Abstract:
Device and method are described for operating a vehicle using a vehicle controller to individually adjust braking forces of the wheels of at least one axle of the vehicle and using a yawing moment compensator to at least partially compensate for a yawing moment of the vehicle resulting from different braking forces of individual wheels of at least one axle by intervening in a steering of the vehicle, the action of the yawing moment compensator on the steering not being performed or only to a lesser degree while the vehicle controller is adjusting braking forces.

Description:
FIELD OF THE INVENTION 
   The present invention relates to a device and a method for operating a vehicle using a vehicle controller for individually adjusting braking forces of the wheels of at least one axle of the vehicle and using a yawing moment compensator to at least partially compensate for a yawing moment of the vehicle due to different braking forces of individual wheels of the at least one axle by intervening in a steering of the vehicle. 
   BACKGROUND INFORMATION 
   Today, braking systems such as hydraulic, electrohydraulic, pneumatic, electropneumatic, or electromechanical braking systems may be increasingly electrically controllable. The electrical control may permit a pressure build-up in the wheel brakes independent of the driver&#39;s braking intent, i.e. of the brake pedal operation by the driver. Such electrical controls of braking systems may be used, for example, for implementing an anti-lock control (ABS, i.e., anti-lock (braking) system) or an electronic stability program (FDR or ESP). 
   A purpose of an anti-lock (braking) system (ABS) may include to prevent the vehicle from slipping due to its wheels locking while braking, in particular on a slippery surface. For this purpose, when the driver operates the brake pedal for an extended period of time, sensors determine whether the individual wheels are locked, and whenever this is the case, the brake pressure on the corresponding wheel brakes is reduced. In such an anti-lock (braking) system, the front wheels of the vehicle may be (but not necessarily) separately and consequently differently controlled, while the rear wheels are controlled together. 
   An electronic stability program (FDR or ESP) is used to monitor steering, braking, and gas pedal inputs by the driver in order to prevent the vehicle from slipping as a result of false inputs. In this context, false inputs are intercepted by targeted braking actions at the individual wheels. 
   Similar to braking systems controlled by electrical controls, steering systems may also be controlled by motor-driven steering systems. In this context, the power of a power source of an electromotor, for example, is able to be superimposed on the steering-wheel power applied by the driver, e.g. using a control element for the superimposed steering action. On the one hand, an effect supporting the steering-wheel power of the driver is able to be achieved. On the other hand, steering signals that increase the driving safety and/or the driving comfort are able to be applied to the steering systems of the vehicle. Such a motor-driven steering system is described in German Published Patent Application No. 40 31 316, for example. 
   A combination of a control of a braking system and of a steering system of a vehicle is described in European Published Patent Application No. 0 487 967 (vehicle having an anti-lock controller). Reference is made to this patent with respect to the entire content. In short, a yawing moment compensation (GMK) for a vehicle equipped with an anti-lock (braking) system (ABS) is described in European Published Patent Application No. 0 487 967. The yawing moment compensation determines a correction steering angle to compensate for the yawing moment of the vehicle occurring when braking on an inhomogeneous roadway (e.g. a μ-split) due to different braking forces on the left or right wheel(s). 
   SUMMARY 
   An object of the present invention may include providing an improved method and an improved device for controlling a braking system and a steering system of a vehicle as well as providing a vehicle having the corresponding device. 
   This objective may be achieved by an example method according to the present invention. In this context, for operating a vehicle using a vehicle controller to individually adjust braking forces of the wheels of at least one axle of the vehicle and using a yawing moment compensator to at least partially compensate for a yawing moment of the vehicle resulting from different braking forces of individual wheels of at least one axle by intervening in a steering of the vehicle, the action of the yawing moment compensator on the steering is not performed or only to a lesser degree while braking forces are being adjusted by the vehicle controller. 
   Hence, the action of the yawing moment compensator on the steering is not performed while the vehicle controller is active. 
   In particular, the vehicle controller may be part of an electronic stability program (FDR or ESP) as described, for example, in the article, FDR—The Operating Dynamics Regulation of Bosch, by A. van Zanten, R. Erhardt and G. Pfaff, Journal of Automobile Technology 96 (1994), 11 pages 674 to 689, and SAE paper 973184, Vehicle Dynamics Controller for Commercial Vehicles, by F. Hecker, S. Hummel, 0. Jundt, K. -D. Leimbach, I. Faye, and H. Schramm. In this context, the vehicle controller may be configured for adjusting the braking forces as a function of the yaw rate of the vehicle and a setpoint yaw rate of the vehicle, in particular as a function of the difference between the yaw rate of the vehicle and the setpoint yaw rate of the vehicle. In this context, the braking forces may be adjusted by calculating the setpoint slip values for the wheels that may be input quantities in secondary control loops. 
   The intervention of the yawing moment compensator in the steering may be reduced by at least one filter. 
   In a further example embodiment of the present invention, the axle may be the front and/or the rear axle. 
   In another example embodiment of the present invention, the action on a steering of the vehicle may be performed using a compensation steering angle determined as a function of the braking forces of individual wheels. 
   In a further example embodiment of the present invention, a compensation steering angle dependent on a difference of separately controlled braking pressures of the front and/or rear wheels may be adjusted at a rear-wheel steering system or may be superimposed on a front-wheel or rear-wheel steering angle in order to at least partially compensate for the yawing moment of the vehicle. 
   In this context, the braking pressures may be used as substitute quantities for the braking forces. 
   In another example embodiment of the present invention, the value of the compensation steering angle may be set to zero in a predefined or variable range of small braking pressure differences, i.e. within a dead zone, and to a value not equaling zero outside of the dead zone. 
   The values for the dead zone may be different for the front and rear axle. 
   In another example embodiment of the present invention, separate partial compensation steering angles may be determined in each case for the front wheels and the rear wheels, the compensation steering angle being determined as a function of the partial compensation steering angles. 
   In a further example embodiment of the present invention, the compensation steering angle may be determined by adding the partial compensation steering angles. 
   In another example embodiment of the present invention, at least one partial compensation steering angle may be determined after the dead zone is exceeded by adding the product of a constant and the initial value of the dead zone and the product of a variable amplification and the initial value of the dead zone. 
   In a further example embodiment of the present invention, the compensation steering angle may be stored when braking forces are adjusted by the vehicle controller. 
   In another example embodiment of the present invention, the stored compensation steering angle may be essentially continuously transferred after the completion of the adjustment of the braking forces via the vehicle controller to an instantaneous compensation steering angle. 

