle ( 2 ) have been newly frosted.  Method according to one of the preceding claims, in which the method is operated with the lowest possible signal amplitude. System for operating a motor vehicle ( 2 ), which has a steering system ( 10 ) with a rack ( 12 ), with the wheels ( 6 . 8th ) on an axis ( 4 ) on which tires are mounted, including associated suspensions of the motor vehicle ( 2 ), wherein the motor vehicle ( 2 ) is moved on a roadway, the system being a control device ( 16 . 16a ), which is adapted to a coefficient of friction or a quantity derived therefrom between a respective tire and the roadway taking into account a driving state, a steering mechanism kinematics, a kinematics of the axle ( 4 ), a behavior of the tires on the wheels ( 6 . 8th ) and / or determine their trailing distance as a characteristic variable, wherein the control device ( 16 . 16a ) is adapted to changing operating parameters of the motor vehicle ( 2 ) and the tire through a learning phase during at least one journey of the motor vehicle ( 2 ), recognize and take into account.

Images