   
     BRIEF DESCRIPTION OF THE DRAWINGS 
       FIG. 1  is a block diagram of a technical field that may be improved by an example embodiment of the present invention. 
       FIG. 2  is a graph diagram for the block diagram in FIG.  1 . 
       FIG. 3  is a graph diagram for the block diagram in FIG.  1 . 
       FIG. 4  is a graph diagram for the block diagram in FIG.  1 . 
       FIG. 5  is a block diagram of a modified technical field that may be improved by an example embodiment of the present invention. 
       FIG. 6  is a block diagram of an example embodiment of the present invention. 
       FIG. 7  is a diagram of the example embodiment in FIG.  6 . 
   

   DETAILED DESCRIPTION 
   In the following, a technical field, which may be improved by an example embodiment of the present invention, is first explained via an example on the basis of  FIGS. 1 through 5 . An example embodiment of the present invention is then described on the basis of  FIGS. 6 and 7 . 
   The present example of a technical field in  FIG. 1  explains the compensation for the brake yawing moment by a rear axle steering for a select-low braked rear axle. 
   The braking pressures in the front wheels supply, in a first approximation, a measure of the used braking force, the difference Δp of the pressures consequently supplies a measure for the brake yawing moment. Rear-axle steering angle δ produces an opposing moment about vertical vehicle axle that compensates for the brake yawing moment given a suitable configuration. The steady-state relationship between δ and Δp is described by proportionality factor k p . 
   Since the brake pressures may be constantly modulated during an ABS braking, a rear axle steering control having only the abovementioned proportionality may react very irregularly. Therefore, a filtering may be provided before the pressure difference is calculated. This difference first overcomes a significant threshold (dead zone) before the control becomes active; this measure may also be intended to prevent steering irregularity in the case of small disturbances. 
   Measured braking pressures P vl  and P vr  are filtered in two stages. 
   Disturbances caused by measuring noise (peaks, A/D errors) are to be suppressed in pre-filter  1 ,  1 ′ by variably restricting the pressure change rate. The increase limit remains at smaller values when there is frequent change of pressure build-up and decrease. Given a change having the same sign over a longer period of time, the increase limit is continuously increased to a maximum value. 
   Decay filters  2  and  2 ′ may be configured for the relationships between ABS control cycles (ABS control cycles with series of pulses) and rear-wheel steering. So that the rear-wheel steering angle does not directly follow the pressure jumps in particular in the pressure reduction phases, a decrease of the filtered braking pressures is only allowed very slowly during the first pressure reduction after a pressure increase phase. After a predefined time (e.g. 100 ms) elapses, the time constant of the low pass filter is switched over so that the filtered value (output of block  2  or  2 ′) approaches the output quantities of pre-filter  1  ( 1 ′) more quickly. 
   The measured pressure as well as the intermediate value and the filtered pressure are shown in FIG.  2 . 
   According to this, the difference of the output quantities of filters  2  and  2 ′ is formed from filtered braking pressures P vlf  and P vrf  in a subtracter circuit  3 , the difference supplying after a dead zone  4  is exceeded input quantities f(Δp) for control amplifiers  5  and  6  whose output signals are added in adder  7  to form steering angle Δ. 
   The control may be made up of a constant proportion
 
δ p   =f (Δ p )· k   p  (Block  5 ).
 
   As a result of the filtering, the dead zone, and the dynamic response (characteristic) of the steering controller, a yawing motion first builds up which is also maintained in the case of an ideal configuration of amplification k p . Therefore, a time-variable proportion is also calculated at the start of the control action: 
    δ v   =f (Δ p )· k   v (Block  6 ) 
   Factor k v  is set to a certain value when the difference of the filtered pressures exceeds the dead zone and then continually decays. 
   Therefore, when the control is switched in, the rear-wheel steering angle is noticeably increased, so that the yaw rate changes its sign and the yaw angle is consequently reduced again. In this case, the driver may no longer need to intervene. Viewed over the entire braking action, the yaw rate only assumes very small values, i.e., the irregularities are largely compensated for by the ABS control cycles. 
   The yawing moment compensation prevents the vehicle from breaking away at low speeds as well as at high speeds. Its support may become clearer as the speed increases. 
   In tests with a fixedly held steering wheel, the track displacement remains quite small, and a yaw angle builds very slowly. 
   As already said above, the measurement used to date of the front wheel brake pressures may also be replaced by an estimation algorithm. One is described in patent application P 4030724.7, which is included in European Published Patent Application No. 0 486 967 as an appendix. In this context, the filtering of the braking pressures is able to be simplified such that blocks  1 ,  1 ′ are eliminated. 
   The front-wheel steering angle may be influenced according to the same principle. Only quantitative differences arise. 
   Given different friction coefficients on different vehicle sides, introducing the time-variable amplification may result in desirable features yet may cause an oversteering behavior of the vehicle when fully braking in a curve. To prevent this, the transversal acceleration of the vehicle may also be taken in to account. However, considering the transversal acceleration as described does not presuppose acquiring the steering angle according to the top branch in FIG.  1 . 
   A correction factor K by , which is multiplicatively linked to the rear-wheel angle (in  12 ), is first determined from measured transversal acceleration b y  via the characteristic curve (block  8 ) shown in FIG.  3 . 
   This characteristic curve causes the compensation to not be influenced (K by =1) in the case of low transversal accelerations, e.g. less than 2 m/s 2 , thereby resulting in a reduction proportional to the transversal acceleration, and causes the compensation to be completely suppressed (K by =0) in the case of a very high transversal acceleration, e.g. above 8 m/s 2 . This characteristic curve is based on the knowledge that in the case of μ-split braking, the occurring transversal accelerations are approximately in the range of +/−2 m/s 2 . 
   Only this characteristic curve may not be sufficient. Fluctuations in the transversal acceleration for values b y &gt;2 m/s 2  (e.g. sign change of b y  during lane change while braking) result in proportional fluctuations of the correction factor and consequently of the rear-wheel steering angle that may be noticeable as an irregularity. In addition, it may be undesirable that these steering-angle fluctuations then effect the b y  signal. A suitable filtering of the correction factor may therefore be required. However, it may be required to ensure that when building up a transversal acceleration, the GMK is quickly reduced. However, during certain driving maneuvers, e.g. lane changes, an intervention may not be performed again too quickly. This may be achieved using two alternative low pass filters  10  and  11  having very different time constants. As such, the transversal acceleration-dependent steering angle correction may have the form shown in  FIG. 1  in blocks  8 ,  9 ,  10 , and  11 . 
   Example values for the time constants of the two alternative low pass filters may be 10 ms and 1000 ms, respectively. 
   Blocks  9 ,  10 , and  11  are to symbolize the following situation. If the transversal acceleration increases and Kb y  becomes smaller, low pass filter  10  having the small time constant becomes active, i.e., output value Kb y  quickly follows the input from block  8  and decreases the steering angle. If however the transversal acceleration decreases and Kb y  consequently increases, Kb y  follows the input value from block  8  but in a delayed manner. 
   These measures may reduce the yawing moment compensation when braking on curves and changing lanes while braking on surfaces having high coefficients of friction. The remaining portions of rear-wheel steering angle δ GMK  from the compensation may no longer have a negative effect on the vehicle performance. 
   The measured transversal acceleration may be replaced by a quantity subsequently formed from the steering angles and the vehicle speed (e.g. tacho signal). When considered in a steady-state manner, the following relationship for the transversal acceleration is able to be derived from the conventional linear single-track model: 
         b     y   ,   stat       =           V   x   2     ⁡     (       δ   v     -     δ   h       )         l   o       ⁢     1     1   +       (       V   X     /     V   ch       )     2               
 
where:
 
   
     
       
             
             
             
           
         
             
                 
                 
             
           
           
             
                 
               V x   
               longitudinal vehicle speed 
             
             
                 
               δv 
               front-wheel steering angle 
             
             
                 
               δh 
               rear-wheel steering angle 
             
             
                 
               l o   
               wheel base 
             
             
                 
               V ch   
               characteristic speed 
             
             
                 
               B y.stat   
               estimated steady-state acceleration 
             
             
                 
                 
             
           
        
       
     
   
   In this context V ch  is made up of the model parameters as follows: 
         V   ch     =       1       m     l   0   2       ⁢     (         l   h       C   v       -       l   v       C   h         )               
 
where
 
   
     
       
             
             
             
           
         
             
                 
                 
             
           
           
             
                 
               m 
               Vehicle weight 
             
             
                 
               l v   
               Distance from center of gravity - front axle 
             
             
                 
               l h   
               Distance from center of gravity - rear axle 
             
             
                 
               C v   
               Slip angle rigidity - front axle 
             
             
                 
               C h   
               Slip angle rigidity - rear axle 
             
             
                 
                 
             
           
        
       
     
   
   Using the parameters of a certain model may result in a value of V ch  of about 20 m/s. 
   In the case of a transient driving maneuver (changing lanes while braking), it turns out that steady-state equation (1), which is adjusted to cornering, may deliver transversal accelerations that are too high. For this reason, a dynamic member (low pass filter having time constant T bys ), which takes the vehicle dynamics into consideration, is connected in series (block  13 ). 
   When implementing equation (1) in the computing device, it offers itself to store the portion 
           V   x   2       l   o       ⁢     1     1   +       (       V   X     /     V   ch       )     2             
 
as a speed-dependent characteristic curve (block  14 ). Equation (1) is consequently reduced to the interpolation of a characteristic curve (in block  14 ) as well as the multiplication of the result by the difference (δ v −δ h ) (in block  15 ). The total transversal acceleration correction consequently may have the form shown in the middle branch in FIG.  1 .
 
   When estimating the transversal acceleration as shown above, rear-wheel steering angle δ h  is included as an input quantity. At the same time, the estimation has a reciprocal effect on part of the rear-wheel steering angle, namely the GMK part. So that no feedback effects are able to occur in this context, only the portion of the rear-wheel steering angle coming from another rear-wheel steering control is taken into consideration as an input quantity of the transversal acceleration estimation. 
   To suppress the amplified turning-in at the end of a curve braking by the yawing moment compensation, an amplification factor K Vx  dependent on the vehicle speed is multiplicatively superimposed. 
   Its example characteristic curve is stored in block  16  and shown in FIG.  4 . Over 50 km/h, for example, the amplification factor remains unchanged at one, and in the range of 50 km/h to 20 km/h, for example, it is continuously reduced to zero. This measure may be less important for μ-split braking, since vehicles having ABS may not show any manageability problems in lower speed ranges. 
   This additional factor K Vx  is multiplicatively considered in multiplier  12 . Therefore, the steering angle for the yawing moment compensation as a whole is:
 
δ GMX   =K   by   ·K   vx ·δ.
 
   A variable dead zone  4 ′ differentiates the block diagram of a modified technical field in  FIG. 5  from that in FIG.  1 . In this context, filtered braking pressures P vlf  and P vrf  are multiplied together by a multiplier  20 . The product of P vlf  and P vrf  is multiplied by a correction factor K th  and added to a predefined limiting value P to  to form a corrected limiting value P toth . 
   The example of a technical field described using  FIGS. 1 through 5  that may be improved by an example embodiment of the present invention starts out from a vehicle having an anti-lock (braking) system (ABS) in which the braking pressures of the rear wheels are not individually regulated. This may often be sufficient for the purposes of a simple anti-lock (braking) system (ABS) so that provision may not be made for an individual control of the braking pressures of the rear wheels for commercially available anti-lock (braking) systems (ABS). Consequently, braking pressure differences may only occur at the wheels of the front axle and may only need to be considered there. 
   Something different may be true for vehicles equipped with an electronic stability program (FDR or ESP). In this instance, within the framework of the electronic stability program, braking pressures of the wheels of both axles may be individually regulated at least intermittently. In this context, different braking pressures may be set in a targeted manner at each wheel of an axle in order to influence the vehicle motion. 
   These conditions are considered in the example embodiment of the present invention shown in FIG.  6 . In this context, the variant from  FIG. 5  having a variable dead zone  4 ′ is presupposed. The present invention may also be used for the variant from  FIG. 1  having a fixed dead zone. In this manner, it may be achieved that yawing moment compensation (GMK) only reacts to braking pressure differences in an anti-lock braking system (ABS) and may not also be dependent on a vehicle controller of a electronic stability program (FDR or ESP). 
   In comparison with the variant in  FIG. 5 , yawing moment compensation (GMK) is expanded in  FIG. 6  by two parts: 
   The first expansion, which is shown in the upper left portion of  FIG. 6 , is used for considering the braking pressure differences of the wheels of the rear axle. For this purpose, another branch was added to the block diagram that may correspond to the top branch in  FIG. 5  (or FIG.  1 ). Therefore, the same components in the representation are designated by the same reference numerals, and only “h” for the rear axle and “v” for the front axle were added. 
   The braking pressures of rear wheels P hl  and P hr  are able to be measured or estimated as described above for the braking pressures of front wheels P vl , P vr . They may then be treated in the same manner as the braking pressures of front wheels P vl , P vr . Consequently, they are filtered in pre-filters and decay filters  1   h ,  1   h ′,  2   h ,  2   h ′. The difference of filtered braking pressures P hlf , P hrf  is determined in a subtracter circuit  3   h . If the difference of filtered pressures P hlf , P hrf  exceeds a dead zone 4, which is dependent on the total pressure level or is predefined in a fixed manner, a partial compensation steering angle δ GMKh  is determined. A steering angle determined from the braking pressures of the wheels of the front axle as described above is added as an additional partial compensation steering angle δ GMKv  to partial compensation steering angle δ GMKh  of the braking pressures of the wheels of the rear axle to form a rear and/or front axle steering angle δ ideal . 
   Mainly the following points differentiate the treatment of braking pressures P hl , P hr  of the rear wheels from the treatment of braking pressures P vl , P vr  of the front wheels: Other parameters may be selected for the filters and the dead zone as well as another value for the constant amplification. Such different parameters may take into account e.g. the different configuration or the different size of the brakes, i.e., a different connection between braking pressure and braking force at the front or rear axle. Furthermore, such different parameters may take into account a possibly different track width of the front and rear axle or different ABS strategies. 
   Moreover, the time-variable amplification of the braking pressure difference (block  6  in  FIGS. 1 and 5 ) may be eliminated. This may be possible since in the case of an ABS action within a electronic stability program (FDR), the braking pressure difference of the rear wheels is regularly controlled such that it only increases slowly. On the other hand, a time-variable amplification of the braking pressure difference of the rear wheels may also be useful and used accordingly. 
   Due to the indicated differences when treating the rear and front braking pressures P hl , P hr , and P vl , P vr , it may be desirable to first form each difference separately as shown in FIG.  6 . Subsequently, partial compensation steering angles δ GMKv , δ GMKv  are added to form total rear or front axle steering angle intervention δ ideal . 
   The thus obtained rear or front axle steering angle intervention δ ideal  may generally correspond to the steering angle for the yawing moment compensation. However, as described above, transversal acceleration b y  and the speed of the vehicle may also be considered. For this purpose, specified correction factors K by  and K vx  are applied to front or rear axle steering angle δ ideal . The thus obtained instantaneous compensation steering angle δ A  is set for yawing moment compensation at the rear axle or is superimposed on a steering angle of the front or rear axle. 
   The second expansion may be used to ensure that yawing moment compensation (GMK) only reacts to braking pressure differences from an anti-lock braking system (ABS) and not as a function of a driving dynamics controller. A signal indicating when interventions of the vehicle controller occur is provided for this purpose. The feature that interventions of the vehicle controller exist may be indicated in electronic stability programs in the form of a flag that is able to assume the values zero and one, for example. Therefore, it may only need to be transmitted to the control of yawing moment compensation (GMK). A selector  50  is provided for processing signal F. 
   This expansion may cause yawing moment compensation (GMK) to be switched off when interventions of the vehicle controller occur. An already applied compensation steering angle δ A  is maintained during a subsequent intervention of the vehicle controller and is then essentially continuously transferred to an instantaneous compensation steering angle δ A . 
   For this purpose, a factor K H  is first formed from flag F of the vehicle controller via a block  52  by a switching-off filter  30 . The value of factor K H  always equals one when flag F is set, i.e. equals one. If flag F zeros, the value of factor K H  tends to zero with a predefined time response. Such a relationship is shown via an example in FIG.  7 . In this example, the value of factor K H  tends to zero in a linear manner in a time Δt. Alternatively, an exponential transition may also be used. 
   With the help of thus obtained factor K H , the front-axle steering angle δ GMK  to be ultimately applied at the steered axle for yawing moment compensation is determined by a block  53  in accordance with the following equation:
 
δ GMK =(1 −K   H )·δ A   +K   H ·δ H 
 
where
 
   
     
       
             
             
           
         
             
                 
             
           
           
             
               δ A  = 
               the intantaneous compensation steering angle in each 
             
             
                 
               case 
             
             
               δ H  = 
               a compensation steering angle maintained during an 
             
             
                 
               intervention of the vehicle controller. 
             
             
                 
             
           
        
       
     
   
   A controllable sample-and-hold member  51  is used to obtain constant compensation steering angle δ H . It is switched such that it assumes in each case instantaneous compensation steering angle δ A  (sample). As long as factor K H  equals zero, sample-and-hold member  51  also outputs this instantaneous compensation steering angle δ A  in each case as an output value (i.e. δ A =δ H ). However, as soon as factor K H  is greater than zero, the value of compensation steering angle δ A  applied last is frozen (hold) and constant compensation steering angle δ H  is consequently generated and output. As soon as factor K H  again assumes the value zero, constant compensation steering angle δ H  is no longer maintained, etc. 
   As long as factor K H  equals zero, i.e., as long as there are no interventions of the vehicle controller, the above equation simplifies to:
 
δ GMK =(1−0)·δ A +0·δ H =δ A 
 
   Therefore, the yawing moment compensations required in each case are performed unchanged in accordance with the above description. 
   As soon as there is an intervention of the vehicle controller, factor K H  equals one. Consequently, the above equation becomes:
 
δ GMK =(1−1)·δ A +1·δ H =δ H 
 
i.e., the compensation angle δ A  last applied before the intervention of the vehicle controller is maintained as a constant compensation angle δ H  and continues to be applied during the intervention.
 
   As soon as the intervention of the vehicle controller is finally completed, factor K H  is continuously transferred rear to the value zero during a time Δt. During this time, constant compensation angle δ H  continues to be maintained and resulting compensation angle δ GMK  is calculated as explained above:
 
δ GMK =(1 −K   H )·δ A   +K   H ·δ H 
 
   In this manner, the compensation angle δ H  maintained during the intervention of the vehicle controller and also applied during this time as resulting compensation angle δ GMK  is continuously transferred to the value of the instantaneous compensation angle δ A  actually needed in each case after the intervention of the electronic stability program (FDR or-ESP) to compensate for the yawing moment. 
   Another possibility for preventing yawing moment compensation (GMK) from counteracting its vehicle controller is to significantly filter instantaneous intervention angle δ A  of yawing moment compensation (GMK) as long as the interventions of the vehicle controller are occuring. Consequently, the driving dynamics interventions in the higher frequency range are not affected by yawing moment compensation (GMK). 
   In comparison with the exemplarily described technical field, the described example embodiment may provide that yawing moment reductions (GMA) to be considered by the anti-lock (braking) system (ABS) integrated in the electronic stability program (FDR or ESP) according to the related art are able to be significantly reduced at the front axle as well as at the rear axle. A transition may also be made to individual ABS interventions at the rear axle already at a higher speed. This may result in a shorter braking distance. Furthermore, other steering actions may superimpose the yawing moment compensation interventions. Measured or estimated braking pressures that may already be available from the electronic stability program (FDR or ESP) may be used as input information for the yawing moment compensation. 
   The above-described example embodiments are only used to improve the understandability of the present invention. They are not intended as a restriction. Therefore, it may be understood that all additional possible example embodiments are within the framework of the present invention. In particular, it may be understood that the present invention also includes a device for implementing the described method and a vehicle equipped with such a device. 
   LIST OF REFERENCE NUMERALS 
   
       
       δδ ideal  Rear-axle and/or front-axle steering angle 
       Δ A  Compensation steering angle 
       δ GMKv , δ GMKv  Partial compensation steering angle 
       ΔP Pressure difference 
       k p  Proportionality factor 
       k v  Factor 
       P vl , P vr  Braking pressures of the front wheels 
       P vlf , P vrf  Filtered braking pressures of the front wheels 
       P hl , P hr  Braking pressures of the rear wheels 
       P hlf , P hrf  Filtered braking pressures of the rear wheels 
       b y  Transversal acceleration 
       K by  Correction factor 
       K by  Amplification factor 
       P tot  Predefined limiting value 
       K th  Correction factor , 
       P toth  Corrected limiting value 
       F Flag 
       K H  Weight factor 
       S/H Sample-and-hold member 
         1 , 1 ′ Pre-filter 
         2 , 2 ′ Decay filter 
         3  Subtracter circuit 
         4  Dead zone or dead zone 
         4 ′ Variable dead zone or dead zone 
         5 , 6  Control amplifiers or block 
         7  Adder 
         8  Characteristic curve of block 
       
         9 
       
         10 , 11  Alternative low pass filters 
         12  Multiplier 
         13  Dynamic member 
         14  Speed-dependent characteristic curve 
         20  Multiplier 
         30  Triggering filter